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Transcript of Proceeding IMAT2012 Bintan
The 5th IMAT, November 12 – 13th
, 2012
i
THE 5th
IMAT
INTERNATIONAL MEETING ON ADVANCES IN THERMO-FLUIDS
NOVEMBER 12 – 13th
, 2012
BINTAN ISLAND - INDONESIA
DEPARTEMENT OF MECHANICAL ENGINEERING
UNIVERSITAS INDONESIA
VERITAS, PROBITAS, IUSTITIA
The 5th IMAT, November 12 – 13th
, 2012
ii
TABLE OF CONTENT
Welcome......................................................................................................................... iii
Organizing Commitee..................................................................................................... iv
Program Summary.......................................................................................................... vi
List of Paper.................................................................................................................... 1
The 5th IMAT, November 12 – 13th
, 2012
iii
WELCOME
It is our pleasure to welcome you, for the 5th
IMAT 2012, International Meeting on Advances
in Thermo-Fluids, to Bintan Island, Indonesia. The 5th
IMAT 2012 is organized by
Department of Mechanical Engineering, University of Indonesia. The history of this
conference has taken place in the following places:
1. 1st IMAT, 2008, UTM, Malaysia
2. 2nd IMAT, 2009, Bogor, Indonesia
3. 3rd IMAT, 2010, Singapore
4. 4th IMAT, 2011, Melaka, Malaysia
The 5th
IMAT will be coming at November 12-13th
, 2012 in beautiful and exotic island of
Bintan, Indonesia, around 45 minutes from Singapore or Malaysia by Ferry. The 5th
IMAT
aims to provide a technical forum that includes keynote lectures, information of researchs at
UI, NUS and UTM, and oral presentation sessions. In addition to the fundamentals of thermal
phenomena and traditional thermal applications, the 5th
IMAT is expected to address the
emerging domains of thermal transport in fishery, building nano-materials, bio-systems,
power generation, microsystems, and energy conversion devices.
All submitted papers will be peer reviewed, and the accepted paper will be published in the
conference proceeding. Selected papers will be offerred for publication in International
Journal.
We wish to provide the most pleasurable time in meeting all participant to talk and discuss
about new researches and applications related to thermal and fluids engineering, and give an
opportunity for the communication and cooperation between the researchers.
Bintan, November 12 – 13th
, 2012
Organizing Commitee
The 5th IMAT, November 12 – 13th
, 2012
iv
ORGANIZING COMMITTEE The 5
th IMAT 2012 is organized by Department of Mechanical Engineering,
Faculty of Engineering, University of Indonesia.
Kampus UI, Depok 16424, Indonesia
Phone : +62 21 727 0032
Fax : +62 21 727 0033
Email : [email protected]
Advisory Board
UNIVERISITAS INDONESIA (UI)
Prof. Dr. Ir. Bambang Sugiharto, M.Eng.
Prof. Dr. I. Made Kartika, Dipl.Ing.
Prof. Dr. Yanuar, M. Eng., M.Sc.
Prof. Dr. Ir. Budiarso, M.Sc.
Prof. Dr.-Ing. Nandy Setiadi Djaya Putra
Prof. Yulianto Sulistyo Nugroho, M.Sc., Ph.D.
Prof. Dr. Ir. Harinaldi, M.Eng.
Assoc. Prof. Dr. Budihardjo, Dipl.Ing.
Assoc. Prof. Dr. M. Idrus Alhamid
Assoc. Prof. Dr.-Ing Nasruddin, M.Eng.
UNIVERSITI TEKNOLOGI MALAYSIA (UTM)
Prof. Ir. Dr. Azhar Abdul Aziz
Assoc. Prof. Dr. Mazlan Abdul Wahid
Prof. Amer Nordin Darus
Prof. Dr. Farid Nasir Hj. Ani
Prof. Dr. Md. Nor Musa
Dr. Jamaluddin Md. Sheriff
NATIONAL UNIVERSITY OF SINGAPORE (NUS)
Prof. Dr. Kim Choon Ng
Prof. Dr. Christopher Yap
Dr. Kandadai Srinivasan
The 5th IMAT, November 12 – 13th
, 2012
v
INTERNATIONAL ISLAMIC UNIVERSITY OF MALAYSIA (IIUM)
Prof. Dr. M.N.A. Hawlader
Organising Committee
Chairman:
Dr. Agus S. Pamitran, ST., M.Eng.
Secretary:
Muhamad Yulianto, ST., MT.
Member:
Ir. Senoadi, MT.
Ir. Ruli Nutranta, M.Eng.
Ir. Supryiadi, M.Sc.
Ir. Arief Surachman, MT.
The 5th IMAT, November 12 – 13th
, 2012
vi
PROGRAM SUMMARY
The 5th
International Meeting on Advances in Thermofluids
Universitas Indonesia at Nirwana Garden – Bintan Island Indonesia
12 – 13 November 2012
DAY 0 (11th
November 2012 – Sundayy)
18.30 – 21.00 Welcoming BBQ Dinner
DAY 1 (12th
November 2012 – Monday)
07.00 – 08.30 Registration
08.30 – 08.40 Opening Ceremony
08.40 – 09.00 Research Info from UI (Prof. Harinaldi)
09.00 – 09.20 Research Info from UTM (Prof. Azhar)
09.20 – 09.40 Research Info from NUS (Prof. Ng KC)
09.40 – 10.00 Tea Break
10.00 – 16.00 PARALEL SESSION DAY 1
ROOM A : HVAC &
Fluid Flow
ROOM B : Heat/Mass Transfer &
Combustion
10.00 – 10.15 IMAT-UI 002 IMAT-UI 003
10.15 – 10.30 IMAT-UI 004 IMAT-UI 016
10.30 – 10.45 IMAT-UI 006 IMAT-UI 018
10.45 – 11.00 IMAT-UI 007 IMAT-UI 020
11.00 – 11.15 IMAT-UI 011 IMAT-UI 024
11.15 – 11.30 IMAT-UI 013 IMAT-UI 027
11.30 – 11.45 IMAT-UI 019 IMAT-UI 038
11.45 – 12.00 IMAT-UI 022 IMAT-UI 039
12.00 – 13.00 LUNCH
13.00 – 13.15 IMAT-UI 026 IMAT-UI 041
13.15 – 13.30 IMAT-UI 028 IMAT-UI 001
13.30 – 13.45 IMAT-UI 031 IMAT-UI 005
13.45 – 14.00 IMAT-UI 033 IMAT-UI 008
14.00 – 14.15 IMAT-UI 034 IMAT-UI 009
14.15 – 14.30 IMAT-UI 035 IMAT-UI 010
14.30 – 14.45 IMAT-UI 036 IMAT-UI 014
14.45 – 15.00 IMAT-UI 040 IMAT-UI 015
Serving Tea Break
15.00 – 15.15 IMAT-UI 042 IMAT-UI 021
15.15 – 15.30 IMAT-UI 012 IMAT-UI 023
15.30 – 15.45 IMAT-UI 017 IMAT-UI 029
15.45 – 16.00 IMAT-UI 025 IMAT-UI 030
16.00 – 16.15 IMAT-UI 032
16.15 – 16.30 IMAT-UI 037
19.00 – 20.30 CLOSING CEREMONY and Sea Food Dinner
at Kellong Restaurant
DAY 2 (13th
November 2012 – Tuesday)
09.00 – 11.00 TOUR
The 5th IMAT, November 12 – 13th
, 2012
vii
List Papers of IMAT 2012
Code Topic Name
IMAT-UI 001
Controlled Auto-Ignition Combustion
In A Two-Stroke Cycle Engine Using
Hot Burned Gases
Amin Mahmoudzadeh Andwari, Azhar
Abdul Aziz, M. F. Muhamad Said, Z. Abdul
Latif
IMAT-UI 002
Determination of Motive Nozzle and
Constant-Area Diameters: Numerical
Study of Ejector as an Expansion
Device in Split-type Air Conditioner
Kasni Sumeru, Henry Nasution, Farid Nasir
Ani
IMAT-UI 003 Effect Of External Particles On Brake
Noise Of Disc Braking System
M.A. Nasaruddin, M.K Abdul Hamid, A.R.
Mat Lazim, A.R. Abu Bakar
IMAT-UI 004
Experimental Study On Replacement Of
HFC-134a By Hydrocarbons In
Automotive Air Conditioner
Mohd Rozi Mohd Perang, Henry Nasution,
Zulkarnain Abdul Latif, Azhar Abdul Aziz,
Afiq Aiman Dahlan.
IMAT-UI 005 Effect Of Fuel Droplets During Early
Stage Of Spherical Flame Propagation
Aminuddin Saat, Mazlan Abdul Wahid,
Malcolm Lawes
IMAT-UI 006 The Use Of Mechanical Ventilation
System In An Electric Car
Intan Sabariah Sabri, Haslinda Mohamed
Kamar, Nazri Kamsah, Md Noor Musa
IMAT-UI 007
Retrofitting R-22 Split Type Air
Conditioning With Hydrocarbon (Hcr-
22a) Refrigerant
Henry Nasution, Zulkarnain Abdul Latif,
Azhar Abdul Aziz, Mohd Rozi Mohd Perang
IMAT-UI 008
Characterization Of Generator With
Palm Oil Biodiesel At Different
Compression Ratio
Belyamin, Alias Bin Mohd. Noor, Mohanad
Hamzah Hussein, Mazlan Bin Said, Mohd
Hafidzi
IMAT-UI 009
The Effect Of Fuel Additives On
Gasoline Heating Value And Spark
Ignition Engine Performance
Zulkarnain Abdul Latif, Azhar Abdul Aziz,
Mohd Rozi Mohd Perang, Normaliza
Abdullah
IMAT-UI 010
Design a Four-Stroke Homogeneous
Charge Compression Ignition (HCCI)
Engine
Mohd Rozi Mohd Perang, Zulkarnain Abdul
Latiff, Azhar Abdul Aziz, Mohamad Azzad
Mokhri
IMAT-UI 011
R22 and Various Mixtures of
R290/R600a as its Alternative in
Adiabatic capillary tube Used in split-
type Air-conditioning System
Shodiya Sulaimon, Azhar Abdul Aziz,
Henry Nasution, Amer Nordin Darus
IMAT-UI 012
Flow Pattern at Pipe Bends on
Corrosion Behaviour of Low Carbon
Steel
Muhammadu Masin MUHAMMADU
IMAT-UI 013 Green Refrigerant for Multi-Circuit Air-
Conditioning System Hayati Abdullah and Alif Jalaludin
IMAT-UI 014
Friction Characteristic of Palm Olein at
Different Operating Temperature using
Four-ball Tribometer
S. Syahrullail, C.I Tiong
The 5th IMAT, November 12 – 13th
, 2012
viii
IMAT-UI 015
Pulse Detonation Engine Research
Development at High Speed Reacting
Flow Laboratory, Universiti Teknologi
Malaysia
Mazlan A. Wahid, A. Dairobi G.,
Aminuddin Saat, Mohsin M. Sies, H.A.
Mohammed, A. N. Darus, Mohd Faizal H.,
M. Ibthisham A., Fairus M. Y. and Z. Lazim
IMAT-UI 016 Pool Boiling Of Nanofluids In Vertical
Porous media Nandy Putra, Ridho Irwansyah
IMAT-UI 017 Analysis of Small Bubble
Characteristics in Alum Solution Warjito, Nurrohman
IMAT-UI 018 Effect Of Hot Air Reservoir in The
Development of Vacuum Freeze Drying
M. Idrus Alhamid, Nasruddin, Engkos A.
Kosasih, Muhamad Yulianto
IMAT-UI 019
Experimental of Cascade Refrigeration
System Using Natural Refrigerant
Mixture Ethane and Carbon Dioxide at
Low Temperature Circuit and
Refrigerant Natural Propane at high
temperature circuit
Nasruddin, M. Idrus Alhamid and Arnas
IMAT-UI 020
Performance Analysis of
Thermoacoustic-standing Wave As a
Power Generation Adi Suryo, Sentosa I
IMAT-UI 021 Factors Affecting Performance Of Dual
Fuel Compression Ignition Engines
Mohamed Mustafa Ali, Sabir Mohamed
Salih
IMAT-UI 022
Solar Air-Conditioning System Using
Single-Double Effect Combined
Absorption Ahiller
Hajime Yabase
IMAT-UI 023
Environmental Protection and Fuel
Consumption Reduction Flameless
Combustion Technology : A Review
Seyed Ehsan Hosseini, Saber Salehirad,
Mohsin Mohd Sies, Mazlan Abdul Wahid
IMAT-UI 024
The Effect Of Geometrical Parameters
On Heat Transfer Of Micro-Channels
Heat Sink
Law Wen Zhe, Amer Nordin Darus.
IMAT-UI 025
Investigation of the Velocity Profiles in
a Ninety-Degree Curved Standing Wave
Resonator with PIV
Normah M.G, Irfan Abd. R, Quenet T, Zaki
Ab.M
IMAT-UI 026 MED+AD Desalination Cycle Muhammad Wakil Shahzad, Kim Choon Ng,
Won Gee Chum
IMAT-UI 027 Kinetics Of Propane Adsorption On
Maxsorb III Activated Carbon
Azhar bin Ismail, Loh Wai Soong, Ng Kim
Choon
IMAT-UI 028
Effects of Natural Ventilations on
Indoor Air
of a Double-Storey Residential House
in Malaysia
Haslinda Mohamed Kamara, Nazri Kamsah
& Kam Jia Liq
IMAT-UI 029
Deposit Forming Tendency of Biodiesel
and Diesel Fuel due to High Pressure
Exposure
Muhamad Adlan Abdullah, Arshad Salema
and Farid Nasir Ani
The 5th IMAT, November 12 – 13th
, 2012
ix
IMAT-UI 030
Numerical Analysis of
Elastohydrodynamic Lubrication with
Non-Newtonian Lubricant
Dedi Rosa Putra Cupu, Adli Bahari, Kahar
Osman, Jamaluddin Md Sheriff
IMAT-UI 031
Latest System Simulation Models in
Heating, Refrigerating and Air
Conditioning Field, and Development
of System Simulator
Kiyoshi Saito and Jongsoo Jeong
IMAT-UI 032 Drag Reduction of Bamboo and Abacca
Fiber Suspensions in Circular Pipe
Gunawan, M. Baqi, Yanuar and Sanlaruska
Faternas
IMAT-UI 033
A Study on The Effect of Exhaust
Gases on The Indoor Air Quality
Onboard Naval Ships
Arman Ariffin and Hayati Abdullah
IMAT-UI 034 Application Of Thermal Energy Storage
For A Lowland Farming House
Haslinda Mohamed Kamar, Nazri Kamsah,
Norull Ahmad Azman
IMAT-UI 035
Transient Model And Entropy Analysis
Of A Lithium Bromide – Water
Absorption Chiller
Ang Li and Prof. Kim Choon Ng
IMAT-UI 036 Characteristics of Sea-water Ice Slurry
for Cooling of Fish
A.S. Pamitran, M. Novviali, H.D.
Ardiansyah
IMAT-UI 037
Fluid Flow Characteristic Of Rounded-
Shape FPSO and LNG Carrier During
Off Loading
Mufti Fathonah Muvariz, Jaswar Koto,
Agoes Priyanto
IMAT-UI 038
Comparison Of Simulation Organic
Renkine Cycle (ORC) System Using
Turbocharger and Cycl Tempo V.5
With Environment Friendly Fluid
Ruly Rutranta, M. Idrus Alhamid and
Harinaldi
IMAT-UI 039 Thermophysical Properties of Novel
Zeolite Materials for Sorption Cycles
Kyaw Thu, Young-Deuk Kim, Baojuan Xi,
Azhar Bin Ismail, Kim Choon Ng
IMAT-UI 040
Review Paper: Sea-water Ice Slurry
Generator and Its Application on
Indonesian Traditional Fishing
A.S. Pamitran, H.D. Ardiansyah, M.
Novviali
IMAT-UI 041
Improving Hydrogen Storage Capacity
on Metal Doping Carbon Nanotubes
Using Molecular Dynamics Simulation
Nasruddin, Engkos A. Kosasih, Supriyadi,
Abdul Jabar
IMAT-UI 042 Performance of Thermoelectric and
Heat Pipes Refrigerator Cooling System
Firman Ikhsan, Ali A. Sungkar, M. Afin
Faisol, M. Zilvan Bey, Nandy Putra,
Saripudin
IMAT-UI 043
Preliminary Study on Length of Candle
Filter Surface on the Flow Pattern in
Freeboard of Fluidised Bed Gasifier
A. Farhan Faudzi, Kahar Osman, Nor
Fadzilah Othman, Mohd Hariffin Bosrooh
IMAT-UI 044
Preliminary Study on the Effect of Type
Distributor Plate on Airflow Pattern in
Bubbling Fluidised Bed
Nofrizalidris Darlis, Kahar Osman, Ab
Malik A. Hamat, Nor Fadzilah Othman,
Mohd Hariffin Bosrooh
IMAT-UI 045 Natural Convection in A Differentially
Heated Cavity Using Splitting Method Ubaidullah S., Kahar Osman
The 5th IMAT, November 12 – 13th
, 2012
1
Controlled Auto-Ignition (CAI) Combustion In A Two-Stroke
Cycle Engine Using Hot Burned Gases
Amin Mahmoudzadeh Andwaria, Azhar Abdul Aziz
b, M.F. Muhamad Said
c, Z. Abdul Latiff
d
Automotive Development Center (ADC), Faculty of Mechanical Engineering
Universiti Teknologi Malaysia (UTM), 81310, Johor Bahru, Malaysia
ABSTRACT
A new combustion concept, which is viewed
increasingly as a probable solution to these issues
is Controlled Auto-Ignition (CAI) Combustion. In
such an engine, a homogeneous mixture of air, fuel
and residual gases is compressed until auto-ignition
occurs. Due to its significantly low temperature
combustion, NOx will be dramatically reduced
while the mixture will be under ultra-lean fuel-air
condition, thus able to achieve high efficiency and
low emission. In the case of two-stroke engine,
problem of poor combustion efficiency and
excessive white smoke emission can be addressed
by the incorporation some features that will
ultimately convert a typical two-stroke engine into
an efficient CAI engine demonstrating the best of
both features. Due to its inherent high internal
residual gas rate in partial load operation, the two-
stroke engine has been the first application to
benefit from the unconventional CAI combustion
process. This paper will concisely discuss the
utilization of hot burned gas for induction thus
imposing a CAI combustion feature onto a
reference two-stroke cycle engine. Among the
features incorporated are the increasing in the level
of Internal or External Exhaust gas Recirculation
(In/Ex-EGR) and cycle-by-cycle uniformity of the
air-fuel ratio (AFR) supplied to cylinder which will
be crucial in creating a suitable temperature within
the engine‘s combustion chamber.
Keywords : Two-Stroke cycle engine, Controlled
Auto-Ignition (CAI), hot burned gas,
auto-ignition temperature, ATAC
(Active Thermo-Atmospheric
Combustion), Homogeneous Charge
Compression Ignition (HCCI),
Exhaust gas Recirculation (EGR)
1. INTRODUCTION
Energy conservation and environmental protection
are exerting rigorous demand on internal
combustion engine developers to further improve
fuel economy and emission reduction. In addition,
concerns about the world‘s finite oil reserves and,
more recently, by CO2 emissions brought about
climate change has led to heavy taxation of road
transport, mainly via on duty on fuel. Over the last
30 years, levels of NOx, CO and HC emissions
from vehicles have been dramatically increased.
This has been motivated by a continually tightening
band of legislation related to emission of these
pollutants.
Two-stroke cycle engines are well known owing to
their light weight, simple construction, less
components, cheap to manufacturing and the
potential to pack almost twice the power-density
than that of a four-stroke engine having similar
capacity [1]. For a longtime, the objective of the
different research works on two-stroke engines
optimization was to eliminate its two main
drawbacks leading to high emissions of unburned
hydrocarbons (uHC) and poor fuel efficiency. The
first one is the unstable running operation
combined with incomplete combustion, especially
at light load. The second one is fuel short circuit at
medium and full load. However due to the short-
circuiting of the fuel before combustion, this has
resulted in deterioration in overall performances
especially poor combustion efficiency and high
white smoke emission problem [2].
For that reason, many researchers have begun to
research the new kind of alternative combustion
intensively. One example of these attempts is the
ATAC (Active Thermo-Atmospheric Combustion)
[3], TS Combustion (Toyota-Soken) [4], ARC
(Activated Radical Combustion) [5], Homogeneous
Charge Compression Ignition (HCCI) [6] and CAI
(Controlled Auto-Ignition) combustion concept,
bulk combustion or low temperature combustion, a
combustion process that has conventionally been
used in two-stroke engines. It has been found that
depending on the engine speed, load ratio and level
of Exhaust Gas Recirculation (EGR) applied, it is
possible to induce Auto-Ignited (AI) combustion in
a two-stroke engine as a result of the mixing of
unburned mixtur
IMAT-UI 001
The 5th IMAT, November 12 – 13th
, 2012
2
e gas introduced into the cylinder and hot residual
(burned) gas [5]. These combustion processes can
reduce emissions of unburned HCs and allow stable
engine operation by lower cyclic variation. Owing
to its inherent high internal residual gas rate in
partial load operation, the two-stroke engine has
been the first application to take benefit of the
unconventional CAI combustion process.
Several issues that must be addressed in order to
implement this combustion process in production
engines include control of the ignition timing and
burning rate, which are determined by the
chemical reactions of the unburned mixture, and
expansion of the stable operating region [7]. In
fact, the most recognized original work on CAI
combustion was motivated by some researchers
desire to control the irregular combustion caused
by the auto-ignition of cylinder charge to obtain
stable lean-burn combustion in the conventional
ported two-stroke gasoline engine [8], [3], [4].
It is known that these drawbacks result from the
new mixture shortcut and irregular combustion in
the part load operation. In the light load operating
range, the residual gas in combustion chamber
increases due to the poor scavenging. Normal
flame propagation is disturbed by the large amount
of residual gas, which generates irregular
combustion [5]. The research works during this
period to study the part load lean two-stroke
combustion have led to discover that the
irregularities of the combustion and the auto-
ignition, which are considered as the weak points
of the two-stroke engine, can be effectively
controlled and managed to get a part load stable
two-stroke combustion process for lean mixtures in
which ignition occurs without spark assistance.
Suitably, remarkable improvements in stability,
fuel efficiency, exhaust emissions, noise and
vibration will be achieved [3], [8].
2. ORIGINATION OF CAI
COMBUSTION AND ITS
FUNDAMENTALS
Although it is generally accepted that the first
systematic investigation on the new combustion
process was carried out by Onishi and Noguchi in
1979, the theoretical and practical roots of the CAI
combustion concepts are attributed to the
pioneering work carried out by the Russian
scientist Nikolai Semenov and his colleagues in the
field of ignition in the 1930s. Having established
his chemical or chain theory of ignition, Semenov
sought to exploit a chemical-kinetics controlled
combustion process for IC engines, in order to
overcome the limitations imposed by the physical
dominating processes of SI and CI engines [9].
2.1 Two-Stroke CAI Combustion Engine
The CAI combustion story has been started with
two-stroke engines. Substantial research work was
performed from the end of the 1960s to the end of
the 1970s in order to solve one of the main
problems of the two-stroke engine that was the
unstable, irregular and incomplete part load
combustion responsible for excessive emissions of
unburned hydrocarbons (uHC). Researchers
performed a lot of investigation work during this
period to study the part load lean two-stroke
combustion. They discovered that the irregularities
of the combustion and the auto-ignition that were
considered as the weak points of the two-stroke
engine could be effectively controlled. The
research‘s objective was managed to get a part load
stable two-stroke combustion process for lean
mixtures in which ignition occurs without spark
assistance. The new combustion process occurring
without flame front was called ‗ATAC‘ for (Active
Thermo-Atmosphere Combustion) [3]. During the
same year, another researchers‘ paper concerning
two-stroke auto-ignition was published. They
named this auto-ignition the TS (Toyota-Soken)
combustion process. They also found that such
combustion occurred similarly without flame front
while showing excellent efficiency and emissions
figures. They were the first to suggest that active
radicals in residual gases could play an important
role in the auto-ignition process [4].
2.2 Basic Principle of CAI Combustion
Similar to a conventional SI engine, in a CAI
engine the fuel and air are mixed together either in
the intake system or in the cylinder with direct
injection. The premixed fuel and air mixture is then
compressed. Towards the end of the compression
stroke, combustion is initiated by auto-ignition in a
similar way to the conventional CI engine. The
figure 1 shows ideal representations of both SI and
CAI combustion processes. In the case of SI
combustion, it is the flame front that separates the
burned gases from the fresh unburned gases and its
velocity controls the combustion heat release. In
the case of CAI, the combustion reactions take
place with multiple auto-ignition sites. Even if the
combustion locally can progress slowly, since it
occurs spontaneously and simultaneously at several
locations within the combustion chamber, the
overall heat release can be as fast or even faster
than with the flame front controlled SI without
generating the typically high combustion
temperatures of the flame front. This could
contribute to explain the CAI low NOx emissions
advantage.
The 5th IMAT, November 12 – 13th
, 2012
3
Figure 1: Spark Ignition (SI) combustion (left)
and CAI combustion (right)
The heat release characteristics of the CAI
combustion can be seen with using figure 1. In the
case of SI combustion, a thin reaction zone or
flame front separates the cylinder charge into
burned and unburned regions and the heat release is
confined to the reaction zone. Thus, flame front
velocity controls the combustion heat release. As it
can be seen in equation 1, the cumulative heat
released in a SI engine is therefore the sum of the
heat released by a certain mass, dmi, in the reaction
zone and it can be expressed as
(1)
where q is the heating value per unit mass of fuel
and air mixture, N is the number of reaction zones.
Figure 2: Heat release characteristics of SI (a)
and CAI combustion (b) [3]
In an idealized CAI combustion process,
combustion reactions take place simultaneously in
the cylinder and all the mixture participates in the
heat release process at any instant of the
combustion process. In other word, the combustion
reactions take place with multiple auto-ignition
sites. Even if the combustion locally can progress
slowly, since it occurs spontaneously and
simultaneously at several locations within the
combustion chamber, the overall heat release can
be as fast or even faster than with the flame front
controlled SI without generating the typically high
combustion temperatures of the flame front. This
could contribute to explain the CAI low NOx
emissions advantage that will be described in the
following section. Regarding equation 2, the
cumulative heat release in such an engine is
therefore the sum of the heat released from each
combustion reaction, dqi, of the complete mixture
in the cylinder, m, i.e.
(2)
where K is the total number of heat release
reactions, and qi is the heat released from the ith
heat release reaction involving per unit mass of fuel
and air mixture. Whereas the entire heating value
of each minute parcel of mixture must be released
during the finite duration spend in the reaction zone
in a SI engine, heat release takes place uniformly
across the entire charge in an idealized CAI
combustion. However, in practice, due to in-
homogeneities in the mixture composition and
temperature distributions in a real engine, the heat
release process will not be uniform throughout the
mixture. Faster heat release can take place in the
less diluted mixture and/or high temperature
region, resulting in a non-uniform heat release
pattern as indicated by the dashed lines [3].
3. TWO-STROKE CAI COMBUSTION
CONTROLLING
In spark ignition mode the combustion can be
rather easily directly controlled by the spark
advance. In the case of CAI combustion, there are a
lot of relevant control parameters with, in addition,
complex interactions between some parameters.
Prior to examining in more detail the main relevant
two-stroke CAI control parameters, it is important
to define what has to be controlled: The
Combustion Timing and The Combustion Heat
Release Rate [10]. A correctly controlled CAI
combustion should have the best combustion
timing for the highest combustion efficiency.
3.1 Mixing Between Fresh Charge and Burned
Gases
Inherently in a two-stroke engine, there is a high
amount of internal EGR at part load. But if this
EGR is well mixed with the fresh charge, as is the
case in a conventional two-stroke engine,
especially at low engine speed, it has almost no
effect on the combustion. What is efficient for
getting CAI is to limit as much as possible the
mixing/Stratification of this internal EGR with the
fresh mixture. In such cases, it is possible to
achieve a temperature gradient within the charge
for the same overall amount of in-cylinder EGR,
which means, the same in-cylinder heat
content.[11], [12], [13].
3.2 The Engine Speed
The engine speed is an indirect CAI control
parameter. It has an indirect effect on the
mixing/stratification between the fresh charge and
the EGR. When the engine speed increases the time
for mixing between the internal EGR and the fresh
charge will be shorter, therefore the internal
stratification and the temperature gradient inside
the trapped charge. This has the final consequence
The 5th IMAT, November 12 – 13th
, 2012
4
of advancing and accelerating the CAI combustion
[5].
3.3 The In-cylinder Flow Velocities
To introduce the same amount of fresh mixture in a
much longer time means that the charge is
introduced more smoothly at significantly lower
velocities. This helps to prevent the dilution of the
fresh inlet mixture in the residual gases which is
favorable for getting CAI [14].
3.4 The In-Cylinder Pressure
The easiest solution for controlling the in-cylinder
pressure is to do it through the compression ratio
(CR). A high compression ratio will then be
favorable to extend the CAI range to the low speed
low loads. Nevertheless, the choice of the highest
compression ratio favorable for CAI is always
limited by the fact that the same engine also has to
be able to run in spark ignition at full load without
knock [15].
3.5 The Overall Temperature
Heating the intake charge increases the overall gas
temperature and has the effect of advancing the
CAI combustion timing and therefore of extending
the CAI combustion range in the low load low
speed region. Similarly, the CAI combustion is also
sensitive to the engine liquid cooling temperature,
which indirectly affects the overall gas temperature
[16].
3.6 The Fuel Formulation
Several researchers have studied the effect of the
fuel formulation on two-stroke CAI combustion.
Results show that running an ATAC-CAI two-
stroke engine with methanol allows significant
widening of the auto-ignition range [17]. More
recently, some researchers tried to find some
correlations between the octane number of several
fuels (research octane number RON and motor
octane number MON) and their effect on the auto-
ignition range.
3.7 Changing of Two-Stroke Engine Design
Among the possible control parameters, the in-
cylinder gas temperature effect obtained by
stratifying hot internal EGR (the EGR and fresh
charge mixture/stratification being controlled by
the in-cylinder flow velocities) and the in-cylinder
pressure are the most relevant for practical
application mainly because of their immediate
response time in transient operation (which, for
example, is not the case of intake air heating). Most
of the technologies that have been developed and
applied to obtain CAI combustion on two-stroke
engines were based on the control of these two
internal temperature and pressure effects. For this
purpose, three main technologies and associated
control devices have been developed which
include: Elongated Transfer Duct, Transfer Duct
Throttling and Exhaust Port Throttling [3, 7, 12].
4. EFFECT OF EXHAUST GAS AS
DILUENT
In order to achieve CAI/HCCI combustion, the
temperature of the charge at the beginning of the
compression stroke has to be increased to reach
auto-ignition conditions at the end of the
compression stroke. This can be done by heating
the intake air or by keeping part of the hot
combustion products (charge dilution) in the
cylinder. Both strategies result in a higher gas
temperature throughout the compression process,
which in turn speeds up the chemical reactions that
lead to the start of combustion of homogeneously
mixed fuel and air mixtures. In-cylinder gas
temperature must be sufficiently high to initiate and
sustain the chemical reactions leading to auto-
ignition processes. Substantial charge dilution is
necessary to control runaway rates of the heat
releasing reactions. Both of these requirements can
be realized by recycling and/or trapping the burned
gases within the cylinder, which the former is
represented as External-EGR and the latter is
known as Internal-EGR, respectively.
The presence of the recycled or trapped burned
gases has a number of effects on the CAI
combustion and emission processes within the
cylinder.
4.1 The Charge Heating Effect
If hot burned gases are mixed with cooler inlet
mixture of fuel and air, the temperature of the
intake charge increases owing to the heating effect
of the hot burned gases. This is often the case for
CAI combustion with high-octane fuels, such as
gasoline and alcohols [18].
4.2 The Dilution Effect
The introduction or retention of burned gases in the
cylinder replaces some of the inlet air and hence
causes a substantial reduction in the oxygen
concentration. The reduction of air/oxygen due to
the presence of burned gases is called the dilution
effect [19].
4.3The Heat Capacity Effect
The total heat capacity of the in-cylinder charge
will be higher with burned gases, mainly owing to
the higher specific heat capacity values of carbon
dioxide (CO2) and water vapor (H2O). This rise in
The 5th IMAT, November 12 – 13th
, 2012
5
the heat capacity of the cylinder charge is
responsible for the heat capacity effect of the
burned gases [19].
4.4 The Chemical Effect
Combustion products present in the burned gases
can participate in the chemical reactions leading to
auto-ignition and subsequent combustion. This
potential effect is classified as the chemical effect.
It should be noted that the chemical effects are
influenced by active species or partially oxidized
hydrocarbons or activated radical [20].
It should be noted that the overall effect of hot
burned gases on the CAI combustion process is to
charge heating effect, to advance the Auto-Ignition
timing and to shorten the combustion. By hot
burned gas incorporation, the initial charge
temperature of the total in-cylinder charge will be
increased owing to the heating effect of hot burned
gases, and the relative air fuel ratio λ will be
reduced as burned gases would be replaced some of
the air [21].
The presence of hot burned gases initially causes
the CAI combustion process to accelerate. Both
experiments and analytical studies have shown that
the overall effect of hot burned gases is to advance
the start of CAI combustion due to their charge
heating effect. Ignition is dominated by the charge
heating effect but the combustion duration is
dominated by the dilution and heat capacity effect.
The maximum rate of heat release is equally
affected by the charge heating effect and by the
combined dilution and heat capacity effect. From a
macroscopic point of view of the heat balance, i.e.
the relationship between the calorific value
supplied in a cycle and the total heat capacity of the
in-cylinder gases; a larger heat capacity will take a
longer time to heat up and the maximum
combustion temperature will be lower. Thus,
combustion of a larger heat capacity generates a
slower heat release while that of a smaller heat
capacity permits a quicker heat release. For high-
octane fuels, like gasoline, alcohols, natural gas,
etc., it will be advantageous to retain burned gases
at as high temperature as possible to promote auto-
ignition of fuel/air mixture, particularly at low load
operations.
CONCLUSION
Two-stroke cycle engines can be more efficient and
clean by CAI combustion mode operation. Hot
burned gas utilization in order to induce this unique
combustion has been always interested as result of
some specific advantages, which are included: the
charge heating effect, the dilution effect, the heat
capacity effect and the chemical effect. In general,
hot burned gases that are used in two-stroke engine
either trapped or recycled, have considerable effect
upon combustion phenomenon and its
characteristics which will be led to induce and
control of the CAI combustion as follow:
Hot burned gases are preferred in most cases in
order to increase cylinder charge temperature
without external heating source
Overall effect of hot burned gases is to advance
the start of CAI combustion due to their charge
heating effect.
Start of ignition is dominated by the charge
heating effect
Combustion duration is dominated by the
dilution and heat capacity effect.
Maximum rate of heat release is equally
affected by the charge heating effect and by the
combined dilution and heat capacity effect.
The larger the heat capacity, the slower the heat
release in combustion.
It is most desired for gasoline, alcohols and
natural gas as high-octane fuels to use the hot
burned gases at as high temperature as possible
to promote auto-ignition of mixture, specifically
at low load.
REFRENCE
[1] J. B. Heywood, E. Sher, and S. o. A.
Engineers, The Two-Stroke Cycle Engine: Its
Development, Operation, and Design: Taylor
& Francis, 1999.
[2] G. P. Blair, and S. P. A. S. P. Committee,
Advances in Two-Stroke Cycle Engine
Technology: Society of Automotive
Engineers, 1989.
[3] S. Onishi, S. H. Jo, K. Shoda et al., ―Active
Thermo-Atmosphere Combustion (ATAC) -
A New Combustion Process for Internal
Combustion Engines,‖ 1979.
[4] M. Noguchi, Y. Tanaka, T. Tanaka et al., ―A
Study on Gasoline Engine Combustion by
Observation of Intermediate Reactive
Products during Combustion,‖ 1979.
[5] Y. Ishibashi, and M. Asai, ―Improving the
Exhaust Emissions of Two-Stroke Engines
by Applying the Activated Radical
Combustion,‖ 1996.
[6] R. H. Thring, ―Homogeneous-Charge
Compression-Ignition (HCCI) Engines,‖
1989.
[7] P. Duret, and J.-F. Moreau, ―Reduction of
Pollutant Emissions of the IAPAC Two-
Stroke Engine with Compressed Air Assisted
Fuel Injection,‖ 1990.
The 5th IMAT, November 12 – 13th
, 2012
6
[8] S. H. Jo, P. D. Jo, T. Gomi et al.,
―Development of a Low-Emission and High-
Performance 2-Stroke Gasoline Engine
(NiCE),‖ 1973.
[9] G. P. Blair, The Basic Design of Two-stroke
Engines: Society of Automotive Engineers,
1990.
[10] P. Duret, A New Generation of Engine
Combustion Processes for the Future?:
Proceedings of the International Congress,
Held in Rueil-Malmaison, France,
November, 26-27, 2001: Editions Technip,
2002.
[11] P. Duret, A New Generation of Two-stroke
Engines for the Future?: Proceedings of the
International Seminar Held in Rueil-
Malmaison, France, November 29-30, 1993:
Éditions Technip, 1993.
[12] Y. Ishibashi, ―Basic Understanding of
Activated Radical Combustion and Its Two-
Stroke Engine Application and Benefits,‖
2000.
[13] N. Iida, Y. Yamasaki, S. Sato et al., ―Study
on Auto-Ignition and Combustion
Mechanism of HCCI Engine,‖ 2004.
[14] P. Duret, J.-C. Dabadie, J. Lavy et al., ―The
Air Assisted Direct Injection ELEVATE
Automotive Engine Combustion System,‖
2000.
[15] K. Tsuchiya, Y. Nagai, and T. Gotoh, ―A
Study of Irregular Combustion in 2-Strote
Cycle Gasoline Engines,‖ 1983.
[16] Y. Ishibashi, and M. Asai, ―A Low Pressure
Pneumatic Direct Injection Two-Stroke
Engine by Activated Radical Combustion
Concept,‖ 1998.
[17] N. Iida, ―Combustion Analysis of Methanol-
Fueled Active Thermo-Atmosphere
Combustion (ATAC) Engine Using a
Spectroscopic Observation,‖ 1994.
[18] R. Stone, Introduction to Internal
Combustion Engines, 3rd Edition: Solutions
Manual, 1999.
[19] J. B. Heywood, Internal combustion engine
fundamentals: McGraw-Hill, 1988.
[20] C. F. Taylor, The Internal-combustion
Engine in Theory and Practice:
Thermodynamics, fluid flow, performance:
M.I.T. Press, 1985.
[21] A. Cairns, and H. Blaxill, ―The Effects of
Combined Internal and External Exhaust Gas
Recirculation on Gasoline Controlled Auto-
Ignition,‖ 2005.
The 5th IMAT, November 12 – 13th
2012
7
Determination of Motive Nozzle and Constant-Area Diameters:
Numerical Study of Ejector as an Expansion Device in Split-type
Air Conditioner
Kasni Sumeru
a, Henry Nasution, Farid Nasir Ani
*
aDepartment of Refrigeration and Air Conditioning
Politeknik Negeri Bandung, Indonesia
Email: [email protected]
Department of Thermodynamics and Fluid Mechanics, Faculty of Mechanical Engineering
Universiti Teknologi Malaysia, Skudai 81310 Johor
*Email: [email protected]
ABSTRACT This paper presents a numerical approach for
determining the motive nozzle and constant-area of
an ejector as an expansion device, based on cooling
capacity of the split-type air conditioner using R22
as working fluid. The use of an ejector as an
expansion device in split-type air conditioner can
improve the coefficient of performance (COP).
Typically, the split-type air conditioner may be
installed on the geographical area with moderate or
high outdoor air temperature using capillary tube.
For this reason, the motive nozzle and constant-
area diameters of the ejector must be designed
according to these conditions. The diameters of the
ejector are crucial in improving the COP. Three
equations are applied in developing the numerical
model on the ejector: conservation laws of mass,
momentum and energy equations. The results
showed that the motive nozzle diameter is constant
(1.14 mm) with variations of the condenser
temperature, whereas the constant-area diameter
decreases as the condenser temperature increases.
Keywords: Condenser temperature, COP, ejector,
expansion device, R22.
1. INTRODUCTION
Split-type air conditioner (AC) typically uses a
vapor compression refrigeration cycle (VCRC).
The air-conditioning system uses approximately
50% of the total energy consumption of a building
[1]. As a result, a small improvement on the
performance of the system will generate a
significant impact on energy saving.
Split-type air conditioner is the most widely used
as residential and commercial air conditioners. This
type of AC generally uses a capillary tube as an
expansion device. Shodiya et al. [2] developed a
numerical model to improve the capillary tube
performance prediction using enthalpy equation
and also included metastability phenomenon to
further improve the performance [3]. The use of
capillary tube generates irreversible process on the
throttling and causes energy loss during expansion
from high pressure to low pressure. Replacing
capillary tube by an ejector as an expansion device
in the AC is an alternative way of improving COP.
Theoretically, the pressure drop in the conventional
expansion devices is considered isenthalpic process
(constant enthalpy). Isenthalpic process causes a
decrease in the evaporator cooling capacity because
of energy loss in the throttling process. To recover
this energy loss, isentropic (constant entropy) is
required in the expansion process. An ejector can
be used to generate isentropic condition in the
throttling process.
Based on literature survey, we have not found any
study on the ejector-expansion refrigeration cycle
(EERC) that investigate determination of diameter
of the motive nozzle and the constant-area, based
on the cooling capacity of the air conditioner using
R22 as refrigerant. Hydrochlorofluorocarbons-22
(R22) is the most commonly used refrigerant in
split-type air conditioner. The objective of the
present study is to obtain the main geometric
parameter of an ejector, namely diameters of the
motive nozzle and the constant-area of ejector and
COP improvement on the split-type air conditioner
using R22 as the working fluid.
2. SYSTEM DESCRIPTION
2.1 Ejector Expansion Refrigeration Cycle
In 1931, Gay patented the ejector as an expansion
device in the refrigeration system. He patented the
ejector to minimize throttling losses in expansion
device on the vapor compression refrigeration
cycle. In 1966, Kemper et al. [4] modified the
Gay‘s patent, by using a pump and a heater to
increase pressure and temperature the liquid stream
before entering motive nozzle. In 1972, Newton
(1972) [5-6] proposed patent to improve previous
patent by Kemper [4]. Newton applied a hot gas
IMAT-UI 002
The 5th IMAT, November 12 – 13th
2012
8
from compressor discharge on the liquid stream
before entering to motive nozzle.
In 1950, Keenan et al. [7] carried out experimental
and numerical investigation and concluded that
there are two types of ejector, namely constant-area
mixing ejector and constant-pressure mixing
ejector. Figure 1 shows a constant-area mixing
ejector which has three sections: a nozzle section, a
constant-area mixing section, and a diffuser. In the
constant-area mixing ejector, the mixing between
the primary and secondary flow occurs in the inlet
of the constant-area.
Figure 1: Constant-area mixing ejector
Figure 2 shows a constant-pressure mixing ejector
which has four sections: a nozzle section, a mixing
section, a constant-area and a diffuser. The mixing
between primary and secondary flow occurs in the
mixing section or suction chamber, before
constant-area.
Figure 2: Constant-pressure mixing ejector.
The constant-pressure mixing ejector has a better
performance than that of the constant-area mixing
ejector [4], as a result, a constant-pressure mixing
ejector is generally used in the various refrigeration
applications, especially in the ejector refrigeration
systems. However, Yapici and Ersoy [8] found that
for the same operating temperature, the constant-
area mixing ejector has higher COP than that of the
constant-pressure mixing ejector. Furthermore, in
the last decade, the constant-area mixing ejector is
widely used in numerical and experimental studies
on the EERC [9-13].
Figure 3 depicts the schematic diagram of the
standard cycle and an ejector as an expansion
device. A capillary tube or expansion valve is used
as an expansion device in the standard cycle,
whereas an ejector is used for an expansion device
in the EERC.
Figure 3: Schematic diagram of the vapor
compression refrigeration cycle:
(a) Standard cycle, (b) Ejector-
expansion cycle.
2.2 COP Improvement on the VCRC
The advantages of the ejector as an expansion
device have been demonstrated by several
researchers. Kornhauser [14] was the first to
perform a thermodynamics analysis of vapor
compression refrigeration cycle using an ejector as
an expansion device. He proposed a one-
dimensional model in his study. He found that the
COPimp was up to 21% over the standard cycle. Liu
et al. [15], Li and Groll [16] and Deng et al. [17]
developed a mathematical model and found that the
COPimp was between 6-14%, 7-18% and 22%,
respectively, over the standard vapor compression
refrigeration cycle using CO2 as a refrigerant.
Takeuchi et al. [18] reported an increase of 45-64%
COPimp for a vehicle refrigeration system. Disawas
and Wongwises [19] and Elbel and Hrnjak [20]
carried out experimental investigation on an ejector
as the expansion device using R134a and CO2
respectively, and reported an increment in the COP
over the standard cycle. Nehdi et al. [9] presented a
numerical analysis to determine the effect of the
geometry of ejectors on system performance using
twenty synthetic refrigerants. They found that the
COPimp over the standard cycle is 22%. A
numerical study using natural refrigerant on the
EERC was also performed by Sarkar [13] and
reported that the maximum COPimp for isobutane,
propane and ammonia are 21.6%, 17.9% and
11.9% respectively. Bilir and Ersoy [11] performed
a computational analysis of the performance
improvement of ejector expansion cycle over
standard cycle. Their computational methods are
similar to that of Kornhauser [14]. Using an R134a
refrigerant, the COP improvement of the expansion
cycle over standard cycle is 10.1-22.34%. They
found also that the COP improvement increases
when the condenser temperature increases. This
means that the use of ejector instead of an
expansion valve is more advantageous in the air-
cooled condensers than that of water-cooled
condensers.
The 5th IMAT, November 12 – 13th
2012
9
Figure 4 illustrates the standard cycle and EERC in
the Ph diagram. The isenthalpic throttling process
is from point 3 to 11, while isentropic throttling
process is from point 3 to 4. Refrigerant flow on
the Ph diagram of standard cycle is point 8, 2b, 3,
11 and 8. There are two flow on the EERC,
primary and secondary flow. The primary flow is
circulated by a compressor through condenser,
ejector and separator (point 1, 2, 3, 4, 10, 5 and 1),
whereas the secondary flow circulates in the
capillary tube, evaporator, ejector and separator
(point 6, 7, 8, 9, 10, 5 and 6). The primary and
secondary flow mixes at constant-area and diffuser
(point 10 and 5).
Figure 4: Ph diagram of the ejector-
expansion and standard cycle.
As shown in Figure 4, the pressure at point 1 is
higher than that of suction pressure in the standard
cycle (point 8). This means that the compressor
work of the ejector expansion cycle is lower than
that of the standard cycle. Based on Figure 4, the
COP of standard refrigeration cycle is calculated
as,
comp
bcomp
ecomp
comp
estd
hhm
hhm
W
QCOP
)(
)(
82
118
(1)
since compe mm , equation (1) becomes,
comp
b
stdhh
hhCOP
)(
)(
82
118 (2)
where ηcomp is the isentropic efficiency of the
compressor which is calculated by an empirical
relation by Brunin et al. [21] as,
suct
disccomp
P
P01345.0874.0 (3)
The COP of the EERC is calcylated as,
comp
comp
ecomp
comp
eej
hh
hh
m
m
W
QCOP
)(
)(
12
78
(4)
Furthermore, the COP improvement of the EERC
over the standard cycle can be calculated by,
tds
tdsej
impCOP
COPCOPCOP
)( (5)
Two parameters viz. the entrainment ratio (ω) and
pressure lifting ratio (Plift) are used to investigate
the EERC performance.
c
e
m
m
(6)
8
1
,
,
P
P
P
PP
oute
outdif
lift (7)
Both quantities should be as large as possible to
obtain optimum COPimp. High ejector pressure
lifting ratio decreases the compression ratio of the
compressor. Increasing the mass entrainment ratio
reduces compressor mass flow rate for a given
cooling capacity. Ejector efficiency increases when
mass entrainment ratio and/or pressure lifting ratio
increase. However, the entrainment ratio cannot be
increased as high as possible, because it will cause
the flow of refrigerant on primary flow to reduce.
A high entrainment ratio will cause the primary
flow as a driven-flow to be weak.
The operation principle of ejector as an expansion
device is similar to the ejector function in other
applications, in which the primary flow from high
pressure induces the secondary flow from low
pressure in the suction section of ejector and brings
to a higher pressure at diffuser. Figure 5 illustrates
the refrigerant flow, pressure and velocity profile
inside an ejector. The motive flow or primary flow
from high pressure is accelerated and expands
through motive nozzle, from point 3 to 4. Very
high speed flow at point 4 causes pressure drop.
The low pressure at point 4 induces fluid from
secondary flow (point 8 to 9). The two flows mix in
the outlet of suction chamber and become one
stream in the constant-area. The mixing stream
flows through point 10 to the diffuser. In the
diffuser, the refrigerant experienced deceleration as
a result of pressure increase (point 5). The
refrigerant at point 5 circulates to separator. The
vapor refrigerant in the separator circulates through
compressor (point 1 to 2) as a primary flow,
whereas the liquid refrigerant from separator
circulates through expansion valve and evaporator
(point 6, 7 and 8), as a secondary flow.
The 5th IMAT, November 12 – 13th
2012
10
Figure 5: Refrigerant flow inside an ejector.
3. METHODOLOGY
Three equations viz. conservation of mass,
momentum and energy as shown in equation (8),
(9) and (10), respectively, are used to develop
thermodynamic model on each part of the ejector.
oooiii auau (8)
iiii umaP oooo umaP (9)
)2
()2
(22
o
oo
i
ii
uhm
uhm (10)
The following assumptions are made in the
calculation of each part of the ejector:
1. There are no heat transfer except in the
evaporator and condenser.
2. Properties and velocities are constant over the
cross section (one-dimensional).
3. The refrigerant condition is in thermodynamic
quasi-equilibrium.
4. There is no pressures drop along the evaporator
and condenser.
5. There is no wall friction.
6. The refrigerant conditions at the outlet of the
evaporator and condenser are saturated.
7. The pressure of exit of nozzle and suction
nozzle at the entrance of the constant-area are
assumed to have the same pressure.
8. Deviation from adiabatic reversible processes
for each section of ejector is calculated by
efficiencies.
In the present study, a constant-area mixing ejector
flow model is used as an expansion device. Based
on the manipulation of equations (1)-(3) and
referring to studies performed by Nehdi et al. [9]
and Sarkar [13], the thermodynamics modeling
produces equations (11)-(20), as shown in the
flowchart depicted in Figure 6. This flowchart was
used to determine the diameter of motive nozzle
(d4) and the constant-area of the ejector (d10). The
properties of the refrigerant were obtained from
REFPROP [22]. Using flowchart (Figure 6) and
properties of refrigerant, several parameters such as
the diameter of the motive nozzle and constant-
area, Plift, COPej and COPimp can be calculated.
Set n,d
Determine Te and Tc
At nozzle outlet:
h4 = h3 -n(h3-h4,is) (11)
u4 = [2(h3-h4)]0.5
44
4u
ma c
(12)
(13)
At mixing chamber:
(14)
(15)
(16)
2)(
1
1 2
108310
uhhh
At diffuser outlet:
2
2
10105
uhh d (20)
Determine P4 and ω
Update P4
1)1( 5 x
COPej, COPimp
d4, d10, Plift
Yes
No
1010
10u
mma ec
10410410 )()( ummumaPP ecc
2
10
4
10
42
10
4
2
44
410 )1(225.0
a
a
a
a
u
PP
1
1
1 4
8
4
10
(17)
4101
1uu
(18)
(19)
0)( 410 PP
Yes
Update ω and P4
No
evapTT 5
Figure 6: Flow chart of the calculation
algorithm of motive nozzle and
constant-area diameter.
4. RESULTS AND DISCUSSION
In this study, the cooling capacity and the mass
flow rate of refrigerant R22 in the split-type air
conditioner was taken based on the experimental
result of Zhou and Zhang [23], namely 2.4 kW and
56.95 kg/h, respectively. To start the iteration using
the flowchart in Figure 6, the values of nozzle and
diffuser efficiencies is chosen as 0.9 and 0.8,
respectively [11]. The influence of condenser
temperature on diameters, COPimp and Plift will be
explained in the subsection below.
The 5th IMAT, November 12 – 13th
2012
11
4.1 Influence on the diameters
Split-type AC may be installed on the geographical
area with moderate or high outdoor air temperature.
For this reason, the motive nozzle and constant-
area diameters of the ejector must be designed
according to these conditions. The dimension of
each section of the ejector is crucial in improving
the COP. The diameter of the motive nozzle is
calculated by using equation (13), while the
diameter of the constant-area is iterated by the
flowchart depicted in Figure 6. The iteration result
is shown in Figure 7. It shows that the diameter of
the motive nozzle is constant, at 1.14 mm, except at
the condenser temperature of 40oC, which is 1.15
mm. The size of the diameter of the motive nozzle
based on the numerical results is similar to those
used in the experiment by Caiwongsa and
Wongwises [24]. They tested three different motive
nozzle diameters 0.8, 0.9 and 1.0 mm, using
R134a. The motive nozzle with a diameter of 1.0
mm yields a higher mass flow rate in the condenser
than the other nozzles. Meanwhile, the smallest
diameter (0.8 mm) yields a lower mass flow rate in
the condenser, producing the highest COPimp. There
is slight difference between these numerical
approaches with the experimental data. This
distinction is caused, among other things, by the
differences in the working fluid and cooling
capacity.
The effect of the constant-area diameter with
increase in condenser temperature is shown in Fig.
7. The figure shows that the diameter of constant-
area decreases with increase in condenser
temperature. For example, diameters of the
constant-area are 2.49, 2.51, 2.48 and 2.46 mm
when the condenser temperatures are 40, 45, 50,
and 55oC, respectively. The figure also shows that
the area ratio (AR), the ratio between the cross-
sectional area of constant-area to motive nozzle
(a10/a4), decreases as the condenser temperature
increases. With similar results in the present study,
Sarkar [13] reported that the ejector AR decrease
with increase in temperature of condenser. The
results of the present study showed that the
decrement of the area ratio of ejector is more
influenced by constant-area diameter than the
motive nozzle diameter, as shown in Figure 7.
Figure 7: Variation of the motive nozzle,
constant-area diameters and area
ratio of ejector, versus condenser
temperature. (Te = 5oC, ηn = 0.9
and ηd = 0.8).
4.2 Influence on the COPimp
According to energy analysis, the COP of a
standard refrigeration cycle decreases as the
condenser temperature increases, as shown in
Figure 8. For the EERC, the COP reduction due to
increase in condenser temperature is lower than
that of the standard cycle. As a result, the increase
of the COPimp of EERC is very significant for high
condenser temperature. For example, as seen from
Figure 8, the COPimp is 14.82% at Tc = 50oC and
becomes 23.03% at Tc = 55oC. These results
indicate that the use of an ejector as an expansion
device is effective for geographical areas which
have high outdoor air temperature.
Figure 8: Variation of the COP and COP
improvement, versus condenser
temperature. (Te = 5oC, ηn = 0.9
and ηd = 0.8).
The 5th IMAT, November 12 – 13th
2012
12
4.3 Influence on Plift
To obtain the maximum COPimp, the Plift value
should be as high as possible. However, according
to the flowchart as shown in Figure 6, the condition
of (1+ω).x5 = 1 must be fulfilled for realistic flow.
As depicted in the Ph diagram in Figure 4, to
reduce the compressor work, P1 should be as high
as possible, in order to obtain the minimum
enthalpy difference (h2 - h1). In this study, the Plift is
calculated by equation (7) with the evaporator
temperature of 5oC. In the saturation condition, for
R22 refrigerant, the evaporator temperature of 5oC
is equal to evaporator pressure of 584.11 kPa. The
iteration results for various P1, T5 and Plift with
different condenser temperature is shown in Table
1.
Table 1: Variation P5 and T5 versus the condenser
temperature, where Te = 5oC, ηn= 0.9 and ηd=0.8.
Tc (oC) 40 45 50 55
P5 (kPa) 590.05 601.52 613.59 626.91
T5 (oC)
Plift
5.44
1.01
5.94
1.02
6.58
1.04
7.28
1.06
Pre
ss
ure
(P
)
Specific enthalpy (h)
1
2
3
4
6
7 8
910
Pc=1533.6 kPa Tc = 40oC
T5 = 5.44oC
Te = 5oC
T4 = 3.44oC
P5 =
590.0
5 k
Pa
Figure 9: Ph diagram of the EERC (Te =
5oC, Tc = 40
oC, ηn = 0.9 and ηd =
0.8).
Figure 9 shows a detail of EERC of split-type air
conditioner using R22 as the working fluid, with Tc
= 40oC, Te = 5
oC on Ph diagram. The temperature
and pressure data are obtained from iteration using
flowchart in Figure 6. It can be seen from Table 1
and Figure 9 that since the value of Plift is small, the
different between T5 and evaporator temperature
(Te) is also small.
5. CONCLUSION
It has been found that the motive nozzle diameters
are much closer to the experimental results
obtained by other researchers. The results of this
study showed that the decrement of the area ratio of
ejector is more influenced by constant-area
diameter than the motive nozzle diameter.
Replacement of the capillary tube with an ejector
as the expansion device on split-type residential
AC using R22 as the working fluid can improve the
COP, particularly at condenser temperature above
40oC. This indicates that the use of an ejector as an
expansion device is recommended for geographical
areas which have high outdoor air temperature. The
authors hoped that this present study will
encourage other researches on EERC because there
are still many approaches that have to be explored
to determine the dimensions of the other geometric
parameters of the ejector in accordance with the
cooling capacity of the refrigeration system
ACKNOWLEDGMENT
The present study was supported financially by
ASHRAE and Universiti Teknologi Malaysia: GUP
TIER 2 Fund No. 00J25 from the Ministry of
Higher Education (MOHE) Malaysia.
REFERENCES
[1] L. P. Lombard, J. Ortiz, and C. Pout, ―A
review on buildings energy consumption
information,‖ Energy and Buildings, vol. 40,
pp. 394-398, 2008.
[2] S. Shodiya, A. A. Azhar, and A. N. Darus,
―Improved refrigerant characteristics flow
predictions in adiabtic capillary tube,‖
Research Journal of Applied Sciences,
Engineering and Technology, vol. 4, pp.
1922-1927, 2012.
[3] S. Shodiya, A. A. Azhar, N. Henry, and A. N.
Darus, ―Numerical simulation of refrigerant
flow in adiabatic capilarry tubes including
metastability phenomenon,‖ Proceeding of
11th Asian International Conference on Fluid
Machinery and The 3rd Fluid Power
Technology Exhibition, IIT Madras, Chennai,
India. pp. 1-14, 2011.
[4] G. A. Kemper, G. F. Harper, and G. A.
Brown, Multiple phase ejector refrigeration
system, US Patent Patent No.3,277,660, 1966.
[5] A. B. Newton, Capacity control for
multiphase-phase ejector refrigeration system,
US Patent No. 3,670,519, 1972a.
[6] A. B. Newton, Control for multiphase-phase
ejector refrigeration system, US Patent No.
3,670,519, 1972b.
[7] J. H. Keenan, E. P. Neumann, F. Lustwerk,
―An investigation of ejector design by
The 5th IMAT, November 12 – 13th
2012
13
analysis and experiment,‖ Journal of Applied
Mechanics, vol. 17, pp. 299-309, 1950.
[8] R. Yapıcı, H. K. Ersoy, ―Performance
characteristics of the ejector refrigeration
system based on the constant-area ejector
flow model,‖ Energy Conversion and
Management, vol. 46, pp. 3117-3135, 2005.
[9] E. Nehdi, L. Kairouani, and M. Bouzaina,
―Performance analysis of the vapour
compression cycle using ejector as an
expander,‖ International Journal of Energy
Research, vol. 31, pp. 364-375, 2007.
[10] S. Elbel, P. Hrnjak, ―Experimental validation
of a prototype ejector designed to reduce
throttling losses encountered in transcritical
R744 system operation,‖ International Journal
of Refrigeration, vol. 31, pp. 411-422, 2008.
[11] N. Bilir, H. K. Ersoy, ―Performance
improvement of the vapour compression
refrigeration cycle by a two-phase constant-
area ejector,‖ International Journal of Energy
Research, vol. 33, pp. 469-480, 2009.
[12] H. K. Ersoy, N. Bilir, ‗The influence of
ejector component efficiencies on
performance of Ejector Expander
Refrigeration Cycle and exergy analysis,‖
International Journal of Exergy, vol. 7 pp.
425-438, 2010.
[13] J. Sarkar, ―Geometric parameter optimization
of ejector-expansion refrigeration cycle with
natural refrigerants,‖ International Journal of
Energy Research, vol. 34, pp. 84-94, 2010.
[14] A. A. Kornhauser, ―The use of an ejector as a
refrigerant expander,‖ In: Proceeding of the
USN/IIR-Purdue Refrigeration Conference.
West Lafayette, IN, USA, pp.10-19, 1990.
[15] J. P. Liu, J. P. Chen, and Z. J. Chen,
―Thermodynamic analysis on trans-critical
R744 vapor compression/ejection hybrid
refrigeration cycle,‖ In: Proceeding of the
Fifth IIR Gustav Lorentzen Conference on
Natural Working Fluid. Guangzhou, China,
pp.184-188, 2002.
[16] D. Li, E. A. Groll, ―Transcritical CO2
refrigeration cycle with ejector-expansion
device,‖ International Journal of
Refrigeration, vol. 28, pp. 766-773, 2005,
[17] J. Q. Deng, P. X. Jiang, T. Lu, and W. Lu,
―Particular characteristics of transcritical CO2
refrigeration cycle with an ejector,‖ Applied
Thermal Engineering, vol. 27, pp. 381-388,
2007.
[18] H. Takeuchi, H. Nishijima, and T. Ikemoto,
World's first high efficiency refrigeration
cycle with two-phase ejector: "ejector cycle",
In: SAE World Conggres. Detroit, MI, USA,
Paper 2004-01-0916, 2004.
[19] S. Disawas, S. Wongwises, ―Experimental
investigation on the performance of the
refrigeration cycle using a two-phase ejector
as an expansion device,‖ International Journal
of Refrigeration, vol. 27, pp. 587-594, 2004.
[20] S. Elbel, P. Hrnjak, ―Experimental validation
of a prototype ejector designed to reduce
throttling losses encountered in transcritical
R744 system operation,‖ International Journal
of Refrigeration, vol. 31, pp. 411-422, 2008.
[21] O. Brunin, M. Feidt, and B. Hivet,
―Comparison of the working domains of some
compression heat pumps and a compression-
absorption heat pump,‖ International Journal
of Refrigeration, vol. 20 pp. 308-318, 1997.
[22] E.W. Lemmon, M.L. Huber, and M.Q.
McLinde, REFPROP, Reference Fluid
Thermodynamics and Transport Properties,
NIST Standard Reference Database 23,
Version 9.0., 2009.
[23] G. Zhou, Y. Zhang, ―Performance of a split-
type air conditioner matched with coiled
adiabatic capillary tubes using R22 and
HC290,‖ Applied Energy, vol. 87, pp. 1522-
1528, 2010.
[24] P. Chaiwongsa, S. Wongwises, ―Effect of
throat diameters of the ejector on the
performance of the refrigeration cycle using a
two-phase ejector as an expansion device,‖
International Journal of Refrigeration, vol. 30,
pp. 601-608, 2007.
The 5th IMAT, November 12 – 13th
2012
14
Effects of External Hard Particles on Brake Noise of Disc
Braking System
M. A. Nasaruddina, M. K. Abdul Hamid
a, A.R. Mat Lazim
a, and A.R. Abu
Bakara
aDepartment of Automotive Engineering, Faculty of Mechanical Engineering,
Universiti Teknologi Malaysia 81310 Johor Malaysia.
Tel : (607) 5534667. Fax : (607) 5566159
E-mail : [email protected], [email protected], [email protected], [email protected]
ABSTRACT
The open design and position of disc brake that is
closed to road surfaces enable contaminants to
enter the brake gap and caused noise and
tribological disturbance at the brake interface.
Contaminants such as dirt and soil can be present
and are expected to influence the occurrence of
brake squeal that produce an annoying sound
during braking action. The objective of this study
was to examine the effect of external hard particles
at different disc sliding speed on generation of
brake squeal using a brake dynamometer. Different
rotational speed of disc brake was selected and the
experiments squeal noise data was collected and
analyzed using the Fast Fourier Transformation
(FFT) analyzer. From the experiments, the
presence of external particle and the rotation speed
of disc brake promotes the generation of brake
squeal phenomenon by changing the surface
roughness and effective contact of brake interface.
Results obtained from the experiment also showed
that higher rotating disc generate higher sound
level meter or squeal frequency and increase
numbers of squeal noise generated.
Keywords : External particles, brake squeal, noise
level, surface roughness, effective
contact
1. INTRODUCTION
Higher frictions have significant impact on noise
characteristics. Chen et al. [1] stated that the higher
coefficient of friction (cof) of the brake pad, the
higher squeal tend to occur. However, high cof
does not necessary be the main reason of squeal but
it plays big role in noise characteristics of brake
systems. Surface roughness is another factor that
can affect the squeal noise. Generally smooth
friction surface may provide more stable contact
between sliding surfaces and cause less system
vibration and noise occurrence. Abdul Hamid [2]
states present of external particle give different
value of surface roughness and it tends to reduce
with small external particle such as silica sand and
dust. However, the surface roughness influences
not only the contact pressure distribution but the
stability and noise of the system. According to
Mario et al. [3], number of instability intensity of
unstable modes generated by high friction model
significantly increased and cof is directly
proportional to squeal propensity. In this work, the
external particle influence on surface roughness,
noise level and vibration of brake pad in generating
squeal was studied.
2. METHODOLOGY
Brake dynamometer testing was utilized to study
the effect of external particle on squeal noise of
disc braking system. Present of different size and
shape of external particle are expected to influence
surface roughness and the occurrence of squeal.
2.1 Test Rig
Figure 1 shows the schematic diagram of the test
rig. The drag type brake dynamometer was used in
this experiment in order to validate brake squeal
performance. The test rig was mounted separately
on the two units of I beam. The disc brake was
mounted on the main shaft and the caliper holder
was located below the caliper. The flange was used
to drive the disc brake instead of the main shaft.
To control the applied pressure, a hydraulic unit
was used to apply pressure to disc brake through
brake pad with maximum pressure of 20 bars. Four
transducers used for this experiment are load
transducer to measure the load during shaft
rotation, accelerometer to measure vibration of the
brake pad during squeal, a microphone and a set of
amplifier to record the squeal sound and its
frequency. Data Acquisition System (DAQ) was
used to collect information from the transducer,
and process it to suitable data for display or storage
outcomes. A small plastic tube was used to direct
the external particle from the container to the gap
of the brake disc. A transparent cover was used to
avoid splashing of the external particle particles
during the experiments.
IMAT-UI 003
The 5th IMAT, November 12 – 13th
2012
15
Figure 1: Test rig of brake noise dynamometer
2.2 Testing Procedures
Tests are carried out using brake dynamometer
equipped with actual pad and disc brake of the
conventional car. The arithmetic surface roughness
of the pad (Ra) was measured before and after the
brake test with a surface roughness machine. For
the external particle, to get the required size and
shape, the external particles were filtered by a
shaker machine. There are two sections in squeal
test procedure. First is bedding in process and
second is drag noise procedure. Bedding in process
is the process to make sure that the brake pad and
disc brake are aligned and for the new brake pad to
have maximum contact between disc and brake
pad. A series of tests was conducted at three
different disc rotational speeds of 50 rpm, 60 rpm
and 70 rpm while the pressure of 10 bars. Each
experiment was supplied with 35g of external road
particles.
3. RESULTS AND DISCUSSIONS
Presence of external particles changed the pad
surface roughness and increased the total effective
contact area and resulted in increasing cof value.
Table 1 shows the minimum and maximum value
of mean deviation of the surface height (Ra) or
average roughness of pad samples scan area. The
surface roughness of pad samples result showed a
significant increasing of Ra value of both sides
during squeal. This experiment reveals that surface
roughness values increase with the external
particles and affect the tendency of squeal noise
occurrence.
Table 1: Surface roughness values of brake pad. Brake pad Ra min (μm) Ra max (μm)
Piston_side
(Before test) 9.511 11.546
Finger_side
(Before test) 10.178 17.523
Piston_side
(After test) 3.761 17.397
Finger_side
(After test) 3.743 17.786
Figure 2: Graph of sound level versus squeal
frequency with external particle at
different speed.
Figure 2 shows the result of sound level versus
squeal frequency with the existence of external
particle at different speed of rotating disc. The
number of squeal generated for each speed was 15
at 70 rpm and 10 squeal was recorded for 50 and 60
rpm. The maximum squeal frequency recorded at
speed of 50 rpm was 81 dB at 4255 Hz which is
lower compared to the speed of 60 rpm that has the
reading of 84 dB at 4462Hz. The conclusion from
the graph is that, with present of external particle,
higher sound level and squeal frequency will be
generated at higher rotating disc speed.
Figure 3: Graph of sound level versus
vibration with external particle at
different speed.
Figure 3 shows the relationship between sound
level meter and vibration for three different speeds.
The average vibration for all speed recorded was
around 0.53 ms⁻² to 0.85 ms⁻² while the average
value for sound level meter is from 79 dB to 82 dB.
The maximum vibration occurred at 1.2 ms⁻² for
sliding speed of 70 rpm, while for 60 rpm
maximum vibration occurred at 0.94 ms⁻² and the
maximum vibration for 50 rpm occurred at 0.75
ms⁻². Lowest value of vibration for all the speeds
was almost the same at about 0.31 ms ⁻². Thus,
higher sliding speed will generate higher sound
The 5th IMAT, November 12 – 13th
2012
16
level meter and produce higher vibration value with
the existence of external particle.
4. CONCLUSION
The external particles effect on brake noise of disc
braking system was investigated using a specially
developed brake test rig with the actual size of
external particle below 400 µm. The following
conclusions can be made:
The external particles increase the surface
roughness of the pad interface and affect the
tendency of squeal noise occurrence.
The number of squeal generated increases
with higher rotating disc speed.
The frequency of the squeal also increases
proportionally with the sound level meter
reading.
The vibration of the brake pad increase with
increasing in speed of the rotational disc
brake.
ACKNOWLEDGMENT The authors would like to thank Universiti
Teknologi Malaysia for supporting this
researchwork under the University Grant Project
(GUP-Tier 2 Vot 00J45).
REFERENCES
[1] F. Chen, F. Tan, F. C. Chen, C. A. Tan, and
R.L. Quaglia, ―Disc Brake Squeal: Mechanism,
Analysis, Evaluation and Reduction/Prevention‖,
SAE-Society of Automotive Engineers, 2006
[2] M. K. Abdul Hamid, ‗Study of the grit particle
size and shape effects on the frictional
characteristics of the automotive braking system.
PhD Thesis. University of Western Australia, 2010.
[3] T.J., Mario, N.Y., Samir, and J., Roberto,
―Analysis of brake squeal noise using finite
element method: A parametric study‖, Journal of
Applied Acoustics, 69, pp. 147-162, 2008.
COPYRIGHT The author confirm that this papers is original, and
has not been published or under consideration for
publication elsewhere.
The 5th IMAT, November 12 – 13th
2012
17
Experimental Study on the Replacement of HFC-R134a by
Hydrocarbons in Automotive Air Conditioner
Mohd Rozi Mohd Perang
a, Henry Nasution
a,b,c, Zulkarnain Abdul Latiff
a,b,
Azhar Abdul Aziza,b
, Afiq Aiman Dahlanb
aAutomotive Development Centre, Universiti Teknologi Malaysia
81310 Skudai, Johor, Malaysia.
Phone: +60 75535447, Fax: +60 75535811 bFaculty of Mechanical Engineering, Universiti Teknologi Malaysia
81310 Skudai, Johor, Malaysia.
Phone: +60 7 5534575, Fax: +60 7 5566159 cDepartment of Mechanical Engineering, Bung Hatta University
25134 Padang, Sumatera Barat, Indonesia.
Phone: +62 751 7054657, Fax: +62 751 7051341
E-mail: [email protected], [email protected], [email protected], [email protected]
ABSTRACT Performance characteristics of the current
automotive air conditioning system have been
evaluated in this experimental study which will
evaluate the power consumption, temperature
distribution and coefficient of performance (COP)
at various internal heat loads and engine speed
using hydro-chlorofluorocarbons refrigerant (HFC-
R134a) and hydrocarbon refrigerant (HC-R134a) as
the working fluid of the compressor. Both
refrigerants will be tested on the experimental rig
which simulated the actual cars as an internal cabin
complete with a cooling system component of the
actual car including the blower, evaporator,
condenser, radiator, electric motor, compressor and
alternator. The electric motor acts as a vehicle
engine, and then will drive the compressor using a
belt and pulley system, as well as to the alternator
to recharge the battery. The rig also equipped with
simulation room acting as the passenger
compartment. The tests have been performed by
varying the motor speed; 1000, 1500, 2000, 2500
and 3000 rpm, temperature set-point; 21, 22 and
230C, and internal heat loads; 0, 500, 700 and 1000
W. As the results, the performance characteristics
of the HC-R134a indicate the positive
improvement of the system compared to HFC-
R134a.
Keywords : Air conditioning, automotive, HC-
134a, hydrocarbon refrigerant,
performance, energy saving.
1. INTRODUCTION
The heating, ventilation and air conditioning
(HVAC) of the automotive is designed to provide
thermal comfort level of the driver and passengers.
Thermal comfort is the crucial things to be
fulfilled. Human thermal comfort is defined by the
American Society of Heating, Refrigeration and
Air Conditioning Engineers (ASHRAE) as the
state of mind that expresses satisfaction with the
surrounding environment (ASHRAE Standard 55)
[1].The function of an air conditioning (A/C)
control system is to modulate the A/C system
capacity to match off the design condition, load
variation and climate change, to maintain the
indoor environment within desirable limits at
optimum energy use during the entire drive.
Automotive A/C component have been going
through a steady evolution since the introduction
of A/C in cars in 1940. From the early days to
today, A/C was a very expensive option in luxury
cars when it is standard equipment in many models
and many different types of systems have been
used [2].
Hydrocarbon (HC) is new findings for refrigerants
by the experts to replace the current HFC-R134a
as the refrigerant for A/C system. HC is used as
refrigerants gases at an early stage, which were
accepted before the emergence of CFCs
(chlorofluorocarbons) and HCFCs
(hydrochlorofluorocarbons). After a long period,
HC refrigerant is no longer in use because of the
flammability characteristic. Thus, CFCs and
HCFCs are being used instead of using HC
refrigerant, which is flammable, not practical and
harmful to a user [3]. However, several studies
have shown that the used of hydrocarbons in A/C
system would improve the performance of the
system [4, 5]. Some applications of the HC can be
found in aerosol filler, the heating fuels in gas
ovens, etc. The HC used in the A/C and freezer as
the working fluid is not yet in common.
The research work will perform the results of the
experimental studies on HC mixtures (HC-134a)
as an alternative refrigerant to the automotive A/C
system in Malaysia. HC mixture used for this work
contains propane (R290), butane (R600), and
isobutane (R600a). The mixture is known as HC-
134a and Wongwises et al. [6, 7], Ghodbane [8]
and Tashtoush et al. [9] are one of the researchers
IMAT-UI 004
The 5th IMAT, November 12 – 13th
2012
18
that involved in automotive HC A/C system. Table
1 shows the properties of the refrigerant used in
this experimental work [6, 10].
Table 1: Properties of refrigerant
HC refrigerant is generally considered as
environmental friendly and has been used again
regarding to the noticeable value of the Global
Warming Potential (GWP) and zero Ozone
Depletion Potential (ODP).
2. EXPERIMENTAL APPARATUS AND
PROCEDURE
In this research work, the performance
characteristics of the automotive A/C system have
been performed via experimental analysis. The test
is performed on the automotive A/C experimental
rig in order to evaluate the power consumption,
temperature distribution and coefficient of
performance (COP). The refrigerants used are
using HFC-R134a and (HC-R134a) as the working
fluid of the compressor. The test is done at various
internal heat loads, temperature setting and engine
speed. Figure 1 illustrates the schematic diagram of
automotive A/C system that has been used in this
work.
Figure 1: Automotive A/C system
experimental rig
Figure 2 shows on the experimental rig which
simulated the actual cars as an internal cabin
complete with A/C system component of the actual
car including the blower, evaporator, condenser,
radiator and compressor.
Figure 2: Experimental test rig
Other components involved in this work are the
electric motor and alternator. The electric motor
worked as a vehicle engine, and then will drive the
compressor using a belt and pulley system, as well
as to the alternator to recharge the battery. The
compressor was run by 12 volt original car battery
The 5th IMAT, November 12 – 13th
2012
19
which continuously charged by the alternator to
ensure the whole range of rotating speeds is like the
actual car speed. The rig also equipped with
simulation room acting as the passenger
compartment.
The temperatures and pressures parameters were
measured by fitted the type T thermocouple
(accuracy ±0.1oC) and flow meter (accuracy of ±1
gr/s) respectively, at several locations as shown in
Figure 1. Then, the current and voltage of the
electric system were measured in order to obtain
the compressor energy consumption. The accuracy
of the current meter is ±1%, and voltage meter is
±1.5%.
120 experiments have been done in this work by
varying the motor speed, internal heat load and
cabin temperature using HFC-R134a and HC-
R134a simultaneously, as the refrigerant for the
A/C system. Varied parameters are listed and
shown in Table 2. Table 2: Varied parameters
Parameter Range of variation
Motor speed, N
(rpm)
1000, 1500, 2000, 2500 and
3000
Cabin temperature
(°C) 21, 22 and 23
Type of
refrigerants HFC-R134a and HC-R134a
Internal heat load
(W) 0, 500, 700 and 1000
The experiment was carried out according to the
following procedures: 2.1 The A/C system was evacuated using a
vacuum pump.
2.2 Refrigerant R134a has been charged into the
system.
2.3 The A/C system was started and let the system
running for 15 minutes.
2.4 The cabin temperature was set at 21°C.
2.5 The speed of the motor was set at 1000 rpm.
2.6 The internal heat load was set by switching on
the bulb with 0, 500, 700 and 1000W
simultaneously. Each data was collected after
15 minutes of the A/C system run.
2.7 Procedures 2.7-2.6 were repeated with the
motor speed of 1500, 2000, 2500 and 3000
rpm simultaneously.
2.8 Procedures 4-7 were repeated with the cabin
temperature of 22 and 23°C simultaneously.
2.9 Procedures 1-8 were repeated with the
refrigerant of HC-R134a.
In collecting the data, the thermostat was set to the
maximum cool position. Data was collected when
the A/C system was considered stable after it ran
for 15 minutes for each test condition.
Thermodynamic properties of the HFC-R134a and
HC-R134a used were taken from the REFPROP
database [10] and from the thermodynamic
properties given by the company respectively.
3. RESULTS AND DISCUSSIONS
The data from the experiment was analyzed by
varying two parameters and unvarying two other
parameters. The analysis was done on measured
parameters to obtain the performance of the A/C
system. The parameters are the coefficient of
performance (COP), power consumption consumed
by the compressor and compression ratio (Cr). COP
of the system is a relationship between the energy
released from the evaporator (refrigerating effect
(Qe)) and the energy required by the compressor
(Wc). COP was calculated by using equation (1)
[11]:
COP=Qe/Wc= (h1-h4)/(h2-h1) (1)
Where,
Qe= Cooling capacity of evaporator, (kJ/kg)
Wc= Compressor power, (kJ/kg)
h1=Enthalpy on suction compressor, (kJ/kg)
h2= Enthalpy on discharge compressor, (kJ/kg)
h3 = Enthalpy on condenser exit, (kJ/kg)
h4 = Enthalpy on evaporator inlet. (kJ/kg)
In order to get the enthalpy through REFPROP
software, refrigeration cycles necessitate to be
illustrated on the P-h (pressure vs. enthalpy)
diagram as illustrated in Figure 3 [12].
Figure 3: P-h diagram of refrigeration system
3.1 Temperature Distribution of System
Figure 4(a), 4(b) and 4(c) show the graphs of
temperature distribution against the internal heat
load (0, 500, 700 and 1000W) at the compressor
speed of 1000, 1500, 2000, 2500 and 3000 rpm and
the set-point temperature of 21, 22 and 23˚C for
HC-R134a and HFC-R134a.
The 5th IMAT, November 12 – 13th
2012
20
(a) At 21
0C
(b) At 22
0C
(c) At 23
0C
Figure 4: Graph of temperature distribution
against load (at compressor speed of
1000 – 3000 rpm).
From the figures above, the temperature
distribution of HC-R134a is always lower (up to
5%) than HFC-R134a. This is because the cooling
capacity of the HC-R134a is higher than HFC-
R134a. At all temperature set-point, the HC-R134a
is nearer to the set-point temperature at compressor
speed of 1000, 1500 and 2000 rpm. This will
indicate the used of HC-R134a is better than HFC-
R134a because the cooler air is distributed faster to
all compartment of the cabin.
3.2 Coefficient of Performance (COP)
Figure 5(a), 5(b) and 5(c) exhibits the graph of
COP against load at variable compressor speed of
1000 - 3000 rpm with refrigerant of HFC-R134 and
HC-R134a. COP of HC-R134a is higher than HFC-
R134a in which is up to 40%. This is a positive
improvement because HC-R134a produced high
COP which is indicating the better performance
than HFC-R134a.
When the compressor speed increases with the
increment of internal heat load and compressor
speed, the COP will decrease. Thus COP will
decrease when compressor work increases at
constant condition of the evaporator heat
absorption.
The 5th IMAT, November 12 – 13th
2012
21
(a) At 21
0C
(b) At 22
0C
(c) At 23
0C
Figure 5: Graph of COP against load (at
compressor speed of 1000 – 3000
rpm).
There is also a relationship between COP of
different temperature set-point and internal heat
load at compressor speed of 1000 rpm. It can be
observed from the figures above that illustrate the
temperature set-point increased, the COP will
increase. Therefore, COP will be high when
compressor work is constant but evaporator heat
absorption is lower.
3.3 Compressor Ratio (Cr)
Figure 6(a), 6(b) and 6(c) illustrate the graph of
compression ratio against internal heat load which
is showing Cr is decreased with the increasing of
internal heat load. And as the temperature set-point
increased, the compression ratio will be decreased.
This shows that the compression ratio of HC-
R134a is better than HFC-R134a. This is an
encouraging improvement as the compression work
increased with less energy needed to cool the same
set-point temperature. The compression ratio will
affect the energy desired for compressor.
The 5th IMAT, November 12 – 13th
2012
22
(a) At 21
0C
(b) At 22
0C
(c) At 23
0C
Figure 6: Graph of compression ratio against
load (at compressor speed of 1000 –
3000 rpm).
3.4 Compressor Power Consumption
Figure 7(a), 7(b) and 7(c) show the relationship
between power consumption and internal heat load
at various speed of compressor (1000 - 3000 rpm).
The compressor power consumption is increased
proportionally with the increasing of internal heat
load. At the temperature set-point of 21, 22 and
23°C, the graphs show the same trend of rising. At
every temperature set-point temperature and with
the increasing of internal heat load, the compressor
needs to work more often and more power will be
consumed. As a result, the compressor uses more
power to compress the refrigerant to higher
pressure to maintain the temperature in the cabin.
The compressor power consumption of HC-R134a
is lower than HFC-R134a which indicates a
positive improvement up to 25% (an average
decreased is 15%).
The 5th IMAT, November 12 – 13th
2012
23
(a) At 21
0C
(b) At 22
0C
(c) At 23
0C
Figure 7: Graph of power consumption against load
(at compressor speed of 1000 – 3000 rpm).
4. CONCLUSIONS
From the discussions above, the HC-R134a
indicates a positive improvement of performance
characteristics compared to HFC-R134a in term of
COP, temperature distribution and power
consumption. Therefore, the HC-R134a is
suggested to be the replacement of current HFC-
R134a used in the automotive industry.
Energy consumption will vary with the changes of
the A/C compressor speed. Whilst the compressor
speed increases, the room temperature is going to
decreases as well as the COP is decreased. As the
results, the consumption of energy will increase
and less value of the energy can be saved and vice
versa.
ACKNOWLEDGEMENTS
Special thanks to Thermodynamic Laboratory,
Faculty of Mechanical Engineering, Universiti
Teknologi Malaysia for their facilitating support in
this study. Their guidance and assistance are highly
appreciated.
The 5th IMAT, November 12 – 13th
2012
24
REFERENCES
[1] ANSI/ASHRAE Standard 55, Thermal
environmental conditions for human
occupancy. American Society of Heating,
Refrigerating and Air-Conditioning Engineers,
Inc., 2008.
[2] Tom Birch, ―Automotive Heating and Air
Conditioning 2nd
Edition‖, Prentice Hall, 2000.
[3] Granryd, E., Hydrocarbons as refrigerants –
an overview. International Journal
Refrigeration, 24:15-24, 2001.
[4] D. Jung and C.B. Kim., Testing of
propane/isobutene mixture in domestic
refrigerator. International Journal
Refrigeration, 2000.
[5] K. Mani and V. Selladurai, Experimental
Analysis of a new refrigerant mixtures as drop-
in replacement for CFC12 and HFC134a.
International Journal of Thermal Sciences,
2008.
[6] Wongwises, S., Kamboon, A., and Orochon,
B., Experimental investigation of hydrocarbon
mixtures to replace HFC-134a in an automotive
air conditioning system, Energy Conversion &
Management; 47:1644-1659, 2006.
[7] Wongwises, S. and Chimres, N., Experimental
study of hydrocarbon mixtures to replace HFC-
134a in a domestic refrigerator, Energy
Conversion & Management; 46:85-100, 2005.
[8] Ghodbane. M., An Investigation of R152a and
Hydrocarbon Refrigerants in Mobile Air
Conditioning, SAE Paper 1999-01-0874, 1999.
[9] Tashtoush, B., Tahat, M., Shudeifat, M.A.,
Experimental study of new refrigerant mixtures
to replace R12 in domestic refrigerators.
Application Thermal Engineering; 22:495-506,
2002.
[10] REFPROP, Thermodynamic properties of refrigerants and refrigerant mixtures, Version
6.1, Gaittherbeurg, MD, National Institute of
Standards and Technology, 1998.
[11] Y.A. Cengel and M.A. Boles.,
Thermodynamics: An engineering Approach
(6 ed.). McGraw-Hill, 2007.
[12] Henry Nasution, et. al., ―Experimental
Evaluation of Automotive Air-Conditioning
Using HFC-134a and HC-134a‖, American
Institute of Physics, 2012
The 5th IMAT, November 12 – 13th
2012
25
Effect of Fuel Droplets During Early Stage of Flame Propagation
Aminuddin Saata,*
, Malcolm Lawesb, Mazlan Abdul Wahid
a,
Mohd Farid Muhamad Saida
a Faculty of Mechanical Engineering, Universiti Teknologi Malaysia, 81310 UTM Johor, Malaysia.
b School of Mechanical Engineering, University of Leeds, LS2 9JT, Leeds, United Kingdom.
*email: [email protected]
ABSTRACT
There are only few experimental data of a
fundamental nature that clearly demonstrate the
similarities and differences in flame propagation and
burning rates between single phase and two phase
combustion. Such data are essential in order to
establish a better understanding of the spray
combustion phenomena and associated processes. In
the present study, experimental investigations of
combustion of droplet and vapour mixtures under
quiescent condition have been conducted in a closed
combustion vessel. Droplet and vapour mixtures or
aerosol mixtures were generated by expansion of
iso-octane gaseous pre-mixture to produce a
homogeneously distributed suspension of fuel
droplets. The aerosol mixtures were ignited centrally
in the combustion vessel and the flame propagation
was recorded by high-speed schlieren photography.
Flame speed was obtained from the measured flame
radius-time data. The effect of fuel droplets in the
early stage of flame propagation was investigated by
comparing the flame structure and flame speed of
gaseous mixtures at identical conditions.
Comparisons between gaseous and aerosol flame
structure have shown quantitatively that the presence
of fuel droplets causes earlier onset of instabilities
and cellularity than for gaseous flames, particularly
at rich conditions. It is shown that the initial growth
of flame propagation in aerosol mixtures was
different than in gaseous mixtures. This difference
was shown to be a function of droplet size and
overall equivalence ratio. It is suggested that these
factors lead to vary the local equivalence ratio which
increases the initial burning rate of lean aerosols, but
decreases that of rich ones.
Keywords : droplet, aerosol, flame speed,
burning rate
1. INTRODUCTION
The combustion of droplets have been studied
extensively because of their relevance in many
practical combustion devices such as automotive
engines, gas turbines, power generation, boiler
and heating system. However, the multiplicity of
dependant parameters involved and the difficulty
in carrying out well-controlled studies to extract
the effects of individual parameters renders
practical spray combustion systems as some of
the most challenging environments to
investigate. There is theoretical and
experimental evidence [1-4] to suggest that
flame propagation through droplet and vapour
mixture, under certain circumstances, is higher
than that in a fully vaporised homogeneous
mixture. Although this may be advantageous in
giving more rapid burning, its effects on
emissions are uncertain. Conversely, it is a
serious disadvantage in the hazard context.
Thus, an understanding of the influence of the
presence of fuel droplets in flame propagation is
crucial for understanding the practical spray
combustion system.
Since the essential aspects of single-droplet
combustion become clear, investigations on
combustion of droplet and vapour mixtures are
necessary for a further approach to spray
combustion because the majority of droplets
burn as a group and interact with one another in
practical environments [5]. Experimental study
on the effect of fuel droplet on propagating
flame was performed by Hayashi and Kumagai
[1], Nomura et al. [2,6] and Lawes et al. [3].
Based on the principle of Wilson‘s cloud
chamber [7], they investigated flame
propagation in a stagnant pre-mixture in which
fuel droplets were monodispersed. Hayashi and
Kumagai [1] found the highest enhancement in
aerosol burning rate occurred at the slighly rich
mixture when the droplet diameter is 30 m.
Nomura et al. [2,6] investigated the influence of
fuel droplets under microgravity conditions and
they concluded that there are two regions in
which flame speed of aerosol mixtures exceeds
IMAT-UI 005
The 5th IMAT, November 12 – 13th
2012
26
that of premixed gases. Lawes et al. [3] observed
that aerosol flames became more unstable
compared to the gaseous flame, and hence had
faster burning rates. However, the enhancement
was not very significant at which the overall
equivalence ratio is indicated at slightly rich
mixture. In a numerical study, Polymeropoulos
[4] predicted a significant burning rate
enhancement for monodispersed aerosol fuel-air
mixtures comprising droplets in the range of 5-
15 m. He claimed that there is still lack of
experimental data within this range to support
his prediction.
In this paper, spherically expanding flames
following central ignition of globally
homogeneous combustible fuel mixtures at near
atmospheric pressures are employed to quantify
the differences in the flame structure and
burning rates of gaseous flames with aerosol
flames. Iso-octane-air aerosol mixtures with
droplet size ranges of up to 20 m, were
generated by expansion of the gaseous pre-
mixture to produce a homogeneously distributed
suspension of fuel droplets. The effect of fine
fuel droplets in the early stage of flame
propagation was investigated by comparing the
flame structure and flame speed of gaseous
mixtures with aerosol mixtures at several range
of mixture equivalence ratio.
2. EXPERIMENTAL APPARATUS AND
TECHNIQUE
The combustion vessel and auxiliary equipment for
the experimental work are shown schematically in
Figure 1. A full description of the system and aerosol
generation technique is presented in [8]. The
combustion vessel, which essentially resembled a
Wilson cloud chamber [7], was a cylindrical vessel of
305 mm diameter by 305 mm long. Optical access
windows of 150 mm diameter were provided on both
end plates for aerosol characterisation and
photography of flame propagation. Four fans, driven
by electric motors, adjacent to the wall of the vessel,
initially mixed the reactants. Two electrical heaters
were attached to the wall of the vessel to preheat the
vessel and mixture to the desired temperature.
Figure 1: Schematic of aerosol mixtures
generation and combustion
apparatus.
Iso-octane-air aerosol mixtures were prepared by a
condensation technique [7], used elsewhere in
combustion studies by [1-3], to generate well defined
and near mono-dispersed droplet suspensions. This
achieved by controlling the expansion of a gaseous
fuel-air mixture from the combustion vessel into the
expansion tank, which was pre-evacuated to less than
1 kPa. The expansion caused a reduction in mixture
pressure and temperature which took it into the
condensation regime and caused droplets to be
formed. The characteristics of the generated aerosol
were calibrated by in-situ measurements of the
temporal distribution of pressure, temperature,
droplet size and number without combustion, with
reference to the time from start of expansion.
Figure 2 shows a typical variation of pressure, P,
temperature, T, droplet size, D, and number density,
ND, of droplets with time from the start of expansion
for a stoichiometric iso-octane-air mixture expanded
at 200 kPa and 303K. The measurement of droplet
mean diameter, D10, was performed using Phase
Doppler Anemometer system. The estimation of
number density, ND was obtained from calculations
based on laser attenuation and droplet size
measurements using the Beer-Lambert Law
correlation given in [9]. The pressure of the mixture,
P, was set at nearly atmospheric condition for all
experiment. The temperature of all mixtures just
before ignition varies from 265 to 293 K. The overall
equivalence ratio, ov, is varied from lean mixture up
to rich mixture. The mean droplet size of aerosol
mixtures was varied up to 20 m by varying the
initial pressure before expansion and time of ignition.
As shown in Fig. 2, the measured temporal variation
of temperature, Tm, initially exhibited a polytropic
relationship, as shown by the dotted curve in Fig. 2,
in which the polytropix index, n, was found to be
1.35. At the start of droplet nucleation, approximately
The 5th IMAT, November 12 – 13th
2012
27
1.6 seconds after expansion start, the measured
temperature departed from that of the polytropic
expansion, in part due to the latent heat of
condensation. This was also evident by the increase
in ND and D10. It is shown that ND is remained nearly
constant during the condensation. During mixture
expansion, the overall equivalence ratio, ov,
remained constant, however, the gas and liquid phase
equivalence ratios, g and l, varied after nucleation
and during droplet growth. Also shown in Fig. 2 are
the standard deviations of the droplet mean diameter,
D. The low values of D indicate the near mono-
dispersed distribution of droplet size that results in
the combustion vessel. Since the expansion of
mixture took place over a period of several seconds
while combustion took place over less than 100 ms,
the far field values of D10 were assumed to be
constant during combustion.
Figure 2: Typical variation of several parameters
with time for iso-octane-air aerosol
mixtures expanded from initial
condition of 200 kPa and 303 K.
The droplet size varies with time during expansion;
hence the effect of droplet size is investigated by
varying the time of ignition after the start of
expansion. The aerosol mixture was ignited at the
centre of the combustion vessel by an electric spark
of about 400 mJ. The flame front was monitored and
recorded through the vessel windows by high speed
schlieren arrangement at a rate of 1000 frame per
second. The flame image was processed digitally to
obtain the flame radius using image processing
software. The laminar flame speed, Sn, was obtained
from the measured flame front radius against time, t,
by Sn = dr/dt. As the burning rates for aerosol flames
was obtained, measurement of burning rates of
gaseous flames at initial pressure and temperature
closer to those of aerosols could provide a more
accurate comparison between gaseous and aerosol
flames. This was achieved by igniting the mixture
during expansion but before the onset of
condensation.
3. RESULT AND DISCUSSION
Figure 3: Typical spherical flame development
of laminar iso-octane-air flames at
ov = 1.2 for gaseous and aerosol
mixtures.
Shown in Fig. 3 are the typical sequences of schlieren
images showing the growth of two expanding laminar
iso-octane - air flames at ov = 1.2. Figure 3a shows a
gaseous flame ignited at 105 kPa and 282 K while
Fig. 3b shows an aerosol flame with droplet diameter
of 14 µm, ignited at 105 kPa and 279 K. The circular
boundary (150 mm in diameter), represents the size
of the optical access windows. The black horizontal
object in each of images is the spark electrode holder.
Since there were negligible differences in initial
pressure and temperature, it is assumed that the
differences in the flame structure are entirely due to
the presence of droplets. The gaseous flame had a
smooth surface and was nearly spherical throughout
the whole period of observation. However, the flame
with droplets shows unstable surface with cracks and
cells that gradually increased throughout flame
growth. Such cellularity has been observed to
increase with increasing ov and with increasing
The 5th IMAT, November 12 – 13th
2012
28
droplet size. Hence, Fig. 3 demonstrates that flame
with fuel droplets are more unstable than those of
equivalent gaseous flames and droplets can cause the
onset cellularities that would not be present in
gaseous flames at similar conditions.
Figure 4:. Variation of flame speed with time for
gaseous iso-octane-air flame with
and without droplets at ov = 1.2.
Figure 4 shows the variation of flame speed with time
for gaseous and aerosol mixtures at an equivalence
ratio of 1.2 at near atmospheric conditions. For
clarity, each plot shows average values from three
explosions with associated error bars. Initially, after
spark ignition, gaseous flame propagates faster than
aerosol flame within the first 8 ms before attaining an
approximately constant value of about 2.5 m/s at later
stage. The aerosol flame developed slightly slower
than gaseous flame, but attained higher, and still
accelerates at later stage. This acceleration was
associated with instabilities, as shown by unstable
surface in Fig. 3, which triggered by the droplets
presence and further enhance the rate of flame
propagation. The mechanism behind unstable flame
behaviour at the later stage is probably related to the
heat loss from the flame in vaporizing the droplets,
accompanied by the local rapid expansion taking
place in the process. This comparison clearly
demonstrates the differences of flame propagation
rate due to the presence of fuel droplets.
It is clear from Fig. 4, that fuel droplets causes
different burning rate either at the early stage or later
stage of flame development. Further investigation is
focused on the effect of droplets in the early stage of
flame development.
(a)
(b)
Figure 5: Effect of droplets in the early stage of
flame propagation in lean mixtures,
ov=0.9. (a) Schlieren images up to 3
ms after ignition, (b) variation of flame
speed with time. Error bars represent
scatter from up to three explosions.
Figure 5(a) shows sequences of schlieren images of
gaseous and aerosol flame kernels for lean mixtures
at an equivalence ratio of 0.9 during first 3 ms after
ignition, while in Fig. 5(b) is the corresponding flame
speed variation with time. At least a minimum of
three experiments were conducted at each condition
and the error band presented in Fig. 5(b) indicates the
variation in flame speed between the experiments.
The circle symbols represent the size of the flame at
3 ms after ignition as shown by the schlieren images
in Fig. 5(a). All the flames were ignited using the
same spark electrode at identical ignition energy. It is
shown in Fig. 5 that initial size of flame kernel and
the rate of flame development is a function of droplet
size, with the gaseous flame having the slowest initial
rate of development and the aerosol flame for a
droplet diameter of 20 m showing the fastest
development. However, this trend reversed for the
rich mixtures.
The 5th IMAT, November 12 – 13th
2012
29
(a)
(b)
Figure 6: Effect of droplets in the early stage of flame
propagation in rich mixtures, ov=1.4. (a)
Schlieren images for up to 3 ms after
ignition, (b) variation of flame speed with
radius. Error bars represent scatter from up
to three explosions.
Figure 6(a) shows sequences of schlieren images of
gaseous and aerosol flame kernels for rich mixtures
at an equivalence ratio of 1.4 during first 3 ms after
ignition, while in Fig. 6(b) is the corresponding flame
speed variation with radius. In contrast to the trend
presented for leaner mixtures in Fig. 5, it is seen in
Fig. 6 that for rich mixtures, the initial flame
development is fastest in the gaseous flame and
slowest in the aerosol flame with 20 m droplet
diameter. It was also observed that as the droplet size
increased, the ignitibility of aerosol mixtures
improved and the probability of ignition and
subsequent flame propagation in aerosol mixtures
was higher than a gaseous mixture at the same overall
equivalence ratio. This observation was particularly
significant for leaner mixtures. However, a
probabilistic study of ignitability of aerosol mixtures
did not form a part of the present study.
Figures 5 and 6 clearly show that the influence of
fuel droplets in the early stages of flame propagation
is significant. In leaner mixtures, increase in the
droplet diameter causes an increase in the flame
speed. However, in richer mixtures, increase in
droplet diameter causes a decrease in the flame
speed. This reversal in trend during initial flame
development can be explained due to droplet inertia
mechanism [10, 11]. For initially quiescent condition,
gaseous fuel-air mixture is nearly stagnant as well as
the fuel droplets movement. The existence of fuel
droplets in the mixture enriches the fuel vapour
density near the reaction zone, leading to an increase
in the local equivalence ratio. In the case of lean
mixtures, this enrichment in fuel vapour density
shifts the local equivalence ratio closer to the
stoichiometric fuel-air ratio and hence the aerosol
flame burns faster than the equivalent gaseous flame.
For richer mixtures, the same mechanism causes the
local equivalence ratio to become even richer and
affects an attenuation in the growth rates which
implies that gaseous flame burns faster than the
aerosol flame, as evidenced by Fig. 6. Furthermore,
larger the droplet size, higher the inertia and hence
stronger the influence. This explains why this effect
gets more pronounced with an increase in the droplet
size. A similar relationship has been demonstrated
experimentally by Nomura [6] in terms of the
formulation of droplet slip velocity in microgravity
combustion.
4. CONCLUSION
The comparison between flame with and without
droplets under laminar conditions have been
experimentally studied in centrally ignited flames in
which aerosol of fuel droplets was suspended in a
mixture of quiescent air and fuel vapour. Inspection
of schlieren images revealed that aerosol flames were,
in general, more unstable than gaseous flames and
this instability increases as the droplet size increases.
The presence of fuel droplets clearly influences
instabilities by causing earlier onset and more rapid
development of cellularity than for gaseous flames.
During the early stages of flame propagation, droplets
are relatively stagnant with respect to the flame. This
inertia of fuel droplets leads to a local enrichment in
the mixture equivalence ratio. Flame propagation
rates in the early stages were found to be consistent
with the altered equivalence ratio in the sense that for
leaner mixtures, droplets enhanced the burning rate,
while for richer mixtures, droplets slowed down the
burning rate.
The 5th IMAT, November 12 – 13th
2012
30
ACKNOWLEDGMENT
The authors acknowledge the financial support from
Universiti Teknologi Malaysia under a Grant
Research No. R.J130000.7724.4P044. The use of
apparatus in the Thermodynamic Laboratory of
Leeds University throughout the study is gratefully
acknowledged.
REFERENCES
[1] Hayashi, S., Kumagai, S., ―Flame propagation in
fuel droplet-vapor-air mixtures‖, Proc. Comb.
Inst. 15:445-452 (1974).
[2] Nomura, H., Koyama, M., Miyamoto, H., Ujiie,
Y., Sato, J., Kono, M., Yoda, S., ―Microgravity
experiments of flame propagation in ethanol
droplet-vapor-air mixture‖, Proc. Comb. Inst.
28:999-1005 (2000).
[3] Lawes, M., Lee, Y., Marquez, N., ―Comparison
of iso-octane burning rates between single-phase
and two-phase combustion for small droplets‖,
Combustion and Flame 144:515-525 (2006).
[4] Polymeropoulos, C.E., Flame propagation in
aerosols of fuel droplets, fuel vapour and air‖,
Combution Science and Technology 40:217-232
(1984).
[5] Annamalai, K, Ryan, W, ―Interactive processes
in gasification and combustion: Part 1 Liquid
drop array and clouds‖, Progress of Energy
Combustion Science, 18:221-295 (1992).
[6] Nomura, H., Kawasumi, I., Ujiie, Y., Sato, J.,
―Effects of pressure on flame propagation in a
premixture containing fine fuel droplets‖, Proc.
Comb. Inst. 31:2133-2140 (2007).
[7] Wilson, C. T. R., ―Condensation of water vapour
in the presence of dust-free air and other gases‖,
Proc. of the Royal Society of London 189:265-
307 (1897).
[8] Saat, A., ―Fundamental studies of combustion of
droplet and vapour mixtures, PhD Thesis, School
of Mechanical Engineering, University of Leeds
(2010).
[9] Bachalo, W.D., Rosa, A. B., Sankar, S.V.,
―Diagnostic for fuel spray characterisation, in
Combustion measurements‖, (Ed. Chigier N.),
Hemisphere, p.229-278 (1991).
[10] Atzler, F., ―Fundamental studies of aerosol
combustion‖, PhD Thesis, School of Mechanical
Engineering, University of Leeds (1999).
[11] Atzler, F., Demoulin, F., Lawes, M., lee, Y.,
―Oscillations in the flame speed of globally
homogeneous two phase mixtures‖, 18th
International Colloquium on the Dynamics of
Explosions and Reactive Systems, paper 83
(2001).
The 5th IMAT, November 12 – 13th
2012
31
The Use of Mechanical Ventilation System in an Electric Car
Intan Sabariah Sabria, Haslinda Mohamed Kamar
b, Nazri Kamsah
c, Md Nor Musa
d
aFaculty of Mechanical Engineering,
Universiti Teknologi Malaysia, 81310 Skudai, Johor, Malaysia
Email : [email protected]
bFaculty of Mechanical Engineering,
Universiti Teknologi Malaysia, 81310 Skudai, Johor, Malaysia
Email : [email protected]
cFaculty of Mechanical Engineering,
Universiti Teknologi Malaysia, 81310 Skudai, Johor, Malaysia
Email : [email protected]
dFaculty of Mechanical Engineering,
Universiti Teknologi Malaysia, 81310 Skudai, Johor, Malaysia
Email : [email protected]
ABSTRACT
The heart of an electric car is its rechargeable
battery pack, which supplies the electric motor with
the energy to move the vehicle. When parked in the
sun, the soak air temperature inside a passenger
compartment can rise up to 60°C. Reducing the air
temperature inside the passenger compartment is
very important not only for passenger comfort but
also to reduce the power consumed by the air-
conditioning (AC) system. This leads to saving in
the battery power. This paper presents the use of
mechanical ventilator to lowering the air
temperature inside the passenger compartment
during parking in a sunny day condition. The
commercial software FLUENT 6.3 is used to
simulate three-dimensional (3-D) air temperature
distributions and air flow field inside the passenger
compartment. The simulated result without
mechanical ventilator is compared with the
experimental data. They show a good agreement
with average deviation of 1.8°C in general. This
study found that the number of mechanical
ventilator has a weak influence on the air
temperature inside the passenger compartment
when they are installed on the rear dashboard.
Keywords : Computational Fluid Dynamics (CFD)
analysis, mechanical ventilation
system, passenger compartment,
electric car, air temperature field, air
flow field.
1. INTRODUCTION
The heart of an electric car is its battery [1]. The
battery of an electric car runs everything such as
operate the lighting, accessories and AC system.
The AC system consumes a lot of energy from the
battery to cool down the air temperature inside the
passenger compartment. The battery power
consumption can be reduced by saving the energy
required by the AC system. During the sunny day,
about 80% of the air temperature inside the
passenger compartment rises during the first 30
minutes
[2]. A car parked in the sun with its windows
closed experiences a greenhouse effect, where the
passenger compartment becomes extremely hot [3].
The greenhouse effect can arise the air temperature
inside the passenger compartment up to 60 - 70°C
[4]. The windows glazing allows about 50 - 70% of
thermal energy entering the passenger compartment
during parking condition [3]. The air and materials
affected by solar radiation reach considerable
temperatures such as about 70°C for the dashboard
when a car parked facing the sun [5]. This will
make a driver feel very uncomfortable when he
enters the car. The AC system is unable to cool
down the air temperature inside the passenger
compartment within a short period of time.
Reducing the air temperature inside the passenger
compartment is very significant not only for
passenger comfort but also to reduce the power
consumed by the AC system. This could lead to
saving the battery power consumption of the
electric car for a better driving range [6]. The air
temperature inside the passenger compartment can
be reduced in many ways such as solar-reflective
glazing, solar-reflective coatings and parked car
ventilation [3]. Reducing the size and the glass
transmissivity of the window are able to decrease
the air temperature inside the passenger
compartment [7]. Solar-reflective glazing reflects
the incidence solar radiation and helps in reducing
the air temperature inside the passenger
compartment. Study found that the solar reflective
glazing reduces the air temperature inside the
passenger compartment by 2.7°C as well as
lowered the instrumental panel temperatures by
IMAT-UI 006
The 5th IMAT, November 12 – 13th
2012
32
7.6°C. This leads to reduce about 11% of AC
system power consumption [8]. The use of solar-
reflective glass in all locations reduced the average
air temperature by 34% of the maximum possible
[3]. The use of solar reflective coatings for opaque
surfaces can also decrease the air temperature
inside the passenger compartment when a car is
parked in the sun [3]. Study found that increasing
the solar reflectance (ρ) of the vehicle‘s shell about
0.5 reduces the air temperature inside the passenger
compartment by about 5 - 6°C [9]. Increasing of
each 0.1 solar reflectance of the shell decrease the
air temperature inside the passenger compartment
by about 1°C [10]. Solar energy that enters a
vehicle heats the interior mass as well as the air of
the passenger compartment. Venting the warm air
and pulling in cooler ambient air by using car
ventilation system can reduce the soak air
temperature inside the passenger compartment [3].
The use of ventilation system on a parked car is
able to reduce the average air temperature by 5.6°C
and the seat temperatures by 5 - 6°C [11].
Study found that by using parked car ventilation
system the average air temperature inside the
passenger compartment is reduced by about 9.3%
[12].
This research aims to quantify the effects of
mechanical ventilator on the soak air temperature
inside the passenger compartment of the electric
car by using a numerical simulation technique.
Three dimension steady state simulation results
inside the passenger compartment are obtained by
using the commercial software FLUENT 6.3. The
numerical simulation results will be compared with
experimental data for the validation purpose. In
addition, the effects of mechanical ventilator
positions inside the passenger compartment are
investigated in terms of air temperature
distributions as well as air flow field.
2. NUMERICAL SIMULATION MODEL,
METHODS AND VALIDATION
2.1 Computational domain and boundary
conditions
The entire inner space of the vehicle is considered
as consists the computational domain and the inner
surfaces of the passenger compartment are defined
as the boundaries of the domain. The numerical
simulation is in a steady state operating condition.
Figure 1 indicates the computational domain based
on the compartment construction of Saga BLM
model. Rear passenger compartment model is
located on the origin position. The passenger
compartment was modelled by using dimensions of
a real car of Proton Saga BLM model where length,
width and height are 2523.30 mm, 1080 mm
and 1240 mm respectively. There are two
separated seats in the front passenger compartment
and a long bench at the rear passenger
compartment. There are four circles representing
the air inlets on the front dashboard and a
mechanical ventilator is located on the center of
rear dashboard. The four inlets are situated
symmetrically about the z-coordinate as shown in
Figure 1.
Both the inlets and mechanical ventilator are
defined as the air-flow inlets and outlet. The
mechanical ventilator is treated as the outlet. The
air velocity at the ventilator is 2.84 m/s and the
direction of the resultant velocity vector was
normal to the surface boundary. Since the air
velocity of the outlet is defined, the air velocity of
the inlets can be calculated by using the
conservation of mass principle. The air velocity at
the four inlets are fixed at 0.83 m/s and turbulent
intensities are 10% [7,15]. The boundary condition
at the bottom surface is set up as a temperature at
300 K. The convection and radiation boundary
conditions are applied to glass and roof surfaces.
Convective heat transfer coefficient for glass
surfaces are fixed at 15 W/m2°C and the thickness
of the glazing is set as 5 mm. Roof is assumed as
opaque wall and the convective heat transfer
coefficient is set as 15 W/m2°C with 12 mm
thickness [14]. The passenger compartment is
assumed to be well sealed, therefore no additional
air-flow inlet as well as outlet should be considered
[7]. No-slip solid wall is applied to the whole car
surfaces. The air flow in the passenger
compartment is assumed to be incompressible with
constant thermophysical properties [7,13].
Figure 1 : Geometric model of Proton Saga
BLM model
The 5th IMAT, November 12 – 13th
2012
33
2.2 Grid generation and numerical methods
Meshing of the computational domain of the model
consists of 955437 tetrahedral cells. The tetrahedral
cell are used due to the complexity of the model
[7,15].
The CFD software FLUENT 6.3 is used for the
simulation to solve continuum, energy and
transport equations numerically with natural
convection effects. The convection term is
discretized by the second-order upwind difference
[7,14,15]. The energy governing equations are
discretized by the finite volume method [7,17]. The
Semi-Implicit Method for Pressure-Linked
Equations (SIMPLE) algorithm [4,7,13-17] is used
for handling the coupling between pressure and
velocity. The standard turbulence model is
adopted in conjunction with the standard wall
function for the near wall region treatment [4,7,13-
17].
2.3 Validation of the numerical model
The numerical model is validated by comparing the
result of simulation with experimental data. During
the experiment, a metallic white color of Proton
Saga BLM model was parked in an open space
under the sunlight from 12 pm – 3 pm. An interior
air temperature and both internal and external
surface temperatures were monitored. The
experiment was conducted under the conditions
where the ambient temperature is 35°C and
incidence solar radiation is 1 kW/m2. The steady
state 3-D simulation result is shown in Figure 2,
where the experimental data are also presented for
comparison. The predicted and measured air
temperature values inside the passenger
compartment from 12 pm – 3 pm are listed in Table
1.
Figure 2 indicates the average air temperature
differences between the predicted data and
measured data are about 1.8°C. The measured data
were affected by many factors such as solar
radiation, material properties of the passenger
compartment and heat losses from the outer
surfaces of the passenger compartment.
Table 1: Predicted and measured air temperature values.
Figure 2 : Predicted air temperature and the
comparison with measured data
inside the passenger compartment.
3. RESULT AND DISCUSSION
3.1 Ventilation System in the Passenger
Compartment
The inlet air temperature is fixed at 36°C. The
ventilator is located on the center of rear dashboard
where the velocity of the ventilator is 2.84 m/s.
Figure 3 shows the steady state of air temperature
distributions inside the passenger compartment
with mechanical ventilation system installed in the
car. At the beginning of the simulation, the average
air temperature inside the passenger compartment
is about 48°C. With the existing of mechanical
ventilator in the passenger compartment, the steady
state average air temperature inside the passenger
compartment is reduced to 44°C. From the figure,
it can be seen that the distributions of the air
temperature inside the passenger compartment is
fairly uniform.
Figure 4 shows the air velocity vectors on a vertical
middle of z-plane. It can be observed that the air
flow field inside the passenger compartment is not
uniform. The air flow field concentrates more at
rear compartment rather than at the front. The
mechanical ventilator may not be at the best
position. The mechanical ventilator should be
placed at different position to obtain more uniform
air flow field inside the passenger compartment.
The 5th IMAT, November 12 – 13th
2012
34
Figure 3 : Steady state of air temperature
distributions at vertical middle of z-
plane (z = 0.54m)
Figure 4 : Velocity vectors distributions at
vertical middle of z-plane (z = 0.54
m).
3.2 Two mechanical ventilators on the rear
dashboard
The steady states air temperature distributions and
velocity vectors inside the passenger compartment
with the two mechanical ventilators are shown in
Figure 5 and 6. The ventilators are located
symmetrically 0.27 m away from the sidewalls on
the rear dashboard respectively. The air velocity at
the ventilators are fixed as 2.84 m/s. The average
air temperature inside the passenger compartment
is 43°C. It can be observed from figure 5 that the
air temperature distribution inside the passenger
compartment is uniform. The uses of two
mechanical ventilators are able to reduce the air
temperature inside the passenger compartment by
4°C.
Figure 6 shows the air velocity vectors of vertical
plane at z = 0.27 m from the sidewall. It can be
observed that the air flow field inside the passenger
compartment is uniform. The results suggest that
the number of mechanical ventilator has strong
influence on the air temperature inside the
passenger compartment.
Figure 5: Steady state of air temperature
distributions at vertical middle of z-
plane (z = 0.54 m).
Air velocity at v =1.11 m/s
The 5th IMAT, November 12 – 13th
2012
35
Figure 6 : Velocity vectors distributions at
vertical plane, z = 0.27 m from
sidewall.
4. CONCLUSIONS
The use of mechanical ventilator is identified as an
efficient way to lowering down the air temperature
inside the passenger compartment. Introducing the
mechanical ventilator located at the middle of rear
dashboard reduces the air temperature inside the
passenger compartment by 3°C. Introducing two
mechanical ventilators located 0.27 m away from
the sidewall of the rear dashboard reduces the air
temperature by 4°C. The simulation results suggest
that the number of mechanical ventilator has a
weak influence on the air temperature inside the
passenger compartment when they are installed on
the rear dashboard.
ACKNOWLEDGMENT
The authors gratefully acknowledge the Universiti
Teknologi Malaysia (UTM) for funding this study
under the project number 00G41 and also UTM-
PROTON Future Drive Laboratory for giving a
space to conduct the experimental work.
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National Renewable Energy Laboratory
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142 (2007 A.D. /1428 A.H.)
The 5th IMAT, November 12 – 13th
2012
37
Retrofitting R-22 Split Type Air Conditioning With
Hydrocarbon (HCR-22) Refrigerant
Henry Nasutiona,b,c
, Zulkarnain Abdul Latiffa,b
, Azhar Abdul Aziza,b
, Mohd Rozi Mohd
Peranga
aAutomotive Development Centre, Universiti Teknologi Malaysia
81310 Skudai, Johor, Malaysia, Phone: +60 75535447, Fax: +60 75535811 bFaculty of Mechanical Engineering, Universiti Teknologi Malaysia
81310 Skudai, Johor, Malaysia, Phone: +60 7 5534575, Fax: +60 7 5566159
email: [email protected]
cDepartment of Mechanical Engineering, Bung Hatta University
email: [email protected]
25132 Padang, Sumatera Barat, Indonesia, Phone: +62 751 7054657, Fax: +62 751 7051341
ABSTRACT
An experimental study to evaluate the energy
consumption of a split type air conditioning is
presented. The compressor works with the fluids R-
22 and HCR-22a and has been tested varying the
internal heat load 0, 500, 700 and 1000 W. The
measurements taken during the one hour
experimental periods at 10-minutes interval times
for temperature setpoint of 20oC. The performance
data considered where the evaporator cooling load,
the condenser heat rejection, the electrical energy
consumption, the refrigeration system temperatures,
and the room temperature. And hence the
Coefficient of Performance (COP) could be
determined. The final results of this study show an
overall better energy consumption of the HFC-22a
compared with the R-22.
Keywords: split air conditioning, R-22 and HCR-
22a, energy saving
1. INTRODUCTION
Hydrochlorofluorocarbons-22 (HCFC22) is the most
commonly used on the split-type air conditioner as a
refrigerant. However, because of issues of
environmental damage, causing the experts try to
use environmentally friendly refrigerants, namely
hydrocarbons. Several studies have shown that the
use of hydrocarbon refrigerants can improve the
performance of air-conditioning systems [1-5].
The negative impact of HCFCs is ozone layer
destruction and global warming. The HCFCs such as
R-22 and R-123 have the impact of the greenhouse
effect or global warming potential (GWP) that is
still relatively high [6]. The positive properties of
hydrocarbon refrigerant are zero ozone depletion
potential (ODP), very low global warming potential,
high miscibility with mineral oil, and non-toxicity.
The main negative of refrigerant hydrocarbons is
their flammability [7]. Although it has been known
that hydrocarbon refrigerants to replace CFC
refrigerants, HCFC, and HFC, but due to economic
reasons and flammability properties, which cause
the development of hydrocarbon refrigerant is
relatively stagnant. Flammability of hydrocarbon
refrigerants becomes which most prominent reason
to weaken the use of this refrigerant. Vehicle fuel is
much more flammable than hydrocarbon
refrigerants, but still safe to be used.
The Montreal Protocol requires the US (United State
of America) to reduce its consumption of HCFCs
by 75% below the US baseline. Allowance holders
may only produce or import HCFC22 to service
existing equipment. Virgin R-22 may not be used in
new equipment. As a result, heating, ventilation and
air-conditioning (HVAC) system manufacturers may
not produce new air conditioners and heat pumps
containing R22. January 1, 2015: The Montreal
Protocol requires the U.S. to reduce its consumption
of HCFCs by 90% below the U.S. baseline. January
1, 2020: The Montreal Protocol requires the U.S. to
reduce its consumption of HCFCs by 99.5% below
the U.S. baseline. A refrigerant that has been
recovered and recycled/reclaimed will be allowed
beyond 2020 to service existing systems, but
chemical manufacturers will no longer be able to
produce R22 to service existing air conditioners and
heat pumps [8].
Hydrocarbons are new findings for refrigerants by
the experts to replace the current HCFC22 as the
refrigerants for air conditioning system. Many
investigators have used a mixture of hydrocarbons
to improve the performance of air-conditioning
system. For example: Wongwises et al. [9] using a
mixture of hydrocarbons (propane-R290, butane-
R600 and isobutene-R600a). Wongwises and
Chimres [10] using a mixture of propane, butane,
and isobutene to replace HFC134a in freezer
applications. The results showed that a mixture of
propane (60%) and butane (40%) as the most
appropriate some other refrigerant. Hammad and
Alsaad [11] using LPG (24.4% propane, butane,
56.4%, and 17.2% isobutane) to replace CFC12
IMAT-UI 007
The 5th IMAT, November 12 – 13th
2012
38
refrigerant for refrigerator. Jung et al. [12] replace
CFC12 with a mixture of propane/isobutane
(R290/R600a) for the refrigerator. Han et al. [13]
study of azeotropic non R32/R125/R161 of the
vapor compression system applications, the results
showed that the cooling capacity and COP are better
than R407c. Park et al. [14] using a mixture of
propylene, propane, HFC152a, and dimethylether as
an alternative refrigerant to replace HCFC22
refrigerant in cooling systems and heat pump
split. The results show that the COP of the mixture
is 5.7% higher than HCFC22. Mani and Selladurai
[5] using the mixture as a replacement refrigerant
CFC12 R290/R600a and HFC134a vapor
compression refrigeration system. According to
experimental results, R290/R600a refrigerant
cooling capacity 19.9% to 50.1% higher than R12
and 28.6% to 87.2% compared to
R134a. R290/R600a mixture COP increased by 3.9-
25.1% compared to R12. Refrigerant R134a has a
slightly lower coefficient of performance of the R12.
With respect to these opportunities, current research
is focused on the temperature distribution, cooling
capacity and coefficient of performance, COP at
various internal heat loads using experimental result.
The experiment is also conducted using two
different refrigerants that are R-22 and HCR-22 to
evaluate the energy consumption of the current split
type air conditioning system.
2. SYSTEM DESCRIPTION AND
PROCEDURE
A schematic diagram of the experimental apparatus
is shown in Figure 1. The air-conditioning was
originally to work with refrigerant R-22 and with
compressor capacity, is 1860 W. Flow meter was
installed on the system to measure the flow of
refrigerant. The objective of the research was to
compare the refrigeration performance of different
refrigerants in terms of a coefficient of performance
(COP), cooling capacity, and power consumption by
the compressor.
The temperatures and pressures of the refrigerant are
measured at various locations in the experimental
set-up as shown in Figure 1. The refrigerant
temperature was measured by type T thermocouple
with an accuracy of ±0.1oC. The pressure was
measured through a tap with a small hole drilled into
the tube in which the refrigerant flows. The flow of
the refrigerants was measured by flow meter with an
accuracy of ±1 gr/s. The compressor energy
consumption was measured by current and voltage
of the electricity in the system. The accuracy of the
current meter is ±1%, and voltage meter is ±1.5%.
The refrigerants R-22 and HCR-22 were charged
after the system have been evacuated by a vacuum
pump. Drop-in experiments were carried out without
any modification on the system. The experiment on
the air-conditioning system was started with
HCFC22. Cooling capacity was obtained by varying
the load on the output of the evaporator. Data were
measured every 10 minutes during one hour. A
commercial refrigeration company prepared the
refrigerant mixtures. Composition of refrigerant
mixtures could not be presented in this research
because of company confidentiality. The data shown
in this study were the result of a mixture of the best.
In collecting data, the thermostat was set to the
maximum cool position. Data was collected on
when the conditions are considered stable, after the
air-conditioning was running 15 minutes for each
test condition. Thermodynamic properties of the
refrigerants were taken from the REFPROP database
[15]. The room setpoint temperatures during the
experiments were 20oC. In each experiment, the AC
system responded to the actual cooling load that
prevailed during the experimental period.
Figure 1: A schematic diagram of the
experimental set-up.
Caption:
1. Inlet compressor temperature, T1 (⁰C).
2. Exit compressor temperature, T2 (⁰C).
3. Exit condenser temperature, T3 (⁰C).
4. Inlet TEV temperature, T4 (⁰C).
5. Inlet evaporator temperature, T5 (⁰C).
6. Supply air temperature, T6 (⁰C).
7. Cabin temperature, T7 (⁰C).
8. Ambient temperature, T8 (⁰C).
9. Inlet compressor pressure, P1 (MPa).
10. Exit compressor pressure, P2 (MPa).
11. Exit condenser pressure, P3 (MPa).
12. Inlet evaporator pressure, P4 (MPa).
To obtain the differences that stand out about the
performance of the air conditioning due to drop-in,
analysis of several parameters is conducted. These
parameters are a coefficient of performance, cooling
capacity of the evaporator, and power consumption
of the compressor. The coefficient of performance
of refrigeration machine is the ratio of the energy
extracted by the evaporator (refrigerating effect) to
the energy supplied to the compressor. The
The 5th IMAT, November 12 – 13th
2012
39
coefficient of performance was calculated with
equation (1).
)(
)(
12
15
hh
hh
W
QCOP
c
e
(1)
where h1, h2 (kJ/kg) are the enthalpy at the
compressor inlet and outlet respectively, h5 (kJ/kg)
is the enthalpy at the evaporator inlet and outlet
respectively, Qe (kJ/kg) is the cooling capacity of
the evaporator, and Wc (kJ/kg) is the compression
work.
The energy saving calculated is expressed in terms
of saving in percentage unit, based on the difference
between energy consumed using R-22 and energy
consumed using HCR-22. The energy consumption
is calculated by multiplying the power consumption
of the compressor by the actual operating hours. The
equations are given as:
100Energy) 22-(R
Energy) 22-(HCR - Energy) 22-(R SavingEnergy
(2)
3. RESULTS AND DISCUSSION
3.1 Compressor Pressure
Figure 2 shows the compressor absolute pressure of
R-22 compared with HCR-22 refrigerants at internal
heat loads 0 - 1000 W. The graph illustrates the inlet
(Pi) and outlet (Po) absolute pressure of the
compressor. The result shows that the HCR-22
indicates the lower absolute pressure at the inlet and
outlet of the compressor compared with R-22.
Figure 2: Compressor absolute pressure of R-22
compared with HCR-22.
3.2 Mass Flow Rate
Figure 3 exhibits the mass flow rate against internal
heat load variation. The mass flow rate of the HCR-
22 refrigeration system is constantly lower
compared to the R-22. The result shows that the
HCR-22 refrigerant is higher than the R-22, the
average difference is 20%.
Figure 3: Mass flow rate of R-22 compared
with HCR-22.
3.3 Room Temperature
Figures 4 shows the relationship between room
temperature and time at various internal heat loads
with R-22 and HCR-22. For conditioned room at a
temperature of 20oC at internal heat load's variation
of 0, 500, 700, 1000 Watts obtained that the room-
temperature distribution at 16-19oC and 18-20
oC for
refrigerant R-22 and HCR-22 respectively. HCR-22
is superior when compared with HCFC22 due to
refrigerant mass flow rate using HCR-22 a lot more
because of the refrigerant density lower than R-22,
so that the ability to cool the room is faster.
Figure 4: Room temperature of R-22 compared
with HCR-22.
3.4 COP
Having obtained the values of enthalpy, the COP
was obtained as in Figure 5. It shows that the COP
for R-22 was lower than that of HCR-22. At the
time of the beginning data collection, this was 15
minutes after the air conditioner was operated, the
COP of R-22 lower than that of HCR-22. The COPs
increase of about 2.84% to 11.98%. The increase of
this COP was lower than that of the study conducted
by Mani and Selladurai [5], i.e. 11.8-17.6%, where
the evaporator temperature was about -8oC, for
freezer purpose. Besides, the drop-in was performed
on a refrigeration system with refrigerant CFC12
The 5th IMAT, November 12 – 13th
2012
40
replaced with hydrocarbon mixture, propane (R290)
and isobutene (R600a). The differences of
evaporator temperature and refrigerants presumably
is caused different the results.
Figure 5: COP of R-22 compared with HCR-
22. 3.5 Cooling Capacity
Increasing COP can be caused by increasing in
cooling capacity at the evaporator and/or by
decreasing the power consumption of the
compressor. To find a more dominant effect,
whether the addition of cooling capacity or decrease
power consumption in the compressor, one can see
it in Figures 6. The result shows that the cooling
capacity of HCR-22 was higher than that of R-22.
Due to the fact that the COP of R-22 lower than that
of HCR-22.
Figure 6: Cooling capacity of R-22 compared
with HCR-22.
3.6 Electrical Energy Consumption and Energy
Saving
Figure 7 shows the relationship between compressor
power consumption and internal heat load. Based on
the data described in the previous section, it was
indicated that, a drop-in from HCFC22 to the HC22
may reduce power consumption of the compressor
and increasing of cooling capacity. However, in
general it can be said that the drop-in from HCHC22
to the HC22 may cause saving of power
consumption. According to of the COP, the average
saving was 9%, and according to of the compressor
power consumption, savings were 15.14%. The
Figure 8 explains the occurrence of saving on the
compressor power consumption significantly.
Figure 7: Compressor power of R-22 compared
with HCR-22.
Figure 8: Energy saving of R-22 compared with
HCR-22.
4. CONCLUSION
Replacement of the R-22 with a HCR-22 as the
working fluid refrigerant of the split-type air
conditioning has been investigated to show the
performance of the system. This indicated that with
hydrocarbon mixtures can further improve the COP
and energy saving. During the experimental test,
HCR-22 mixtures were found to be safe. However,
care should be taken when using R-290/R-600/R-
600a mixture in an air-conditioning system.
Based on this study, the following conclusions were
drawn.
1. The absolute pressure of the HCR-22 at the inlet
and outlet of the compressor is always lower
compared to the R-22. The average difference of
the inlet and outlet are 5.77% and 18.08%
respectively.
2. The mass flow rate of the HCR-22 claims the
average difference is 20% compared to the R-22.
The result shows that the HCR-22 is lighter and
lower in specific gravity than the R-22. This is
the positive improvement for the HCR-22
refrigeration system by reducing the mass flow
of the system as this is relevant to the reduction
The 5th IMAT, November 12 – 13th
2012
41
of the inlet and outlet absolute pressure of the
compressor.
3. The refrigeration system work with both
refrigerants show the room temperature was
increased as the internal heat loads increased.
The average gap of the HCR-22 and R-22 is
8.7%. The room temperature of the HCR-22
exhibits the positive improvement as its
temperature is close to to the temperature setting
(200C) compared to the R-22. The average
difference of room temperature between the
HCR-22 to the temperature setting is 1.95% and
for the R-22 is 9.8%. The less percentage reveals
the nearest temperature to the temperature
setting, and it indicates the best temperature of
the refrigeration system.
4. The evaporator cooling load of the HCR-22
indicates more heat was absorbed into the system
compared to the R-22. The average difference is
50.2%. This is a good improvement of heat
absorption.
5. The COP of the HCR-22 refrigeration system
shows an encouraging improvement compared to
the R-22. The average difference is ranging from
2.84 to 11.98%.
6. The electrical energy consumption of the HCR-
22 is less compared to the R-22 and it explains
the positive improvement of electrical energy
saving of the system. The average difference is
15.14%.
ACKNOWLEDGMENT
The present research was supported financially by
Redto Green (M) Sdn. Bhd., Automotive
Development Centre – UTM, Universiti Teknologi
Malaysia : Fundamental Research Grant Scheme
(FRGS) No.78686 from the Ministry of Higher
Education (MOHE) Malaysia and appreciation to
Refrigeration Laboratory, Faculty of Mechanical
Engineering, Universiti Teknologi Malaysia for
facilitating support for this research. Their guidance
and assistance are gratefully acknowledged.
REFERENCES
[1] R.W. James and J.F. Missenden, ―The use of
propane in domestic refrigerator,‖ International
Journal Refrigeration, vol. 15, pp. 95-100,
1992.
[2] S. Devotta and S. Gopichand, ―Comparative
assessment of HFC 134a and some refrigerants
as alternative to CFC 12,‖ International Journal
Refrigeration, vol. 15, pp. 112-118, 1992.
[3] J. Dongsoo, K. B. Chong, L. H. Byoung, and L.
W. Hong, ―Testing of a hydrocarbon mixture in
domestic refrigerator, in Symposia AT-96-19-
3,‖ ASHRAE Transactions, pp. 1077-1084,
1996.
[4] D.S. Jung, C. Kim, K. Song, and B. Park,
―Testing of Propone/Isobutene mixture in
domestic refrigerator,‖ International Journal
Refrigeration, vol. 44, pp. 517-527, 2000.
[5] K. Mani and V. Selladurai, ―Experimental
analysis of a new refrigerant mixtures as drop-
in replacement for CFC12 and HFC134a,‖
International Journal of Thermal Sciences, vol.
47, pp 1490-1495, 2007.
[6] UNEP, Montreal Protocol on Substances that
Deplete the Ozone Layer, Final Act, New York:
United Nation Environmental Program, 1987.
[7] R.G. Richards and I.R. Shankland,
―Flammability of alternative refrigerants,‖
ASHRAE Journal, vol. 34, pp. 4, 1992.
[8] EPA,‖ What you should know about
refrigerants when purchasing or repairing a
residential A/C system or heat pump,‖
(available online
http://www.epa.gov/ozone/title6/phaseout/22ph
aseout.html [accessed on 08/10/2012]), 2012.
[9] S. Wongwises, A. Kamboon, and B. Orochon,
―Experimental investigation of hydrocarbon
mixtures to replace HFC-134a in an automotive
air conditioning system,‖ Energy Conversion &
Management, vol. 47, pp. 1644-1659, 2006.
[10] S. Wongwises and N. Chimres, ―Experimental
study of hydrocarbon mixtures to replace HFC-
134a in a domestic refrigerator,‖ Energy
Conversion & Management, vol. 46, pp. 85-
100, 2005.
[11] M.A. Hammad and M.A. Alsaad, ―The use of
hydrocarbon mixtures as refrigerants in
domestic refrigerators,‖ Applied Thermal
Engineering, vol. 19, pp. 1181–1189, 1999.
[12] D. Jung, C.B. Kim, B.H. Lim, and H.W. Lee,
―Testing of a hydrocarbon mixture in domestic
refrigerators,‖ ASHRAE Transactions, vol. 3,
pp. 1077–1084, 1996.
[13] X.H. Han, Q. Wang, Z.W. Zhu, and G.M. Chen,
―Cycle performance study on R-32/R-125/R-
161 as alternative refrigerant to R-407C,‖
Applied Thermal Engineering, vol. 27, pp.
2559-2565, 2007.
[14] K.J. Park, T. Seo, and D. Jung, ―Performance of
alternative refrigerants for residential air
conditioning applications,‖ Applied Energy, vol.
84, pp. 985–991, 2007.
[15] E.W. Lemmon, M.L. Huber, and M.Q.
McLinde, REFPROP, Reference Fluid
Thermodynamics and Transport Properties,
NIST Standard Reference Database 23, Version
9.0., 2009.
The 5th IMAT, November 12 – 13th
2012
42
Characterization of Generator with Palm Oil Biodiesel
at Different Compression Ratio
Belyamina,b
, Alias bin Mohd. Noorc, Mohanad Hamzah Hussein
d,
Mazlan bin Saide, Mohd Haffidzi
f
aTransportation Research Alliance
Universiti Teknologi Malaysia, 81300 Skudai, Johor Bahru Tel : +60137675038
E-mail : [email protected]
bMechanical Engineering Department
Politeknik Negeri Jakarta, Depok 16425 Tel : +60137675038
E-mail : [email protected]
cTransportation Research Alliance
Universiti Teknologi Malaysia, 81300 Skudai, Johor Bahru Tel : +60197281873
E-mail : [email protected]
dMechanical Engineering Faculty
Universiti Teknologi Malaysia, 81300 Skudai, Johor Bahru Tel : +60137327098
E-mail : [email protected]
eTransportation Research Alliance
Universiti Teknologi Malaysia, 81300 Skudai, Johor Bahru Tel : +60137797846
E-mail : [email protected]
fTransportation Research Alliance
Universiti Teknologi Malaysia, 81300 Skudai, Johor Bahru Tel : +60126942442
E-mail : [email protected]
ABSTRACT
Experiment to determine exhaust gas emission and
combustion characteristics of a compression
ignition generator was carried out. The experiment
used single cylinder four strokes direct injection
engine which was fueled with diesel and palm oil
methyl ester of B2 (blends 2% palm oil methyl
ester with 98% diesel on a volume basis), B5, B7
and B10. The experiment was conducted at a fixed
engine speed of 3000 rpm and 50% load with
variety compression ratios of 16:1, 18:1, and
20:1and 22:1. Optimum compression ratio,
influence of compression ratio on specific fuel
consumption and thermal efficiency were
examined. Palm oil methyl ester produce better
output when the engine operate with variable
compression ratio. Specific fuel consumption and
NOx decrease and thermal efficiency increase when
using highest compression ratio.
Keywords : palm oil methyl ester , variable
compression ratio, four stroke, direct
injection
1. INTRODUCTION There is a global increase in the investigation on
the application of alternative fuel sources for daily
use, such as biodiesel and alcohol. This is due to
the fact that petroleum products are becoming very
scarce and expensive and also the price of
petroleum products is always on the high side.
There is also an awareness of air pollution caused
by the extensive use of conventional fuel in an
internal-combustion engine.
In the past two decades, vegetable oil such as
mahua oil, sun flower, seed oil, waste cooking oil,
and palm oil have been used as a substitute to
diesel in an internal-combustion engine [1, 2, 3].
These papers study the performance and emission
characteristic of generator fueled with biodiesel or
its blend. It shows that biodiesel can substitute
fossil fuel in an internal-combustion engine with or
without engine modification. It is also economical
and competitive compare to pure diesel. When the
fuel is waste palm oil, no engine modification is
required.
IMAT-UI 008
The 5th IMAT, November 12 – 13th
2012
43
In addition, biodiesel has lower sulphur, aromatics
contents, and net carbon dioxide (CO2) emission
[4]. It has better lubricity and biodegradability and
less toxic relative to fossil diesel [5]. Bio diesel can
be used readily since it can be mixed at any
proportion with diesel. This enables it to be applied
immediately in diesel power generator without
much modification. Considering exhaust emissions,
reference [6] reported that the use of bio diesel
results in lower emissions of unburnt hydrocarbons,
carbon monoxide, smoke and particulate matter
beside some increase in emissions of NOx. A
number of researchers have investigated vegetable
oil-based fuels [2, 3, 7, and 8]. Reference [3] had
concluded that vegetable oil can be safely burnt for
a short period of time in a diesel engine. However,
the use of raw vegetable oil for extended period of
time may result in severe engine deposits, piston
ring sticking, injectors choking, and thickening of
the lubricating oil.
This experiment is to determine exhaust gas
emission and combustion characteristics of
a compression ignition generator using B2, B5, B7
and B10.
2. METHODOLOGY
Experiment to examine combustion characteristic
and exhaust gas emission was conducted on a
single-cylinder diesel engine four stroke for variety
CR (Compression ratio) and biodiesel formulation,
B2, B5, B7 and B10. Variety of CR is achieved by
changing the cylinder head gasket thickness. CR
increase when gasket thickness is reduced.
The exhaust gases emissions such as NOx, CO, CO2
and exhaust smoke density was measured by using
the emission analyzer. All these results were
discussed to find the effect of changing the CR and
fuel formula.
Power generated was loaded by set of lamps.
Ampere and volt of lamps were measured by
ampere meter and voltmeter as alternative to torque
measurement.
The generator set specification is as shown in Table
1. whereas fuel property is shown in Table 2.
Table 1. Engine test specification
Type Yanmar l70N6-MTRIYJ made in Italia 4 stroke, vertical cylinder diesel
generator engine
No. of cylinder 1
Bore x stroke 78 x 67mm
Displacement 0.320L
Combustion system Direct injection
Cooling system Forced air by flywheel fan
Maximum engine speed 3600(rpm)
Starting system Electric start/Recoil start
Max Rated output(KW] @3600rpm 4.9
Compression ratio 20:1
Table 2. Fuel test properties
Specification Pure diesel B2 B5 B7 B10
Specific gravity
@60f0
0.8448 0.8448 0.8453 0.8458 0.8473
Density @15c0kg/m
3 844.3 844.3 844.8 845.3 846.9
A p I 36.0 36.0 35.9 35.8 35.7
Viscosity 28c0 7.123 6.897 6.893 6.881 6.573
Viscosity 40c0 5.17 5.131 5.063 5.048 5.019
Viscosity 100c0 1.781 1.671 1.667 1.638 1.579
Viscosity index 243.35 250.57 263.73 266.74 272.66
Pour point c0 3 3 3 3 3
Flash point c0 84.0 83 86 86 87
gross calorific
value(kJ/kg)
45652.0 45801.3
45589.0
45235.9
45011.9
The 5th IMAT, November 12 – 13th
2012
44
3. RESULT AND DISCUSSION
3.1 Specific Fuel Consumption
Fuel consumption, FC, is calculated by Equation 1.
FC = V / t (1)
Specific Fuel Consumption is then calculated by
Equation 2
SFC = FC/ P (2)
The variation of Specific Fuel Consumption (SFC) with
Compression Ratio (CR) is given in Figure 1. It can be
observed that the SFC is a clear indication of efficiency
with which the engine develops power. The smaller SFC
indicates the more effective use of fuel to generate power.
For all fuels tested, the SFC decreased with an increased of
CR. This is due to increase of temperature in combustion
chamber, leading to complete combustion. It has been
observed that the maximum SFC of B2 reduced by 0.96%
at CR 22:1 relative to CR 20:1 and the other fuels. This
can be because B2 has high caloric value compare to other.
At CR 18 and 16, SFC B5 and B7 have the highest rise of
SFC by 1.1%, 1.26% respectively. Minimum SFC mean
efficient use of fuel. This happen in CR 20 to 22. B2
always have the lowest vaue of SFC which mean the best
one in term of fuel efficiency. At higher percentage of
blends, the SFC increases due to decrease in calorific
value.
Figure 1.Variation of specific fuel consumption with compression ratio for different fuel blends
3.2 Thermal Efficiency.
Thermal efficiency is calculated by Equation 3
ɳth = P/ Qin (3)
Qin is calculated by Equation 4
Qin = m CV (4)
The thermal efficiency (ɳth) of the engine is considered one
of the most important criteria for evaluating the
performance of the engine. It indicates the combustion
effectiveness of the engine. The ɳth is defined as the actual
work per cycle divided by fuel chemical energy (fuel
calorific value). Figure 2 shows the variation in ɳth with
CR for blends fuel tested .The ɳth was found to be lower
for biodiesel at all blends than diesel for all specified CR.
This might be due to lower fuel heat value and higher fuel
consumption of the bio diesel blends to produce the same
power.
In this figure, it appears that the optimum ɳth of B2 occur
at CR 22:1. This may be due to the fuel calorific value and
low SFC. At CR18 and CR16 the ɳth reduce by 11%, 17%
for B5, B7 respectively compare to CR20, it was also
observed that increasing the CR more beneficial when
biodiesel is used rather than diesel. Due to their low
volatility and high viscosity, biodiesel perform relatively
better at higher compression ratios [9].
3.3 Exhaust Gas Emission
3.3.1 Nitrogen Oxide
The variation of Nitrogen Oxide (NOx) with respect to CR
for different blends and constant load is shown in Figure 3.
NOx emission for diesel and other blends increase when
the CR is increased. The augmentation in the biodiesel
ratio in the fuel blend increased NOx emissions by 1.07%,
1.12%, 1.16% and 1.18% for B2, B5, B7 and B10,
respectively, the reason for higher NOx emission for
blends is higher peak temperatures. This figure shows an
increase in NOx by 1.3% at CR 22 while the NOx
decreased by 17.7% and 30% at CR 18:1 and 16:1
respectively as compared to the CR 20. The changes in
NOx resemble up to some extent to exhaust temperature
which is related to an increase in CR.
The 5th IMAT, November 12 – 13th
2012
45
Figure 2.Variation of thermal efficiency with compression ratio for different fuel blends
Figure 3.Variation of Oxide Nitrogen with compression ratio for different fuel blends
3.3.2 Carbon Monoxide
Figure 4 illustrate the variation of CO for variety fuel
blends with respect to VCR. From this figure it seen that
the specified blends produce less CO emission than diesel
for every CR at applied load, it might due to increase the
cylinder temperature. Therefore engine temperature lead to
better combustion process and might cause less CO
emission. The CO decrease by 29.1%, 10.87%, 4.2% and
0.5%
when the compression ratio increase from 20:1 to 22:1 for
diesel,B2,B5,B7 and B10 respectively, This could be
because biodiesel provide more oxygen to the combustion
chamber. This lead to the more complete combustion. The
other reason is that the percentage of CO decreases due to
rising temperature in the combustion chamber. physical
and chemical properties of the fuel, air–fuel ratio, the
effects of fuel viscosity on spray quality will be expected
to cause CO emission increase with vegetable oil
fuels [10].
The 5th IMAT, November 12 – 13th
2012
46
Figure 4.Variation of Carbon Monoxide with compression ratio for different fuel blends
3.3.3 Carbon Dioxide
The variation of CO2 with VCR is shown in Figure 5.
From this figure, it can be observed that the CO2 increases
by around 18% than CR20. In decreasing CR, blends of
fuel increase its CO2 emission by 4%, 4.5%, 8% and 11%
for the B2, B5, B7 and B10 respectively. This is due to the
high oxygen content of blends. Higher amounts of CO2 is
an indication of complete combustion of fuel in the
combustion chamber. It also relates to the exhaust gas
temperature. CO2 emissions of the fuel blends slightly
increase by increasing the load for specified compression
ratios due to complete combustion.
3.3.4 Smoke Density
The variation of Smoke density emission with VCR at
constant load is shown in Figure 6. In this figure, it was
shown that the smoke density increased by 35% and 60%
when the CR decreased to 18:1 and 16:1 respectively. At
lower CR the temperature is lower. Incomplete
combustion in the combustion chamber then lead to more
smoke exhausted from the engine. The smoke decreased
when the blends percentage is increased. These figures
show that the smoke was reduced significantly by around
9%, 12%, 17% and 22% for B2, B5, B7 and B10 than
diesel. In addition, it was found that the maximum
reductions were around 20% at CR 22 for all fuel blends
than CR20, this is due to the increase of inside temperature
of the combustion chamber and because palm oil contains
more oxygen which improves the combustion process. At
the end this will decrease the smoke. The Smoke is emitted
from diesel engines because of the incomplete combustion
in the combustion chamber.
Figure 5. Variation of Carbon Dioxide with compression ratio for different fuel blends
The 5th IMAT, November 12 – 13th
2012
47
Figure 6. Variation of Smoke with compression ratio for different fuel blends
4. CONCLUSION
An experimental was conducted on direct injection diesel
engine generator evaluate the performance, combustion
and exhaust emission at different blends and compression
ratio. Thermal efficiency, ɳth of the generator and CO2
emission increases with an increase in the CR whereas
SFC decreased. The use of biodiesel in general reduce the
CO emission and increase CO2 which indicate better
combustion performance. Performance of the B2 is
superior compare with Diesel. Optimum CR to provide
optimum SFC and Thermal efficiency is 20 and above
ACKNOWLEDGMENT
The authors are grateful to the Ministry of Higher
Education and Universiti Teknologi Malaysia for grant
GUP project under Vot Q.J130000.2609.00J35
NOMENCLATURE
B2 mixture of 2% biodiesel and 98% diesel
B5 mixture of 5% biodiesel and 95% diesel
B7 mixture of 7% biodiesel and 93% diesel
B10 mixture of 10% biodiesel and 90% diesel
CO carbon monoxide
CO2 carbon dioxide
CR compression ratio
CV Calorific Value of fuel
D 100% diesel
FC Fuel consumption
m mass flow rate of fuel
NOx nitrogen oxides
O2 oxygen
P Power generated
Qin Heat input
t time taken to consume the fuel
SBFC specific brake fuel consumption
VCR variable compression ratio V volume of fuel used
ɳth brake thermal efficiency
REFERENCES
[1]. S.C.A. ALMIEDA, C.R. BELCHIOR, M.V.G. NASCIMENTO,
L.S.R. VIEIRA AND G. FLEURY, ―PERFORMANCE OF A DIESEL
GENERATOR FUELLED WITH PALM OIL‖, FUEL, VOL 81,
PP.2097-2102, 2002.
[2] A.SRIVASTAVA, R.PRASAD, ―TRIGLYCERIDES-BASED DIESEL
FUELS‖, RENEWABLE SUSTAINABLE ENERGY REV., VOL 4,
PP.111–33, 2000.
[3] F.KARAOSMANOGLU, G.O.KURT, ¨ ZAKTAS T .LONGTERM CI
ENGINE TEST OF SUNFLOWER OIL‖, RENEWABLE ENERGY,
VOL.19, PP.219–21, 2000.
[4]. M.Z.SULAIMAN, F.M.ISA, ―THE EFFECT OF DIFFERENT
GASOLINE BLENDS DOPED WITH USED ENGINE OIL ON THE
FORMING TENDENCY OF SIMULATED IN TAKE VALVE
DEPOSITS‖, PROCEEDINGS OF THE INSTITUTION OF MECHANICAL
ENGINEERS, PP.213, PART D, 1999.
[5] H.Z.GOETLLEN, M.ZIEJEWSKI, K.R.KAUFMAN, G.L. PRATT,
―FUEL INJECTION ANOMALIES OBSERVED DURING LONG-BURN
ENGINE PERFORMANCE TEST ON ALTERNATE FUELS‖, SAE
TECHNICAL PAPER SERIES. SOCIETY OF AUTOMOTIVE
ENGINEERS, NO.852089, 1985.
[6] M.S.GRABOSKI AND R.L.MCCORNIMIK, ―COMBUSTION OFF
AT AND VEGETABLE OIL DERIVED FUELS IN DIESEL ENGINES‖,
PROG ENERGY COMBUST SCI ., VOL.24, PP.125–64, 1998.
[7] A.ISIGIGUR, F.KARAOSMANOGLU, H.A.AKSOY,
F.HAMDULLAHPUR, L.O.GULDER, ‖PERFORMANCE AND
EMISSION CHARACTERISTICS OF A DIESEL ENGINE OPERATING
ON SUNFLOWER SEED OIL METHYLESTER‖, APP BIOCHEM
BIOTECHNOL, VOL.NO.45/46, PP.93–102, 1994.
[8] R.ALTIN, S.CETINKAYA AND H.S.YUCESU, ―THE POTENTIAL
OF USING VEGETABLE OIL FUELS AS FUEL FOR DIESEL
ENGINES‖, ENERGY CONVERSION MANAGEMENT . VOL.42,
PP.529–38, 2001.
The 5th IMAT, November 12 – 13th
2012
48
[9] H.RAHEMAN AND S.V. GHADGE, ―PERFORMANCE OF DIESEL
ENGINE WITH BIODIESEL AT VARYING COMPRESSION RATIO AND
IGNITION TIMING‖, FUEL, VOL.87, PP. 2659–2666, 2008.
[10].K.MURALIDHARAN AND D.VASUDEVAN, ―PERFORMANCE,
EMISSION AND COMBUSTION CHARACTERISTICS OF A VARIABLE
COMPRESSION
The 5th IMAT, November 12 – 13th
2012
49
The Effect of Fuel Additives on Gasoline Heating Value and
Spark Ignition Engine Performance: Case Study Z.A. Latiff
a, Azhar Abdul Aziz
b, M.R. Mohd Perang
c, N. Abdullah
d
Automotive Development Centre (ADC)
Faculty of Mechanical Engineering
Universiti Teknologi Malaysia (UTM)\
81310 Johor Bahru, Malaysia
[email protected] [email protected]
ABSTRACT
Today fuel additives had been used widely for the
enhancement of fuel economy and engine
performance. Fuel additives are substance that acts
as catalysts for the completeness combustion of
fuel in order to increase the heat released and hence
the work output will be improved. The purpose of
this paper is to investigate the effect of the
additives on fuel heating value and engine
performance. In this study, three different additives
available in the market have been chosen to
determine the effect on heating value and engine
performance when mixed with fuel. Two types of
test were conducted, namely the calorific value and
engine performance test. The first test was
conducted using a bomb calorimeter with test
method in accordance with the DIN 51900 and
ASTM D240. The later test was done using engine
test bed and with the agreement of BS 5514 (Parts
1 to 6), Reciprocating Internal Combustion
Engines: Performance, and SAE 1349 Standard
Engine Power Test Code. The study shows that fuel
additives can cause a standard fuel to have higher
heating value up to 5%. As for the engine
performance, the engine brake thermal efficiency
and brake mean effective pressure were increased
up to 8% and 10% respectively. The specific fuel
consumption can be reduced up to 9%.
Keywords : Engine performance, Fuel additive,
Fuel economy
1. INTRODUCTION
Fuel additives are compounds formulated to
enhance the quality and efficiency of the fuels used
in motor vehicles. This additive can be
incorporated in the fuel itself as needed. The fuel
additive is sold as a separate product and
consumers may use to improve or maintain the
performance of their engines.
One of the main advantages of the fuel additives is
to improve the engine performance. With some fuel
additives added, the product is claimed to boost the
octane level of the fuel, providing the engine with
more power with the same of fuel used. Thus with
this improvement one have the ability to travel
more in other words better mileage.
Along with preventing carbon build-up in the
combustion chamber, fuel additives are claimed to
also enhance proper lubrication of working
components. This particular benefit means less
wear and tear due to less surface friction on the
moving parts, which translates into lower and less
frequent repairs during the life of the vehicles.
2. IMPORTANCE OF FUEL ADDITIVE
Most people agree that some of the basic fuel
additives found in many gasoline products today
are of some benefit, there is some issues regarding
the use of over the counter additives. Proponents
claim the product boost the protection offered by
the gasoline products and make a significant
difference in how well a vehicle performs.
Opponents claim that over-the-counter additives for
fuel provide no extra benefits and in fact could
damage the engines if nor used properly.
A number of experimental investigations have been
reported with a wide variety of metal additives to
improve the fuel properties and the engine
performance, as well as to reduce emissions. The
effect of calcium, barium, iron, and nickel
naphthenates have been studied, concluding that
calcium and barium most efficiently reduce soot,
by both suppressing soot formation and enhancing
soot oxidation [1]. Based on experimental
investigations, Gürü et al. [2] concluded that
manganese, as a fuel additive, has a greater effect
in the reduction of the freezing point of the fuel,
than copper, magnesium, or calcium. Emission
measurements with manganese as a fuel additive
demonstrated that O2 and CO could be decreased
by 0.2% and 14.3%, respectively, SO2 emission
could be reduced, and the overall impact of all
these effects was found to lead to an increase of
0.8% in the net operating efficiency.
Valentine et al. [3] experimentally observed that
bimetallic platinum and cerium diesel fuel borne
catalyst reduces the engine emissions and improves
IMAT-UI 009
The 5th IMAT, November 12 – 13th
2012
50
the performance of the diesel particulate filter.
Shi et al. [4] reported that the particulate matter
emission decreases with increasing oxygenate
content in the fuels, but nitrogen oxides emissions
increase. De et al. [5] experimentally observed that
the presence of ethanol and ethyl ter-butyl ether
(ETBE) significantly alters the characteristics of
volatility and reduces the cetane number, impairing
the fuel‘s performance in engine tests. The effect of
methanol-containing additive (MCA) on the
emission of carbonyl compounds generated from
the diesel engine was studied by Chao et al. [6] and
it was observed that the emission factors for some
of the carbonyl compounds with the use of MCA
are higher than the values for those without the use
of MCA.
Metal oxides such as those of copper, iron, cerium,
and cobalt have been extensively used as fuel
additives. The effect of cerium on the size
distribution and composition of diesel particulate
matter has been studied by Skillas et al. [7],
indicating a reduction in the accumulation mode,
but an increase in ultrafines. Lahaye et al. [8]
studied the effect of cerium oxide on soot
formation and postoxidation and observed that the
soot yield is not affected significantly by the
presence of cerium oxide in the fuel for given
oxygen content. Based on experiments,
Jung et al. [9] observed that the addition of cerium
to diesel fuel causes significant changes in the
number concentration of particles in the
accumulation mode, light off temperature, and the
kinetics of oxidation. Even though the 4 oxidation
rate increased significantly with the addition of
cerium to the fuel, the dosing level was found not
to have much influence [10, 11].
3. HEATING VALUE
The heating value or calorific value of a substance,
usually a fuel or food, is the amount of heat
released during the combustion of a specified
amount of it. The calorific value is a characteristic
for each substance. It is measured in units of energy
per unit of the substance.. Heating value is
commonly determined by the use of a bomb
calorimeter.
The quantity known as higher heating value (HHV)
(or gross calorific value or gross energy or upper
heating value) is determined by bringing all the
products of combustion back to the original pre-
combustion temperature, and in particular
condensing any vapor produced. This is the same
as the thermodynamic heat of combustion since the
enthalpy change for the reaction assumes a
common temperature of the compounds before and
after combustion, in which case the water produced
by combustion is liquid.
The quantity known as lower heating value (LHV)
(or net calorific value) is determined by subtracting
the heat of vaporization of the water vapor from the
higher heating value. This treats any H2O formed as
a vapor. The energy required to vaporize the water
therefore is not realized as heat. Gross heating
value accounts for water in the exhaust leaving as
vapor, and includes liquid water in the fuel prior to
combustion. This value is important for fuels like
wood or coal, which will usually contain some
amount of water prior to burning.
The water vapor produced by combustion,
recovering heat which would otherwise be wasted.
Most applications which burn fuel produce water
vapor which is not used and thus wasting its heat
content. In such applications, the lower heating
value is the applicable measure. This is particularly
relevant for natural gas, whose high hydrogen
content produces much water. The gross calorific
value is relevant for gas burnt in condensing boilers
and power plants with flue gas condensation which
condense.
3.1 Bomb Calorimeter Test
A calorimeter is a device used for calorimetry, the
science of measuring the heat of chemical reactions
or physical changes as well as heat capacity. The
word calorimeter is derived from the Latin word
calor, meaning heat. Differential scanning
calorimeters, isothermal microcalorimeters,
titration calorimeters and accelerated rate
calorimeters are among the most common types. A
simple calorimeter just consists of a thermometer
attached to a metal container full of water
suspended above a combustion chamber.
A bomb calorimeter is a type of constant-volume
calorimeter used in measuring the heat of
combustion of a particular reaction. Bomb
calorimeters have to withstand the large pressure
within the calorimeter as the reaction is being
measured. Electrical energy is used to ignite the
fuel; as the fuel is burning, it will heat up the
surrounding air, which expands and escapes
through a tube that leads the air out of the
calorimeter. When the air is escaping through the
copper tube it will also heat up the water outside
the tube. The temperature of the water allows for
calculating calorie content of the fuel.
Fuels such as coal and oil, are traded based on the
calorific value of the material. Regulations have
been established with regards to the total calorific
content of the coal, the quality or purity of the coal,
and the classification of the coal. The gross
calorific value of the coal is also used to evaluate
the effectiveness of the beneficiation process in use
at the plant. ISO 1928, ASTMD 5865, BS1016, and
DIN 51900 are the most common methods for
determining the gross calorific value of coal and
coke.
Liquid fuels such as gasoline, kerosene, diesel, and
gas turbine fuels are also tested by bomb
The 5th IMAT, November 12 – 13th
2012
51
calorimeter. The heat of combustion (HOC) of the
fuel will provide a measure of the energy available
from a fuel. The mass heat of combustion (essential
for airplanes and hydrofoils) and the volumetric
heat of combustion (essential for automobiles and
ships) can also be determined with the HOC value.
ASTM D240 and D4809 are common methods for
this sort of work.
4. ENGINE PERFORMANCE
An internal combustion engine is a work producing
device that converts chemical energy into heat
followed by conversion into mechanical energy.
Chemical energy contained in the fuel is released as
heat by burning the fuel inside the combustion
chamber to produce gas at high temperature and
pressure. The high pressure gas acts to move the
piston and other related mechanisms inside the
engine to produce mechanical work.
Some parameters have to be obtained to evaluate
the behavior and performance of an internal
combustion engine such as speed, brake load and
fuel consumption. The experimental data have to be
analyzed to determine brake power, brake mean
effective pressure, specific fuel consumption and
brake thermal efficiency.
4.1 Engine Performance Test
The test was done using engine test bed and with
the agreement of BS 5514 (Parts 1 to 6),
Reciprocating Internal Combustion Engines:
Performance, and SAE 1349 Standard Engine
Power Test Code.
5. GASOLINE AND ADDITIVES
MIXTURE
Three type of additives were selected in this study
and the details are given as follows:
Additive A – Power additive which contains
mixture of oil, dimenthyl heptanes, trimethyl
etc.
Additive B – Cleaner additive with high purity
of polyetheramine (PEA) detergent; and
Additive C – Power additive with concentrated
biodegradable formula.
The mixture preparation of gasoline and
additive was down according to the proportion
suggested by the additive manufacturer.
6. RESULT AND DISCUSSION
Table 1 shows the energy value of fuel mixed with
different additive concentration for each sample,
established by using bomb calorimeter test.
Table 1: Fuel energy for different additive concentration
Additive Additive
Concentration (%) A B C
0 42146.5 40714.4 39983.4
5 42313.2 41425.4 42473.0
10 43101.7 42416.3 42710.8
15 43680.0 43078.0 42826.6
20 43861.8 44066.2 43804.4
Figure 1 shows the profile of energy content with
different additive concentrations. The trend shows
increasing in energy posed by the fuel mixture as
the concentration increases. It also shows additive
A have the highest increased in energy value then
followed by C and B. In term of percentage
increased, additive A, B and C shows 5.42, 3.93
and 3.88% respectively when compared to the
standard fuel energy value.
For engine performance test the preparation of fuel
used was based on the standard mixing procedure
recommended by the additive manufacturers.
Figure 2 shows the brake power against engine
speed. The test was done at part throttle opening.
The graph show the increased in brake power at
low speed and high load compared to standard fuel
used. While at high speed, low load the parameter
shows no significant difference. Generally additive
B has the highest average brake power increment of
9.51% and followed by additive A 8.44%. Additive
C in contrast shows a negative improvement except
at low speed.
Figure 1: Energy content at different additive
concentration
Figure 2: Brake power against engine speed
The 5th IMAT, November 12 – 13th
2012
52
Specific fuel consumption shows the economic
point of view for fuel usage and shown in Figure 3.
Based on the same power output, additive B shows
attractive saving followed by additive A and C.
Average saving made are 9.04, 9.06 and 5.9%
respectively for additive A, B and C.
Figure 3: Specific fuel consumption against
engine speed
Brake mean effective pressure is a measure of an
engine's capacity to do work that is independent of
engine displacement. From the experiment done the
measures is shown in Figure 4. Conclusively there
are no major changes made by the additives as far
as brake mean effective pressure is concerned.
Figure 4: Brake mean efective pressure against
engine speed
The ability of an engine to convert thermal energy
to available work can be shown in Figure 5 as
brake thermal eficiency. The thermal energy
obtained from the combustion of air-fuel mixture
represent 100% of energy input to the engine. The
experimental results exhibit additive B posed the
most potential of poducing work, moderately by
additive A and least work produced by additive C
compared to the standard fuel used.
Figure 5: Brake thermal efficiency against
engine speed
7. CONCLUSION
From this experimental study the following
conclusions are made:
Additives mixed with fuel increased the
energy content of the fuel.
The additives used have the effect on
engine performances as discussed.
The value of energy content has
significant effect on the engine
performance.
REFRENCES [1] N. Miyamoto, H. Zhixin, A. Harada, H.
Ogawa, and T. Murayama, Characteristics of
Diesel Soot Suppression with Soluble Fuel
Additives. SAE Technical Paper 871612,
1987.
[2] M. Gürü, U. Karakaya, D. Altiparmak, and
A. Alicilar, Improvement of Diesel Fuel
Properties By Using Additives, Energy
Conversion And Management. Volume. 43,
no. 8, pp. 1021–1025, 2002.
[3] J. M. Valentine, J. D. Peter-Hoblyn, and G.
K. Acress, Emissions Reduction and
Improved Fuel Economy Performance From
A Bimetallic Platinum/Cerium Diesel Fuel
Additive At Ultra-Low Dose Rates. SAE
Technical Paper 2000-01-1934, 2000.
[4] X. Shi, Y. Yu, H. He, S. Shuai, J. Wang, and
R. Li, Emission Characteristics Using
Methyl Soyate—Ethanol—Diesel Fuel Blends
On A Diesel Engine. Fuel, Vol. 84, no. 12-
13, pp. 1543–1549, 2005.
[5] E. W. De Menezes, R. Da Silva, R. Cataluña,
and R. J. C. Ortega, Effect Of Ethers And
Ether/Ethanol Additives On The
Physicochemical Properties Of Diesel Fuel
The 5th IMAT, November 12 – 13th
2012
53
And On Engine Tests Fuel. Vol. 85, no. 5-6,
pp. 815–822, 2006.
[6] H.-R. Chao, T.-C. Lin, M.-R. Chao, F.-H.
Chang, C.-I. Huang, and C.-B. Chen, Effect
Of Methanol-Containing Additive On The
Emission Of Carbonyl Compounds From A
Heavy-Duty Diesel Engine, Journal of
Hazardous Materials B, 73, pp. 39–54, 2000.
[7] G. Skillas, Z. Qian, U. Baltensperger, U.
Matter, and H. Burtscher,. Influence of
Additives on the Size Distribution and
Composition of Particles Produced By Diesel
Engines. Combustion Science and
Technology, 154,. 259–273, 2000.
[8] J. Lahaye, S. Boehm, P. H. Chambrion, and
P. Ehrburger, Influence Of Cerium Oxide On
The Formation And Oxidation Of Soot,
Combustion and Flame, 104,199–207, 1996.
[9] H. Jung, D. B. Kittelson, and M. R.
Zachariah, The Influence Of A Cerium
Additive On Ultrafine Diesel Particle
Emissions And Kinetics Of Oxidation.
Combustion and Flame, 142, 276–288, 2005.
[10] B. Stanmore, J. F. Brilhac, and P. Gilot, The
Ignition And Combustion Of Cerium Doped
Diesel Soot, SAE Technical Paper 01-0115,
1999.
[11] K. Pramanik, Properties and Use of Jatropha
Curcas Oil and Diesel Fuel Blends In
Compression Ignition Engine. Renewable
Energy 28, 239–248, 2003.
The 5th IMAT, November 12 – 13th
2012
54
Design a Four-Stroke Homogeneous Charge Compression
Ignition (HCCI) Engine
Mohd Rozi Mohd Perang
a, Zulkarnain Abdul Latiff
a,b, Azhar Abdul Aziz
a,b, Mohamad
Azzad Mokhrib
aAutomotive Development Centre, Universiti Teknologi Malaysia
81310 Skudai, Johor, Malaysia, Phone: +60 75535447, Fax: +60 75535811
email: [email protected], [email protected], [email protected] bFaculty of Mechanical Engineering, Universiti Teknologi Malaysia
81310 Skudai, Johor, Malaysia, Phone: +60 7 5534575, Fax: +60 7 5566159
ABSTRACT
This research is to study the operation of the four-
stroke HCCI engine. The design and analysis works
have been performed using computer software
which is GT-Power and Solidwork to study on the
engine performance simulation work and 3-D
modelling on the combustion chamber designed
respectively. The design is based on 4-cylinder
passenger car, 2000 cc and a four-stroke cycle
engine. The compression ratio used is 10. The fuel
used is ethanol in which the air-fuel ratio (AFR) is
9. The parameters selected have typical range of
value based on the previous study and research
done. With the use of GT-power, the analysis will
consider two parameters which are the cam timing
angle and the injection timing angle to get the
optimum result for the HCCI engine. The typical
angle of cam timing angle is between 2600 – 270
0
since this is the moment of the compression cycle
of the engine. For the injection timing angles, the
angles that will be studied for this project are 50, 0
0,
-50, -10
0,-15
0 and -20
0 relative to Top Dead Centre
(TDC). The objective is to obtain the maximum
torque and brake power when the engine speed is in
between 4000 rpm to 5000 rpm and 6000 rpm to
7000 rpm respectively. Finally, the optimum
conditions for the engine to perform better are at
2640 of cam timing angle for the valve and at -5
0
before TDC for the injection timing angle. The
maximum torque and brake power achieved is
37.60 Nm at 4000 rpm and 23.46 kW at 7000 rpm.
Keyword : Homogeneous Charge Compression
Ignition (HCCI); combustion
characteristics; ethanol fuel; gt-power
1. INTRODUCTION
The Internal Combustion Engine (ICE) is the
pioneer of the engine maker in the automotive
industry closely related to spark ignition (SI) and
compression ignition (CI) engine. Each of them
posses their own advantages and disadvantages.
Two-stroke engine is known for higher weight-to-
power ratio and four-stroke engine provide better
combustion and less pollution. The industry is
aimed to enhance the combustion behaviour of the
ICE by applying the HCCI engine mode that can
improve the engine operator. Thus, all big
automotive companies in Europe and Japan such as
General Motor (GM) and Toyota invest in a
research on HCCI engine that can be implemented
in car for future.
It is a challenge for the engineers nowadays to
develop an engine that have better performance
without neglecting the environmental issue. For the
current research technology, HCCI engine is
leading the new innovative technology in ICE.
HCCI is one of the ICE process as an alternative to
the conventional SI and CI engine combustion
process. The HCCI phenomenon is a combination
of the best features of both SI and CI engine which
is to produce an engine operating with very low
emission and high efficiency respectively. HCCI is
similar with SI in term of premixed charge usage
and also performs auto ignition to initiate the
combustion as practiced by CI [1]. In addition,
another advantage of HCCI engine is it can
consume any type of fuel. HCCI engine is the new
trend under extensive research nowadays. The
reason why HCCI engine is under the spotlight
today is the ability of the engine to satisfy better
performance without having to sacrifice the
environmental issues.
In the HCCI engine, the fuel and air should be
blended homogenously before the combustion
starts and that mixture will be auto-ignited due to
the increase in engine cylinder temperature and
pressure from compression stroke [2]. The
combustion process of HCCI is totally different
with SI and CI engines. HCCI combustion
capability is performed by controlling the
temperature, pressure, and composition of air and
fuel mixing to spontaneously ignite in the engine.
Since ignition occurs in HCCI engine is by auto-
ignition of the air/fuel mixture, the selection of fuel
will have an important impact on both engine
design and control system. All of the HCCI engine
concepts will be developed around the requirement
of making the fuel auto-ignition at the correct
timing to achieve good combustion phasing, and a
IMAT-UI 010
The 5th IMAT, November 12 – 13th
2012
55
rapid bulk burn to obtain high fuel economy and
minimize emissions.
In this study, alcohol based-fuel, ethanol was used
as the alternative fuel to operate the HCCI 4-stroke
engine mode. Currently, ethanol is one of the
renewable energy which widely used at Brazil and
United States [3]. Ethanol was created from the
natural agriculture feedstock. It is produced from
subsistent crops such as sugar cane, potato and corn
[4]. Besides that, ethanol can be produced through
cellulosic waste. This source offers promise
because cellulose fibres, a major component in
plant cells walls, can be used to produce ethanol in
large quantities.
The paper will focus on the design of combustion
chamber by identifying the cam and injection
timing angle in order to obtain the optimum result
for the HCCI engine. The research is based on the
simulation analysis via Gt-Power. The objective is
to obtain the maximum torque and brake power
when the engine speed is in between 4000 rpm to
5000 rpm and 6000 rpm to 7000 rpm respectively.
2. SIMULATION ANALYSIS
The simulation involved throughout this research is
GT-Power as to study on the engine performance.
The 3-D modeling of combustion chamber design
will be performed by CAD drawing, Solidwork.
Figure 1 shows the flow of the research activity.
Figure 1: Research activity flow process
Initially, some parameters of the combustion
chamber were calculated manually. Then, those
values were computed in the GT-Power, which was
modelled earlier according to the scope.
Subsequently, the software will solve the problem
and execute the analysis. The results were
presented in the form of tables and graphs.
Finally, the optimum condition of the design will
be decided based on the results by providing
relevant reasons and evidences. In addition, the
schematic diagram of ECU is also included since it
is very important and has significant value for this
research.
For manual calculation, it focused on the
parameters of the combustion chamber such as the
size of the bore, the length of the stroke and the
swept volume. Before that, some specifications
were established in order to limit the scope and
guide the analysis process. The specifications for
this study are based on the spark ignition engine,
passenger car and four stroke cycles.
Hence all the parameter can be summarized as the
following;
Table 1: Summary of the parameters for the engine
Parameters Symbol Value
Bore B 0.086m
Stroke S 0.086m
Swept volume Vd 4.996 x 10-4
m3
Total swept volume
(4 cylinder)
Vd,total 1.998 x 10-3
m3
Clearance volume Vc 5.55 x 10-5
m3
Compression ratio rc 10
Crack radius/offset a 0.043m
Connecting rod r 0.15m
r/a ratio R 3.5
Air-fuel ratio AFR 8.99
Engine cycle - 4-stroke
Crack angle ɵ 00-720
0
Intake valve - 1
Exhaust valve - 1
Fuelling system - IDI (indirect
injection)
Start of injection
angle
- 50, 0
0, -5
0, -10
0,
-150, -20
0
Cam timing angle - 2600 - 270
0
2.1 Method of using the GT Power
The method using GT-Power software were,
2.1.1 Run the software. Make sure that the interface
of the program as shown in Figure 2. The format of
the file is *.gtm. The red circle is the Template
Library and the blue circle is the Project Area. The
Template Library contains all the available
templates that can be used in GT-Power
applications whilst the Project Area model can be
created by our virtual engine according to our
design.
No
No
Yes
Yes
Literature review
Parameter setting
(using Gt-Power)
Satisfy?
3D Modeling using
Solidword
Satisfy?
Final engine design
The 5th IMAT, November 12 – 13th
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56
Figure 2: The interface of GT Power for modelling
the designed engine
2.1.2 Some of the templates in the Template
Library will be used in this project. Drag the
template into the Project Area according to the
project requirement. The templates that needed are
as shown in Table 2.
Table 2: Templates used for the project
Flow Folder
EndEnvironment
EngCylinder
Pipe
InjAF-RatioConn
InjProfileConn
OrificeConn – def(object)
ValveConn
Ethanol – Vapc (object)
N2 – vapc (object)
O2 – vapc (object)
Air (object)
Mech Folder
EngineCrankTrain
2.1.3 Next, fill in the template parameter with
appropriate value according to the project
objectives. The default value can be applied to
some of the parameters as it is compatible for all
engine designs. However, some parameters such as
size of the bore, length of the stroke, compression
ratio and others as shown in Table 2 are required to
be filled in the templates. Those parameters need to
be computed accordingly to the manual calculation
earlier from the design. Table 2 exhibits the value
that needed to define according to the design. The
start of the injection angle and cam timing angle
are manipulated parameters which will be adjusted
in order to get the optimum result.
2.1.4 Then components can be arranged according
to its sequence. The procedure is to drag the
template into the Project Area. Automatically it
will convert the value into an object. The sequence
is as illustrated in Figure 3.
Environment – intake runner – intake port – intake
valve – cylinder – IDI injection (above cylinder) –
exhaust valve – exhaust port – exhaust runner –
environment
Figure 3: The sequence of template in the project
Area
Next, make the connection between the objects.
There is a button named Create Link at the toolbar
which can do the linkage (Figure 4).
Figure 4: The location of the icons and its
purpose
2.1.5 Finally, the schematic diagram as Figure 5
will be obtained;
Figure 5: The schematic diagram of the
designed engine in the GT Power
2.1.6 Then, the study case for this project can be
set up and Run/Case Setup was selected to run the
simulation program. The engine speed varies from
0 rpm to 10 000 rpm. Next, the analysis can be
started by clicking on Run/Start Simulation. The
result can be viewed in GT-Post. This software can
show the data in the form of tables and also graphs.
Create Link
The 5th IMAT, November 12 – 13th
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57
3. ANALYSIS SETUP
Two important parameters will be manipulated in
this project to generate a new parameter (Table 1)
that will be fixed in order to obtain the optimum
result for the engine. The first parameter is the cam
timing angle and the second is the injection timing
angle.
The cam timing angle is the angle when the intake
valve will be lifted to permit fresh air get into the
combustion chamber. The typical angle is in
between 260⁰ - 270⁰ since this is the moment of the
compression cycle of the engine.
The injection timing angle is the moment of the
crank angle when the injector will spray the fuel
into the combustion chamber. The fuel and air are
assumed to have enough time to mix
homogeneously inside the combustion chamber
before the charge combust to produce power. The
injection timing angles that will be studied for this
project are 50, 0
0, -5
0, -10
0, -15
0 and 200
relative to
Top Dead Centre (TDC).
The objective of this work is to get the highest
torque when the engine speed runs between 4000
rpm to 5000 rpm. This is because it will help the
vehicle to accelerate faster for the short period at
the earlier time, for instance to overtake another
vehicle in front. In addition, the maximum brake
power desires to be achieved during the engine
speed operates between 6000 rpm to 7000 rpm. The
reason is the vehicle will reach the maximum speed
when the load is constant with stable driving
condition. As reference, some important data such
as maximum brake power and maximum torque for
several types of engines are collected in order to
compare and validate our result. Table 3 shows the
data.
Table 3: The maximum brake power and torque for
several engines
Parameter
Engine
4B11
(Mitsubi
shi
Lancer)
N74
(Ford)
N34
(Ford)
R4 FSI
(Volkswa
gen Golf)
Capacity (cc) 1998 1998 1998 1984
Bore (mm) 86.0 86.0 86.0 82.5
Stroke (mm) 86.0 86.0 86.0 92.8
Compression
ratio 10:1 10.3:1 9.8:1 10.5:1
Maximum
power (kW) 114 @
6000 pm
110
@630
0 rpm
101 @
6300
rpm
147 @
6600 pm
Maximum
Torque (Nm)
198 @
4250
rpm
190 @
4500
rpm
175 @
4200
rpm
280 @
4700 rpm
4. RESULTS AND DISCUSSION
Table 4 exhibits the result on the cam timing angle
analysis. It shows that the maximum torque and
power decreased when the engine speed and cam
angle increased from 2500 - 5000 rpm and 2600 -
2700
respectively. However the engine speed values
for all maximum brake power are not in the
targeted range, which is between 6000 rpm and
7000 rpm. The targeted engine speed for maximum
torque is in between 4000 - 5000 rpm. Finally the
cam timing angle selected is 2640 since the angle
produced the highest maximum torque at 4000 rpm,
which is within the targeted range. Table 4: The result of various cam timing angles
Cam
timing
angle
(o)
Maximum torque
(Nm) @ Engine
speed (rpm)
Maximum brake
power (kW) @
Engine speed (rpm)
260 41.52 @ 2500 24.38 @ 8000
261 40.81 @ 3000 24.31@ 8500
262 40.35@ 3000 24.23 @ 8500
263 39.26 @ 3000 24.17 @ 8500
264 37.60 @ 4000 24.10 @ 8500
265 36.81 @ 4500 24.03 @ 8500
266 36.39 @ 4500 23.96 @ 8500
267 35.93 @ 4500 23.88 @ 8500
268 35.45 @ 4500 23.81 @ 9000
269 34.76 @ 5000 23.72 @ 9000
270 34.32 @ 5000 23.64 @ 9000
Table 5 shows the data for the injection timing
angle. Obviously the maximum torque and
maximum brake power are increased as the fuel
was injected earlier before reaching the Top Dead
Centre. There is no problem for the maximum
torque produced by each injection timing angle
since the engine speed is within the targeted range
except for 5⁰. However, there is a problem to
define the optimum condition since the maximum
brake power at required engine speed for all
injection timing angle are not in targeted range
Thus, more details analysis should be done on it.
Table 5: The result of various injection timing angles
Injection
timing
angle (⁰)
Maximum torque
(Nm) @ Engine
speed (rpm)
Maximum brake
power (kW) @
Engine speed (rpm)
5 33.11 @ 3000 19.61 @ 7500
0 35.46 @ 4000 22.01 @ 8000
-5 37.60 @ 4000 24.10 @ 8500
-10 39.29 @ 4000 26.07@ 9500
-15 40.38 @ 4000 27.57 @ 9500
-20 40.82 @ 4000 28.43 @ 9500
The 5th IMAT, November 12 – 13th
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Table 6 exhibits the detail result of brake power for
various injection timing angles(ITA).
ITA (o)
Max
BP
(kW)
@
rpm
BP (kW) @ rpm Differe
nce
(kW)
6000 6500 7000
0
22.01
@
8000
20.67 21.70 21.87 0.14
-5
24.10
@
8500
22.14 22.97 23.46 0.64
-10
26.07
@
9500
23.33 24.38 24.89 1.18
-15
27.57
@
9500
24.17 25.22 25.92 1.65
-20
28.43
@
9500
24.60 25.71 26.48 1.95
The difference between maximum brake power at
8000 – 9500 rpm and brake power at 7000 rpm
were increased as the injection timing angle gets
earlier. The angle of 00 and -5
0 will be considered
since the difference is lower and can be accepted.
On the other hand, the results for the angle from -
100 to -20
0 need to be eliminated from the
consideration because the difference is too big,
which is more than 1kW. Finally, between 00 and -
50, the injection timing angle of -5
0 is chosen as the
best angle since it produces more brake power
compared to 00.
As the conclusion, the optimum condition for the
engine operation to perform the best output is at
2640 of cam timing angle and at -5
0 of injection
timing angle.
4.1 Brake Torque Result
Figure 6 illustrates the graph of brake torque
against engine speed. The maximum brake torque
occurred at 37.6 N.m when the engine speed runs at
4000 rpm. This is for one cylinder operation. For
four cylinder engine, the total brake torque is 150.4
N.m. The value is not too far compare to engine
N34 (175 N.m) since it is a basic engine. However,
the difference with other engines such as
Mitsubishi Lancer (198 N.m) and Golf
Volkswagon (280 N.m), it is too far as they are
high performance engine.
Figure 6: Graph of brake torque against engine
speed.
4.2 Brake Power Result
Figure 7 shows the graph of brake power of the
engine during the optimum condition for various
engine speeds. The observation can be made that
the brake power achieved the maximum around
8000 rpm and sustain until 10000 rpm. From the
previous analysis, the maximum brake power of the
engine is 23.46 kW at 7000 rpm. Thus, the power
produced by the whole engine, which consist of
four cylinders is 93.84kW. The value is acceptable
since this is just a basic engine and does not equip
with other features such as turbocharged (Golf
Volkswagen) and MIVEC (Mitsubishi Lancer).
Figure 7: Graph of brake power against engine
speed (at optimum condition)
4.3 The Indicated Specific Fuel Consumption
(ISFC) Result
Figure 8 exhibits the graph of ISFC against engine
speed. Typically, a car consumes less fuel during
cruising condition (4000 to 6000 rpm). This is
because the vehicle moves under steady-state
condition.
The slope of the curve decreased and then starting
to increase from 6250 rpm to 10000 rpm. The
minimum ISFC that can be achieved for the engine
is 0.51 kg/kWh at 5500 rpm. Initially, the ISFC is
higher in order to overcome the inertia effect of the
vehicle. Besides that, it is not recommended to
drive a vehicle with high rpm as it is consuming
more fuel and not economic.
The 5th IMAT, November 12 – 13th
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59
Figure 8: Graph of ISFC against engine speed
(at the optimum condition).
4.4 The P-V Diagram
P-V diagram is the curve to show the relationship
between pressure and volume in the engine
cylinder [5]. Basically, the diagram consists of two
loops, which are upper and lower loop [5]. The
upper loop represents the power loop produced
during compression and expansion cycle.
Meanwhile, the below loop is the pumping loop,
which come from the intake and exhaust cycle.
Figure 9 shows the net power produced by
subtracting the power loop to pumping loop. From
the P-V diagram of the engine, the net power
produced is positive, since the area of power loop is
bigger than pumping loop. The positive net power
shows that the work is transfer out from the
cylinder to the crank shaft. The general idea is to
get higher net power by having bigger power loop
and minimized the pumping loop.
Figure 8: P-V diagram of the engine (at
optimum condition)
4.5 The P-Ө Diagram
The P-Ө diagram shows the relationship between
the pressure rises in the engine cylinder relative to
the crank angle [5]. It is very significant because
the data provide the highest pressure exerted on the
piston at specific crank angle. Figure 9 shows the
maximum pressure at momentary crank angle.
Figure 9: P-Ө diagram of the engine (at the
optimum condition)
In the engine, the piston is among the most critical
part that needs to be focused. For this engine, the
highest pressure exerted on the piston is 6 MPa at
170. Most of the pressure rises during the
compression and power cycle. The peak occurred
after the combustion take place.
4.6 Cycle Event
Cycle event is another important diagram for the
engine cycle that gives significant value. It
represents the process position at specific crank
angle for the engine such as the moment of valve
lift during the intake and exhaust cycle, injection
timing and etc [5]. Figure 10 illustrates the moment
for the intake valve, exhaust valve and fuel injector
to operate.
Figure 10: the event cycle diagram of the
engine
The intake valve will start open at 410 after TDC
and close at -460
before TDC in the first cycle. At
this moment, the fresh air will be induced into the
combustion chamber. Then, at -50 before TDC, the
fuel injector will start to spray the fuel. Lastly, the
exhaust valve will start open at 1250 after TDC and
close at 3980 at the next cycle.
4.7 ECU
Electronic Control Unit (ECU) is an integrated part
in the engine consists of sensors, microcontroller
The 5th IMAT, November 12 – 13th
2012
60
and actuators in order to assist the engine to
perform better. The basic idea is a sensor will
detect particular characteristic and send a signal to
the microcontroller. The microcontroller will give a
response according to the programmed set up
earlier. At the end, the signal will send to the
actuator to execute.
Figure 11 shows the map of the ECU. The first
sensor is the air-temperature sensor. This
component is installed in the air-intake track and
sense the air intake temperature. If the temperature
is not hot enough for the charge combust
efficiently, the EGR will be executed [6]. Thus, it
will increase the incoming air.
Figure 11: The map of the ECU of the engine
The second sensor consists of several components,
which is Manifold Absolute Pressure (MAP),
Throttle Position Sensor (TPS) and crack angle
encoder. These components will cooperate with
each other to determine the amount of fuel supplied
and injection timing [6]. So, it will give accurate
fuel needed and injection timing according to the
condition of the vehicle.
If a driver desire to accelerate, he/she will step
more on the acceleration pedal. Then, the butterfly
valve of the throttle plate (gasoline engine) and
injection pump (diesel engine) will operate since its
influence the acceleration of the vehicle.
Conventionally, the connection between pedal and
that particular component is by mechanical system,
which is Bowden cable or linkage.
5. CONCLUSIONS Finally, the optimum conditions for the engine to
perform better are at 2640
of cam timing angle for
the valve and at -50 before TDC for the injection
timing angle. The maximum torque and brake
power achieved is 37.60 Nm at 4000 rpm and 23.46
kW at 7000 rpm.
REFERENCES
[1] Mingfa Yao, Zhaolei Zheng, Haifeng Liu.,
―Progress and Recent Trends in Homogeneous
Charge Compression Ignition (HCCI)
Engines‖. Energy and combustion science 35,
398-437.
[2] T. Aoyama, et. al, ―An Experimental Study on
Premixed-Charge Compression Ignition
Gasoline Engine‖, Society of Automotive
Engineering, SAE 960081, 1996.
[3] Kohtaro Hashimoto, ―Effect of Ethanol on
HCCI Combustion‖, SAE of Japan, JSAE
20077106, 2007.
[4] Lu Xingcai, et. al, ―Experimental Study on the
Auto-Ignition and Combustion Characteristics
in the Homogeneous Charge Compression
Ignition (HCCI) Combustion Operation with
Ethanol/n-heptanes Blend Fuels by Port
Injection‖, Fuel 852622-2631, 2006.
[5] Heywood, John B., ―Internal Combustion
Engine Fundamentals‖, Massanchusetts : Mc-
Graw Hill, 1998.
[6] Hirschlieb, et. al., ―Engine Control. In Ronald
K. Jurgen. (ed) Automotive Electronics
Handbook‖. New York : McGraw Hill, 1999.
[7] Yunus A. Cengel, Michael A. Boles.(2007).
Thermodynamics : An Engineering Approach
Sixth Edition (SI Unit). s.l. : McGraw-Hill
Higher Publication.
[8] Mack, J. Hunter , Salvador M. Aceves, Robert
W. Dibble., ―Demonstrating Direct Use of Wet
Ethanol in a Homogeneous Charge
Compression Ignition (HCCI) Engine‖.
Energy 34, 782-787, 2008.
The 5th IMAT, November 12 – 13th
2012
61
R22 and Various Mixtures of R290/R600a as its Alternative in
Adiabatic capillary tube Used in split-type Air-conditioning
System
Shodiya Sulaimona,b
, Azhar Abdul Aziza, Henry Nasution
a,c, Amer Nordin Darus
a
aAutomotive Development Centre (ADC), Faculty of Mechanical Engineering
Universiti Teknologi Malaysia (UTM),Skudai, Johor, Malaysia.
Email:[email protected], [email protected], [email protected] bDepartment of Mechanical Engineering, Faculty of Engineering
University of Maiduguri (UNIMAID), Maiduguri, Borno, Nigeria. cDepartment of Mechanical Engineering
Bung Hatta University
ABSTRACT
Conventionally, (hydro chlorofluorocarbon)
HCFC22 is used as a working fluid in small vapor
compression refrigeration system with capillary
tube as expansion device. According to Montreal
Protocol, HCFC22 must be phased out owing to its
high ozone depleting potential (ODP). Several
natural substances including ammonia, carbon
dioxide, water and hydrocarbon (HC) such as
propane (HC290) and iso-butane (HC600a) and
their mixtures have immerged as close substitute.
Literature showed that pure HC refrigerant may not
be suitable enough because of the difference in
operating pressure and volumetric cooling capacity
when compared with HCFC22. The main objective
of this study is to theoretically investigate different
ratios of HC refrigerants HC290/HC600a mixtures
flowing through adiabatic capillary tube using
homogenous model. In this study, the percentage of
HC600a was varied from 0 to50 % in a step of 5%.
The pressure at the two extreme ends and
temperature along the capillary tube, using
HCFC22 refrigerant, which was used as
benchmark, was experimentally determined in the
air-conditioning (AC) system. Comparing the
model results with the experimental data showed
that HC refrigerants HC290/HC600a in ratio
70%/30% gave 2.65% minimum error and thus it
can be used as a substitute to HCFC22 in the split-
type AC system.
Keywords:Capillary tube, Split-type air
conditioner, HCFC22, HC290/HC600a,
Homogenous model
1. INTRODUCTION As a result of the increase in public concern about
the depletion of ozone layer and global warming of
some refrigerants such as chlorofluorocarbon
(CFC)‘s and HCFC‘s, a lot of refrigerants testing is
going on in search of the best alternative for these
non-eco-friendly refrigerants. For decades now, the
hydro fluorocarbon (HFC) refrigerants have been
chosen as an alternative to HCFC‘s and CFC‘s
because of their zero ODP despite their high global
warming potential (GWP). In addition, some of the
advantages associated with the HFC refrigerants‘
include the vapor pressure similarity with the
CFC‘s and HCFC‘s, their stability and non-
flammability. However, the problem of the GWP of
the HFC refrigerants has forced the scientists to
look for a more eco-friendly refrigerants such as
ammonia, carbon dioxide, water and hydrocarbons
(HC) like propane, butane, iso-butane and their
mixtures. The HC refrigerants are preferred
because of their advantages such as their zero ODP,
low GWP, non-toxic, highly miscible with mineral
oil and their higher performance when compared
with the CFC‘s and HCFC‘s refrigerants. As a
result, many refrigeration and air-conditioning
systems are using HC refrigerants and safety
precautions are made concerning the leakage of the
refrigerant from the system.
All investigations conducted using HC refrigerants
mixtures showed that these refrigerants have higher
coefficient of performance (COP) and energy
efficiency compared with the CFC‘s and HCFC‘s.
Wongwises et al. [1] investigated the behavior of
HC refrigerant mixtures of HC290, HC600 and
HC600a in automotive air-conditioning system
which was formerly using refrigerant R134a. They
concluded that all the HC mixtures yielded the
COP higher than R134a. Nasution, H [2]
conducted an experiment using a mixture of HC
refrigerants HC290/HC600/HC600a on split-type
air-conditioning system. The AC system is
formerly designed to use HCFC22 as a working
fluid. The result showed that the COP of the HC
mixture is higher than the HCFC22 by 9% and
saving of energy consumption of about 16%. Akash
and Said [3] replaced R12 refrigerant with liquefied
petroleum gas (LPG) composing of 30% R290,
55% R600 and 15% R600a in a household
refrigerator. Their result showed that the LPG
refrigerant gives the best performance compared
with R12. Likewise, Jung et al. [4] replaced R12
IMAT-UI 011
The 5th IMAT, November 12 – 13th
2012
62
with mixture of R290 and R600a having mass
fraction of 60% and 40% respectively. They
reported that the energy efficiency and the COP
increase by 4% and 2.3% respectively.
The expansion device used in a split-type air-
conditioner is capillary tube. This expansion device
is the heart of the system [5] because the proper
design of the tube will improve the performance of
the system and thus the behavior of the alternative
refrigerants in capillary tube needs special
attention. Many investigation have been reported,
both experimental [6-8] and theoretical [9-12] on
the characteristic flow of alternative refrigerants in
capillary tube. Undoubtedly, most of the
experimental and theoretical studies on capillary
tube used hydro fluorocarbon (HFC) like R407C
and R410A as alternative to R22 and HC
refrigerants are rarely used. As a result of the
excellent properties of HC mentioned above, the
world attention is now highly concentrated on HC
as a refrigerant.
Literature showed that pure HC refrigerant may not
be suitable as substitute to the non-eco-friendly
refrigerants due to different in operating pressure
and volumetric cooling capacity between the
HCFC‘s and the pure HC [13]. Pure HC290 has the
highest saturation pressure and pure HC600a has
the lowest saturation pressure as shown in Figure 1.
Though, the saturated pressure of pure refrigerant
HC290 is closer to refrigerant HCFC22 (Figure 1),
however, when considering their specific heat
capacity (cooling capacity), it can be seen that they
are farther apart as shown in Figure 2. Therefore,
appropriate composition mixture of HC290 and
HC600a will be required to substitute HCFC22 in
the split-type air-conditioner. The objective of this
study is therefore, to theoretically explore the flow
behavior of different composition mixtures of
HC290 and HC600a in the capillary tube using an
improved homogenous model which has not been
used by other researchers and determine a possible
composition that can replace HCFC22 refrigerant
in a split-type AC system.
-30 -20 -10 0 10 20 30 40 50 60 70
0
0.5
1
1.5
2
2.5
3
Temperature (oC)
Sa
tura
tio
n P
re
ssu
re
(M
pa
)
HC600a
HCFC22
HC290
Figure 1: Saturation pressure against
Temperature
-30 -20 -10 0 10 20 30 40 50 60 70
0.8
1
1.2
1.4
1.6
1.8
2
Tempearture (oC)
Sp
ec
ific h
eat c
apa
city (k
j/kg
-k
)
HC600a
HCFC22
HC290
Figure 2: Specific heat capacity against
Temperature
2. MATERIAL AND METHODS
2.1 Experimental work
The main aim of the experimental work is to collect
experimental data from the capillary tube in the
commercially purchased split-type air-conditioning
system which is used to compare the theoretical
model results. The air-conditioner worked with
HCFC22 as working fluid having cooling capacity
of 2.47kW. Flow meter is installed on the system to
measure the refrigerants flow rate with an accuracy
of ±0.1
g/s. Likewise, the temperature of the
refrigerant are measured at equal interval along the
capillary tube by connecting eight temperature
sensors (K-type thermocouple with an accuracy of
±0.1oC) to a computer through a data logger. The
pressure of the refrigerant is measured at the two
extreme ends of the capillary tube using pressure
transducer with an accuracy of ±0.1Mpa. The
capillary tube of length 1.30 m and diameter
1.327mm which is from the manufacturer of the
AC was used.
To collect the experimental data, the refrigerant
HCFC22 was charged into the system after the air
in the system has been evacuated using vacuum
pump. When the system is on, it takes about 30 to
45mins before the system could attain steady state.
The pressure of refrigerant at the two extreme ends
of the capillary tube and temperature along the
capillary tube is displayed on the computer. The
experiment is repeated five times and the average
value is taken.
2.2 Model Description
The detailed description of the two-phase
homogenous model is presented in [14]. As shown
in Figure 3, the capillary tube is connected between
the condenser and evaporator. The flow in the
capillary tube is modeled by dividing the flow into
subcooled liquid single phase, metastable phase
The 5th IMAT, November 12 – 13th
2012
63
and two-phase liquid-vapor regions. The flow in
the capillary tube is based on one dimensional
homogenous two-phase flow assumptions. The
model is also based on the fundamental equations
of mass, momentum and energy and that of
Wongwises et al. [10] with modifications.
1 2 3 4 5Condenser Evaporator
Subcooled liquid
region
Metastable
region
Two-phase
region
Capillary tube
Figure 3: Adiabatic capillary tube with three
regions
2.2.1 Single-phase flow region
The liquid single-phase length is given
in eq. (1)
Where, p1 and p3 are pressures at points 1 and 3
respectively, G is mass flow per unit area , D is
inner diameter of capillary tube, k is coefficient of
entrance loss and its value is 1.5 as given by Zhou
and Zhang, [15]. Single-phase friction factor is
calculated from Colebrook and Churchill formula,
given in eq. (2) and (3) respectively.
Where
and are wall roughness and refrigerant viscosity
respectively
2.2.2 Metastable flow region
The inception of vaporization does not take place at
the real saturated vapor pressure Ps, but actually
takes place at a later point Pv, downstream from
thermodynamic saturated position. This process is
called metastability. Chen et al. [16], quantified this
metastable flow and propose a correlation, given in
eq. (5), to predict this delay in vaporization.
Where =
is level of subcooling, is critical
temperature, are pressure at points 3 and
4 respectively and is temperature at point 3.
and are specific volumes of liquid and
liquid/vapor phases respectively.
The length of metastable is given by
Where metastable friction factor, was
evaluated using Colebrook formulation as shown in
eq. (2).
2.2.3 Two-phase region
The differential equation of the two-phase length of
the capillary tube is given in eq. (7)
The two-phase friction factor, can be
conveniently evaluated using Bittle and Pates [17]
correlation given in eq. (8)
Where
The viscosity models used in calculating the two-
phase viscosity are given by Cicchitti et al.[18],
Dukler et al.[19], McAdam et al. [20] and Lin et al.
respectively as follow:
To evaluate the elemental two-phase length of the
capillary tube, eq. (7) is discretized given in eq.
(14)
Summation of the elemental lengths results in the
two-phase length given in eq. (15)
Thus, the total length of the capillary tube, L, can
be written as
All thermophysical and thermodynamics properties
are taken from the REFPROP [21] computer
program, version 8 (2007) which are developed in
the function of pressure.
The 5th IMAT, November 12 – 13th
2012
64
3. RESULTS AND DISCUSSION
In this section, model verification is made and
using the model, various ratios of the HC mixtures
are compared with the standard refrigerant,
HCFC22. The model operating parameters and
their ranges are as follows: pressure inlet, 0.4 – 2.0
Mpa, mass flow rate, 0.005 – 0.018 g/s, capillary
tube diameter, 0.20 – 1.8 mm, degree of
subcooling, 1.0 – 12.0oC.
The present model results gave a good agreement
when compared with the experimental data of
Fiorelli et al. [22] with respect to the shape of
pressure distribution as shown in Figure 4.
Although, Colebrook‘s and Churchill‘s
formulations were used to determine the single-
phase friction factor and the three viscosity models
mentioned in this study was also used in the
simulation, however, the Colebrook and Dukler et
al. correlations‘ combination gave the best
prediction with an average error of 1.75%.
0 0.5 1 1.50
0.2
0.4
0.6
0.8
1
1.2
1.4
1.6
1.8
2x 10
6
D i s t a n c e f r o m C a p i l l a r y I n l e t ( m )
C a
p i l l a
r y
T
u b
e P
r e
s s
u r
e (
P a
)
N u m e r i c a l r e s u l t
E x p e r i m e n t a l d a t a o f F i o r e l l i e t a l .
Figure 4: Comparison of the numerical results
with experimental data of Fiorelli et
al. [22].
3.1 Comparison of Different ratios of the
HC with the Model The developed model was used to simulate
different ratios of HC290 and HC600a. The
percentage of HC600a was varied from 0 to 50% in
a step of 5% and comparison with experimental
data using refrigerant HCFC22, the benchmark
refrigerant, is reported. From the results obtained,
the best composition lies between 75% propane,
25% iso-butane (HC1), 70% propane, 30%
isobutene (HC2) and 65% propane, 35% iso-butane
(HC1).
The sample of these HC mixtures and the
experimental data with respect to pressure
distribution along the capillary tube, up to choked
condition, is shown in Figure 5. From the Figure, it
can be seen that the rate of decrease in pressure of
HC1 is higher than HC2 and HC3. This could be
attributed to the fact that the specific volume and
velocity of HC1 is higher as a result of lesser
amount of HC600a present in HC1 compared with
HC2 and HC3. Consequently, the mach number of
HC1 is increased as a result of its high velocity,
leading to its earlier choke flow compared with
HC2 and HC3. From the Figure, refrigerant HC1,
HC2 and HC3 at pressure 2.2, 2.5 and 2.8 bar have
their choked flow at length 1.30, 1.32 and 1.34 m in
the capillary tube respectively. Therefore, pressure
drop per unit length for HC1, HC2 and HC3 are
1.62, 1.92 and 2.10 bar/m respectively for a given
mass flow rate and capillary tube inner diameter.
Thus, HC2 refrigerant is the best composition,
having pressure drop per unit length of 1.92 and the
minimum error of 2.78% when compared with the
benchmark refrigerant, HCFC22.
0 0.2 0.4 0.6 0.8 1 1.2 1.4
0
2
4
6
8
10
12
14
16
18
x 10
5
Pin = 1.636 Mpa
Tsub. = 4.5oC
m = 25g/sec.
D = 1.752 mm
E=2.592*10-4
D i s t a n c e f r o m C a p i l l a r y I n l e t ( m )
C a
p i l l a
r y T
u
b e
P
r e s s u r e
( P
a
)
HC1
Experimental data
HC2
HC3
Figure 5: Comparison of various HC mixtures
with the experimental data
The variation of mass flow rate of HC1, HC2, HC3
and HCFC22 with respect to pressure inlet for
diameter 2.00 mm and inlet subcooling of 10.5oC
under choked flow condition is shown in Figure 6.
It can be seen from the Figure that HC1 has the
lowest mass flow rate for the flow condition and
HC3 has the highest mass flow rate. However, HC2
and HCFC22 have almost the same mass flow rate
confirming that HC2 can be substituted for
HCFC22.
4.92 4.94 4.96 4.98 5 5.02 5.04 5.06 5.08 5.1 5.12
x 10
5
0.009
0.01
0.011
0.012
0.013
0.014
0.015
0.016
0.017
L = 1 . 3 0 m
T s u b . = 1 0 . 5 o C
D = 2 . 0 0 mm
E = 2 . 5 9 2 * 1 0 - 4 m m
P r e s s u r e I n l e t ( P a )
M a
s s f l o
w
r a
t e
( K
g
/ s )
HC1
HC2
HC3
HCFC22
Figure 6: Comparison of refrigerant mass flow
rate of HC1, HC2, HC3 and
HCFC22 with pressure inlet at inlet
subcooling 10.5oC and diameter 2.00
mm
The comparison of mass flow rate of HC1, HC2,
HC3 and HCFC22 with respect to pressure inlet for
diameter 1.742 mm and inlet subcooling of 4.5oC
under choked flow condition is shown in Figure 7.
The 5th IMAT, November 12 – 13th
2012
65
Figure 6 and 7 are similar in respect of their graph
profile, however, the respective mass flow rates of
Figure 6 are higher than Figure 7. This can be
attributed to the increase in their subcooling and
diameter of capillary tube. A higher subcooling will
have more refrigerant in liquid phase in the
capillary tube and liquid phase refrigerant offers
lesser resistance to flow. Figure 7 also reveals that
the mass flow rate ratio of HCFC22 to HC1 is
about 0.51, with HC2 is about 0.92 and with HC3
is about 0.49.
4.92 4.94 4.96 4.98 5 5.02 5.04 5.06 5.08 5.1 5.12
x 10
5
0.006
0.008
0.01
0.012
0.014
0.016
0.018
P r e s s u r e I n l e t ( P a )
M a
s s f l o
w
r a
t e
( K
g
/ s )
L = 1 . 3 0 m
T s u b . = 4 . 5 o C
D = 1.752 mm
E=2.592*10-4
HC1
HC2
HC3
HCFC22
Figure 7: Comparison of refrigerant mass flow
rate of HC1, HC2, HC3 and
HCFC22 with pressure inlet at inlet
subcooling 4.5oC and diameter 1.742
mm
4. CONCLUSION
In this present work, a theoritical model that was
developed was used to explore the behavior of HC
HC290 and HC600a refrigerant mixtures, in order
to determine the possible HC composition ratio that
can replace HCFC22 refrigerant in the split-type air
conditioner. Different ratios of HC HC290 and
HC600a was simulated and the results compared
with the experimentally determined data using
refrigerant HCFC22, the bench mark. The result
showed that HC2 (70% HC290 and HC600a 30%)
refrigerant is the best composition, having pressure
drop per unit length of 1.92 and the minimum error
of 2.78% when compared with refrigerant
HCFC22. The mass flow rate ratio of HCFC22
with HC2 is about 0.92 signifying their closeness.
With the result obtained, it can be concluded that
HC2 refrigerant is a close substitute to HCFC22 in
the split-type air conditioner.
ACKNOWLEDGMENT
The present study was financially assisted by
research university grant (RUG) program: Tier 2
Cost Centre Code: QJ130000.7124.01J85,
Universiti Teknologi, Malaysia. The financial
support of Educational Trust Fund (ETF), Nigeria
is also acknowledged.
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2012
67
Review paper on:
Flow Pattern at Pipe Bends on Corrosion Behaviour of Low
Carbon Steel Muhammadu Masin Muhammadu
1 , Jamaluddin Muhammad Sheriff
2 and Esah Binti Hamzah
3
1Department of Mechanical Engineering, Federal University of Technology, Minna
2,3Faculty of Mechanical Engineering, Universiti Technologi, Malaysia
[email protected], [email protected] and [email protected]
ABSTRACT
Most importantly the identification of positions or
sites, within the internal surface contact areas
where the maximum corrosion stimulus may be
expected to occur, thereby allowing better
understanding, mitigation, monitoring and
corrosion control over the life cycle. Some case
histories have been reviewed in this context, and
the interaction between corrosion mechanisms and
flow patterns closely determined, and in some cases
correlated . Since the actual relationships are
complex, it was determined that a risk based
decision making process using selected ‗what‘ if
corrosion analyses linked to ‗what if‘ flow
assurance analyses was the best way forward.
Using this in methodology, and pertinent field data
exchange, it is postulated that significant
improvements in corrosion prediction can be made.
This paper outlines the approach used and shows
how related corrosion modelling software data such
as that available from corrosion models Norsok
M5006, and Cassandra to parallel computational
flow modelling in a targeted manner can generate
very noteworthy results, and considerably more
viable trends for corrosion control guidance. It is
postulated that the normally associated lack of
agreement between corrosion modelling and field
experience, is more likely due to inadequate
consideration of corrosion stimulating flow regime
data, rather than limitations of the corrosion
modelling.
Keywords: ALARP (As Low As Reasonably
Practicable) Co2 corrosion,
corrosion resistant alloy (CRA),
decision gates, erosion-corrosion,
life cycle performance, risk basis
1. INTRODUCTION
The situation of flow separation, for example inside
a sudden expansion in the pipe, turbulence is
moved downstream from the objective of
separation (1). There is no simple relation
involving the bulk flow parameters as well as the
local near-wall, hydrodynamic, mass transfer and
erosion-corrosion conditions as well as the latter
ought to be determined either experimentally [1,8,
13] or by record simulation both by [1,8,13]. This
paper focuses the bond between flow pattern and
corrosion behaviour at pipe bends as well as the
advances in using turbulence models for the record
simulation of erosion-corrosion throughout
yesteryear decade.
The development of water pipeline infrastructure in
a few part of World remains rapid growth within
the last few years. If you have been new projects
plus much more tie backs into existing systems.
This helps it be crucial that you acknowledge,
identify, and develop, the critical associations
between corrosion conjecture and flow systems at
pipe bends. Most substantially the identification of
web sites where maximum corrosion stimulus can
be expected thus enabling better understanding,
minimization, monitoring and corrosion remedies
for the whole existence cycle. Some situation
histories are actually examined in this particular
context, as well as the interaction between
corrosion systems and flow designs carefully
determined for a number of design campaigns. The
specific associations are complex, too for basic
reasons some risk based making choices procedure
using ―what if‖ corrosion analyses connected with
―what if‖ flow assurance analyses was considered
our advice [15-23].
The experience based on formerly looked into work
by [24] and [25] has generated an instantaneous
final results of corrosion rate of low carbon steel
loss and fluid flow designs. A modification of flow
pattern can lead to significant improvements in
tube existence by remaining from impinging flow.
The mechanism in the internal corrosion rates aren't
fully understood but appears being linked strongly
towards the particulate matter inside the flow [24-
29]. Even if pollutants are simply inside the
micrometre size range, they could still deviate from
fluid streamlines to make sure that glancing flow
can lead to abrasion in the protective surface layer
[31-37]. Rapid positioned on by corrosion-erosion
will result when the layer is soft in compliance
while using particulate matter.
By using this methodology, and area data
exchange, it's thought significant enhancements in
IMAT-UI 012
The 5th IMAT, November 12 – 13th
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69
corrosion behaviour are achievable. This paper
outlines the approach used and shows how relevant
corrosion modelling software data for example that
supplied by corrosion models available Norsok
M5006, and Cassandra Software [15,30] to parallel
computational flow modelling within the specific
manner can generate very significant results, and
sometimes unforeseen trends for corrosion control
guidance. The methodology offers good semi-
quantitative arguments for justification and a
variety of references are really used.
Having considered the basic characteristics of flow
in curved open channels are fairly understood,
theoretically and numerically the configuration of
the side wall has significant effects on the flow
structures [29].
Since the streamlines ought to be curved inside the
same sense since the pipe itself there is a radial
pressure gradient, so pressure is bigger round the
outer wall in the pipe compared to the related point
round the inner wall.
The mechanism which supplies birth for the
stagnation pressure gradient is not considered
incorporated within the secondary flow
phenomena, although there's evidence of mutual
interaction involving the two inside the turning
passage to ensure that as result causes turbulent
which in turns involve some unwanted effects
round the pipe bend.
This paper is tried to give consideration towards the
approaches and techniques familiar with identify,
develop, and verify critical associations between
corrosion behavior and flow programs systems.
Used this frequently reduces with a practical
interpretation in the outcomes of flow on corrosion
turbulence systems.
2. Turbulence Models and methods
The task of turbulence models is always to provide
equations that will enable calculation in the
Reynold stresses,liuj, as well as the turbulent
diffusion fluxes,lm5uj, which arise when the time-
averaged equations for turbulent flow and mass
transport are acquired within the immediate
equation [14]. The k- , turbulence models [15]
which are currently [16] broadly useful for the
computation of business flows are eddy viscosity
model which be a consequence of the concept
recommended by Boussinesq and assumes caused
by turbulence round the mean flow can be
considered through viscosity. The turbulent
viscosity, µ,t is made the decision within the kinetic
energy of turbulence, k, which is rate of
dissipation,
µt = Cµ ƒµ ( k2)/ ) (1)
The effective viscosity is provided by
Deff = µ + µt (2)
similarly the effective diffusivity is given by
Deff = µ/ t + µt/ (3)
Where is the turbulent S chmidt number
[13].The conservation equations for mass,
momentum, kinertic energy of turbulence which is
dissipation, and species, m, might be witten in the
general form. For axisymmetrical flow in 2D round
co-ordinates [8]:
( ) + ( ) = ( ) + (r
) + (4)
Where
The values of , general diffusion coeffients and
, the source terms.
Flow separation with recirculation and
reattachment for example inside a sudden
expansion and understanding in the concentration
area enables the rate of mass transfer being
calculated [1,13] with the expansion. The
calculation in the concentration area near the wall
requires utilizing a low Reynolds number [LRN] k-
model since the mass transfer boundary layer is
deeply embedded within the viscous sublayer. The
concentration is required at y ~ .1, insidewithin all
the viscous sublayer where the mass transport is
diffusion controlled, to have the ability to calculate
the wall mass transfer rate. The 2nd requirement is
because of the bit of turbulence inside the vjscous
sublayer obtaining a significant effect on mass
transfer within the high Schmidt amounts Sc ~
1000 frequently familiar with aqueous mass
transfer. Low reynold number (LRN) models
utilize turbulence damping functions [13,32] as the
easiest way of modelling wall-boounded flow with
warmth/mass transfer under separated flow
conditions [9]. The option wall function (WF)
closure places the initial computation mode inside
the logarithmic law region (30 < y+ < 150) and
bridges over the important viscous sub-layer.
Furthermore, the motion of a dispersed particulate
phase within a turbulent flow field can be modelled
by either a Lagrangian or Eulerrian approach [34].
He further lamented that in Lagrangian models a
large number of individual particle trajectories are
calculated in the flow domain whereas in the
Eulerian approach the particles are treated as a
second fluid. From an erosion modelling standpoint
the direct calculation of particle/wall interaction
statistics, impact frequency, angle and velocity with
the Lagrangian approach is an advantage: at least in
dilute particulate suspensions. As pointed out by
Nesic [8], the Eulerian approach would be more
appropriate in concentrateed suspensions.
Also [10] review and revealed that particle
breakage may cause quality control problems,
whereas wall erosion increases equipment
maintence costs and environmental burden, and
causes loss of productivity and a requirement to
replace damaged components. Under normal
operating conditions, erosion rates in pipe bends
are much higher than those in straight pipe sections
due to local turbulence and unsteady flow
The 5th IMAT, November 12 – 13th
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70
behaviour. Lamented that therefore, the time for
failure piping systems is often dependent on the life
of pipe fittings (valves, bends, elbows, flow
metres), flow pumps, turbines, and compressors
[56-59]. It is, therefore imperative to improve the
protection of these components from solid particle
erosion by better understanding the physics of
particulate flows.
Many investigators [58-61] have carried out both
physical and numerical modelling of the erosion of
pipe bends, elbows, tees and related geometries.
Since the early 1990‘s, computational fluid
dynamics (CFD) has been widely used for solid
particle erosion prediction in curved pipes and
ducts, with various analytical, semi-empirical and
empirical models having been developed. In 1979
[67] provided a critical review of some of the
erosion models that had been developed since
Finnie (1960) proposed the first analytical
approach, and found 28 models that were
specifically for solid particle-wall erosion. The
authors [66,67] reported that 33 parameters were
used in these models, with an average of five
parameters per model. These parameters influence
the amount of of material eroded from a target
surface and the mechanism of erosion. The review
revealed that each model equation was the result of
a very specific and individual approach, hence it is
clear that no single equation exists that can be used
to predict wear from all known standard material or
particle parameters, and that some reliance on
experimental measurement will always be required
to provide empirical constants necessary in the
various erosion models. The following review is
limited to models that have been used in CFD-
based erosion modelling and which have received
wide usage in applications to erosion in pipes bends
and pipe fittings.
In 1987 [69] developed an empirical erosion model
for AISI 1018 steel. Mclaury (1993-1996) extended
this model for aluminium and used it to predict
particle erosion resulting from both direct and
random impingements in two-dimensional
geometries. This erosion model was developed at
the Erosion-corrosion Research Centers. In 1996
[71] predict erosion in elbows, plugs, tees, sudden
contractions, and sudden expansions for pipes with
circular cross-sections, and was subsequently
referred to as erosion -corrosion Research Centres
models. Also in 1973 [52] applied this model to
investigate the effects of elbow radius of curvature
on erosion rates in circulate pipe. Meanwhile,
[60] developed mechanistic models for predicting
erosion in elbows based on the E/CRC model. In
1998 and 2001, [67] and [68] used a commercial
CFD code to model fluid-solid flow and added
routine to predict erosion on particle impact using
the E/CRC models. The [67] and [68] also
modelled the erosion in oilfield control valves
using a commercial CFD code, accouting for both
deformation and cutting erosion. They obtained a
good agreement for the predicted wear rates and
wear locations in pipe bends with the experimental
gas-solid erosion results of Bourgoyne, added [63
and 64] the stochanstic rebound model of [61 and
62] and the E/CRC model [56] to a commercially
available CFD code and investigated the relative
erosion severity in elbows and plug tees found in
oilfield geometrics. Numerical simulations showed
that particle rebound behaviour played an important
role in determing the motion of the particle [65 and
71].
The performed their erosion research on 90 degree
elbows and bends of circular cross-section. The
fluid phase was modelled using a simple modefied
mixing-length model. The predicted fluid axial
velocity was validated against the experimental
data of [56], with erosion modelled using the
E/CRC model [68]. They compared their predicted
penetration rates with the experimental data of [65],
obtaining good qualitative agreement but poor
qualitative agreement between the predictions and
data. The poor agreement occurred because most of
the data available were from erosion experiments
with high particle rates. The authors also found that
erosion in long radius bends was reduced when the
carrier phase was changed from liquid to gas. They
further reported that the effect of the squeeze film,
secondary flows and turbulent flow fluctuations
may all play important roles in erosion prediction
when the carrier fluid is a liquid.
In 1991 studied [4]the local erosion in chokes and
determining the local fluid velocity and particle
impingement. Also, in 2006 [21] used a
commercial CFD code coupled with an in-house
particle tracker to predict fluid particle flow in a
full 180 bend. He further implemented two erosion
models, namely those of finite (1961) and [ 1 and
65] investigated erosion-corrosion problems in U-
bends. However, there was no comparison was
made with experimental data ; hence, the validity
of the model could not be ascertained. Attempted to
account for the shape of wear scars in predicting
the life of pneumatic conveyor bends undergoing
erosive Wear. However, these Authors did not use
the shape of the scar to alter the computational
mesh used in the fluid phase calculations.
Fan and co-workers [55] used Eulerian-Lagrangian
approaches with the empirical restitution
coefficient of [63] to model particle rebound
velocities, and subsequently employed the semi-
empirical erosion equation of [64] to study erosion
in a vertical-to-horizontal bend. The authors used
this technique to predict the erosion of tube banks
in heat exchangers (65 and 66), to study protection
techniques against tube erosion (67), and to
investigate anti-erosion in 900 bends (68, 69, 70
and 71). In all the studies, the authors used the
standard k- turbulence model, except in [62]
where the authors used large eddy simulation LES
The 5th IMAT, November 12 – 13th
2012
71
to obtained flow and turbulence field predictions.
The LES solution was not, however, validated
against previous experimental or numerical results.
In 1994 and 2004 [69 and 70] also used the semi-
empirical erosion equation of [63] in 1998 to
estimate the erosion rate in tube banks.
In 2007 [70] conducted a number of CFD-based
erosion modelling investigations using a
commercial CFD code. The erosion equation
developed at the University of Tulsa (Albert, 1994;
Edwards et al, 2001; McLaury 1993, 1996;Mclaury
and [49] in 2000; [50] in 1995; [64] in 1995 and
[54] in 2006, were included in the through user-
defined-function. Particle trajectories were
validated against the authors‘ experimental data for
liquid-solid flows obtained using laser diagnostics.
The entire CFD -based erosion authors‘ modelling
procedure was then validated by comparison with
the authors‘ experimental data obtained in 900
standard elbows with air flows, measured using a
sensitive electrical-resistance probe. In 2009 [66]
investigated particle motion in the near-wall region
using a commercially available CFD code,
modifying the code to account for particle size
effects in this region before and after particle
impact. For turbulent flow in a 900 bend, their
results showed that the near-wall modifications and
turbulent particle interactions significantly affect
simulation results when compared with
experimental data.
In 2009 [66] applied an Eulerian-Lagrangian
approach with particle-particle interaction and a
erosion model to simulate solid particle movement
as well as the particle erosion characteristics of a
solid-liquid two-phase flow in a choke. The authors
used the standard k- model to treat the turbulence,
the discrete particle hard sphere model to
accommodate inter-particle collisions and the semi-
empirical correlations of [65] in 1979 to study anti-
erosion effects.
Despite all this work, there is continued interest in
pipe wall erosion modelling because the prediction
of erosion, in particular, is of value in estimating
the service life of pipe bends systems, as well as in
the identification of those locations in a particular
pipe geometry most prone to erosion. In this
review, [55] a study on three-dimensional
computational fluid dynamic model of erosion is
developed to investigate the erosion of both the
concave and the convex walls of cross-sectioned
ducts of different bend geometries and orientations
due to particle collisions with the wall surfaces.
Results are discussed in terms of eroded depth and
the location of primary and secondary wear, and
are compared with available experimental data
[67].
In practice after corrosion risk appraisal, this can be
reduced to:
Total CA = Uniform CA + erosion allowance +
localized (pitting) allowance. The
predictive corrosion rates can be based on the
predicted temperature profiles and the published
top of line corrosion correlations for example, the
Nyborg-Dogstad correlation [69]. There are few
published such [equation 5] can be used on an
iterative basis, to determine top of line activity
[50,67].
Corrosion rate = 0.004*Rc*Cfe*(12.5 - 0.09*7) (5)
Where:
Corrosion rate: given in mm/y, Rc: Condensation
rate g/m2/s, Cfe: Saturation iron levels in
condensation (difficult variable, 50 - 200ppm often
select), and T: Temerature in degree centigrade (0c)
respectively.
3. Flow Accelerated Corrosion
The key factor factor response to determine flow
faster corrosion (FAC) might be the oxide film on
structure surfaces, which evolves consequently of
corrosion and, simultaneously, controls the
corrosion rate within the role like a protective film
[47]. The primary parameters to discover FAC they
fit into material parameters, flow dynamics
parameters and atmosphere parameters (48).
Metallic ions, mainly ferrous ions (Fe2 ), are
released for the water within the boundary layer
where a number of within the supersaturated Fe2
ions become oxide pollutants and in addition they
deposit over the metal surface being magnetic
oxide layer [48]. The oxide layer plays an
important role in preventing further relieve Fe2
(corrosion reaction). The thickness inside the
boundary layer is impacted by flow dynamics of
people processes, oxygen concentration (O2) inside
the boundary layer plays an important role for
oxidizing magnetite to hematite, which contributes
to achieving much greater corrosion resistance
[49].
Generally, subsea pipelines flow assurance covers
all multiphase transport phenomena. Diligent
design techniques, understanding and capabilities
are very important to make sure safe, continuity of
fluids transport from reservoir to topsides
processing plant. The main areas involve steady
condition and transient multiphase flow hydrates,
sand, oil, emulsions, wax, scale and corrosion
phenomena. The interaction between
corrosion/scaling and flow assurance can therefore
be instrumental in determining true production
rates, and software packages such as the Scand
energy or Non-destructive test types codes can
certainly give reliable predictive modelling.
Delivering an exact and reliable advantages of flow
assurance and corrosion modelling is regarded as as
as one of the primary challenges facing the today.
The overriding performance fingerprint is often
best known to as tub curve and also to help
facilitate this better, corrosion needs to be
recognized to love an operating hazard. Once that's
The 5th IMAT, November 12 – 13th
2012
72
recognized the benefits of a soundly planned and
faithfully applied corrosion management strategy
becomes self apparent along with a pronounced
reliance upon important water subsea infrastructure
and tie back [25,27, and 30].
The beginning of issues with corrosion integrity
throughout early existence is important, though mid
existence is often better handled since area
existence is extremely frequently well below design
existence and for your reason options and time can
be expected to favour planned retrofit as needed.
Existence extension beyond the situated on out
zone is generally more tightly related to older
researches whereupon original design life‘s are
actually exceeded and continuing production
needed [28,31].
4. Enabling Presumptions
Uncover the dominant corrosive species usually
CO2, with defined or understood water chemistry
content.
Review, verify, and prioritize or assume the
dominant flow programs, e.g. single or multiphase
flow, stratified, slug, annular flows etc. forecasted
while using existence cycle. This might require
review of the steady condition and non-steady
activities or transient flow situations. Frequently
the flow assurance report examines the likely
situations and phone connection, and with this
judgment [37].
Validate the H2S souring inclination because this
greatly impacts cracking and corrosion behaviour.
As being, this is often less influenced by flow
programs provided the most effective scales remain
intact.
Corroborate dissolved chloride and oxygen levels,
and the existence of aggressive organic species for
example acetic/formic chemicals additionally for
their types.
Assume the bottom corrosion is uniform but
validated by real-time inspection and monitoring
and via analysis of coups/probes, and deposits.
However anticipate to witness localized corrosion
when unsteady conditions occur.
Confirm threat and chance of biofilm formation
assuming pattern of growth follows laminar or fluid
stagnation sites, identify microbe species and MIC
within the existence cycle.
Corroborate the extent of sand production (steady
and episodic), along with the effect on materials
degradation. Determine sand concentration, and
particulate dimensions, and review interactions
with inhibitor performance [53].
Determine the role of small-stagnant zone
corrosion under primary flow conditions per small
locations where pollutants and biofilms may
proliferate [37].
Establish extent of pigging and inhibitors to both
create sustainable inhibitor films and repair such
inhibitor films even at high flow rates under
sanding nd erosion conditions [54].
5. Impact of Temperature and Flow
Temperature and flow regime are carefully linked
since CO2 corrosion is dynamics and very mindful
to electro-chemical and physical fluctuations (for
example changing P, T, V). Generally steady
condition (P, T, V) Conditions frequently promote
protective film compaction as well as for your
reason passivation, and low corrosion rates. Lower
temps < 1200F (50
0C) tend to promote patchy
corrosion with softer multi-layer iron carbonate
scales providing some barrier protection increasing
up to 140 to 1600F (60-70
0C). Above these
temperatures damaging localized corrosion is
observed as films lose stability and spall off giving
rise to galvanic ‗mesa‘ attack. Though there is
evidence of a down turn in the plateau after 800C
for certain cases. In reality the project design basis
usually insists on a maximum value for
temperature, as it does for other critical parameters
such as pressure, materials characterization (yield
stress, hardness, toughness, etc.). Regarding flow,
the production rates can be influenced by flow
regimes such as slug flows and annular gas flows.
This can prove critical for vertical risers connecting
the flowline to the offshore structure or topsides.
The geometry can act almost like a ‗specification
break‘ whereupon the flow regime shifts largely
due to the effects of gravity as the flowline
transforms from a horizontal to a vertical member.
The sag bend at the touch down zone can become a
high risk corrosion component warranting greater
degree of corrosion control, such as a thicker
section, increased local CA, internal coating epoxy
or increased monitoring routines etc.
Furthermore, as lamented earlier most likely
probably the most challenging conjecture is less
whether localized corrosion will occur but much
more likely at this point you request , generally
where? An thorough study on the expected flow
regime curves and maximum flow velocities might
help help with that judgment [25, 26]. It's also
likely that whenever flow line localized corrosion
starts, then it is more probably being self
propagating (auto catalytic) and is less influenced
by modifications inside the majority conditions
though more jobs are needed in this region (42,45
50).
For operating profiles examined in that way, for
example stratified flow, annular flow, slug flow
and bubble flow, it should be assumed that essence
turbulent as with line using the typical flow line
Reynolds number, but realize that within each flow
regime you will observe complex inter-facial
behavior, including laminar, stratification and
turbulence. The level of smoothness of people
activities isn't necessarily expected, however
The 5th IMAT, November 12 – 13th
2012
73
reliable online monitoring results might help
minimization and control, given to seize control of
the feelings at a great choice.
6. Internal Corrosion Direct Assessment
Harmful internal corrosion in addition to failures
have happened on pipelines carry gas /liquid
specified being dry [37]. The process referred to as
internal corrosion direct assessment (ICDA)
remains designed to look at the corrosion impact of
short-term upsets on pipeline integrity. The process
is anticipated to boost pipeline integrity, reliability,
and public safety. ICDA might come to terms with
enhance the assessment of internal corrosion in gas
transmission pipeline and help ensure pipeline
integrity. The process is primarily positioned on
gas transmission lines that normally carry dry
gas/liquid but they're affected from temporary
upsets of wet gas/liquid water (or electrolyte). The
understanding basis must be readily transferrable
together with other media liquids.
7. Single phase flow
An low Reynolds number (LRN), k-€ model has
lately been positioned on the calculation of pipe-
wall mass transfer rates in the sudden change or
expansion [5], an immediate constriction [1] and
flow round the groove [8]. They pointed out, in
addition mass transfer rates are really calculated for
that pipe wall in the sudden bend where small
patches in the protective ‗rust‘ film were assumed
to possess been removed. The above mentioned
pointed out stated results indicate the mass transfer
regions of erosion- corrosion processes in flow
pattern conditions may be satisfactorily laboured
with through turbulence models.
A much more intractable issue is an chance to
calculate protective film removal under single
phase aqueous flow conditions. The partial
elimination of a protective surface film is
frequently the precursor to rapid corrosion and
component failure. For instance failures in copper
piping in apartment structures frequently connect to
rapid corrosion in the sudden difference in the
geometry in which the normally protective film
remains broken using the enhanced turbulence.
Recent findings have proven that although point
about this kind of corrosion happens near the outlet
of 900 bends the film breakdown and subsequent
failure began inside the sudden step in which the
downstream pipe was soldered towards the elbow.
The idea of an essential shear stress for eliminating
protective layers remains asked for because the
small stresses involved wouldn't be sufficient to
robotically remove a surface oxide film. As pointed
out by Launder, B. E., 1998 pressure fluctuations
unlike velocity fluctuations don't vanish inside the
wall. In 1991, [4] has recommended that pressure
fluctuations inside the wall, in flow pattern, raise
the overall shear stress and could produce
mechanical damage.
8. Liquid/Solid flow
The presence of solid particles enhances the
destruction of protective films giving rise to
increased corrosion rates and may add to the
overall metal loss by the mechanical erosion of the
underlying metal. As with single phase flow these
destructive effects are more pronounced under flow
pattern condition.
In 1991 [4] have developed a predictive model for
localized erosion- corrosion under flow pattern
conditions based on the application of a two phase
flow version of an low Reynolds number (LRN), k-
model of turbulence. The motion of the particles
was predicted by means of a Lagrangian
Stochastic-Deterministic (LSD) model. The model
which was applied to various pipe geometries
including a sudden expansion, constriction and a
groove was based on an oxygen-mass-transfer
controlled corrosion model with the assumption
that the particles removed the protective rust film,
and an erosion model based on the cutting wear
erosion equations [58].
The successfully applied a two phase k- model to
the numerical simulation of uniform CO2 erosion-
corrosion under separated flow conditions at a
sudden expansion. The particles were modeled by
an Eulerian approach [62].
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2012
74
9. Summary of previous works
Year(s) Researcher(s) Method(s) Flow Pattern
Reynolds
number
Size(D=depth/L=length)
1966 Burgraff O. Numerical:
modified
relaxation
Up to 400 Square bend (D/L = 1)
1979 Benjamin A., et
al
Numerical:
ADI Scheme
for linear
system and
Newton-like
method for
non-linear
system
Up to 10000 D/L =1 and 2
2004 Yao H., et al Numerical:
Finite
difference for
unsteady 3D
incompressible
Nevier-Stoke
equation
1000 to 10000 D/L = > 0.1
2005 Cheng M., et al Numerical:
Lattice
B0ltzmann
Range of 0.01
to 5000
D/L range from 0.1 to 7
2006 Thierry M., et
al
Experimental:
Doppler
Velocimetry
1150 < Re <
10670
0.5 < D/L = 2
2009 Ozalp A., et al Experimental:
PIV
1230, 1460, and
1700
Rectangular, Triangular
and Semi-circular with
D/L = 2
2011 Muhammad R.
M.
Experimental:
PIV ( tape as
fluid) and
Numerical: 2-
D fluent
23144, 32963
and 39275
D/L =extend surface,
X = 0, 5, 10, 15, 20cm
2012 Muhammadu
M. M.
Experiment:
PIV (Seawater
as fluid) and
Numerical: 3-
D fluent
On going On going
Corrosion Behaviour
1960 Bitter J. G.A. 2-D model the result could not ascertained
1993 to
1996
Fan J., et al E/CRC and
Eulerian-
Lagrangian 3-
D models
Invesgate Concave and Convex walls of
square duct, and no significant results
1990 Nesic S., et al K- model Determined gas flow pattern near-wall
turbulence intensity
The 5th IMAT, November 12 – 13th
2012
75
(under gas
flow)
1991 Nesic S. K- model to
low Reynolds
number
Determined wall-pipe mass transfer rates
1993 Postlethwatte
S., et al
2-D (under
disturbed flow)
Determined local mass rates and s
Products at sudden expansion
2009 Binder S., et al Norsok M506
(Multiphase
gas /liquid
flow)
Generate noteworthy results but lack
agreement b/w modelling and field data
2011 Mamat M. F. Immersion and
Salt spray tests
Determined corrosion rate of the welded
and unwelded joint
2012 Muhammadu
M. M.
Non-
destruction test
(NDT) and X-
ray test
To determine the effects of fluid flow at
pipe bends on the corrosion behaviour of
low carbon steel
Further Work
Further work is required to clarify the stability of
corrosion behaviour un- der the condition of
turbulence fluid flow at pipe bend where are
amenable to both experiment observation and
numerical simulation in relation to Reynolds
number or particle image velocimetry (PIV).
Also, identification of positions or sites, within the
internal surface contact areas where the maximum
corrosion stimulus may likely to occur, thereby
allowing better understanding, mitigation,
monitoring and corrosion control over the life
cycle.
Conclusion
Corrosion and its interactions with the flow
phenomena is a complex discipline, neither one
dominates the other. Corrosion modelling results
are often found not to agree with field data. The
discrepancy is often blamed on the inadequacy of
the models. However, this article or review has
concluded that the differences are better explained
by the rationalization of the effects of flow pattern
on the base and localized corrosion rates. The core
postulate is that the lack of agreement between
corrosion modelling data and field experience is
due more to inadequate account of corrosion
stimulating flow regimes, rather than limitations of
the modelling: thus reverting the onus for corrosion
prediction away from corrosion modelling to the
flow regime side.
The arguments are new and on going and a
paradigm shift in thinking is needed to explore
further the links between flow pattern and
corrosion behaviour. In other word, a
computational fluid dynamic model coupled to a
Lagrangian particle tracking routine and a number
of erosion models have been used to predict the
solid particle erosion in square cross-section bends
for dilute particle-laden flow. The results obtained
were clearly affected by uncertainties in the
empirical restitution coefficients. However, the
results obtained do demonstrate the ability of the
CFD techniques employed to predict erosion
location and the depth of material eroded, and
hence their usefulness in providing estimate of the
service life of pipe systems as well as in the design
of mitigation measure.
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2012
78
Friction Characteristic of Palm Olein at Different Operating
Temperature using Four-ball Tribom S. Syahrullail, C.I Tiong
Faculty of Mechanical Engineering
Universiti Teknologi Malaysia, UTM Skudai
81310 Skudai, Johor, Malaysia
Tel : (607) 5534661. Fax : (607) 5566159
E-mail : [email protected]
ABSTRACT
One of the main disadvantages of vegetable oil is its
poor performance at high temperature. In this
experimental work, the performance of refined,
bleached and deodorized (RBD) palm olein was tested
at different operating temperatures using four-ball
tribometer, following the procedure of ASTM D 4172.
The result produced by RBD palm olein was
compared with the result by additive free paraffinic
mineral oil. The result showed that the RBD palm
olein had lower coefficient of friction compared to the
paraffinic mineral oil. However, the wear scars on the
ball bearings surface lubricated with RBD palm olein
were larger compared to those lubricated with
paraffinic mineral oil.
Keywords : Four-ball tribometer, friction coefficient,
wear scar diameter.
1. INTRODUCTION Lubrication is a technique applied to reduce the wear
caused by contacting surfaces either in close proximity
or moving relative to each other. A substance called
lubricant is interposed in between the contacting
surfaces to help it to lessen the friction caused by the
pressure due to the weight of the the load. Proper
lubrication is the most effective method to control
friction and wear.
Tribology is science of friction, wear and lubrication.
Tribology is very important in modern machinery
since it consists of sliding and rolling surfaces. In fact,
tons of lubricants are used every year in the industry.
In 2004, 37.4 million tons of lubricants were used
worldwide. The percentages included 53% automotive
lubricant, 32% industrial lubricant; including related
specialities, 5% marine oil and 10% process oil [1].
The total consumption of industrial lubricant included
37% of hydraulics
oil, 7% industrial gear oils, 31% other industrial oils,
16% metal working fluids and 9% greases. Researches
in this field can ensure greater efficiency, better
performance, less breakdown and significant saving
[2]. Due to high demand, the researchers all over the
world try to come out with the alternative for mineral
oil based lubricant, such as vegetable oil. The
vegetable oil is so far the best choice because it is
made from vegetables and the most environments
friendly. Public awareness about the importance of the
conservation of the environment also encourages the
use of vegetable oil as lubricant. The public has been
made aware that the lubricants used in machines can
contaminate the soil and underground water supply.
Therefore, the role of vegetable oil as a lubricant is
crucial and the potential is very high.
Nowadays, lubricants based on vegetable oil have
started to replace the role of mineral based oil as
industrial lubricants. This is due to the fact that the
mineral oil lubricants are very dangerous in many
applications because they are not readily
biodegradable and are toxic. Global environment
awareness also encourages researchers to produce
environmental-friendly lubricants. Non-toxic and
biodegradable lubricants have become a major issue
especially when the lubricants involve contacts with
soil, crops and ground water. Biodegradability is the
ability of a substance to be decomposed by the action
of the bacteria into CO2, water, mineral compound
and bacterial bodies. There are some factors affecting
the biodegradability such as molecular structure,
chemical properties and environmental conditions [3].
Some other properties of vegetable oil such as high
viscosity index, good lubricity, high flash point and
low evaporative loss [4] are measurable characteristics
to choose the best lubricant. As a result, there has been
a major interest in developing all sorts of lubricants,
including greases and hydraulic fluids, based on
vegetables oils such as rapeseed oil, castor oil and
palm oil [5]. Rapeseed oils have excellent lubricating
properties, load carrying capacity, and corrosion
protection properties compared to mineral oil.
Vegetable oils or natural oils are not only limited for
use in the food industry, but there are also various
IMAT-UI 014
The 5th IMAT, November 12 – 13th
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79
applications of vegetable oil, such as the use as
lubricants. The examples of the fields of application
for environmental-friendly lubricants are [6]: outboard
two strokes engine oils, chain saw and saw frame oils,
hydraulic oils for forest and agricultural equipment,
lubricants for sewage treatment plant, water weir plant
and lock gate, lubricant for food machinery, metal
working and metal forming processes and internal
combustion engine and hydraulic system.
However, vegetable oils have poor performances at
high temperature and poor oxidation properties. In
general, vegetable oils have a structure of triglyceride
with three different fatty acids. In this structure, there
are saturated and unsaturated fatty acids. The higher
the concentration of unsaturation, the poorer the
oxidation stability will be. These fatty acids are
effective in improving the capability of vegetable oil
to create the boundary lubrication condition along the
process. When two surfaces move towards each other,
the movement will generate heat. The increment of
heat would break the double bond (unsaturated bond)
in vegetable oils and create a possibility to other
elements to have chemical reaction with the vegetable
oils structure. At the same time, the existence of
oxygen molecules in vegetables oils chain would react
with other metals and lead to oxidization of the metal.
In this experimental work, a refined, bleached and
deodorized (RBD) palm olein was tested using a four-
ball tribometer at different operating temperatures
following the procedure of ASTM D 4172. The results
produced by RBD palm olein were compared with
results by the additive free paraffinic mineral oil. It
was found that RBD palm olein showed a lower
coefficient of friction.
2. EXPERIMENTAL METHOD
The experimental work was done using a four-ball
tribometer as shown in Figure 1, following the
procedure of ASTM D 4172. The normal load was
fixed to 40kg and the speed was set at 1200rpm. The
test was conducted for one hour. Test lubricants were
heated up to 75°C before starting the experiments. The
details of the test procedure had been described in
previous publications [7]. The standard steel balls used
in this experiment are made from AISI E-52100
chrome alloy steel, with the diameter of 12.7 mm,
extra polish (EP) grade 25, hardness 64 to 66 HRC
(Rockwell C Hardness). Four new balls were used for
each test. Each time before starting a new test, the
balls were cleaned with acetone and wiped dry using a
fresh lint-free industrial wipe.
The test lubricant used was RBD palm olein. RBD is
the abbreviation of refined, bleached and deodorized.
Palm olein is the liquid fraction obtained from the
fractionation of palm oil after crystallization at a
controlled temperature. In this research, a Malaysian
Standard of MS 816:1991 of RBD palm olein was
used. This type of palm olein has been refined and
contains less free fatty acid. The results obtained from
the experiments using RBD palm olein were compared
with the results from the experiment which used
paraffinic mineral oil. Each test used 10ml of
lubricant.
3. RESULTS AND DISCUSSION
3.1. Coefficient of Friction
Antifriction ability is one of the important
characteristics of lubricants in order for the
mechanical system to run smoothly and reduce the
maintenance cost. In this experiment, the performance
for RBD palm olein (PO) and paraffinic mineral oil
(PMO) were evaluated in term of coefficient of
friction (COF) using a four-ball tribometer. The results
in Figure 1 shows the COF for RBD palm olein and
paraffinic mineral oil at different operating
temperatures. In this experiment, the operating
temperatures were set from 55˚C to 85˚C with gradual
10˚C increment. It was observed that the trend of COF
for RBD palm olein and paraffinic mineral oil
increased as the operating temperature increased. At
the operating temperature of 55˚C, the COF for RBD
palm olein and paraffinic mineral oil were similar,
with 0.066 and 0.068 respectively. At the operating
temperature of 85˚C, the COF for RBD palm olein and
paraffinic mineral oil increased to 0.08 and 0.086
respectively. The result showed that the coefficient of
friction for both oils increased with the increase of the
operating temperature. The coefficient of friction for
RBD palm olein (PO) was lower compared to
paraffinic mineral oil (PMO). It was due to the
presence of fatty acid molecules in vegetable oil that
stuck well on the surface and created a boundary
lubrication condition [8,9]. The increment of operating
temperature might have lead the boundary lubricant
breakdown due to lower viscosity [10].
The 5th IMAT, November 12 – 13th
2012
80
0.04
0.05
0.06
0.07
0.08
0.09
55 65 75 85
Temperature (Celcius)
Co
eff
icie
nt
of
fric
tio
n
PO
PMO
Figure 1: Coefficient of friction for RBD palm
olein and paraffinic mineral oil under
different operating temperature.
3.2. Wear Scar Diameter
The wear scar diameter (WSD) of three stationary
bearing balls was measured using a CCD microscope
for each lubricant in this experiment, and plotted as
Figure 2.
0.5
0.55
0.6
0.65
0.7
0.75
0.8
0.85
0.9
55 65 75 85
Temperature (Celcius)
Wear
scar
dia
mete
r
PO
PMO
Figure 2 : Wear scar diameter measured for
bearing ball lubricated with RBD
palm olein and paraffinic mineral oil
under different operating
temperature.
The result showed that the RBD palm olein had almost
similar or slightly higher value of WSD compared to
paraffinic mineral oil. This could have been attributed
by the chemical attack on the rubbing surfaces of fatty
acid in RBD palm olein [11]. Both test oils showed a
slight reduction of WSD value with the increase of
operating temperature.
3.3. Wear Scar Surface Roughness
The surface roughness profile of wear scar surface of
bearing balls lubricated with RBD palm olein and
paraffinic mineral oil were measured using a Mitutoyo
surface roughness profiler. Figure 3 shows the surface
roughness, Ra of wear surface for RBD palm olein and
paraffinic mineral oil under different operating
temperatures. From Figure 3, it can be seen clearly
that the wear surface roughness of bearing ball
lubricated with RBD palm olein became smoother;
however, those lubricated with paraffinic mineral oil
had more wear surface roughness with the increase of
the operating temperature. For RBD palm olein
lubricant, the highest surface roughness value was
2.313µm, at the operating temperature of 55°C. For
paraffinic mineral oil, the highest surface roughness
value was 3.250µm, at the operating temperature of
85°C. From the results in Figure 3, it can be
understood that the wear surface roughness for RBD
palm olein after experiments had become smoother
with the increase of the operating temperature but the
wear surface profile paraffinic mineral oil became
worse with the increase of the operating temperature.
The change in surface roughness value mainly resulted
from the change of wear mechanisms as suggested by
Hiroyuki [12] and the reduction of lubricant viscosity.
0
0.5
1
1.5
2
2.5
3
3.5
55 65 75 85
PO
P2
Figure 3 : show the surface roughness for the
wear surface of bearing ball lubricated
with RBD palm olein and paraffinic
mineral oil respectively after the
experiment.
The 5th IMAT, November 12 – 13th
2012
81
(a) PO at 55°C
(b) PMO at 55°C
(c) PO at 65°C
(d) PMO at 65°C
(e) PO at 75°C
(f) PMO at 75°C
(g) PO at 85°C
(h) PMO at 85°C
Figure 4 : wear worn surface lubricated with PO and PMO under different temperature.
The 5th IMAT, November 12 – 13th
2012
82
3.4. Worn Surface Observation
The worn surfaces of balls bearing were observed and
captured using a CCD microscope. All the worn
surfaces are shown in Figure 4 with lubrication of PO
and PMO respectively, under different operating
temperatures. PO and PMO represent the RBD palm
olein and additive free paraffinic mineral oil in this
experiment. The worn surfaces of PO and PMO were
compared mutually at specific operating temperature.
For the wear surface with lubrication of PO at
temperature of 55°C until 85°C, we could clearly see
the scar or grooves on the worn area. The scars or
grooves became narrower with the increase of
operating temperature. The largest wear groove with
lubrication of PO was found at the temperature of
55°C. This also indicated that the wear surface was
the roughest compared to others. The surface became
smoother with the increase of operating temperature
by showing narrower grooves on the surface.
However, at the temperature of 85°C, the wear
surface with lubrication of PO became rough and an
adhesive wear could be observed just like others. The
best wear worn surface was at the temperature of
75°C for PO by giving the smoothest surface. On the
other hand, the wear surface with lubrication of PMO
showed the opposite trend of result compared to PO.
The wear worn surface became worse with the
increase of operating temperature in this experiment.
At the operating temperature of 55°C until 65°C, the
wear groove could clearly be seen on the worn
surface lubricated with PMO. The wear scar or
grooves became unclear with the increase of
operating temperature, thus it could be observed that
material transfer or remover occurred on the worn
surfaces at the operating temperature of 75°C onward.
4. CONCLUSION
The tribological properties of RBD palm olein have
been investigated using four-ball tribometer at
different operating temperatures. The results of RBD
palm olein are compared with the result by additive
free paraffinic mineral oil. It can be concluded that
RBD palm olein gives better lubrication
performances based on the lower coefficient of
friction compared to paraffinic mineral oil. However,
the wear with the use of RBD palm olein is slightly
higher compared to the wear with the use of
paraffinic mineral oil. This problem could be solved
by the addition of anti-oxidant agent. The wear scar
on the ball bearings with lubrication of RBD palm
olein are smooth, showing that less metal-to-metal
contact occurred.
ACKNOWLEDGMENT
The authors wish to thank the Faculty of Mechanical
Engineering at the Universiti Teknologi Malaysia for
their support and cooperation during this study. The
authors also wish to thank Research Management
Centre (RMC) for the Research University Grant
(GUP) from the Universiti Teknologi Malaysia,
Fundamental Research Grant Scheme (FRGS) from
the Ministry of Higher Education (MOHE) and E-
Science Grant and ERGS from the Ministry of
Science, Technology and Innovation (MOSTI) of
Malaysia for their financial support.
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Pulse Detonation Engine Research Development at High Speed
Reacting Flow Laboratory, Universiti Teknologi Malaysia Mazlan A. Wahid, A. Dairobi G., Aminuddin Saat, Mohsin M. Sies, H.A. Mohammed,
A. N. Darus, Mohd Faizal H., M. Ibthisham A., Fairus M. Y. and Z. Lazim
High-Speed Reacting Flow Laboratory - HiREF
Department of Thermofluids
Faculty of Mechanical Engineering
Universiti Teknologi Malaysia
81310 Skudai, Johor Darul Takzim
MALAYSIA
E-mail : [email protected]
ABSTRACT
Pulse detonation is a propulsion technology that
involves detonation of fuel to produce thrust more
efficiently with mechanical simplicity than currently
available engine systems. Detonation combustion
modes in pulse detonation engines (PDE) is known to
be more efficient in creating high power simply
because the energy release in detonation modes of gas
combustion are several magnitude higher than in
deflagration combustion modes. PDE research
program has been started at High Speed Reacting
Flow Laboratory (HiREF), Universiti Teknologi
Malaysia (UTM) since 2005. The studies began with
single pulse detonation study for detailed investigation
of detonation characteristics of various fuels and
followed by development of repetitive PDE engine,
performance study and augmentation of thrust using
various types of ejectors. This paper summarizes the
laboratory facilities, research activities conducted and
the output from all of the related research programs.
Keywords : High Speed Reacting Flow Laboratory
(HiREF), pulse detonation engine,
combustion, high speed reacting flow
1. INTRODUCTION
Pulse detonation engine (PDE) is a new type of
propulsion technology with advantages on thermal
efficiency, higher thrust impulsive and simplicity
mechanical design [1-4]. Researchers from various
institutions also suggest that the PDE have the
potential to power aircraft on subsonic, supersonic and
hypersonic speed [5-7]. PDEs could also be used to
power tactical aircraft, air and ship-launched missiles,
unmanned aerial vehicles, power generation, and a
wide range of stand-off munitions [8,9]. However
extensive research and development are indeed
essential before PDE can be successfully applied on
those practical applications.
High Speed Reacting Flow Laboratory or in short
HiREF, has been established to focus on the study of
various discipline of sustainable combustion, high
speed reacting flow and heat transfer. In 2005 HiREF
embarked on pulse deflagration and detonation
combustion research studies and since then the group
has earned various invention awards as well as several
patents on pulse combustion technology. The
advantages and potential inherited by PDE as the new
propulsion engine concept has driven HiREF members
to focus on such promising area especially in the era
when fuel efficiency and sustainability has been a real
concern in today's depleting fossil fuel scenario. Since
then HiREF laboratory has been evolved stage by
stage to meet the PDE research requirement.
The research programs on PDE initially began with
fundamental study on detonation characteristic using
shock tube to study the characteristic of various fuels.
The studies continued with the development of the
repetitive PDE and some applications of PDE. The
team successfully operate the PDE at a reasonable
repetitive rate with substantial thrust. HiREF team is
currently busy in upgrading the fuel-oxidizer injection
system, higher energy ignition system, Deflagration to
Detonation Transition (DDT) mechanism as well as
improving the firing sequencing system.
Augmentation of PDE thrust also been studied by
employing ejectors of various geometry placed at the
open end of the PDE tube.
2. HIREF LABORATORY AND
COMPUTING FACILITIES
HiREF laboratory has been designed, modified and
equipped in meeting the PDE research requirement
and precise data collection purpose with safety as the
priority. The HiREF research facility is comprised of a
research office furnished with computing facility and a
laboratory that is equipped with soundproof room,
damping chamber, exhaust system, control room, fuel
and oxidizer supply system. Fuel is stored at the
IMAT-UI 015
The 5th IMAT, November 12 – 13th
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85
outside of the laboratory and delivered through
certified stainless steel pipeline and brass tubing.
Figure 1 shows the schematic diagrams of the HiREF
laboratory. Description of the HiREF laboratory
facilities is discussed next.
Figure 1: Schematic diagram of HiREF
laboratory arrangement
2.1 Soundproof Room
The dimension of the soundproof room is 8‘ x 11‘ x
14'. The room is made of double layers wooden walls
where polyurethane foam is fitted in between them.
Soundproof room is equipped with damping chamber,
exhaust fan, and fuel pipeline. The PDE setup with all
the necessary instrumentation is placed in this room
during experiment. Most of the experiments are
conducted inside this room as a confine area to
prevent detonation shock wave, noise, heat or any
unwanted event from harming the researcher.
Researcher can observe, operate and gather all PDE
experimental data from a control room that is
connected to the sound proof room through a blast
proof observation window.
2.2 Damping Chamber
Pulse detonation engine producing noise level with
more than 100 db [10]. The sound need to be damped
and this is achieved with the design and installation of
damping chamber in the soundproof room. This
chamber is also function to absorb shock waves and to
extract the exhaust gas from the chamber. It also helps
prevent any unburned mixture from spreading to the
open space in the laboratory in order to prevent
unexpected accident. The damping chamber is made
of large stainless steel silencer housed in a concrete
cylinder. The outer concrete cylinder formed as a rigid
structure for the silencer and act as an absorber to the
silencer due to the impact from the shock wave. Steel
silencer is a perforated metal plate rolled into
cylindrical shape. Polyurethane foam in the inner liner
of the concrete cylinder is function to further absorbs
the sound produce during the detonation event. Figure
2 shows the assembly of damping chamber concrete
cylinder with the silencer.
Figure 2: Damping chamber design
2.3 Control Room
Researcher control and gather performance data for
the PDE during operation using two respective
computers located in a control room. A dedicated
control room served to protect the operator from PDE
shock wave and noise produced by PDE during
operation. Figure 3 shows the schematic diagrams of
control room in the HiREF laboratory.
Figure 3: Schematic of control room
2.4 Measurement System and Devices
Measurements are an essential part of the PDE
experiments. Typical data captured for PDE analysis
are pressure, temperature, force, and acceleration. All
digital data acquired from various transducers such as
9” wall
Observation
windows
Sound
absorption
layer
4” wall
The 5th IMAT, November 12 – 13th
2012
86
pressure transducers, thermocouples, load cell and
accelerometers are fed through data capturing systems
comprises of a software and acquisition card. Labview
7.1 software is used to control and command the PDE
operation. Separate computer using the same
programming language is used to write, store and read
back the data for analysis. PCI-6133 acquisition card
is used to read the signal from all of the transducers
for each data type in terms of voltage before it will be
stored in the hard disc. This card then connected to
connector block and the signal conditioner. Signal
conditioner receives the signal from a transducer in
analog signal type that is in voltage. This signal
converted to digital signal that later sent to the
connector block that communicated with date
acquisition card. Figure 4 shows the schematic
diagrams of the data-acquisition connections. In the
figure, it also shows the location of pressure
transducer 1 and 2 (PT1 and PT2), load cell and
accelerometer.
Figure 4: Schematic diagram of PDE data
acquisition system
In the previous study, accelerometer and load cell
were used to measure the acceleration and thrust
respectively. The study on the suitability of using the
accelerometer on PDE thrust measurement have been
performed to fulfill the data validation requirement
[11, 12].
2.5 Cooling System
Experiments on supersonic combustion at a high
repetitive rate will generate heat sometimes up to
thousands of degree Celsius. The heat produces from
the combustion process need to be managed properly
since excessive heating will cause damage to the
transducers. The maximum operating
temperature for pressure transducer is only around
250°C whereas temperature during detonation
combustion may easily reach more than 1000°C.
Temperature higher than allowable operating
temperature will most likely cause damage to the
electronic part of the transducers.
The PDE instrumentation cooling system, such as
shown in Figure 5, includes a cooling station and
transducers cooling adaptors. This station functions
as liquids storage, supply, circulations and control.
The cooling station consists of water storage tank,
water pump, radiator, fan and also the control panel.
Cooling adapter is served also as the mounting for
transducer to provide cooling to the transducer. This
adaptor was custom made from stainless steel material
for the corrosion resistant. In the system cooling
fluid is circulated by 0.5 horsepower water pump with
40 milliliters per minute of water flow rate.
Figure 5: Picture of cooling system station.
3. SINGLE PULSE DETONATION
SETUP
Single pulse detonation setup was constructed in order
to study the details characteristic of high-speed
reacting shock waves for various fuels [13]. The
detonation tube is made of stainless steel with internal
diameter of 100 mm, damping chamber, deflagration
to detonation transition (DDT) section, soot film
section, gas filling system, a data acquisition system
and ignition control unit. The tube was designed to be
able to withstand static pressure loading of 100 bars
with safety factor of 2.49.
Water
tank
Water pump
Control panel
Radiator
The 5th IMAT, November 12 – 13th
2012
87
Figure 6: Schematic of single pulse detonation
setup
Figure 7: Single Pulse Detonation
The single pulse detonation tube was constructed in
three modular sections such as shown in Figure 6 and
7. Each section is 0.5 m long and has an internal
diameter of 100 mm and an outer diameter of 112 mm.
Section A and C has ports for plumbing and
instrumentation while section B only has ports for
instrumentation. Each section was connected to each
other using flanges. The flanges are 164 mm in
diameter and 5 mm thick and are held together by
eight M10 bolts. Silicon glue and gasket paper were
used to seal small gap between flanges and
connections when they are connected. The close end
part of the tube, which is located at the section A, is
attached with a plumbing valve, transducers and spark
plug. The open end where the detonation wave
exhaust into the atmosphere was inserted inside a
damping chamber to reduce sound that is created by
the transmitted wave. The damping chamber was
placed onto a detachable support. The detachable
support was tightened to the main structure using four
M12 bolts.
This facility was also used to study the characteristics
of reacting shock waves for biogas in comparison to
several other gaseous fuels [13] and experimental
study of confined biogas pulse detonation combustion
[14]. Biogas was comprised of 65% methane with
35% carbon dioxide. The oxygen concentration in
the oxidizer mixture was diluted with nitrogen gas
at various percentage of dilution. Computational as
well as the experimental studies of biogas and natural
gas fuel characteristics has also been performed in
HiREF [15, 16]. From both research programs it can
be concluded that the biogas and natural gas were not
sensitive to detonation propagation compared to other
gaseous like propane due to the lower calorific value.
4. REPETITIVE PULSE DETONATION
ENGINE SETUP
Repetitive Pulse Detonation Engine been developed
by HiREF team and the setup is shown in Figure 8.
The dimension of PDE tube was 50 mm in inner
diameter and 600 mm in length. Obstacles with
blockage ratio of 43 percent was used as an flame
accelerator inside the tube. Such tube size was chosen
since it will provide enough space for the inlet device
to be mounted on the tube surface. Propane and
oxygen was used as fuel and oxidizer. In the
experiments the cell size (λ) for propane-oxygen was
determined to be 1.3 mm [17].
Purging system uses solenoid valves with the
capabilities to supply compressed air up to a pressure
of 800 kPa. The solenoid coil was powered by 24Volt
DC power supply circuit and controlled by transistor-
transistor logic (TTL) signal from a PDE control
circuit. This PDE been designed to operate by using
gaseous propane as fuel and pure oxygen as oxidizer.
LO-Gas injector is the manufacturer for injector and
pressure regulator. This injector was designed to work
on 12Volt DC supply voltage with minimum pressure
as low as 100kPa to maximum pressure of 300kPa.
The injector‘s assembly came with 4 in line units for
oxidizer injection and 2 in line units for the fuel
injection. The injectors are mounted not directly to the
detonation tube to prevent overheating and failure due
detonation pressure wave and heat generated. These
injectors are attached to a custom-made mounting and
a 6 mm stainless steel pipe is used to deliver the
injected mixture to the detonation tube. One-way
valves are used to prevent the flame from flashing
back to the fuel-system tank. 12 VDC MSD Digital
DIS- 4 was chosen as an ignition system. This device
can provide four sparks of 105-115 mJ each. The
mechanism of filling, purging and ignition was
controlled by using LabView. Figure 9 shows the
sequences of the filling, ignition and purging. The
injection process was set for 50% duty cycle, 15%
duty cycle for ignition process and 20% for purging
process.
The 5th IMAT, November 12 – 13th
2012
88
Figure 8: Repetitive Pulse Detonation Engine In order to determine the characteristics and
performance of the PDE, the complete system has to
be mounted on a strong support structure. The support
structure must be rigid enough to be able to withstand
the vibration and force exerted by the impulse
generate during the test running. The tube mounting
also designed to have free movement on a railing so
that the amount of thrust generated can be measured.
Figure 9: PDE sequence
The PDE support structure was built from mild steel
angle iron with a dimension of 50mm x 50mm with
the thickness of 5mm combine with rectangular
hollow with a dimension of 50mm x 25mm and the
thickness of 3mm. The main support structure is
assembled by using arc welding to get rigid structure
to resist shock by the impulse cause by the detonation
event. The detonation tube is mounted on a freely
moving structure that is installed on a railing. This
structure has a sliding mechanism at the base so that it
will reduce the friction to the movement. This
requirement is needed to measure thrust by using load
cell. During the test, this structure will move backward
and compressed a load cell located between these
structures to the thrust wall.
Figure 10 shows the thrust recorded by the load cell at
frequency 5 Hz for 5 seconds. It can be seen that the
average values of the thrust spike by the PDE was
between 480 N and 520 N at these operating
frequencies. The average thrust spike calculated from
these results is 512.7 N. Figure 11 below shows the
pressure profile at frequency 5 Hz. The average values
of pressure spike were around 8 to 14 bars. The rise
time of the pressure is within the order of 10-4
s. In
order to further understanding the phenomena occur
inside of the PDE chamber, simulation had be
conducted detail in the literature [18].
Figure 10: Repetitive PDE thrust signal
Figure 11: Repetitive PDE pressure signal
5. THRUST AUGMENTATION
Installation of ejectors has shown some improvement
on the thrust generated by the PDE [19]. Experimental
test was conducted at HiREF to study thrust
performance on 5 Hz PDE and the ejector setup is
shown in Figure 12. This research utilized propane as
a fuel, oxygen as oxidizer and air as purge gas. The
effects of four different ejectors dimension, shape and
axial location on augmentations were investigated and
the ejector photo is shown in Figure 13. All ejectors
have similar length and inlet diameter which are 400
mm and 130 mm respectively. The length chosen for
all ejector configurations was at ejector length-to-
ejector diameter, LAugmentor/DAugmentor= 3.08. 5Hz
operation frequency of PDE shown the wave velocity
The 5th IMAT, November 12 – 13th
2012
89
was in range of 1500 m/s to 1800 m/s (~5-6 Mach)
and the thrust is about 9-14 bar is produced. The
ejector was found to be very sensitive to the axial
position. All the ejectors give improve performance of
thrust and the upstream position provided the best
performance due to the thrust augmentation.
Figure 12: Ejector arrangement on PDE
Figure 13: Ejector geometries; from left diverge,
converge, convergent-divergent and
straight ejector
6. FUTURE WORK
HiREF research group is currently becoming more
active in acquiring research grant in various discipline
of sustainable combustion, high speed reacting flow
and heat transfer. In 2012 the group has secured about
six new research grants and hence HiREF laboratory
will undergo few more upgrading in term of facilities.
One of the proposed research aims is to understand
fundamentally on the process of high speed reactive
flows especially for safety and accidents prevention.
The transition from deflagration to detonation will be
observed in unconfined congested environments. Such
environments are analogous to that in factories,
chemical process plants, forests, and warehouses. The
congestions (i.e. obstacles) will cause high-speed
turbulent deflagrations to form shock waves at the
flame front, which is the main characteristic of
supersonic combustion. In contrast, detonation also
occurs spontaneously when the ignition energy is
sufficient to enforce a shock wave at the onset of
combustion. In year 2012 the group has successfully
increases the number of publications in high rank
journals with cumulative impact factor of 30 [20-32].
ACKNOWLEDGMENT
The authors would like to acknowledge Universiti
Teknologi Malaysia for providing laboratory facility
and fund granted under Research University Grant
(RUG) scheme. The financial support provided by the
Malaysian Ministry of Higher Education (MOHE)
under Vote 79299 for PDE reasearch and laboratory is
highly acknowledged. Special thanks also to Ministry
of Science and Technology of Malaysia (MOSTE) for
providing various type of funding to support various
research programs at HiREF. We also want to
acknowledge the combustion and thermodynamic
laboartory staff for their help and advice on
experimental research program at HiREF laboratory.
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2012
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Pool Boiling of Nanofluids in Vertical Porous Media
Ridho Irwansyah, Nandy Putra
Heat Transfer Laboratory
Department of Mechanical Engineering
Unviersity of Indonesia, Depok, 16424
E-mail: [email protected]
Abstract:
The development of electronic components such as
microprocessor requires a better thermal
management system to overcome the high heat flux
produce by the component. The method to absorb
the heat produce by the microprocessor is still use
the conduction or either natural or free convection
which still in a single phase heat transfer. One of
heat transfer method that suitable for a high heat
flux application is pool boiling which has a two
order of magnitude higher than of a single phase
heat transfer and does not require a pump to move
the fluid. In this study has been conducted the pool
boiling experiment with four different porous
media surface which are sintered copper 300 µm
and 400 µm, copper screen mesh and stainless steel
screen mesh with four different fluid which are
H2O-Al2O3 1%, 3% and 5%. The sintered copper
400 µm has shown a better heat transfer
performance compared to the other porous media.
The H2O, H2O-Al2O3 5% has shown a performance
no better than H2O-Al2O3 1% and 3%.
Keywords: Nanofluids, Sintered Copper, Screen
Mesh, Pool Boiling
1. Introduction Technological developments towards
miniaturization of electronic components require
methods for better thermal management. The
microprocessor is one example of electronic
components with rapid growth, the development of
which was followed by an increase in waste heat
generated when the microprocessor works. The
typical cooling system for electronic components
based on conduction and single phase forced or
natural convection already inadequate to handle
such a high performance electronic components
[1].
One method of heat transfer that is often used in
the cooling system on the electronic components is
pool boiling. This is due to the high heat transfer
capabilities and the process does not require a
pump to move the working fluid [2]. Pool boiling
chosen due to its ability to move heat two times
better compared to the single-phase heat transfer in
conventional cooling methods [3]. There are
several method to enhance the heat transfer ability
of pool boiling, one of the common methods is
additive for liquids. In this case the addition of
nano size solid particle was required (10-100 nm).
The addition of nanoparticles to the base fluid
tends to increase the thermal conductivity of
working fluids [4, 5]. Several researchers have
done the research about boiling heat transfer, they
found that the boiling of nanofluids increase the
critical heat flux (CHF) of the boiling process
compared to the base fluid. Hyungdae Kim et.al
found that the using of Al2O3 and TiO2 nanofluids
increase the CHF of the boiling process into 170%
[6]. S.M You et.al found that the using of Al2O3
nanofluids at 0-0.05 g/l concentration increase the
CHF into 200% [7]
The other method to enhance the heat transfer
ability in pool boiling is the modification of heater
surface. Y. Takata et.al found that the coating of
TiO2 nanoparticles on the surface of the heater
improve the critical heat flux (CHF) compared to
the heater without nanoparticle coating [8]. Other
researchers have done another method to modified
the heater surface with coating with additional
nanoparticle layer on the heater surface [9, 10]
Shoji Mori and Kunito Okuyama found that the
using of porous media on the surface of the heater
increase the CHF value 2.5 times compared to the
heater without porous media[2].
The purposes of this research are to conduct the
pool boiling experiment of H2O-Al2O3 nanofluids
in vertical porous media and compare the result of
the nanofluids to the base fluid.
2. Experimental Setup The H2O-Al2O3 nanofluids in 1%, 3% and 5% of
volume concentration were used in this research.
The average size of the particle was 20 nm.
Distilled water (H2O) was used as the base fluid.
There are two steps to produce the nanofluids; the
first step was the stirring between the mixture of
nanoparticles and base fluid by utilize the magnetic
stirrer for 15 minute, and later the mixture were
added to the ultrasonic processor for one hour in a
high frequency process.
The porous media that being used in this research
were made from copper powder (particle size 300
µm 400 µm) and screen mesh (copper and stainless
steel) which has 40 mesh number, the porous media
is depicted in fig. 1. The SEM (Scanning Electron
Microscope) images of 300 μm copper powder are
depicted in fig.2. The particle was magnified 100
times and 300 times.
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Fig. 1. Copper screen mesh, stainless steel screen mesh and sintered copper porous media
(a) (b)
Fig.2. SEM images of 300μm copper powder (a) 100 times magnifying (b) 300 times magnifying
1
2
8
4
6
5
3
9
7
(a) Experimental setup
150 mm
130 mm
20 mm40 mm40 mm
(b) Thermocouple position on the heater surface
1. Circulating thermostatic bath
2. DC power supply
3. Condenser
4. Main heater
5. Auxiliary heater
6. Wall thermocouple
7. Fluid thermocouple
8. Computer
9. Data acquisition
Fig.2. The experimental setup and detail for thermocouple position
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300 W DC electric heater was used in this research.
The length and diameter of the heater was 150 mm
and 6 mm respectively and made of stainless steel.
The heater was connected to the adjustable DC
power supply. The auxiliary heater was used during
the early stage of the experiment. The temperature
measurement was done at 5 points which are 3
point at the heater surface and 2 points in the fluid.
5 type K thermocouples were connected to the NI
9201 data acquisition system and the error of the
thermocouple was 0.1 %. The boiling vessel was
made of glass. The thickness, height and diameter
of the glass were 6 mm, 200 mm and 115 mm
respectively. At the top of the boiling vessel a
condenser made from copper connected to the
circulating thermostatic bath. The temperature of
the bath were maintained at 25 0C. The detail of
experimental setup and thermocouple position of
the heater is depicted in fig. 3.
The thermal conductivity of the nanofluids was
measured using the KD2 Decagon method, the
same method was also use by [11, 12, 13]. The
thermal conductivity of the fluids from the
measurement result was shown in table 1.
Table 1. Thermal conductivity of fluids
Fluids
Thermal
Conductivity
[W/m.K]
Thermal
Conductivity
Improvement
[%]
H2O 0.56 -
H2O-Al2O3 1% 0.67 19.64
H2O-Al2O3 3% 0.69 23.21
H2O-Al2O3 5% 0.72 28.57
16 experiments were done during this research. 4
different fluids were tested for every porous media.
The surface roughness of the porous media that
made from sintered copper was measured before
and after the boiling process. The matrix of the
experiment was shown in table 2.
Table 2. Matrix of the experiment
Porous Media Fluid
Sintered Copper 300 µm H2O
Sintered Copper 400 µm H2O+ Al2O3 1%
Screen Mesh Copper H2O+ Al2O3 3%
Screen Mesh Stainless Steel H2O+ Al2O3 5%
3. Result and Discussion
3.1 Boiling heat transfer in sintered copper
porous media
The boiling curve of 300 μm sintered copper under
the variation of fluids is depicted in fig. 4. It can be
observed that boiling with the H2O-Al2O3 1%
volume concentration provides higher heat transfer
compared to other fluids. It can be seen that the
H2O-Al2O3 1% produce the lowest temperature
difference between the heater wall and the fluid
which was 3.74 K when the highest heat flux was
applied. It can be noted that the range of the heat
flux was between 1.8-36 kW/m2. The temperature
difference between the heater wall and the fluid of
water, 3% and 5 % nanofluids were 3.95 K, 4.42 K
and 4.2 K respectively.
Fig.4. Boiling curve of 300 μm sintered
copper under the variation of fluids
The boiling curve of 400 μm sintered copper under
the variation of fluids is depicted in fig. 5. The
additions of 3% nanoparticles to the base fluid
provide a better heat transfer compared to the other
fluids with the temperature difference between the
heater wall and the fluid was 0.22 K. In this
experiment it was found that the 1% and 3%
nanofluids provide a better heat transfer compared
to base fluid (H2O). The temperature difference
between heater wall and fluid for water, 1% and
5% nanofluids were 0.49 K, 0.47 K and 0.52 K
respectively.
Fig.5. Boiling curve of 400 μm sintered
copper under the variation of fluids
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3.2 Boiling heat transfer in screen mesh porous
media
The boiling curve of stainless steel screen mesh
porous media is depicted in fig. 6. The 1% H2O-
Al2O3 nanofluids consistently shows a better heat
transfer compared to other fluid until the heat flux
reach 30 kW/m2 with the temperature difference
was 3.67 0C. When the highest heat flux 26 kW/m
2
applied, the base fluid shows the better heat
transfer compared to other fluid. When the highest
heat flux applied, the temperature difference
between heater wall and fluid for water, 1%, 3%
and 5% H2O-Al2O3 nanofluids were 1.89 K, 2.62
K, 2.99 K and 3.16 K respectively. It means that
the base fluid provide a better heat transfer
compared to the nanofluids.
Fig.6. Boiling curve of stainless steel screen
mesh
The same phenomenon also observed by Sarit K.
Das et.al [14, 15] where they found that the
addition of nanoparticle to the base fluid deteriorate
the heat transfer.
The boiling curve of copper screen mesh was
depicted in fig. 7. The 1% and 3% H2O-Al2O3
consistently shows a better heat transfer compared
to the 5% H2O-Al2O3 and base fluid. The 1% and
3% H2O-Al2O3 have almost a similar heat transfer.
The temperature difference between the heater wall
and the fluid of The 1% and 3% H2O-Al2O3 were
0.57 K and 0.54 K.
Fig.7. Boiling curve of copper screen mesh
3.3 Surface roughness measurement
Table 3 shows the surface roughness of 300 μm
sintered copper surface before and after the boiling
process with nanofluids. The measurement results
show the changes in the surface roughness after
boiling of nanofluids. The surface characteristic of
the porous media were measured using the
profilometer.
Table 3. The surface roughness of sintered copper
porous media before and after boiling Measurement
Point
Ra Before
Boiling [µm]
Ra After
Boiling[µm]
1 14.38 5.46
2 11.7 7.36
3 12.18 7.1
4 13.18 11.22
The addition of nanoparticles deposit on the surface
of the porous media can be one of the causes in the
decrease of the surface roughness, since the
nanoparticles has a smaller particle diameter
compared to the powder that being used as the
porous media.
4. Conclusion
From the result of the experiment, it can be
concluded that :
The 5% H2O-Al2O3 shows a heat transfer
that no better than the other fluids. The 5%
H2O-Al2O3 always shows a higher
temperature difference of the heater wall and
fluid compared to other fluids.
The 400 µm sintered copper consistently
shows a better heat transfer compared to
other porous media.
The decreased of the surface roughness of
the porous media was found after the boiling
process. It can be said that the additional
layer of nanoparticles appear on the surface
of the porous media. This phenomena was
possible due to the smaller size of the
nanoparticle compared to the porous media.
5. Acknowledgement
The author would like to thank DRPM UI for
funding this research. The first author would like to
thank the Department of Mechanical Engineering,
University of Indonesia through the IMHERE
(Indonesian-Managing Higher Education for
Relevance and Efficiency) for the scholarship.
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6. References
[1] B. Pulvirenti, A. Matalone and U. Barucca,
"Boiling heat transfer in narrow channels with
offset strip fins: Application to electronic
chipsets cooling," Applied Thermal
Engineering, vol. 30, pp. 2138-2145, 2010.
[2] S. Mori and K. Okuyama, "Enhancement of
the critical heat flux in saturated pool boiling
using honeycomb porous media," International
Journal of Multiphase Flow, vol. 35, pp. 946-
951, 2009.
[3] Z. Xu, Z. Qu, C. Zhao and W. Tao, "Pool
boiling heat transfer on open-celled metallic
foam sintered surface under saturation
condition," International Journal of Heat and
Mass Transfer, vol. 54, pp. 3856-3867, 2011.
[4] S. S. Murshed, C. N. d. Castro, M. Lourenco,
M. Lopes and F. Santos, "A review of boiling
and convective heat transfer with nanofluids,"
Renewable and Sustainable Energy Reviews,
vol. 15, pp. 2342-2354, 2011.
[5] R. A. Taylor and P. E. Phealan, "Pool boiling
of nanoflids: Comprehensive review of
existing data and limited new data,"
International Journal of Heat and Mass
Transfer, vol. 52, pp. 5339-5347, 2009.
[6] H. KIM, J. Kim and M. H. Kim, "Effect of
nanoparticles on CHF enhancement in pool
boiling of nano-fluids," International Journal
of Heat and Mass Transfer, vol. 49, pp. 5070-
5074, 2006.
[7] S. You, J. Kim and K. Kim, "Effect of
nanoparticles on critical heat flux of water in
pool boiling heat transfer," Applied Physics
Letters, vol. 83, pp. 3374-3376, 2003.
[8] Y. Takata, S.Hidaka, J. Cao, T. Nakamura, H.
Yamamoto, M. Masuda and T. Ito, "Effect of
surface wettability on boiling and
evaporation," Energy, vol. 30, pp. 209-220,
2005.
[9] B. Stutz, C. H. S. M. M. d. F. d. Silva, S.
Cioulachtijan and J. Bonjour, "Influence of
nanoparticle surface coating on pool boiling,"
Experimental Thermal and Fluid Science, vol.
35, pp. 1239-1249, 2011.
[10] E. Williamson, E. Forrest, J. Buongiorno, L.-
W. Hu, M. Rubner and R. Cohen,
"Augmentation of nucleate boiling heat
transfer and critical heat flux using
nanoparticle thin-film coatings," International
Journal of Heat and Mass Transfer, vol. 53,
pp. 58-67, 2010.
[11] W. N. S. H. R. R. I. Nandy Putra, "Thermal
performance of screen mesh wikc heat pipes
with nanofluids," Experimental Thermal and
Fluid Science, vol. 40, pp. 10-17, 2012.
[12] M. Kao, C. Lo, T. Tsung, Y. Wu, C. Jwo and
H. Lin, "Copper-oxide brake nanofluid
manufactured using arc-submerged
nanoparticle synthesis system," Journal of
Alloys and Compounds, Vols. 434-435, pp.
672-674, 2007.
[13] X. Wei, H. Zhu, T. Kong and L. Wang,
"Synthesis and thermal conductivity of Cu2O
naofluids," International Journal of Heat and
Mass Transfer, vol. 52, no. 19-20, pp. 4371-
4374, 2009.
[14] S. K. Das, N. Putra and W. Roetzel, "Pool
boiling of nano-fluids on horizontal narrow
tubes," International Journal of Multiphase
Flow, vol. 29, pp. 1237-1247, 2003.
[15] S. K. Das, N. Putra and W. Roetzel, "Pool
boiling characteristics of nano-fluids,"
International Journal of Heat and Mass
Transfer, vol. 46, pp. 851-862, 2003.
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Analysis of Small Bubble Characteristics in Alum Solution
Warjitoa, Nurrohman
b
aMechanical Engineering Departement,
University of Indonesia
Kampus Baru UI Depok, Indonesia 16424
[email protected] bMechanical Engineering Departement,
University of Indonesia
Kampus Baru UI Depok, Indonesia 16424
ABSTRACT
Waste of batik processing can raise Total
Suspended Solid (TSS) that exceeds water quality
standard. One technique that can be used to reduce
TSS is bubbles flotation. The effectiveness of
bubbles flotation depends on three parameters, i.e.
the probability collision between bubbles with
particles, the particles stick to the bubbles surface
and the particles carried by bubbles. These
probabilities are greatly influenced by the bubble
characteristics, i.e. diameter, rise velocity along
column, and terminal velocity. Understanding the
dynamics of bubbles is necessary in order to
increase the effectiveness of separation of the
flotation process. The purpose of this research was
to study characteristics of small bubbles (bubbles
with diameter of 0.2-1 mm), which rise in a liquid
column. Experimental set up was a column made of
an acrylic pipe with inner diameter of 90 mm and
length of 2000 mm. Small bubbles were generated
by copper cathode. The dynamics of bubbles were
observed using a video camera. Videos images
were processed using image processing software.
The results showed that at height of 500 mm from
cathode tip bubbles in average have reached its
terminal velocity. It has been proven that effect of
alum surfactant can reduce the bubbles terminal
velocity.
Keywords : bubbles, batik, waste, flotation,
velocity, alum
1. INTRODUCTION
Water is essential for living thing. There will be no
life on earth if there is no water. Industries that
develop rapidly cause reduction in availability of
clean water. These reduction is caused by pollution
of industries‘ waste. One of the pollution is waste
of batik processing.
Batik is one of spesific characteristics of
Indonesian. People in the world have admitted that
batik is one of Indonesian culture heritages.. This
has been proven with authentication by UNESCO.
Day after day batik industries develop. This
development makes waste of batik processing
increases. The waste of batik processing causes
increasing Total Suspended Solid (TSS) of water,
therefore exceed the water quality standard [1]. A
technique for decreasing TSS is needed therefore,
at least, the TSS does not exceed the water quality
standard. One of the techniques that can be used is
bubbles flotation.
Microbubble is bubble which is less than 200 µm in
diameter. Bubbles with more than 200 µm but less
than 1 mm in diameter are called ―small bubble‖,
whereas bubbles with more than 1 mm in diameter
are called ―large bubble‖. Bubble flotation is a
process conducted by producing bubbles from
bottom of water column contaminated by waste
particles. Bubbles will rise because of Bouyancy
force and collide with the particles. These particles
stick to the bubble and will be carried by the
bubbles to water column surface. When these
particles gather in the surface column, it can be
separated from the water easily.
The efficiency of the bubbles flotation depends on
three parameters, i.e. the probability of collision
between bubbles with particles, the particles stick
to the surface of the bubbles and the particles
carried by the bubbles [2]. The higher value of
these probablities, the higher the efficiency of the
flotation process. Microbubble prefers to be used
because it has larger surface area therefore the
probability of collision between bubbles and
particles can be increased.
Today, the aplication of bubbles in flotation is still
troubled by the understanding of the bubbles
characteristics. Therefore, it is necessary to do a
research that studies the characteristics of bubbles
rise in liquid column. These charactersitics include
diameter, velocity profile along liquid column, and
terminal velocity of the bubbles. This research was
to study the characteristics of ―small bubbles‖ rise
in liquid column with addition of surfactant, i.e.
alum or aluminum sulfate.
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2. THEORITICAL ASPECT
2.1 Bubble
Bubble is a particle whose dispersed phase is gas.
Bubble shape can be divided by three, i.e.
spherical, ellipsoidal and cap. Bubble motion can
be rectilinear, zigzag and spiral [3].
2.2 Dimensionless Numbers
The movement of the bubbles is usually in
terms of dimensionless numbers, i.e.:
(1) Reynods Number
Reynods number is ratio of inertia force to
viscous force.
Re = ρ.v.D/µ (1)
(2) Eotvos Number
Eotvos number is ratio of gravity force to
surface tension force.
Eo = g.Δρ.D/σ (2)
(3) Weber Number
Weber number is ratio of inertia force to
surface tension force.
We = ρf.vb2.D/σ (3)
(4) Morton Number
Morton is ratio of viscous force to surface
tension force.
Mo = g.µ4.Δρ/ρ
2.σ (4)
Where ρ, v, D, µ, Δρ and σ is liquid density, bubble
velocity, bubble equivalent diameter, liquid
dynamic viscocity, density difference between gas
inside bubble and liquid where bubble moves and
bubble surface tension [3].
2.3 Surface Tension
Surface tension works on surface plane, normal or
perpendicular to each line that works on surface
and its magnitude is the same at every point.
Surface tension will decrease at a certain
temperature on the surface of two substances and
when temperature rises [4]. Bubble surface tension
is defined as:
σ = pσ.R/2 (5)
where pσ and R are the surface tension pressure
and bubble radius [5]
2.4 Terminal Velocity
At the first, bubble will accelerate after leaving the
tip of the cathode. At some point, it will experience
terminal velocity. It is when the speed constant
when the influence of body force (gravity) and drag
force equal to the buoyant force. Sam et al. (1996)
has conducted research on single bubble velocity
and characterize a three-stage velocity as shown in
Figure 1 [6].
Figure 1: Velocity stages of bubbles rise in
liquid column
Hadamard-Rybczynski have formulated terminal
velocity in creep flow as follows:
u∞ = (6)
where u∞, g, r, Δρ, μ, and κ are the terminal
velocity, the acceleration of gravity, the bubble
radius, the difference in density bubble where the
bubble moves with the fluid, dynamic viscosity and
dynamic viscosity ratio of the particles to fluid. The
above formula applies to bubble with mobile
surface [7].
In addition to the above formula, Stokes (1880)
formulated an equation for bubble terminal velocity
where the surface does not move (immobile) as:
(7)
with u∞, de, ρl, ρg, and μl are the terminal velocity,
the equivalent diameter of the bubble, liquid
density where the bubble moves, the density of the
gas inside the bubbles and the dynamic viscosity of
the liquid where the bubbles move [8]. Davies-
Taylor (1950) have formulated an equation for
large bubble where the dynamic surface tension
and viscosity can be neglected:
(8)
with u∞ and de are the terminal velocity and the
equivalent diameter of the bubble [8].
2.5 Effect of Surfactant
Researchs have proven that the properties and
behavior of bubbles will change significantly in the
contaminated fluid. Effect of surfactant used to be
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compared in the term of terminal velocity of the
bubble. Surfactants tend to inhibit movement of the
bubble internal circulation thereby reducing the
terminal velocity. With a low terminal velocity, the
probability of attachment of particles can be
increased so that the flotation process efficeiency
increases.
2.6 Electrolysis
Hydrogen bubbles can be produced by separating
hydrogen and oxygen from water molecules by
electrolysis method. Electrical energy can be used
to separate hydrogen. It is made by supplying
electric current to an electrolitic cell. Thus,
electrolysis is the course of events of a chemical
reaction by an electric current. The tool consists of
electrolytic cell containing electrolyte (solution or
melt), and two electrodes (anode and cathode). In
the anode, oxidation happens while in the cathode
reduction happens. In an electrolysis experiment,
the reaction at the cathode depends on the tendency
of reduction.
2.7 Flotation
Bubbles attached to the hydrophobic particles and
carry the particles to the surface of the liquid, in
which particles are removed by using skimming
equipment. There are two classifications based on
the size of the bubble flotation. In the microbubble
flotation, bubbles are usually used 10-70 µm in
diameter. The bubbles are often generated by
depressurization of the dissolved air of liquid
(water dissolved liquid). This process is called
dissolved air flotation (DAF). In dispersed air
flotation, bubbles used was 1 mm.
The main key of flotation process is in bubbles-
particles capture which commonly known as a
series of three subprocesses. Total capture
efficiency is usually known as the product of three
successive steps, i.e. the collision efficiency, the
sticking efficiency, and the efficiency of stability
the bubble-particle aggregates [9].
3. EXPERIMENTAL METHOD
Experimental scheme is shown by Figure 2.
Figure 2: Experimental equipment scheme
3.1 Flotation Column
Column flotation was made of acrylic pipe with
inner diameter 8.4 cm, thickness 0.6 cm and height
of 200 cm. Water jacket was used to overcome the
optical distortions with dimension of 26.2 mm x
26.2 cm x 200 cm.
3.2 Surfactant and Water
The surfactant used was alum or aluminum sulfate
to the level of 100 grams per liter of water. Water
which was used to fill the acrylic pipe was drinking
water from AQUA. Meanwhile, water used to fill
the water jacket was tap water.
3.3 Bubbles and Surfactant Producing
Equipment
The bubbles were produced by electrolysis. The
electrodes are copper wire. For cathode, the wire
diameter was 0.1 mm, whereas for the anode the
wire diameter was 0.2 mm. DC Power Supply was
KPS3030DA of ATTEN instrument.
3.4 Camera Mechanisms
The camera was a Nikon D5000 with a Nikor 60
mm lens AF f/28D. Guide ways was used to trace
the bubbles as shown in Figure 3.
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3.5 Image Processing
Data processing was done using ImageJ software to
measure the size of the bubble and Frame Shot
software to determine the time required for the
bubble to move.
3.6 Lighting
For the lighting in order that the produced bubbles
images are good, the back lighting technique by
placing diffusive paper on the water jacket on the
opposite side of the installed camera was used.
3.7 Tools Set Up
At first, both edges of the electrodes were sanded to
remove the insulation. One edge was connected to
a DC Power Suply whereas the other edge was
inserted into the column flotation. The cathode is
inserted from the bottom of the column to generate
bubbles and the anode inserted from the top of the
column. Then the water from the AQUA mixed
with surfactant and acrylic was inserted into the
acrlylic pipe. Whereas tap water put into the water
jacket. Cameras mounted on one side of the water
jacket on the guide ways. To determine the size of
the bubble, some lines were given on the diffusive
paper. To determine the position of the bubble,
water jacket was marked on certain points. To
determine the distribution of bubbles, the camera
was placed at 12.5 cm, 25 cm, 75 cm, 125 cm, and
160 cm from the tip of the cathode to capture the
bubble at these points. The focus of the camera lens
was set manually whereas the exposure and shutter
speed was adjusted automatically.
Figure 3: Guide ways
3.8 Data Collection Procedures
To get the data distribution of bubbles, the camera
was placed at 12.5 cm, 25 cm, 75 cm, 125 cm and
160 cm. Once the camera was mounted on one
point, DC Power Supply switch was turned on for
about a second and then turned off quickly. The
generated bubbles were quite a lot and the picture
was taken for the image bubbles at the top, middle
and bottom. Then the stored image data were saved
to be processed using ImageJ software.
To get the bubble velocity data, the camera was
placed at the level of a 12.5 cm and was set to be
ready to record. Then DC Power Supply switch
was turned on then immediately shutted down. The
traced bubbles was the top one. It is because the
bubbles are more likely affected by the other
bubble on top of it. Then the data were stored to be
processed using the Frame Shot software.
4. RESULTS AND DISCUSSION
4.1 Bubble Size Distribution
From the data processing, for the distribution
of bubbles, it was obtained result as shown in
Figure 4.
Figure 4: Bubble size distribution graph
From the graph of Figure 4, it can be seen that
the higher the position of the bubble, the greater
diameter. This is because the hydrostatic pressure
decreases. The resulting bubbles ranging from 0.27
mm to 0.4 mm. All of them are small bubbles.
4.2 Bubble Rise Velocity
Bubble rise velocity was varied with voltage 3 and
7.5 volts. The result for voltage of 3 volts was
shown in Figure 5. As for the voltage of 7.5 volts,
the result was shown in Figure 6. From these data it
can be seen that most of the line tends to be
straight. This means that the bubble has reached
terminal velocity. For some graphs, the velocity at
a height of 25 cm is steeper than most of the other
points. This shows that the bubbles on a height of
25 cm still deaccelerate to reach terminal velocity
or by Sam et al. (1996) are in stage 2 to reach the
stage 3. For the graph with a voltage of 7.5 volts
bubble looks more in line with the results shown by
Sam et al. (1996). This is because the diameter of
the bubble bigger so it is easier to visualize it. From
the data processing the velocity of the average
bubble either to the voltage of 3 volts and 7.5 volts
The 5th IMAT, November 12 – 13th
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101
at a height of 25 cm were in the second stage, in
which the declaration occurs. While at a height of
50 cm, the average bubble has reached its terminal
velocity (stage 3).
Figure 5: Bubble rise velocity with voltage of
3 volts
Figure 6: Bubble rise velocity with voltage of
7.5 volts
4.3 Terminal Velocity
4.3.1 Comparison Between Generated Bubbles
From the data available, it can be determined the
estimated terminal velocity of the bubble
generated, i.e. by averaging velocity values at the
points that tend to be straight. From the data
processing, Figure 7 shows a graph of the results of
the bubble terminal velocity with voltage of 3 and
7.5 volts. It can be seen that the average terminal
velocity of bubbles generated by voltage of 3 volts
are smaller than those indicated by voltage of 7.5
volts. This is because the bubbles produced by a
voltage of 7.5 volts were bigger than the bubbles
generated by the 3 volts.
Figure 7: Graph of terminal velocity vs
diameter voltage of 3 volts and 7.5
volts 4.3.2 Effect of Surfactant
Graph of Figure 8 is the result of experiments
conducted by Huang et al. (2011) for a comparison
of the value of the terminal velocity of small
bubble in pure water to the results of this study [9].
The data experimental results by Huang et al. are
indicated by the circle. From experimental result by
Huang et al., the terminal velocity for bubbles with
the same size in this study was lower than those in
the result by Huang et al.. This shows the effect of
alum or aluminum sulfate contaminant in this
study. As can be seen from graph of Figure 8, for
pure water with bubbles of 0.5 mm in diameter, it
has terminal velocity about 10 cm/s, whereas the
results from this study as shown in the graph of
Figure 7 only ranged from 2 cm/s to 2.5 cm/s.
The 5th IMAT, November 12 – 13th
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102
Figure 8: Graph of voltage terminal velocity vs
diameter ratio
4.3.3 Comparison between Experimental Result
and Some Equations
Terminal velocity occurs when the drag and
bouyancy force are equal. Therefore, it can be
derived a simple equation of terminal velocity of
the bubble in such conditions by incorporating the
existing parameters. The equation in this question
is:
FB = Fd (9)
With the bouyancy force (FB) is formulated as:
FB = ρ.V.g (10)
with ρ is the density of the fluid where the bubble
moves, V is the volume of the bubble and g is the
acceleration of gravity. Whereas the drag force Fd
is formulated as:
Fd = CD.1/2.ρ.UT2.A (11)
where CD is drag coefficient and A is the surface
area in this case the value is equal to πR2, where R
is the radius of the bubble. By incorporating the
existing values into the equation, the equation of
terminal velocity of the bubble obtained is:
u∞ = 0.258 (12)
with a value judgment of CD based on graph CD vs
Re for solid sphere as shown in graph of Figure 9
[10].
Figure 9: Graph of CD vs Re for solid sphere.
From the graph of Figure 9, CD value that was
taken was 20 due to Re obtained from the results of
the study has value of 22 in average. By inserting
the CD values, Equation (12) was obtained. The
graph of Figure 10 is a comparison between the
experimental results, equation (7), (8) and (12) for
a voltage of 3 volts and the graph of Figure 11 for a
voltage of 7.5 volts.
Figure 10: Graph of comparison between
experimental result and some
equations for voltage of 3 volts.
Figure 11: Graph of comparison between
experimental result and some
equations for voltage of 7.5 volts.
The 5th IMAT, November 12 – 13th
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From graph of the Figure 10 and 11, either bubbles
produced by 3 volts or 7.5 volts, both of them are
between graph of equation (8) (Davies-Taylor) and
equation (12). However, for a voltage of 3 volts the
graph is more likely to approach the equation (12)
whereas for a voltage of 7.5 volts tends to be right
in between them. This is because bubbles generated
with a voltage of 3 volts are smaller than that
produced by a voltage of 7.5 volts. This is
reasonable because the Davies-Taylor equation is
for the application of a large bubble. Bubble graph
experimental results are all above the graph of
equation (12). This suggests that there may be
some motion on the surface of the bubble (mobile
surface) so that the value of its terminal velocity
greater than equation (8) where equation (8) is for a
solid sphere where there is no motion on the
surface (immobile surface) or maybe it was mobile
but the value of Re taken was too small for a
voltage of 7.5 volts so that the graph of
experimental result a little bit away from the graph
of equation (12). Stokes equations are far away
from the experimental results. It has been
mentioned that Stokes equation accurates for very
small bubbles, such as microbubbles.
5. CONCLUSION
A study on the characteristics of small bubble
in a solution of alum has been performed and
analyzed. From the analysis and data processing it
was shown that the bigger bubble the bigger its
terminal velocity. It was known that the average
bubble has reached its terminal velocity at a height
of 50 cm above the cathode. It has also been proven
that the effect of surfactant alum can reduce the
terminal velocity of bubbles. With the low value of
the terminal velocity, the efficiency of flotation
process can be improved.
REFERENCES
[1] F. Astuti, ―Pengolahan Limbah Cair Industri
Batik dengan Koagulan dan Penyaringan
(Studi Kasus di CV. Batik Indah Rara
Djonggrang)‖, Postgraduate thesis
Environtmental Science Program
Interdisciplineary Group, Gadjah Mada
University, 2004.
[2] K. A. Matis, Flotation Science and
Engineering, CRC Press, Inc., 1995.
[3] R. Clift, J. R. Grace, and M.E. Weber,
Bubbles, Drops, and Particles. New York:
Academic Press, Inc., 1978.
[4] C. Brücker, ―Structer and dynamics of the
wake of bubbles and its relevance for bubble
interaction‖, Phys. Fluids, 11, pp. 1781-179,
1999.
[5] T. G. Leighton, The Acoustic Bubble.
California: Academic Press, Inc., 1997.
[6] A. A. R. Mehrabadi, ―Effects of frother type on
single bubble‖, Postgraduate Thesis Department of
Mining, Metals and Materials Engineering,
McGill University, Montreal, Canada, 2009.
[7] L. Parkinson, R. Sedev, D. Fornasiero, and J.
Ralston, ―The terminal rise velocity of 10–100
µm diameter bubbles in water‖, Journal of
Colloid and Interface Science 322, pp. 168–
172, 2008.
[8] A. R. M. Talaia, ―Terminal velocity of a
bubble rise in a liquid column‖, World
Academy of Science, Engineering and
Technology, 2007.
[9] Z. Huang, D. Legendre, and P. Guiraud, ―A
new experimental method for determining
particle capture efficiency in flotation‖, Journal
of Chemical Engineering Science 66, pp. 982 –
997, 2011.
[10] J. D. Anderson Jr., Fundamentals of
Aerodynamics, Third Edition, New York: The Mc-
Grawhill Companies Inc., 2001.
The 5th IMAT, November 12 – 13th
2012
104
Effect of Hot Air Reservoir in the development of Vacuum freeze
drying M. Idrus Alhamid
1, Nasruddin
2, Muhamad yulianto
3
Mechanical Engineering Department University of Indonesia
Kampus UI Depok 16424, INDONESIA
Ph. +62 21 7270032, Fax. +62 21 727003
Email: [email protected] 1, [email protected]
3
ABSTRACT The Objective of this work is to know effect of
inserting hot air from reservoir to the process of
vacuum freeze drying. Tentacle of jelly fish as
sample with constant weight 50 g and placed at
container which isolated, the samples were freeze
dried with condition at experiment varying between
inserting and without inserting hot air at
temperature 27oC. The result of experiment shows
that while inserting hot air into vacuum freeze
drying make pressure rise in until pressure reach 40
mbar. And this phenomena make material
evaporation and this event cant be done in vacuum
freeze drying. And when without hot air reservoir
the pressure can reach 3.5 mbar and the subimation
can be done in this process. Vacuum freeze drying
process without hot air reservoir need time 12.5
hours and for vacuum freeze drying with hot air
reservoir need time 10.5 hour to drying 50 g of
jelly fish tentacle. From this experiment can be
concluded that for vacuum freeze drying with
inserting hot air need more ability of vacuum pump
specially in flowrate and ultimate vacuum.
Keywords : Vacuum Freeze Drying, Hot Air
Reservoir, Tentacle of Jelly Fish.
1. INTRODUCTION
Freeze vacuum drying is a dehydration process in
which water removed by sublimation of ice from
frozen materials directly at low pressure (Vacuum
pressure). Freezing and sublimation process starts
from outside surface and then to the material
recedes[1]. The others researcher also describes
that Vacuum Freeze drying (VFD) is an optimal
drying technology method because it maintains the
structure, nutrients, and color of the original
substance [2, 3, and 4]. Vacuum freeze drying
process consists of three processes that are freezing
process, primary drying and secondary drying [5].
In freezing process, actually use vacuum freezing
method. It is a freezing process based on the rapid
evaporation of moisture from the surface and
within the products due to the low surrounding
pressure below saturation pressure of the product.
This methode have a problem because of Vacuum
freezing caused high evaporation and mass loss,
this phenomenon for any product make a serious
damage [6]. In the other side, to reduce energy
consumption in vacuum freeze drying process,
manny of researcher has been declare the
innovation to solve this problem, that are : Adding
energy from electrical heating at lower and top
possition [7], adding energy from micowave to the
system [8, 9, 10], adding energy from infra red
radiatio [11], and the news one is adding energy
from heater from condenser‘s heat loss [12]. About
combination vacuum freezing and internal freezing
to reduce mass loss due to evaporation and also
using heater from condenser heat loss also have
been described to solve the problem in vacuum
freeze drying, and have a diagram can be seen at
figure 1 [13].
Figure 1. Diagram P-T of vacuum freeze
drying using internal freezing and
heater from condenser heat loss
There is one methode which is never done by other
research, using hot air reservoir to reduce energy
consumption. Fress air with high temperature
which entering the drying chamber will be
increasing the molar mass of air at drying chamber
and due to this phenomena create differences
mollar mass between material and air at drying
chamber and the moisture content the material will
be evaporate.
Due to the problems above, the objective of this
research is to analyze the efeect of hot air reservoir
at vacuum freeze drying process. The material in
this research using jelly fish (scyphomedusae) as
basic ingredient of medicine.
IMAT-UI 018
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2. MATERIAL AND METHODE
2.1. Sample Preparation
Before tested using vacuum freeze drying, jelly fish
is first stirred as seen in figure 2 and checking the
initial moisture content and weighed the Initial and
final mass measured using digital balance with
range 0–600 g and accuracy of 0.4%. For each
experiment, 50 g of samples was used on vacuum
freeze drying process.
Figure 2. Initial condition of jellyfish tentacles
2.2. Experimental Set Up
A compact vacuum freeze drying with internal
cooling and hot air reservoir from condenser heat
loss was designed, built and installed at
Department of Mechanical Engineering, University
of Indonesia (Fig 2 and 3). It consist of 3 Insulated
cylindrical, drying, cold trap and reservoir chamber
with a diameter and a length of 25 cm. Cold trap is
a cascade refrigeration with refrigerant at High
stage R22 and at low stage mixture between
HCR22 80% and CO2 20%. Cold coil installed at
drying, cold trap and reservoir chamber. Heat from
condenser‘s heat loss also installed at drying
chamber with same coil. To shut on and shut off
cold and heat coil using shut off valve. A pressure
transmitter PTX 1400 with an accuracy of 0.4%
and measure range 0-1600mbar was used to
measure pressure in the drying chamber. A tray of
material used from Teflon and insulated to keep the
heat transfer during process. The temperature at
drying chamber, cold trap, material, cascade
refrigeration system and reservoir were monitored
by thermocouple type K with accuracy 0.4%. Dial
pressure used to monitor pressure in cascade
refrigeration system. The thermocouple and
pressure transmitter connected to Data Acquisition
which is have number of slot 4, total power 15 w
and operating temperature -20oC until 55
oC.
Figure 3. Experimental Set Up
Figure 4. Schematic of experiment 1. Vacuum
pump , 2. Coldtrap, 3. Drying
Chamber, 4. Compressor at LS stage,
5. PHE, 6. Expansion valve (Needle
Valve), 7. Check Valve, 8.
Evaporator HS, 9. compressor HS,
10. Evaporator HS, 11. Expansion
valve HS (Needle Valve), 12.
Material tray, 13. Capilarry tube, 14.
Flow Meter, 15. Hot air reservoir,
16-26. Thermocouple, 27-28.
Pressure Transmitter
3. RESULT AND DISCUSSION
3.1. Temperature characteristics
The characteristic of product temperature during
vacuum freeze drying process can be seen at Figure
4. Product temperature devided into 3 region, that
is freezing region, sublimation region dan
secondary drying region as the other research
mentioned [14]. At process vacuum freeze drying
with internal cooling has longer time to take
material freeze than the vacuum freezing and this is
also make the energy consumption more higher
than vacuum freezing. As can be seen at the figure
1, primary drying occurs at constant pressure and
temperature. For vacuum freeze drying with
internal freezing and heating at room temperature
40oC, primary drying occurs at product temperature
-10oC. For Vacuum freeze drying with hot air
reservoir, primary drying occurs at product
temperature 1oC. At figure also can be seen that
vacuum freeze drying with internal cooling and
The 5th IMAT, November 12 – 13th
2012
106
heat from condenser heat loss take time longer than
vacuum freeze drying with hot air reservoir.
Figure 4. Product temperature during vacuum
freeze drying
3.2. Phase change During Process
At figure 5 can be seen the change phase of product
during vacuum freeze drying process. For process
with internal cooling and heat from condenser heat
loss product temperature reach he solid phase until
product temperature -20oC without mass loss due to
evaporation and after that change phase to gas
region by the sublimation process at temperature -
20oC and pressure of 3 mbar. This process is
occurs in all vacuum freeze drying prosess as
general. For process with hot air reservoir the
product temperature can not reach the ice / solid
region due to the pressure rise in after entering hot
air from reservoir. After that the product
temperature change phase from liquid to gas phase
at temperatur 1oC and pressure of 20 mbar, and
this process is evaporation not sublimation. This
process can not called as vacuum freeze drying but
vacuum drying.
Figure 5. Phase change diagram of product
during process
3.3. Cold trap Temperature
At figure 6 can be seen the temperature of coldtrap
at vacuum freeze drying process with internal
cooling and heating and also hot air reservoir. For
vacuum freeze drying process with internal cooling
the cold trap temperature has fluctuate graphing at
begining due to the refrigerant devided in to 2 part
for cold trap and drying chamber (double
evaporator) and also process of flash point. The
temperatur in this variation can rech of -25oC. For
vacuum freeze drying process with hot air reservoir
the temperature of cold trap can reach temperatur -
45oC, becuse in this process without internal
freezing and the function of coldtrap still on single
evaporator. All refrigeration condition at this
experiment is same condition, that is for High
Stage using refrigernt R22 and for Low Stage
mixture between HCR 22 (80% mass) and CO2
(20% masss)
Figure 6. Profile of cold trap temperature
3.4. Final Product and moisture content
Figure 7 and table 1 can be seen the final product
of vacuum freeze drying with variation internal
freezing and heater (A) and also hot air reservoir
(B). Product of vacuum freeze drying with internal
cooling and heater has color very white and smooth
this result due to the process of vacuum freeze
drying at sublimation process . For misture content
of material is approximately 0%. For process with
hot air reservoir the final product has color more
dark and moisture content is 0.15%
Figure 7. Final Product jelly fish
The 5th IMAT, November 12 – 13th
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107
Table 1. Mositure Content Final Product
No Variation Initial
mass
(g)
Final
Mass
(g)
Mositure
Content
1 Vacuum
freeze
drying with
internal
freezing
and heating
50 2.289 0.18%
2 Vacuum
freeze
drying with
Hot air
reservoir
50 3.25 2.1%
CONCLUSION :
1. Vacuum freeze drying at the process
devided 3 region, freezing, primary
drying / sublimation and secondary
drying
2. Adding hot air to the system can be
reducing drying time but the process
become evaporation not sublimation
3. Adding hot air to the system make the
moisture contaent higher than vacuum
freeze drying with internal cooling and
heating
ACKNOWLEDGMENTS
The authors acknowledge the financial support
from Ministry of Higher Education Indonesia
(DIKTI) by Strategis Nasional Grant with no
contract : 3398 / H2.R12/HKP.05.00/2012
Reference :
[1] A.B. Edinara, R.M. Filho, E.C.V. De Toledo,
Freeze drying process : real time model and
optimization, Elsavier International Journal
Chemical Engineering and Process 43 (2004)
1475-1485
[2] George.J, P., A,K,Datta., 2002. Development
and validation of heat and mass transfer
models for freeze-drying of vegetable slices.
Elsevier Journal of Food Engineering (52),
89-93
[3] Chakraborty,R,.,A,K,Saha.,P, Bhattacharya.
2006. Modeling and simulation of parametric
sensitivity in primary freeze-drying of
foodstuffs. Elsevier Separation and
Purification Technology (49), 258-263
[4] Ghio, S., A,A, Barresi, G, Rovero.2000. A
comparison of evaporative and conventional
freezing prior to freeze-drying of fruits and
vegetables. IChem Journal., 0960-3085
[5] George-Wilhelm Oetjen., Peter Haseley.,
2004. Freeze Drying Second, Completely
Revised and Extended Edition. WILEY-VCH
Verlag GmbH&Co. KGaA, Weinheim ISBN:
978-3-527-30620-6
[6] Jackman, Patrick., Da-wen Sun., Jivun
Zheng., 2007. Effect of combined vacuum
cooling and air blast cooling on processing
time and cooling loss of large cooked beef
joints. Elsevir International Journal of Food
Engineering. (32). 266-271.
[7] Belyamin, Tambunan, A.H., Hadi, K.,
Purwadaria, & M.I. Alhamid, 2007, ―The
application of freeze vacuum and heating
from top and bottom on vacuum freeze
drying‖, Journal of Agricultural Engineering,
Association of Agricultural Engineering
Indonesia, Vol. 21, 235-248
[8] Wang, Rui., Min, Zhang.,Mujumdar. 2010.
Effects of vacuum and microwave freeze
drying on microstructure and quality of potato
slices. Elsevier Journal of Food Engineering
(101), 131-139
[9] Duan, X., Zhang, M., Li,X., Mujumdar, AS.,
2008b. Ultrasonically enhanced osmotic
pretreatment of sea cucumber prior to
microwave freeze drying. Drying Technology
26 (4), 420-426
[10] Huang, Lue-lue., Min, zhang., M,Mujumdar.,
Rui, X,L. 2011. Comparison of four drying
methods for re-structured mixed potato with
apple chips. Elsevier Journal of Food
Engineering (103), 279-284
[11] Chakraborty, R., M. Bera., P.
Mukhopadhyay., P. Bhattacharya. 2011.
Prediction of optimal condition of infrared
assisted freeze-drying of aloe vera (Aloe
barbadensis) using response surface
methodology. Separation and purification
technology. 80 (2011) 375-384.
[12] Nasruddin., M. Idrus Alhamid., Engkos A
Kosasih., M. Yulianto. 2011. Effect of Freeze
Vacuum Drying and Heating from
Condenser‘s Heat Loss on Drying Rate and
Microstructure of Aloevera. Research Journal
of Applied Sciences 6 (5) : 335 - 343
[13] M. Idrus Alhamid, Muhamad Yulianto,
Nasruddin, Engkos A. Kosasih. 2012
Development of a Compact Vacuum Freeze
Drying for Jelly Fish (Schypomedusae).
Jurnal Teknologi UTM-Malaysia 58 (2012)
Supl 1, 25 – 32. ISSN 0127-9696
[14] May, JC.. 2004. Freeze-Drying /
Lyophilization of Pharmaceutical and
Biological Products (Second Edition, Revised
and Expanded). Marcel Dekker, Inc. ISBN :
0-8247-4868-9
The 5th IMAT, November 12 – 13th
2012
108
Experimental of Cascade Refrigeration System Using Natural
Refrigerant Mixture Ethane and Carbon Dioxide at Low
Temperature Circuit and Natural Refrigerant Propane at High
Temperature Circuit
Nasruddin*, M. Idrus Alhamid, Darwin R.B. Syaka and Arnas
Refrigeration and Air-Conditioning Laboratory, Mechanical Engineering Department –
Faculty of Engineering - University of Indonesia, Kampus UI Depok, 16424, Indonesia *E-mail: [email protected]
ABSTRACT Medicine and biomedical research activities require
cold storage (cold storage) to store biomedical
specimens such as, for example, stem cells (stem
cells), sperm, blood and other organs. During
storage, to prevent the specimen from damage
required a special cold storage reaches -80oC [1].
Using single cycle refrigeration machine can only
reach -40oC, and performance deteriorates below -
35oC drop in pressure associated with evaporation.
Thus, to reach lower temperatures, use cascade
refrigeration machine [2]. During this low-
temperature circuit cascade refrigeration systems still
use refrigerants that contain ozone-depleting or
global warming (CFCs and HCFCs). To overcome
this, a mixture of carbon dioxide and ethane
azeotropis a promising alternative refrigerants.
Simulation studies and experiments indicate a mixture
of carbon dioxide and ethane were able to achieve the
minimum temperature to -80oC [4-7]. With the mass
ratio 70% R170 and 30% R744 circuit at low
temperature refrigeration systems and uses a
capillary tube expansion device 0.054 inch diameter
with a length of 6 meters and 3 meters then use an
electric heater as the cooling load. Cooling load is
given by the variation of 90 W, 120 W and 150 W at a
cabin in the low temperature circuit. From the
experiment will be known characteristics of cascade
refrigeration system with refrigerant mixture and will
get the parameter data to make cascade refrigeration
machine.
Keywords : Cascade, Refrigerant, Ethane, Co2,
Capillary tube
1. INTRODUCTION
Cascade refrigeration system consists of at least two
refrigeration systems that work independently. Two
refrigeration systems are connected in cascade heat
exchanger where the heat is released in the condenser
circuit low temperature (low temperature circuit /
LTC) is absorbed from the evaporator temperature
circuit (high temperature circuit / HTC) [3]. During
this time at a low temperature circuit used CFC
refrigerants such as R13 or R503 banned for
damaging the ozone layer. Meanwhile, HFC
refrigerants such as R23 alternatives although it does
not contain ozone-depleting substances, but the cause
of global warming. So, look for alternative
refrigerants directed on natural refrigerants and one of
which is carbon dioxide [5].
Carbon dioxide has an advantage because it is not
toxic, not flammable (non-flamable), easy to get, no
ozone depletion potential and very low global
warming [5]. However, the high pressure and
temperature triple preclude the use of carbon dioxide
when used for low-temperature circuit [6]. The
solution to overcome this shortcoming is to mix
carbon dioxide with other natural refrigerants are
hydrocarbons. To overcome this, a mixture of carbon
dioxide and ethane azeotropis a promising alternative
refrigerants. Simulation studies and experiments
indicate a mixture of carbon dioxide and ethane were
able to achieve the minimum temperature to-80oC [4-
7].
If the alternative refrigerant is used in a refrigeration
system, each component of the system must be
redesigned for reliability and high efficiency. In this
particular cascade refrigeration system at low
temperature circuit using the capillary tube expansion
device. The capillary tube is a tool used in the
expansion generally makes a small cooler as air
conditioning, refrigeration and cold storage, because it
is cheap, simple and reliable [8].
This study aims to develop a low temperature cold
storage for applications in the biomedical field using a
mixture of carbon dioxide and ethane refrigerants that
have high energy efficiency and safe that has a low
flammability and non-toxic for use in low-temperature
circuit in cascade refrigeration system.
2. METHODS
Cascade refrigeration machine consists of two
refrigeration circuits, a circuit called the refrigeration
circuits of high temperature (high temperature circuit /
HTC) that will contain environmentally friendly
refrigerant propane (R290) and the low temperature
circuit (low temperature circuit / LTC). which will be
filled with the refrigerant mixture R744/R170 the
azeotropic composition.
IMAT-UI 019
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Refrigerant is a closed cycle of cascade refrigeration
system comprising a compressor, oil separator,
kondesnser, PHE, filter dryer, expansion valve,
evaporator, and accumulator, can be seen in Figure 2.
At the position of the high temperature refrigerant
discharge circuit and oil through the oil separator and
only high-pressure refrigerant gas phase can exit to
get to the condenser. Refrigerant is cooled by the
ambient temperature so that the refrigerant out of the
condenser has a temperature equal to the ambient
temperature. After going through the condenser
refrigerant dryer filter before heading into the
expansion valve. Refrigerant is expanded according to
the temperature needed to cool the cascade condenser
or heat exchanger between the high-temperature
circuit with low temperature circuit. Heat exchanger
that is used is a type of PHE (plate heat exchanger).
Out of PHE refrigerant back to the compressor.
Temperatures were measured at high temperature
circuit only four positions are out compressor,
condenser exit, entry and entry kompreser PHE.
Temperature measurement using a thermocouple type
K with the value of reading ± 0.14% accuracy. While
pressure is measured only on the position of the
compressor discharge and suction. Measurement of
pressure using a pressure transmitter type Druck PTX
1400 readings with an accuracy of ± 0.15%.
Figure 1. Scheme diagram of refrigeration
cascade system At the position of the low temperature circuit
refrigerant compressor exit directly to the condenser
to be cooled to ambient temperature, it aims to ensure
that oil enters the oil separator in the liquid phase so
as not to be drawn into the whole system. Once out of
the oil separator directly cooled condenser refrigerant
cascade (PHE) so that the temperature is low. Then
the refrigerant through the filter dryer before heading
capillary tube to be expanded according to the desired
temperature. At this low temperature circuit system
desired temperature is -80oC. Very low temperature
refrigerant goes to the evaporator. Evaporator is inside
the cabin where the cabin is made from a fan to blow
air into the evaporator. Then the refrigerant flow to
the accumulator before getting back into the
compressor to be pressed. Temperature measurement
at low temperature circuit is positioned in the
compressor exit, entry PHE, PHE exit, before the
capillary tube, evaporator entry, exit evaporator,
inside the cabin and into the compressor. Temperature
measurement using a thermocouple type K with the
value of reading ± 0.14% accuracy. For the
measurement of the pressure placed on four parts:
compressor exit, enter the capillary tube, out of the
capillary tube and into the compressor. Measurement
of pressure using a pressure transmitter type Druck
PTX 1400 readings with an accuracy of ± 0.15%.
The air inside the cabin is heated by a 500 watt heater
is blow by a fan. Heater is controlled by a dimmer to
create a stable electrical current into the heater that
will be used as the cooling load. Pictures of the cabin
can be seen in Figure 3. The flow of air in the cabin is
measured by anemometer. The temperature is
measured with a thermocouple type K. The data
results from performance refrigeration system are
pressure, temperature and mass flow rate. The data
obtained will be used as the design and operating
parameters of cascade refrigeration system.
T
TEvaporator
Heater
Figure 2. Scheme of cabin
3. RESULTS AND DISCUSSIONS
The results of this data collection illustrate the
character of the cascade refrigeration system circuit
high temperature and low temperature circuits.
Refrigeration system has a small capacity so it can be
put in the room. Refrigerant substitute experiments
were carried out only with the modification to the
capillary. The experiment was started with Ethane in
LTC to set up the base reference for further
comparisons with new mixture under identical
working conditions while HTC was kept at same
condensing and evaporating pressure. The diameter of
capilary tube is 0.054 inch and its length is 3 meter
dan 6 meter.
The 5th IMAT, November 12 – 13th
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110
Figure 3. Variation of cooling load (heater) with
the refrigerant temperature at the
evaporator inlet in LTC.
In Figure 3. It can be seen the influence of the cooling
load on the evaporator entrance temperature low
circuit temperatures. The addition of cooling load
resulting evaporator temperature inlet increases.
Using a capillary tube 0.054 inch and a length of 6
meters evaporator inlet temperature is -80.6oC when
given cooling load by 90 W and the temperature
continues to increase to-78.4oC when the cooling load
is added to 150 W. The same capillary with a length
of 3 meters and temperatures higher evaporatornya-
70.5oC when the cooling load is given at 90 W and
continue to increase with the cooling load is added.
the evaporator inlet temperature by using a capillary
tube 6 meters is lower than the capillary length 3
meters and it is shown that the longer of capillary tube
the lower evaporator inlet temperature. The
evaporator inlet temperature affects the temperature in
the storage room or cabin and can be seen in Figure 4
below.
Figure 4. Relation of evaporator inlet
temperature with room storage
temperature.
The temperature of the storage room have the same
tendency to the evaporator inlet temperature at low
circuit temperatures when increasing the evaporator
inlet temperatures rise due to the cooling load
increases. At LTC using a capillary tube with a length
of 6 meters shown in figure 4. temperature of storage
room when load 90 W is -75.7oC and the evaporator
inlet temperature is -80oC and then the cooling load
was increased to 150 W resulted in a storage room
temperature is 71.9oC and the evaporator inlet
temperature is -78.4oC. The capillary tube length 3
meters has a trend similar to the capillary tube length
of 6 meters where the temperature difference between
the evaporator inlet temperature and the storage room
temperature increasing when the cooling load is
increases.
In figure 5 seen influences of the cooling load at the
evaporator inlet pressure that has the same tendency
to influence the cooling load on the evaporator inlet
temperature continues to rise when the cooling load is
increases. The capillary tube length 6 meters has a
expanssion pressure lower than the capillary tube 3
meters.
Figure 5. Variation of cooling load (heater) with
the refrigerant pressure at the
evaporator inlet in LTC.
Figure 6 below shows the effect of the cooling load
against the mass flow rate refrigerant in the LTC.
Where the mass flow rate of refrigerant adjusts
cooling load that increases with the mass flow rate
refrigerant increase in both LTC using a capillary tube
with a length of 6 meters and a length of 3 meters. In
Figure 7 below shows the relationship between the
evaporator inlet temperature and the evaporator inlet
pressure at LTC.
Figure 6. Variation of cooling load (heater) with
the refrigerant mass flow rate in LTC.
Figure 7. Relation of evaporator inlet
temperature with evaporator inlet
pressure.
The 5th IMAT, November 12 – 13th
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111
In Figure 8 below shows the effect of the cooling load
of the discharge temperature. Where the trend is
continues to increase when the cooling load is
increases.
Figure 8. Variation of cooling load (heater) with
the refrigerant temperature at the
discharge in LTC.
The relationship between the discharge temperature
and the discharge pressure can be seen in Figure 9
below, the capillary tube between 6 meter and 3 meter
delta temperature difference seen when the cooling
load was increase very different where delta
temperature capillary tube with a length of 6 meters
which is approximately 1oC while capillary length 3
meter has delta temperature is 0.1oC.
Figure 9. Relation of discharge temperature with
discharge pressure.
At high temperature circuit has a tendency similar to
the low circuit temperature which increases the
temperature when the cooling load in the low circuit
temperature increases. Graph influences of the cooling
load at HTC can be seen in figure 10 and 11 below.
Figure 10. Variation of cooling load (heater) with
the refrigerant temperature at the
cascade condenser in HTC.
Figure 11. Variation of cooling load (heater) with
the refrigerant temperature at the
discharge in HTC.
4. CONCLUSION
The experimental results using mixture refrigerant of
ethane and carbon dioxide at low temperature circuits
and propane at high temperature circuit and then using
0.054 inch diameter capillary tube and by varying the
cooling load using an electric heater and getting the
cascade refrigeration system characteristics are:
• Pressure, temperature and mass flow rate at low
temperature circuit will increase when the length of
capillary tube is lower and enhanced cooling load.
• By using a capillary tube 0.054 inch and a length of
6 meters at LTC it will produce 1.9 bar expansion
pressure and the evaporator inlet temperature is -
80.6°C, the cabin or storage room temperature is -
75.7oC and mass flow rate is 6.2 x 10
-4 kg/s with
cooling load is 90 W.
ACKNOWLEDGMENTS
This research was supported by Hibah Bersaing 2011
and Hibah Kompetensi Tahun 2012, Direktorat
Jenderal Pendidikan Tinggi, Kementerian Pendidikan
Nasioal, Republik Indonesia.
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Rahadiyan, 2009, Utilization of CO2/Ethane
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System and Applications (SI), American
Society of Heating, Refrigerating, and Air-
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Refrigeration, 29 (2006):1100-1108;
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G.G, 2007, carbon Dioxide for supermarkets,
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[11] Cox.N, 2007, Working towards more
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Condenser In CO2/NH3 Cascade Refrigeration
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Refrigeration, ICR07-B2-454, Beijing, 2007
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Teknik Universitas Indonesia, Depok.
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2012
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Performance Analysis of Thermoacoustic-Standing Wave As a Power Generator
Surjosatyo A, Sentosa I Mechanical Engineering Department, Universitas Indonesia, Kampus Baru UI, Depok, Indonesia
Email : [email protected]
ABSTRACT
This study is relating to analyze performance of
thermoacoustic-standing wave. Stirling cycle
thermoacoustic engine is developed conventional
stirling engine. This system is more efficient than
ordinary stirling engine because does not use a
moving piston[7]
. The engine uses thermal power to
generate acoustic power. It consists mainly of
three parts: a thermodynamic part consisting of a
stack, two heat exchangers, and a thermal buffer
tube; an acoustic network consisting of an acoustic
compliance and an inertance; and a resonator.
When thermodynamic part heated, it will generate
sounds. The sounds will flow along cylinder tube.
Some aspects can be analiyzed to determine
performance of tharmoacoustic-standing wave.
The effect of temperature difference, stack
geometry, stack position determine performance of
the thermoacoustic-standing wave. Some research
show that acoustic power will increase with
increasing of temperature at hot heat exchanger.
And optimal position and geometry of stack will
generated optimal acoustic power.
Keywords: Thermoacoustic-standing wave, stack,
sound, acoustic power
1. INTRODUCTION
Thermoacoustics is a field that studies the
conversion of heat to acoustic energy. Research on
thermoacoustics can be dated back to the late 19th
century. Modern research of thermoacoustic
systems is largely based on the work of Rott,
Steven Garrett and Greg Swift, in which linear
thermoacoustic models were developed to form a
basic quantitative understanding and numeric
models for calculation. The thermoacoustic engines
contains no moving parts yet the acoustic
stimulation of heat flux and the generation of
acoustic work, point to some type of timed phasing
of thermodynamic process. This phasing in
thermoacoustic engines is due to the presence of
two thermodynamic media fluid and stack plate.
the temperature and pressure oscillations induce
sound waves. The combination of all such process
produces an affluent ‗‗thermoacoustic‘‘ effects.
The applications of thermoacoustics are widely
spread in the gas mixture separation, natural gas
liquefaction, heat pumps, pulse tube,
thermoacoustic regenerator and thermoacoustic as a
electrical generator.
The experiment about performance of
thermoacoustic standing-wave engine has been
studied. Hariharan, P. Sivashanmugam, S.
Kasthurirengan Influence of stack geometry and
resonator length on the performance of
thermoacoustic engine. The results obtained from
the experiments are in good agreement with the
theoretical results from DeltaEc.
2. THERMOACOUSTIC COMPONENT
The open end standing wave thermoacoustic engine
consists of parts such as heat exchangers, stack,
resonator, and working fluid.
2.1. Heat exchangers
The function of heat exchangers in a
thermoacoustic engine is to transfer heat from an
external source to the working fluid in the sealed
resonator chamber and they are used to maintain
the temperature gradient across the stack. The
active heat exchange takes place between the
working fluid and a series of closely spaced,
parallel plates with their surfaces aligned with the
direction of the wave propagation and positioned at
either end of the stack. The heat exchanger should
provide high heat transfer coefficient and low
acoustic power dissipation to the thermoacoustic
side. The hot heat exchanger supplies heat to hot
end of the stack and ambient heat exchanger
extracts heat from other end of stack. The blockage
ratio is considered as same as that of stack so plate
size and spacing used for heat exchanger is
identical to that of stack for the present system.
This allows the gas parcels to move freely from
heat exchanger to stack. With the assumption of
same heat transfer coefficient and temperature
difference between solid plate and the working
fluid, the hot heat exchanger requires more heat
transfer area compared to ambient heat exchanger.
So the length of hot heat exchanger is chosen as
twice the length of ambient heat exchanger. The
optimum PS and length of heat exchanger, which is
equal to the peak to peak displacement of the
working gas is given by the following expression:
y0 = 2la/A
lc = sin(kl)
2.2. Stack
Stack is the heart of standing wave engines, where
the thermoacoustic cycle is generated. It provides
solid heat capacity and large cross sectional area to
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maintain a good thermal contact between gas and
solid stacks. They are finely divided in small
parallel channels with hydraulic radius comparable
to thermal penetration depth. The stack is placed at
a certain location in the resonator, where the
magnitude of local acoustic impedance |Z| is larger
than qga/A, because the magnitude of gas velocity
amplitude is relatively small to reduce the viscous
dissipation of viscous power. Acoustic impedance
is defined as the ratio of complex pressure to
volumetric flow rate. A good stack should
minimize the ordinary heat conduction along
temperature gradient and viscous dissipation of
acoustic power. The minimum thickness of the
stack plate should be 8ds, where ds is solid thermal
penetration depth, which is defined as
𝛿s =
the above equation is used to calculate the plate
thickness used for the present system which states
that the thermoacoustic effects are optimal if the
plate thickness is in the range of 6–8ds. The
thermal penetration depth is defined as the layer
around the stack plate where the thermoacoustic
phenomenon occurs. It is measured perpendicular
to the direction to the motion of gas and it gives
approximately the distance that the heat can diffuse
through the gas. Gas thermal penetration depth is
𝛿s =
by placing heat exchangers at either side of the
stack, heat can be moved so that the temperature
difference across the stack is created. As a result
sound wave can be induced. In order to maximize
the hydrodynamic heat flow, the stack material
with large Ksqscs in comparison to Kgqgcp is
favorable. Successful operation of a standing wave
engine requires an imperfect thermal contact
between the gas and the stack which is obtained
when the spacing between the plates is roughly two
to four times of thermal penetration depth of gas.
Viscous penetration depth is defined as the
thickness of the layer of fluid around the stack plate
that is restrained in its movement under the
influence of viscous forces. Within this layer,
viscous dissipation is responsible for the loss of
kinetic energy, so that the fluid layer of thickness
dv in the vicinity of each stack plate contributes
less to the thermoacoustic effect.
2.3. Resonator
The resonance tube is one of the key components
of a thermoacoustic engine. A smooth, linear
cylindrical resonator pipe without steps,
misalignments and abrupt transitions should be
used to avoid unwanted eddying or non-linear
pressure variations that would greatly complicate
the analysis. Resonance frequencies are mainly
determined by the length of the resonator.
Prolongation of resonance tube may leads to
decrease of working frequency and increase of
stacks hot end temperature with the same heating
power. The velocity amplitude increases from the
heater to the water cooler with a certain length of
the resonance tube, because the heater is closer to
the velocity node. On the other hand, when the
resonance tube is prolonged, the relative location of
the thermoacoustic core shifts nearer to the velocity
node so the velocity amplitude in the
thermoacoustic core decreases [18]. For lowest
dissipation, resonator should provide sufficient
inertance and compliance, thereby maintaining
resonance frequency, while simultaneously
minimizing the acoustic power dissipation. For
thermoacoustic engine the resonance frequency can
be estimated from
|U1|2 - |p1|
2
The above expression gives the acoustic power
dissipation per unit length of the channel, due to
thermal and viscous processes at the channel walls.
To avoid the thermal relaxation losses, tube
material with the combination of smallest possible
Ksqscs and gas with largest possible combination
of Kgqgcp should be selected. Easiest way to
decrease the acoustic power losses is to decrease
the surface area of resonant tube walls. In order to
decrease the acoustic power loss, k/4 wavelength
resonator has been chosen for the present study.
The equation for normalized acoustic power
dissipated in the quarter wavelength resonator is
estimated
ΔĖ2n,r ≈ -
The energy dissipated in the resonator is
proportional to wall surface area of resonator.
2.4. Working fluid
The choice of working fluid especially gas for a
thermoacoustic engine is an important aspect to be
considered as it affects power and efficiency. Gas
properties play an important role in determining the
onset temperature difference. The lightest gases
have highest sound speeds and high thermal
conductivity which will give highest powers due to
high thermal penetration depth, since heavier gases
condense or freeze at low temperatures or exhibit
non-ideal behavior. Gases with high ratios of
specific heats and low Prandtl numbers are well
suitable for thermoacoustic devices. These
The 5th IMAT, November 12 – 13th
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115
properties can be optimized by the use of a mixture
of light and heavy noble gases. The optimum
mixture of gases in a thermoacoustic engine
depends upon the application and design goals. For
thermoacoustic engines, gases with Prandtl nearer
to one must be used to attain the minimum onset
temperature difference [21]. As the present
experimental setup is an open end thermoacoustic
primemover, air is chosen as working fluid.
3.a. Experimental Set-up
Figure 3 shows the experimental set-up of
thermoacoustic-standing wave. The experiments
are perform the influence of temperature difference
toward acoustic power generated.
Subject Description
Tube Diameter
Tube Lenght
Tube Material
Stack(Regenerator)
Material
Stack Lenght
Amplifier Diameter
Amplifier Lenght
12.5mm
150 mm
Glasses
Steel
35mm
100mm
280 mm
Heater
Sound
Meter
Ambient Heat
Exchanger
Figure 3. Thermoacoustic Devices set-up for
measurment
The sound oscilation will generate due to
temperature differences. This oscilation will be
measured with sound meter and temperature in
both hot side and cold side are measured using
termocouple by national intsrument.
3. RESULT AND DISCUSSION
3.a. Acoustic Power
Acoustic power is measured using sound meter.
The data which will converted into acoustic power
is sound intencity (I). The formula is:
β = 10 log
where β is intencity degree(dB), I is sound
intensity(W/m2) and Io is ambient sound
intensity(10-12
W/m2). Further an acoustic power P
(Watt) will obtained by multiplication between an
is an area which through by sound A (m2) and
sound intencity I. Given by equation:
P= I.A
The acoustic power could to ceonverted into
electrical energi using piezoelectricity. ‗Piezo‘ is a
Greek term meaning to apply pressure to, or to
press. Piezoelectric, therefore, refers to the way in
which certain materials can generate a current
when pressure is applied to them. Piezoelectric
charge coefficient values were in the range 1-100
pico coloumb / Newton. Natural Piezoelectric
materials such as: quartz (Quartz, SiO2), berlinite,
tourmaline and salt Rossel. Made of piezoelectric
material are: Barium titanate (BaTiO3), Lead
zirconium titanate (PZT), Lead titanate (PbTiO3)
and so on. The phenomenon of the piezoelectric
effect can be described as follows:
Figure 4. The phenomenon of the piezoelectric
effect (A) before gived pressure or
electrical field. B) giving electrical
field, lenght will increase. (C) giving
reverse field , lenght will decrease.
(D) giving pressure , induksi
polarisation and out tension
happened.
3.b. Figures and Tables Until present, the correlation between acoustic
power obtained and time as shown in figure 5
a)
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b)
Figure 5.a) Intencity degree versus time and b)
Acoustic power output wersus time
Figure 5 shows that intencity degree is a parameter
of sound generated. Intencity degree increases with
after hot side of tube flamed and stable after 75
seconds with about 78 dB. Power output will also
increase after flamed with average of 0.225mW .
The increasing of intencity degree and power
output because of temperature diferences. Within 1
until 75 seconds, temperature differences between
hot side and cold side was unstable. After 75
seconds, the temperature difference was obtained
stable so that the degree of intensity will also be
stable at that time.
Further axperiment was focused influencing of
temperature differences toward acoustic power
obtained. The temperature difference will
influences sound intencity generated. Figure 6
shows the temperature difference in every passing
time. Figure 6 shows also that temperature
differences increases which each passing seconds
and will stable after about 150 seconds. This
research will be adjusted with the thermoacoustic
device to be built.
Figure 6. Temperature hot side and cold side at
stack1]
4. CONCLUSION
A simplified theory gives a satisfactory estimation
for the intencity degrees and power acoustic
obtained. With possible applications of more
optimal stack materials and design, the
thermoacoustic efficiency of the engine can be
further increased. For more accurate determination
of acoustic parameters a pressure measurement
inside the resonator is desirable. Other directions
for the system improvement include a reduction of
the heat leak through the stack holder. The
integration of the engine with compact and efficient
combustors using electroacoustic transformers can
open a possibility to develope power systems pf
thermoacoustic.
5. ACKNOWLEDGEMENT
This experiment was supported by IMHERE
program, Mechanical Engineering of Universitas
Indonesia from 2011 to 2013. The author also
appreciate all students for their help with this
experiment.
5. REFERENCES
1] N.M. Hariharan, P. Sivashanmugam, S.
Kasthurirengan.2011.Influence of stack
geometry and resonator length on the
performance of thermoacoustic engine
2] Matveev, I,K., Najmeddin, S.T, Richards,
C.D.2008. Small Scale Thermoacoustic
Demonstrator. Proceedings of PowerMEMS
2008+ microEMS2008, Sendai, Japan.
3] Swift, GW. 2002. Thermoacoustics: A
Unifying Perspective for Some Engines and
Refrigerators. Sewickley, PA, Acoustical
Society of America
4] Trapp, C. Andrew., Zink, Florian. 2011.
Thermoacoustic heat engine modeling and
design optimization. Applied Thermal
Engineering;2518e2528
The 5th IMAT, November 12 – 13th
2012
117
5] Garrett, SL. 2005. Acoustic laser kit
instruction
6] Backhaus, Scott., Swift, Greg. 2002. New
Verieties Of Thermoacoustic. LA-UR-02-
2721, 9th International Congress on Sound
and Vibration.
7] Gardner, Catherine., Lawn, Chris. 2009.
Design of a Standing-Wave Thermoacoustic.
The sixteenth International Congress on
Sound and Vibration.
8] Garrett, S.L., Backhaus, Scott.. 2000. The
Power of Sound. The Scientific Research
Society.
9] Trapp, Andrew C., Zink, Florian. 2011.
Thermoacoustic heat engine modeling and
design optimization. Applied Thermal
Engineerin: 2518e2528
The 5th IMAT, November 12 – 13th
2012
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Factors Affecting Performance of Dual Fuel Compression Ignition
Engines
Mohamed Mustafa Ali a, Sabir Mohamed Salih
b
a Mechanical Engineering Dep., Faculty of Engineering Sudan University of Science and
Technology – Khartoum - Sudan
Tel : (00966)509370959. Tel : (00249) 911270089
E-mail : [email protected] b Mechanical Engineering Dep., Faculty of Engineering Sudan University of Science and
Technology – Khartoum - Sudan
Tel : (00249)912133761
E-mail : [email protected]
ABSTRACT
Compression Ignition Diesel Engine use Diesel as
conventional fuel. This has proven to be the most
economical source of prime mover in medium and
heavy duty loads for both stationary and mobile
applications. Performance enhancements have been
implemented to optimize fuel consumption and
increase thermal efficiency as well as lowering
exhaust emissions on these engines.
Recently dual fueling of Diesel engine has been found
one of the means to achieve these goals. Different
types of fuels are tried to displace some of the diesel
fuel consumption.
This study is made to identify the most favorable
conditions for dual fuel mode of operation using
Diesel as main fuel and Gasoline as a combustion
improver. A single cylinder naturally aspirated air
cooled 0.4 liter direct injection diesel engine is used.
Diesel is injected by the normal fuel injection system,
while Gasoline is carbureted with air using a simple
single jet carburetor mounted at the air intake. The
engine has been operated at constant speed of 3000
rpm and the load was varied.
Different Gasoline to air mixture strengths
investigated, and diesel injection timing is also varied.
The optimum setting of the engine has been defined
which increased the thermal efficiency, reduced the
NOx % and HC%.
Keywords : Dual Fuel Combustion; Thermal
Efficiency; Exhaust Emissions;
Mixture Strength; Injection Timing.
1. INTRODUCTION
Different combustion strategies were used to improve
thermal efficiency and reduce exhaust emissions of
diesel engines. It was started by optimizing the
combustion of diesel fuel alone through turbocharging
and high fuel injection pressures with electronically
controlled fuel quantity and injection timing. But still
this created unwanted exhaust emissions which was
attempted to control using EGR or through exhaust
after treatment including catalytic converters and urea.
That is to reduce NOx, HC, and CO as well as
particulate matter.
Recently the approach has been changed and more
thinking is towards changing combustion strategy by
fuel design, thus changing of fuel characteristics and
control of start of ignition [1]. This will lead to
introduction of another fuel into the engine either by
direct blending of diesel fuel or by partial pre-mixing
of fuel into air induction. In all cases this will require
different ignition delay than that of diesel fuel and
change the combustion pressure and temperature as
well as end gas properties. These approaches are
summarized as follows:
1- Partially Premixed Combustion (PPC): In-cylinder
fuel blending of gasoline with diesel fuel with
parameter sweeps included gasoline-to-diesel fuel
ratio, intake charge mixture temperature, in-cylinder
swirl level, and diesel start-of-injection timing. This
resulted in improved thermal efficiency and reduced
NOx and particulate matter (PM).[2,6 &8].
2- Use of fuel additives in a Reactivity Controlled
Compression Ignition (RCCI) combustion through
addition of the cetane improver di-tert-butyl peroxide
(DTBP) to pump gasoline. Unlike previous
diesel/gasoline dual-fuel operation of RCCI
combustion, this used a single fuel stock (gasoline) as
the basis for both high reactivity and low reactivity
fuels. The strategy consisted of port fuel injection of
gasoline and direct injection of the same gasoline
doped with a small volume percent addition of DTBP.
This resulted in higher thermal efficiency and NOx
and PM within the emission limits [3].
3- Fumigation of diesel engine using alcohol. Here
alcohol is either port injected or just introduced using
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a simple carburetor by modifying the air intake. The
burning of alcohol with injected diesel fuel replaced
diesel and at the same time reduced combustion
temperature which resulted in less NOx and PM
[4,5&7].
Gasoline is also used to fumigate diesel engines to
improve their efficiency and reduce emissions. In this
study gasoline will be introduced in the air intake by
adding a single jet carburetor specially designed to
provide extra lean mixture. Different jet sizes were
used to vary the mixture strength, and the effect on
efficiency and emissions is evaluated.
Another control factor affecting the performance is the
diesel start of injection timing. This is also
investigated in order to define the trends and rules
which will lead to better performance of diesel engines
under dual fuel mode of operation.
2. ENGINE MODIFICATION AND DUAL
FUELING
A single cylinder diesel engine with specifications
mentioned in table 1, has been modified by adding a
carburetor into its air intake. The engine is directly coupled
to AC alternator which generates electricity at constant
rotational speed of 3000rpm.
Table 1. Diesel Engine Specification
Type Single Cyl. Direct Injection
Cooling Air cooled engine
Bore x Stroke 86x72mm
Compression ratio 19:1
Max Output 4.5kW
Speed 3000rpm
The carburetor was designed to suit the engine
displacement and at the same time, the geometry of
the air intake. Different fuel jet sizes have been tried
and the most suitable sizes with regards to emissions
were found, fuel jets of 0.25mm and 0.50mm. Both
were found to give lean mixture to ensure enough
oxygen remain for burning diesel fuel.
Fuels are supplied through two different tanks and
gasoline flow rate has been measured using a rota-
meter while diesel consumption was measured with a
digital weigh machine. A (Kane) 5-gas analyzer was
used to measure exhaust gas analysis (NOx, CO, CO2,
O2), and HC was measured using Horiba gas analyzer.
An electric load has been applied and varied to cover
all engine operating output range. The test rig (Fig.1)
was used to test the engine performance using diesel
fuel alone for baseline performance, and also the
performance when introducing gasoline with different
mixture strengths. The start of injection timing was
also varied to obtain a trend for the tendency of diesel
engine emissions.
Figure 1: Description of test rig setup.
Four tests at constant speed and variable load are
carried out on the engine as follows:
1. Baseline performance using diesel fuel alone.
2. Dual fueling with gasoline using 0.25mm jet.
3. Dual fueling with gasoline using 0.50mm jet.
4. Dual fueling with gasoline using 0.50mm jet
and with retarded start of diesel injection timing.
The constant engine speed is always maintained by the
speed governor which controls the diesel fuel quantity. The engine rotational speed has been set to 3000rpm
throughout the test period.
3. RESULTS
The test has covered all the load range, but for purpose
of comparison for this study, point of maximum load
will be investigated for comparison. Figure 2 shows
variation of brake thermal efficiency with different
operating mode. Using diesel fuel only, efficiency at
maximum load is 24.5% while for 0.25mm and
0.50mm fuel jets, efficiencies are 24% and 27%
respectively. With 0.50mm jet and injection timing
retarded, the efficiency is 26%. Emission gases of
NOx and HC are shown in figures 3 & 4.
Figure 2. Comparison of Brake Thermal
Efficiency for different fuel modes
and injection settings
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120
D: Diesel fuel alone.
D+G-I: Diesel and gasoline with 0.25mm jet.
D+G-II: Diesel and gasoline with 0.50mm jet.
D+G-III: Diesel and gasoline with 0.50mm jet and
injection timing retarded.
Figure 3, shows comparison of NOx emission for
different modes of operation.
At maximum load NOx was highest with diesel alone
(800ppm), while decreased to lowest level (620ppm)
when operating with 0.25mm gasoline jet. ). 0.50mm
jet gave moderate decrease in NOx with its two
different injection timing settings.
Figure 4. Shows trend for HC emission using
different operating modes.
Significant change has been recorded, not only in the
amount of HC emission, but also the trend of variation
with load as shown in figure 4. That was obtained by
injection timing retardation. HC decreased from
130ppm with diesel to 90ppm using 0.50mm jet and
injection retarded. The HC is decreasing with load
increase. This is a reverse trend to that of diesel
operation.
Figure 5. Variation of CO with different fuel and
injection settings.
As in figure 5, CO% in exhaust gases is highly
reduced throughout the operating loads for gasoline-
diesel mode of fueling using the 0.25mm jet. Level
became as low as 0.01% by volume at maximum load.
0.50mm jet with retarded injection showed also
decrease of CO% but at upper half of load range,
while 0.50mm jet without retardation gave the highest
CO%.
4. DISCUSSION
Results are showing comparison of performance and
emission parameters with varying operating settings.
Super lean gasoline-air mixture using 0.25mm jet gave
best results in terms of CO (0.01%) and
NOx(620ppm) emissions, but at a sacrifice of decrease
in efficiency (4%less). Extra lean mixture using
0.50mm jet gave the best increase in thermal
efficiency (10%more than diesel) and a decrease in
NOx to 720ppm. HC and CO are above diesel ones.
Retarding the injection timing helped to decrease (HC)
throughout the operating load range, with a negative
gradient. HC decreased to only 90ppm at maximum
load. Super lean mixture helped to reduce combustion
temperature and hence reduced NOx and CO, while
increased HC due to quenching effect. Extra lean
mixture have better tendency to ignition by diesel
injection, and therefore resulted in increased thermal
efficiency and lower NOx, but again it quenching
effect produced more HC. The strategy used to retard
the injection timing seem to be favorable for the
combustion of the mixed charge, and resulted in lower
HC levels keeping at the same time higher thermal
efficiency and lower NOx in comparison with diesel
fuel operation.
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2012
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5. CONCLUSION
From these results, factors affecting performance of
dual fuel operation of diesel engines can be stated as
follows:
1- Mixture strength of the added gasoline fuel should
be optimized. Super lean mixtures will result in lower
efficiency and higher HC levels, while richer mixtures
can improve efficiency but increase HC
2- To have control over HC with higher mixture
strengths (still lean), the injection timing retardation
showed a positive effect and reversed the curve trend
for HC emission keeping at the same time the gain in
both thermal efficiency and NOx and CO reductions.
6. ACKNOWLEDGMENT The authors would like to acknowledge the support of
Sudan University of Science and Technology (SUST),
department of mechanical engineering for facilitating
this research.
7. REFERENCES
[1] Gautam Kalghatgi, Leif Hildingsson, Bengt
Johansson ―Low NOx and Low Smoke
Operation of a Diesel Engine Using
Gasolinelike Fuels‖ - Journal of Engineering for
Gas Turbines and Power-Transactions of the
ASME 2010 Volume 132 Issue 9 –
[2] Scott Curran, Vitaly Prikhodko, Kukwon Cho,
Charles Sluder, James Parks, Robert Wagner
- Oak Ridge National Laboratory
Sage Kokjohn, Rolf Reitz - Univ of Wisconsin
―In-Cylinder Fuel Blending of
Gasoline/Diesel for Improved Efficiency and
Lowest Possible Emissions on a Multi-
Cylinder Light-Duty Diesel Engine‖SAE
paper number 2010-01-2206
[3] Derek Splitter, Rolf Reitz, Reed Hanson -
Univ of Wisconsin Madison ―High Efficiency,
Low Emissions RCCI Combustion by Use of
a Fuel Additive‖ SAE paper number 2010-
01-2167
[4] C. Sundar Raj*,1, S. Arul2 and S.
Senthilvelan3- University, Chennai‖ Some
Comparative Performance and Emission
Studies on DI Diesel…‖ The Open Fuels &
Energy Science Journal, 2008, 1, 74-78
[5] Kent Ekholm, Maria Karlsson, Per Tunestål,
Rolf Johansson, Bengt Johansson, Petter
Strandh,‖ Ethanol-Diesel Fumigation in a
Multi-Cylinder Engine‖SAE International
Journal of Fuels and Lubricants, 1:1, pp. 26-
36, April 2009.
[6] Leif Hildingsson, Bengt Johansson - Lund
Univ., Gautam T. Kalghatgi ,Andrew J.
Harrison - Shell Global Solutions UK ―Some
Effects of Fuel Autoignition Quality and
Volatility in Premixed Compression Ignition
Engines‖SAE paper 2010-01-0607
[7] M. Abu-Qudais, O. Haddad, M. Qudaisat-
Jordan University of Science and Technology
―The effect of alcohol fumigation on diesel
engine performance and emissions‖ Energy
Conversion & Management 41 (2000)
[8] Vittorio Manente, Bengt Johansson- Lund
University Faculty of Engineering ―Gasoline
Partially Premixed Combustion -…….‖
ISBN: 978-628-8144-3
The 5th IMAT, November 12 – 13th
2012
122
Figure.1: Energy Use Composition in
Buildings
For Air Conditioning: 43%
Air Cnonditioning
Machinery
Others
Water
Transport
Others
Air
Transport
Others
Hot
Water
Lighting
Electric
Device
Ventilation
Water Supply
& Drainage
Elevator
By HP of The Energy Conservation Center, Japan
Solar Air-conditioning System Using Single-Double Effect
Combined Absorption Chiller
Hajime Yabase
Kawasaki Thermal Engineering Co.,Ltd., Engineering Office,
Kusatsu, Shiga, Japan
Tel ::+81-77-563-1111, Fax:+81-77-564-4353
E-mail: [email protected]
ABSTRACT
Since further energy saving for global environmental
protection becomes a matter of urgency, promotion of
introduction of renewable energy sources is required
for realization of low-carbon society. We developed a single-double effect combined
absorption chiller for "Solar air-conditioning system"
in 2010. This chiller is composed of a highly-efficient
gas absorption chiller as a main machine which are
equipped with a solar heat recovery unit comprising a
heat recovery heat exchanger and special condenser. It
enables low temp. solar hot water at 75ºC under
operation at the cooling rating of load factor: 100%. And we constructed the demonstration plant in Japan.
We confirmed that the solar heat priority usage
function and gas-based backup function operate
properly and overall system functions normally. In
summer, fuel gas reduction by 10% could be achieved
and the results as estimated were obtained. Keywords : Absorption chiller, Solar heat, Solar air-
conditioning system, solar collector, gas-based
backup function, Demonstration plant
1. INTRODUCTION
Absorption chillers are units to supply chilled water using gas and oil as fuel. In Japan, absorption chillers
have been widely used for industrial and commercial
central air-conditioning because they contribute to
electric-load leveling in summer because of capable of
cooling using little power, and use water having zero
ozone depletion potential (ODP) as refrigerant.
Meanwhile, the global warming issue has worsened
markedly in recent years, which causes us to be
confronted with the urgent task of realization of low-
carbon society. As shown in Figure 1, in case of
Japan, power for air-conditioning accounts for 43% of
total power consumption used for office buildings and
absorption chillers are also strongly required saving-
energy.
Under these situations, a solar cooling system which
performs cooling by introducing hot water obtained
from solar heat into absorption chillers using thermal
energy as driving source has received increasing
attention and undergone promotion of development
toward practical use recently. This is because this
cooling system is capable of using solar heat whose
reserve amount is much abundant and whose energy
conversion efficiency is higher among renewable
energy for air-conditioning application with high
power consumption rate in industrial and commercial
fields.
We has developed a single-double effect combined
absorption chiller exclusively designed for the solar
cooling system and launched in August 2010[1][2],
and we constructed the demonstration plant of this
system in Japan. we report the outline and the
performance of the chillers and demonstration plant.
2. SINGLE-DOUBLE EFFECT COMB-
INED ABSORPTION CHILLERS
IMAT-UI 022
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2.1 Use and Problem of Solar Heat
Among renewable energy, solar energy is rich in
reserve, which undergoes promotion of application,
however, it is almost applicable to solar battery
(photovoltaic energy) but solar energy has been not
picked up as the method for using as heat source so
much. However, as shown in Figure 2, photovoltaic
energy generation is as low as approx. 10% in
generating efficiency and in case that solar energy is
picked up as hot water at approx. 90ºC, the energy
conversion efficiency is as high as 40% and the high-
end evacuated tubular type reaches 50%.The system
using solar hot water as driving source is applicable
only to absorption chillers practically. The
conventional air-conditioning system is shown in
Figure 3. However, solar heat is unstable heat source which is
easily influenced by weather and it is difficult to use it
according to fluctuating air-conditioning loads. Solar
thermal air-conditioning system has been tried to be
diffused since 1980‘s, however, they have been
familiarized fully. The reasons are shown as follows:
(1) As shown in Figure 3, in the solar thermal air-
conditioning system, in addition to the absorption
chiller, the backup boiler and accompanying
machines are required, which causes the system
composition to be complicated and the investment
efficiency using renewable energy is not
expected.
(2) It is necessary to control the solar hot water and
backup system according to fluctuations in solar
heat and air-conditioning loads, however, it is
difficult to control and establish an optimal
control system to use solar heat efficiently, and it
is necessary to familiarize local operators to learn
as well.
(3) The double-effect type which is mainly used as an
absorption chiller, which requires heat at 120ºC
or more as driving source. In case that solar heat
at approx. 90ºC is used, only the single-effect
system functions. In case that the backup system
functions, even if fuel is used, the efficiency is
low because of single effect system, the effect of
introduction by renewable energy is not expected
so much.
2.2 Improvement from Conventional System
In consideration with the problems in the above-
mentioned solar thermal air conditioning, we
developed single-double effect combined absorption
chillers for using solar hot water preferentially
in combination with backup heat source such as
gas, oil, etc. (hereinafter referred to as the Solar
Absorption Chillers). The aspect is shown in
Figure 4.
The features are shown as follows:
Figure 2: Solar energy conversion
efficiency
From NEDO homepage
Figure 3 ventional Solar Thermal Air-Conditioning
System
Figure 4: Aspect of Solar Absorption
Chiller
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(1) Since these solar absorption chillers are equipped
with generators driven by solar hot water based on
direct-fired absorption chillers, a backup system
are unnecessary to be prepared, No. of
composition elements are reduced, which
simplifies the air-conditioning system
(2) These solar absorption chillers control so as to use
solar hot water preferentially based on driving by
fuel such as gas, oil, etc. In addition, control of
loading is performed according to fluctuations in
air-conditioning loading.
(3)When driving by fuel, double-effect operation is
performed, the same efficiency as absorption
chillers which are currently diffused is obtained,
which allows saving energy operation because
renewable energy is used.
2.3 Outline of Solar Absorption Chillers
Solar Absorption Chillers are composed of highly-
efficient gas absorption chillers with COP1.3 (gross
calorific value) as main machines which are equipped
with a solar heat recovery unit comprising a heat
recovery heat exchanger and special condenser.
Figure 5: Cycle flow-diagram of Solar Absorption Chiller
Figure 6: Principle of Gene-Links
Figure 7: Principle of Solar Absorption Chillers
The 5th IMAT, November 12 – 13th
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As shown in the cycle-flow diagram in Figure 5, solar
heat hot water is used for heating and regenerating
absorbing solution at the heat recovery heat
exchanger. Using refrigerant generated during this
regeneration process for cooling enables the amount of
fuel used for the high temp. generator to be reduced.
Generally, solar energy collectors have characteristics
in which the smaller the difference in temperature
between the collection temperature and outside air
temperature is, the higher the collection efficiency is,
therefore, it is necessary to allow solar absorption
chillers to use even low temp. hot water to increase the
efficiency of the overall system.
As chillers which are capable of reducing fuel
consumption by introducing hot water, exhaust heat
introduction absorption chillers (Gene-Link) can be
considered, however, Gene-Links are products
designed to use exhaust heat hot water at stably high
temperature (83 to 90ºC) obtained by cogeneration
systems, etc. and cannot use low temp. hot water.
The reason is that Gene-Links are composed as shown
in the principle drawing in Figure 6, in which
refrigerant vapor generated at the heat recovery heat
exchanger and refrigerant vapor generated at the low
temp. regenerator are condensed in the condenser of
the base absorption chillers. Therefore, the saturated
temperature of the heat recovery heat exchanger is
restricted by the saturated temperature of the
condenser of this absorption chiller body, which
prevents the log-mean temperature difference to
collect low temp. hot water from being maintained.
Consequently, as shown in the principle drawing in
Figure 7, in Solar Absorption Chillers, a condenser
exclusive for the heat recovery heat exchanger is
newly provided to separate the heat recovery unit and
the base absorption chiller and a structure to initially
introduce cooling water to the special condenser is
employed, which reduces the pressure in the heato
recvery unit and maintains the log-mean temperature
difference to collect low temp. hot water.
This chillers enable low temp. hot water at 75ºC under
operation at the cooling rating (load factor: 100%,
cooling water temp: 32 ºC) or even lower temp. hot
water depending on loading conditions and cooling
water conditions to be used.
2.4 Performance of Solar Absorption Chillers
The performance of Solar Absorption Chillers is
shown in Figure 8 and 9. Figure 8 shows the heat
recovery amount in each cooling load factor and
Figure 9 shows the combustion gas consumption
amount in this case. The cooling water inlet
temperature is set to 32ºC at 100% load and 27 ºC at
0%, and proportional values at 0 to 100%.
Figure 8 shows that the heat recovery amount in case
of hot water at 75ºC is 0.45kW/RT when the cooling
load factor is 100 %, which increases as the load
factor decreases, and reaches the maximum amount of
1.37kW/RT when the load factor is approx 30%. In
this case, in a loading area with approx. 30% load
factor where the heat recovery reaches the largest
amount, cooling operation only by hot water without
use of combustion gas is possible, therefore, the heat
recovery amount in proportion to load is obtained in a
loading area with approx. 30% and lower load factor.
Figure 8: Heat recovery rate of Solar
Absorption Chillers
chiller-heater
Table 1: Specifications of system
Figure 9: Fuel gas consumption of Solar
Absorption Chillers
chiller-heater
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2.5 Performance of Solar Absorption Chillers
The performance of Solar Absorption Chillers is
shown in Figure 8 and 9. Figure 8 shows the heat
recovery amount in each cooling load factor and
Figure 9 shows the combustion gas consumption
amount in this case. The cooling water inlet
temperature is set to 32ºC at 100% load and 27 ºC at
0%, and proportional values at 0 to 100%.
Figure 8 shows that the heat recovery amount in case
of hot water at 75ºC is 0.45kW/RT when the cooling
load factor is 100 %, which increases as the load
factor decreases, and reaches the maximum amount of
1.37kW/RT when the load factor is approx 30%. In
this case, in a loading area with approx. 30% load
factor where the heat recovery reaches the largest
amount, cooling operation only by hot water without
use of combustion gas is possible, therefore, the heat
recovery amount in proportion to load is obtained in a
loading area with approx. 30% and lower load factor. Further, in case where the hot water is 90ºC, the heat
recovery amount becomes 1.69kW/RT at 100%
cooling load factor and reaches 2.42kW/RT at the
most and the cooling load area operatable only by hot
water is expanded to approx. 57%, therefore, it is
found that heat amount can be more effectively used
even if the temperature of hot water increases.
Figure 9 shows that in case where the hot water is
75ºC, the combustion gas consumption amount can be
reduced by approx. 9% at 100% cooling load factor
compared to the case where hot water is not
introduced. Since the above-mentioned decrease in
combustion gas consumption amount has a
proportional relationship with the heat recovery
amount shown in Figure 8, it becomes larger as the
load factor decreases, as mentioned above,
combustion gas is not required in a area with 30% or
less of load factor.
Further, it is found that in case where the hot water is
90ºC, the combustion gas consumption amount can be
reduced by approx. 32% at 100% cooling load factor.
3. SOLAR COOLING SYSTEM 3.1 Outline of Solar Cooling System
The demonstration plants are installed in our factory
located in Kusatsu City of Shiga Prefecture, Japan.
This system was completed in Dec. 2010 and started
to undergo full-sized verification test in Feb. 2011.
The flow diagram of this system is shown in Figure
10.
Solar heat (hot water at 75ºC to 90ºC) is introduced
into the Solar Absorption Chiller. In addition, if solar
heat is insufficient, the backup system to compensate
for the energy through gas is available.
Evacuated glass tube type solar energy collectors
which are highly efficient in a high-temp area at 75ºC
to 90ºC is used for the solar energy collector. 160
sheets of collectors (260m2) which satisfy the exhaust
heat recovery amount (0.6kW/RT, 126kW *cooling
water at 31℃) during rated operation in case of solar
heat hot water of Solar Absorption Chiller at 75ºC
were installed on the roof of the office.
The hot water storage tank is provided to absorb the
difference of flow rate between the solar energy
collector and Solar Absorption Chiller and serves as a
temporal cushion if solar radiation fluctuates
suddenly.
The radiator is provided to prevent hot water from
boiling by excessive heat collection by operating when
collected solar heat cannot be used on holidays, etc.
3.2 Feature of Solar cooling system
collector
Table 1: Specifications of system
Heatin
g
Coolin
g
Hot water storage
tank (1m3)
Radiator
Heat
exchanger
for heating
Cooling
→
absorption
chiller/heater
pum
p
pump
Gas
inpu
t
↓ Heating
Figure 10: Schematic diagram of system
Figure 11: Aspect of Collector
The 5th IMAT, November 12 – 13th
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127
3.2.1 Lowering the temperature of solar heat
usable area
Generally, solar energy collectors have characteristics
in which the smaller the difference in temperature
between the collection temperature and outside air
temperature is, the higher the collection efficiency is,
therefore, it is necessary to allow solar absorption
chillers to use even low temp. hot water to increase the
efficiency of the overall system.
As section 2 already described, in the Solar
Absorption Chiller in this system, solar heat at 75ºC
(rated) and approx. 60 ºC (under partial loading) can
be used by employing a hot water heat exchanger
optimized for use of solar heat and improving the flow
of cooling water.
3.2.2 Simplifying and downsizing the solar system
The auxiliaries (Figure 12) such as pumps, etc. to
supply hot water obtained from the solar energy
collector to the Solar Absorption Chiller are required,
however, those items are simplified as much as
possible while considering packaging with the Solar
Absorption Chiller.
Packaging after reflecting these verification results can
reduce the details of work at site and costs for building
up the system as one of targets.
3.2.3 Controlling and Monitoring System
When introducing the solar cooling system, it was
necessary to specially build up the control functions
such as starting/stopping the heat collecting facility
and excessive solar heat collection of the solar energy
collector, however, these control functions are
assembled into the Solar Absorption Chiller in this
system and control of the overall system is enabled.
Assembling the control functions into the Solar
Absorption Chiller maximizes the saving energy effect
based on the use of solar heat by strengthening the
linkage with the control functions of the Solar
Absorption Chiller in addition to eliminating the needs
of the special control equipment.
Some examples of linkage control with the Solar
Absorption Chiller added to this system are shown as
follows:
(1) Backup linkage at the Solar Absorption Chiller
(2) Interlocking control in response to change in the
temperature setting
Figure.13: Monitoring System
Figure 12: Collection of heat facility
The 5th IMAT, November 12 – 13th
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Further, the system is equipped also with the
information collection function to monitor
the system operation status and energy-
saving effect.
Figure 13 shows a part of the monitor system
to display the collection data. The monitor
system enables daily and hourly collection
amount and efficiency to be displayed, which
contributes to the grasping of operation
conditions and review for improvement of
system.
As mentioned above, when building up the
solar cooling system, cost reduction is a big
issue. Assembling the control system and
monitor system into the Solar Absorption
Chiller greatly serves to reduce the cost for
building up the system.
3.3 Evaluation status and results
3.3.1 Operation conditions
Figure 14 shows the operation data on May
20 and Figure 15 shows the operation data on
June 28. From the data, it was confirmed that
the solar heat priority usage function and gas-
based backup function operate properly and
overall system functions normally.
Because of operations with comparatively-
low loads on the conditions where the
maximum temperature was 28.4ºC and the
air-conditioning loading factor was 23% on
May 20, the gas amount could be reduced by
25%. Meanwhile, the maximum temperature
was 34.5 ºC and the air-conditioning loading
factor was as high as 60% on June 28,
however, the gas amount could be reduced by
11%.
Cooling operation starts in late May,
therefore, the monthly reduction rate shows
the data only in June, however, reduction by
10% could be achieved and the results as
estimated were obtained.
3.3.1 Effect of Solar Absorption Chiller
It was confirmed that hot water obtained
from the solar energy collector is constantly used at
75ºC or less and can be used even at approx. 60 ºC
during low-load operations.
In the actual system, the effect could not be quantified
because fluctuation in solar radiation and load should
be considered, however, use of Solar Absorption
Chiller developed exclusively for use of solar heat can
reduce the hot water temperature from the solar energy
collector more than use of conventional exhaust heat
introduction type absorption chiller(Gene-Link),
therefore, it was confirmed that this system increased
the collection efficiency of the solar energy collector
and improves the efficiency of overall system.
3.3.3 Improvement points
When changing to the low-load operation mode where
the refrigerant pump of the chiller activates the
start/stop control, introduction of hot water is turned
Figure 14: Operation data of cooling (20th
May)
Figure 15: Operation data of cooling (28th July)
The 5th IMAT, November 12 – 13th
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129
on and off interlinking with start/stop of the refrigerant
pump, however, even in operable load only by solar
heat hot water, it is confirmed that the pick-up
temperature increases due to output delay when
turning on and backup control by combustion
activates. We plan to review the control to minimize
the delay and add it to the system.
4. CONCLUSION
(1) The Combination of the heat recovery exchanger
and special condenser of Solar Absorption
Chiller enables the machine to be operatable even
if the temperature decreases up to 75ºC and does
not require the backup by combustion gas in an
area with 30% or less of load factor
(2) In case where the temperature of hot water is 90
ºC, the reduction rate of combustion gas at 100%
load factor becomes 32%, therefore, the
performance increased in comparison with 26%
of the conventional Gene-Link. In addition, the
cooling load area operatable only by hot water
was expanded to approx. 57%.
(3) From the verification of the demonstration plant
of solar cooling system, the usefulness of the
Solar Absorption Chiller developed exclusively
for use of solar heat was confirmed.
(4) Further, we make sure that our built-up system is
useful to make it easier to introduce a solar
cooling system. By commercializing the system
into which improvement points during
verification were fed back in the future, we aim
to make the solar cooling system to be
recognized as a useful solution tool for global
warming problem and promote them. .
REFERENCES
[1] Hyodo, Y., 2011, Solar Absorption Chillers using
solar heat for cooling of Kawasaki Thermal
Engineering Co., Ltd, Clean Energy, vol.20, no.3,
pp.5-9.
[2] R. Kajii, H. Yabase, M. Ohta, 2011, Development of
Solar Absorption Chillers-Heaters, Trans. Of the
JSRAE, vol.28, no.3,not require the backup by
combustion gas in an area with 30% or less of
load factor
The 5th IMAT, November 12 – 13th
2012
130
Environmental Protection and Fuel Consumption Reduction by
Flameless Combustion Technology: A Review
Seyed Ehsan Hosseini, Saber Salehirad, Mazlan Abdul Wahid, Mohsin Mohd Sies,
Abuelnuor Abdeen Ali Abuelnuor
Faculty of Mechanical Engineering, Universiti Teknologi Malaysia, 81310 UTM Skudai,
Johor, Malaysia, January 2012 Tel: +60176830504 Email: [email protected]
ABSTRACT In recent years global fuel consumption has
increased in the world due to modernization and
progress in the standard of living. The
conspicuous rate of carbon dioxide and nitrogen
oxide released to the environment and fuel
resources are depleted day by day due to
inconsiderate fuel consumption. Requirement for
efficient use of any kinds of fuel has become the
other concern due to the oil crisis and limitation
of fuel resources. In combustion process, the
abatement of pollutants often associates with
efficiency loss. In the other word, high efficiency
and low pollutant which are the main
requirements of combustion are not fulfilled by
the existing combustion. During the development
process of new combustion technology, a
particular
focus was on low NOx burners and engines.
Today, flameless combustion has received more
attention because of its low NOx emission and
significant energy saving.
Generally, compatibility between high
performance and low NOx emission has been
observed by preheated air application and
changing the combustion characteristics from
traditional flame to flameless mode. Although, in
flameless mode the oxidizer is diluted and low
concentration of oxygen can be seen, combustion
is still sustained if the air is preheated higher than
the fuel self-ignition temperature. This aims to
review the concepts and the applications of
flameless combustion and gathers useful
information to understand the necessity of
transient from traditional flame mode to
flameless combustion.
4. Introduction
1.1 General concept of flameless oxidation
Heat generation and power production by
combustion of hydrocarbon container fuels are
the main goals of combustion [1, 2]. One of the
best ways to achieve the higher efficiency and
low pollution in combustion is using regenerative
burners with flameless combustion. Recycled
burned gases have been applied in this
technology to make the preheated air lean in
order to achieve low- NOx emissions and a
reasonable thermal efficiency [3]. This
technology, emerged from 1990, and has been
successfully applied, specially, in metallurgy and
steel industries of some developed countries.
Flameless combustion which is known as
Flameless Oxidation (FLOX) in Germany [4],
also known as High Temperature Air
Combustion (HiTAC) In Japan [5], Moderate and
Intensive Low oxygen Dilution (MILD)
combustion in Italy [6, 7] or Colorless
Distributed Combustion (CDC) [8], Low NOx
Emission Injection in the US is a new
combustion system which accomplishes low NOx
emissions and high efficiency among several
techniques. Postponed mixing of fuel and air and
flue gas utilization in the flame zone are the
fundamentals of FLOX [9]. The application of
high temperature air combustion has been
investigated experimentally [10-15] and
numerically [16-22]. Fundamentally, flameless
combustion is identified by aspects of turbulence
and chemistry strongly [23]. The characteristics
of flameless oxidation for various gas type fuels
like methane or ethane [24], also for mixtures of
gaseous hydrocarbons and hydrogen [25-27], and
biogas [28-30] has been analyzed. In addition,
this method has been applied for liquid fuels [31-
IMAT-UI 023
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34]. Moreover, solid fuels flameless combustion
has been experimented successfully [35, 36].
1.2. Industrial applications of flameless
combustion
Flameless combustion is suitable for different
industrial procedures that need a uniform high
temperature profile inside the furnace [37]. The main
industrial applications of flameless combustion now
concern the metallurgy area for which the major issue
is energy efficiency. For the other industrial sectors,
the issues are sometimes different, but for such as
glass-making and cement industry [38, 39], waste
treatment [40], petrochemicals, gas turbines [41, 42]
or industrial boilers [43], it is very likely that this new
combustion mode will find its place, in the short or
medium term. The main reasons for development of
this technology in industries can be cited as decreasing
the NOx emissions, increasing the heat transfers, and
rising the duration of the equipment, which are mostly
damaged by very high heat flux.
2. Flameless formation
Preheating of the reaction air and burnt recycled gases
inside the chamber are the basic terms of flameless
combustion. The impacts of recycled flue gases under
highly preheated air conditions (from1200 to 1600 K)
were investigated by Katsuki and Hesegawa [44]
where they found that the flameless combustion
generates by high velocity of reaction air. Dally et al.
[45] stipulated that the flame structure starts to change
when the level of oxygen decreases and it happens at
high Reynolds number for air jet and low oxygen
concentration. In order to understand the transition
from conventional combustion to flameless
combustion, distribution of the axial temperature in
the chamber should be measured regularly during the
experiment versus time [1]. The peaks of temperature
are located in the center of the furnace at the burner
level, far away from the burner position [53]. Recycle
ratio (Kv) of the combustion chamber is the most
important factor which describes the efficiency of a
flameless burner. This ratio can be described as
Kv=Me / (Ma+Mf ) [44]. In the recent equation Me is
the exhaust gases flow rate which is recirculated into
fuel and air before reaction, Mf is the fuel flow rate,
and Ma is the combustion air flow rate. In HiTAC
burners, several combustion modes exist, which are
strongly depend on the average temperature of the
chamber. In addition main factor for determining the
regime of combustion is the volume of recycled flue
gases back into the inlet air jets as shown in Fig.1.In
this figure, temperature is plotted versus Kv for
hydrocarbon fuels. It shows the conversion process of
traditional flame to the flameless combustion and the
conception of the flameless oxidation procedure.
Wünning &Wünning experimentally achieved the
relationship between Kv and temperature of the
furnace for different modes of combustion [4].
Fig. 1 –T- diagram for the conversion of
conventional flame to colorless oxidation mode:
A, traditional regime; B, conversion; C, MILD
combustion; D, no reaction zone.
(i), (ii), (a), (b) lines are showing the boarders of
MILD combustion were achieved experimentally
and represent the path of air heating, increasing
the recycle ratio, path of cooling; and diluting
path respectively.
Zone A. At this zone because of the low jet velocity
(low Kv level) the combustion chamber is working in
conventional flame mode and flame is stable. In the
flameless combustor, the preheating section is
provided to avoid the reaction from quenching. Thus,
the temperature of the chamber is raised to the
amounts greater than fuels‘ self-ignition temperature
by traditional flame (normally up to1000 K).
Zone B. When the temperature inside the combustion
chamber is high enough and more than the auto-
ignition value for the fuel, by enhancing the entrance
momentum of the reactants, the amount of Kv is
increased. Therefore, concentration of the Oxygen in
the air declines and fuel velocity in the reactant side
increases. As a result, the flame pales and the average
temperature of the furnace decreases, this region also
can be called as instability zone.
Zone C. The rate of flue gas is much more than the
incoming air for reaction with the fuel, as a result, the
flame becomes invisible and inaudible and the
reaction zone spreads to the downstream regions of the
combustion chamber. It is possible for much higher
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recycle rates (the high Kv) which causes dilution of
oxygen concentration in reaction [46].These
specifications lead to clean flameless combustion [47].
But, in conditions that Kv is greater than 19 the clean
flameless combustion cannot be obtained, particularly
for the coke oven gas. In the flameless combustion
region, the recycle ratio (Kv) is larger than 2.5, and the
chamber temperature reaches the values up to 1100 K
[1].
Zone D. As it mentioned, in zone C the temperature of
the furnace decreases during flameless mode, but in
zone D the furnace temperature reaches near to a
critical value by intense heat transfer inside the
combustion chamber. Therefore, the flame lifts off,
and eventually, if the temperature is not sufficient,
blows out. As an instance, for methane, the limitation
is less than 1300 K, and for coke oven gas (H2/CH4
60/40 % by vol.) it is around 1180 K [46].
For preventing of the flame front formation, great
values of Kv are required. The jet velocity should be
greater than the velocity of flame propagation. For
example, adding hydrogen gives better range of
stability of flame, and causes an increase in the speed
of the laminar flame propagation about six times with
respect to the methane/air flame. It means when the
fuel is supplied from a single nozzle, larger jet
momentum needs at the constant heat input [48].
Some investigations declared initial jet speed of the
fuel, amount of Oxygen in oxidant air, and the ratio of
density of the fuel inlet to density of the ambient gas
are three basic factors which define flame volume
[49]. In flameless mode reactants are fuel and air
which are highly diluted by an amount of inert flue
gases, and the temperatures within the furnace are
higher than the auto-ignition temperature of the fuel.
In these circumstances conventional flame is not
stable and the flame lifts off due to strong shear
motion caused by gas recirculation (Kv).
Consequently, uniform temperature distribution
appears along the combustion chamber and flux of the
net radiation increases by around 30% [50]. Therefore,
temperature uniformity and the chemical type fields
are the main aspects of the flameless combustion
method. For non-premixed fuel and air jets, there is a
critical rate for Kv which the flameless combustion
does not occur below this amount of recirculation
ratio. Experiments also confirmed this claim [51].
Recirculation of flue gas inside the combustion
chamber makes the reaction oxidizer becomes diluted
and decreases the concentration of oxygen.
Consequently, high efficiency and low thermal NOx
formation are achieved by flameless combustion
method [52]. Fig.2 illustrates the shape of a FLOX
furnace working in traditional flame and flameless
modes.
Fig.2 -Schematic diagram of a FLOX burner
firing in flame and flameless condition.
3. Air preheating process
3.1 Required equipments
In the combustion furnace when air and fuel mixed
together as reactants, it requires some heat to occur
combustion. In order to stabilize flames, combustion
products flow should be recirculated behind a pilot
flame. In these conditions combustion takes place
anywhere in the furnace, therefore, this kind of
combustion is called as highly preheated air
combustion and it is different with conventional
preheated air combustion where there is no self-
ignition.For instance auto-ignition of normal air and
natural gas happens when the temperature of
preheated air goes up around 1100 K. Injection of fuel
and air is done by special nozzle whose configuration
plays crucial role in flameless combustion furnace,
because the burner geometry sets the intensity of
turbulence and exhausting recirculation to reach
flameless combustion. In flameless combustion, air
preheating is the main key to achieve higher
efficiency. As a result, the energy of exhaust gases
which have high temperature is transmitted to the
reaction air in regenerative and recuperative heat
exchangers [4]. Regenerators and recuperators imbibe
excess heat from the product gases and use it again by
increasing the temperature of inlet air. Highly
preheated air increases peak temperature of the flame
in conventional combustion with a great impact on the
formation of NOx , however in flameless oxidation
higher air preheat temperature is desirable. Using
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recuperative and regenerative burners are two main
ways for saving energy. They can be applied by
considering the heat exchanger area and the required
preheating rate. In recuperators and regenerators,
energy transmitted from the exhaust gases back to the
inlet reaction air. Normally, recuperators and
regenerators are heat exchangers and located outside
of the combustion chamber to absorb part of enthalpy
of the hot flue gases for preheating the combustible air
which is called as the secondary air [53, 54].
3.2 Recuperator
In the recuperative burners the air can be preheated
near to 1000 K but the regenerator can increase the
temperature of combustion air up to 1300 K. However,
in traditional combustion burners, NOx formation will
certainly enhance due to rise of the preheating
temperature, although combustion intensity maybe
enhances. S.E Hosseini et al [55] calculated NOx
formation in methane traditional and flameless
combustion computationally. It has been stated that
the rate of NOx formation declines in diluted and
preheated oxidizer conditions in flames combustion,
however the rate of NOx constitution increases in
conventional combustion when applying preheated
oxidizer. Recuperative systems normally are used in
the steel industries because they can be used in
industrial burners for direct heating also for indirect
heating in combination of radiant tubes. The following
burner arrangements could be used for different
applications as shown in fig.3 [4]. Empirical
outcomes on a 300-kW setup showed that using
recuperative flameless combustion burners instead of
conventional burners causes drastically reduction of
NOx production, by 1400 ppm, from 1500 ppm to the
amounts less than 100 ppm [39].
Fig.3 Recuperative burner for use in radiant tubes
3.3 Regenerator
After 1990, high efficient regenerators were applied in
order to increase the temperature of combustion air to
very high values, also large amounts of flue gases
recirculation back to the flame was performed for
creating proper ambient with diluted Oxygen. High
temperature amount of oxidant (between 873 and 1273
K) was achieved by a honeycomb type regenerative
preheater which had become very hot in means of flue
gases from natural gas conventional combustion
before the experiment [56]. Basically, regenerative
burners are using the exhaust gases heat to preheat the
combustion air which would be evacuated to the
ambient in the traditional combustion mode. In the
burners which are working in flameless combustion
systems, air inlets and fuel gas outlets are around the
fuel nozzle. Each one of these inlets or outlets contains
a honeycomb regenerator which is made of ceramic
and during the exhaust cycle imbibes heat from flue
gases and relegates this heat during the firing cycle
into the combustion air. Fig.4 depicts the shape of the
burner and air and fuel inlets and outlets placements.
Fig. 4-Schematic of the HiTAC furnace: (a)
furnace Dimension,(b) Configuration of
fuel and air inlets and burned gas outlets.
These regenerators consist of in 6 pairs and divided
into two groups which are specified by intervals. First
category (three pairs) preheats the combustion air with
a switching time of 10 seconds, and the next group
works as an exhaust-gas storage bed and heat
extractor. The temperature of the inlet air increases to
the values of recirculated flue gas temperature when it
goes through the honeycomb regenerator before
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reaction occurs. Fuel is entered persistently by the
same nozzle, while the air nozzles change from inlets
to outlets and repeat during a defined time space. This
performance causes the single flame to be constituted.
Since all regenerators are located around the fuel-jet
nozzle position of the flame, the temprature is nearly
constant during defined time interval. Around 80% of
the flue gases are recirculated inside of the burner, the
other 20%passes through the chimney which is located
on the wall of the furnace [57]. The honeycomb heat
regenerators highly affect (nearly 88%) the process of
combustion and recovers energy which is existed in
the burnt gases by amounts of about 72%.
Experiments also accept these values and endorse
these facts. Moreover, honeycomb regenerator reaches
a periodic steady state performance very fast in
compare to the other kinds of heat regenerators like a
randomly packed bed of solid storing materials.
Generally, the amount of temperature which is needed
for reaching to flameless combustion circumstances
can be steadily supplied only after two minutes. Fig.5
shows a heat regenerator burner [58, 59].
Fig.5-High temperature air combustion system
In oxyfuels flameless combustion [60], air can be
replaced by pure oxygen and shows very good
performance in the steel industries; also it covers all
the specifications of flameless air combustion. Kumar
et al. [61] demonstrated experimentally that flameless
combustion can be constituted without preheated air
when high rates of Kv (larger than three) is used and
great values of volumetric heat load ( greater than 10
MW/M3). Same outcomes were reported by
Krishnamurthy et al. [62] even when using Oxygen
instead of the air for reaction. Their experiments,
performed by using high amounts of velocities (near to
the speed of the sound) in order to obtain a high rates
of Kv [63].
4. Utilization of different fuels in flameless
combustion
4.1 Natural gas and biogas
A.F. Colorado at el. [30] compared the performance of
a flameless combustion furnace using burner run with
natural gas and biogas at 20KW. Also 20% by volume
excess air during flameless combustion of biogas and
natural gas, 74% and 85% of the total flow rate of flue
gas was conducted to the chamber through
regenerators respectively. To ensure the auto-ignition
temperature of biogas and natural gas is fulfilled,
walls mid temperature sustained up to 870°C. During
the operation with natural gas and biogas the
preheated air was 680°C and 537°C respectively. They
concluded that the system performance was the same
in both experiments. Also, produced CO and NOx
were less than 16 ppm and 3 ppm, respectively. The
efficiency of biogas was 2% lower. They also
mentioned that for the same gases value passed via the
regenerators, the biogas combustion products showed
the high density values of CO2, so specifications of
radiation, capability of imbibitions, and the amount of
heat capacity increased. These aspects lead to higher
rates of heat exchange from the produced flue gases to
honeycomb regenerators. Fig.6 depicts their
experiment in detail.
Fig.6- Experimental setup for natural gas and
biogas flameless combustion
Fuel dilution with N2 or CO2 decreases the NOx
production and causes the flame to extinction.
This proofs that premixing of the fuel flow with
flue gases shows better results on the flameless
combustion regime formation without applying
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of higher fuel jet momentum [45]. Effuggi at el.
[24] performed some experimental tests using
biogas as an equimolar mixture of CH4/N2.The
conclusions declared that the probability of
applying flameless combustion with low calorific
value fuels without jeopardizing its proficiency
for NOx reduction. The characteristics of biogas
show that biogas is a kind of a diluted fuel itself;
therefore it‘s lower heating value (LHV) declines
when the amount of N2 or CO2 rises. Burner
thermal efficiency decreases due to increase the
volume of inert gas which should be heated up by
the fuel inflammation. Consequently, biogas
composition is a main item in designing of the
burner to determine the value of thermal
efficiency approximately. Flameless oxidation is
impressive in abatement of NOx production
(amounts less than 15 at 3% O2), and
prevents constitution of soot in rich fuels
utilization. Therefore in flameless combustion the
type of fuels and inside temperature of the
chamber do not play conspicuous effects on
production of pollutants. CO emission in rich fuel
condition is inevitable in every types of
combustion [24]. Derudi et al. [25] stipulated the
low calorific value biogas flameless combustion
can be sustained by reaching to high recirculation
values ( more than 5) and chamber temperature
larger than 800°C.
4.2 Solid fuels
Stadler at el. [36] utilized a high velocity of the inlet
oxidizer to make the needed recirculation of the
product gases which are then entrained into the fresh
gases and the coal. Schematic design of the burner is
illustrated in Fig. 7. Air flow transfers the coal into the
furnace by the velocity around 10 m/s. Back recycled
high temperature flue gases makes the combustion air
lean and permeate straightly among the air entrances
to the coal jet due to the configuration of the nozzles,
as indicated in Fig.7 [36].
Fig. 7.Sketch of burner design and flow scheme
5. Environmentally sound characteristics of
flameless combustion
In recent decade, more stringent laws have been
ordained to cope with environmental issues and global
warming. . In combustion process the reaction occurs
between the fuel and the oxidizer to release heat
(thermal energy) as the required factor for electricity
generation. Also, a lot of emissions such as unburned
hydrocarbon (UHC), dioxide carbon (CO2), mono
oxide carbon (CO), nitride oxide (NOx), soot,
particulate matter (PM) are usually released to
atmosphere during combustion process. These
undesirable pollutants can jeopardize the environment
while the rate of their production increases due to
rapid industrialization. NOx (NO2 + NO) is usually
formed in presence of nitrogen and oxygen in very
high temperature conditions. Atmosphere can be
compromised by raising NOx formation in industrial
sectors. Particularly, acid rain, ozone depletion and
smog are the main consequences of more NOx
constitution [64]. Thermal NOx, prompt NOx and N2O
intermediate NOx formation are mentioned as the most
important NOx formation mechanisms. Nitrogen and
oxygen can react inside the combustion furnace in
extremely high temperature according to the reactions
which are called Zeldivich formulation [65]. This so-
call thermal NOx formation can be accelerated
exponentially at temperatures more than 1500oC [66].
Thermal NOx is suppressed in flameless combustion
due to low resident time, low oxygen concentration
and moderate temperature inside the chamber. In
conventional combustion the efficiency of chamber
increases significantly by using preheated air in
combustion process. However, NOx formation
augments drastically. In flameless NOx formation is
kept at very low level due to exhaust gas recirculation
(EGR) application [67]. These exhausted products are
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conducted into the fresh reactants inside the furnace;
therefore high peak temperature is omitted. As a
result, thermal NOx mechanism is suppressed, and
other insignificant NOx formations methods are
remained. N2O-intermediate NOx formation
mechanism is dominance method for NOx formation
in Hitac due to its moderate temperature and lean fuel
condition. Fig 8 shows a summary of flameless
combustion specifications and NOx formation
mechanisms
Fig8. Flameless combustion specifications
and NOx formation mechanisms
6. Conclusion
Clean combustion and energy efficiency are important
aspects of biofuel and fossil fuel consumption. Since
NOx is an important factor in the constitution of acid
rains and photo chemical smog, control of NOx
production is considered in combustion burners
designing. It has been proven that flameless
combustion systems can be applied vastly in different
industries. Energy savings by the amount of around
30%, also significant reduction of CO230%, reduction
in dimension of the furnace in addition of reduction in
pollutant emissions by 25%in compare of
conventional combustion can be cited as some benefits
of the flameless combustion. The aspects of flameless
combustion introduce it as a new technology to cover
the break among the seemingly paradoxical necessity
of hoarding energy
with respect to the production lower rates of NOx.
Flameless combustion method has been
considered as the optimum combustion technology
because it really saves energy and produces low level
of harmful emissions.
Furthermore, flameless oxidation has more salient
privileges as follow:
1.
2. Concusion
3.
1. The combustion region is extended over the whole
furnace; therefore, thermal gradients can be easily
controlled by applying flameless oxidation which
leads to prevent the formation of hot spots in the
combustion chamber. Therefore the temperature is
uniform in flameless combustion.
2. In combustion of low calorific value (poor quality)
fuels, flameless combustion shows better
performance. Because FLOX technique is able to
reduce emissions and energy consumption
significantly, this method can be considered as one
of the best applicable and useful combustion technologies in the international combustion
community.
3. As a result of reaction air preheating more energy
saving is expected.
4. The geometry of burner is simple; also thermal
efficiency of the burner in flameless operation is
improved. Outer surface temperature of furnace is
more uniform, which shows homogeneous
circumstances compare to the conventional mode,
this uniformity conditions are so useful for many
industrial applications particularly, steel factories.
5. The reaction happens in farther distances from the
face of the burner; therefore, the nozzle of burner
is not damaged. There is no ignition for flameless
combustion, peak of the temperature is low and the
flame is not audible. Moreover, the safety factor of
the colorless combustion system is better than
conventional flame combustion.
6. FLOX can be used for different types of fuels like
gaseous, liquid, and solid fuels in a variety of
conditions, with or without preheating air and fuel,
for stoichiometric, lean, or rich reactions also for
diffusion, partially premixed, or nonpremixed
combustion.
7. The chamber dimensions will be smaller in the
same capacity in compare of conventional
combustion.
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140
THE EFFECT OF GEOMETRICAL PARAMETERS ON
HEAT TRANSFER OF MICRO-CHANNELS HEAT SINK
Law Wen Zhea, Amer Nordin Darus
b
Faculty of Mechanical Engineering
Universiti Teknologi Malaysia, 81310 Skudai, Malaysia [email protected];
ABSTRACT
A three-dimensional numerical simulation on heat
transfer of trapezoidal microchannel was
conducted. Three sets of hydraulic diameter with 4
types of angles each were simulated using
commercial CFD package, FLUENT. The
numerical model was validated using available
literature. The effects of geometrical parameters
(hydraulic diameter, height, width) were
comprehensively studied. The best Reynolds
number was found by considering the Nusselt
number and pressure difference in channel.
Channel hydraulic size optimization was carried
out using thermal resistance as objective function
and pumping power as constraint. For channels of
the same hydraulic diameter, lower bottom-to-top
width ratio yields higher Nusselt number and lower
pressure difference. In addition, for channels of the
same hydraulic diameter, lower height-to-top width
ratio yields higher heat transfer performance.
Channels with 0.17 bottom-to-top width ratio and
0.41 height-to-top width ratio are found to be the
best in terms of Nusselt number and pressure
difference. By evaluating the Nusselt number and
pressure difference, the best Reynolds ranges from
140-180 for 100micron channel, 140-170 for
200micron channel and 141-167 for 300micron
channel. Finally, by evaluating the thermal
resistance and channel pumping power, the optimal
channel size is 300micron.
Keyword : Microchannel; Trapezoidal;
Optimization; Pressure Drop; Heat
Transfer.
1. INTRODUCTION
With the advent of advanced Integrated Circuits
and MicroElectroMechanical Systems, the
demands on heat removal and temperature control
in modern devices require new techniques for
providing high cooling rates. As shown by
Tuckerman and Pease [1], decreasing liquid
cooling channel dimensions to micro scale will lead
to increase of heat transfer rate.
Pfahler et al. [2] tested rectangular three
microchannels. The results showed that for
relatively
large cross section, friction coefficient increased
with increasing Reynolds number. Peng and
Peterson [3] investigated single-phase convective
heat transfer and flow characteristic of water in
small rectangular
channels. The experimental data revealed that
geometric configuration significantly affect
convective heat transfer as well as flow
characteristic. Qu et al. [4] studied heat transfer
characteristic in trapezoidal silicon microchannels.
The result showed that experimentally determined
Nusselt number is much lower compared to
numerical analysis. The authors attributed this
difference to the effects of surface roughness and
proposed a modified roughness-viscosity model to
interpret the experimental data. Qu and Mudawar
[5] investigated pressure drop and heat transfer
characteristic of a single-phase microchannel heat
sink. It was found that higher Reynolds number
reduces outlet temperature and temperature within
heat sink but at the expense of greater pressure
drop. Li et al. [6] performed three-dimensional
numerical simulations on the laminar convective
heat transfer in microchannel with non-circular
cross-section. It was found that heat transfer
intensity in trapezoidal micronchannel was higher
than compared to triangular microchannel.
Wu and Cheng [7] performed experiments on 13
different trapezoidal silicon microchannels. The
Nusselt number increases almost linearly at low
Reynolds number.
Chen et al. [8] presented a paper on optimum
thermal design of microchannel heat sinks using
the annealing method. The paper found that larger
flow power and smaller substrate thickness
provides lower thermal resistance. Tonomura et al.
[9] conducted a shape optimization of
microchannels using CFD and adjoint method. The
study concluded that the adjoint method can be
used to formulate computationally feasible
procedures for channel shape optimization. Wang
et al. [10] conducted a multi-parameters
optimization for microchannel heat sink using the
inverse problem method. The study concluded that
increased pumping power reduced the overall
thermal resistance of the design.
By referring to literature, most of the optimization
process are complicated and requires extensive
mathematical work. As far as the author is concern,
IMAT-UI 024
The 5th IMAT, November 12 – 13th
2012
141
few or none of the published paper considers
pressure difference and the best Reynolds number
that give minimum thermal resistance in their
optimization process.
2. OBJECTIVES The present work is dedicated to study the effects
of geometrical parameters on heat transfer of
microchannel heat sink and to determine the
optimal cross sectional area of microchannel in
terms of heat transfer performance and pressure
difference.
3. MATHEMATICAL FORMULATION
3.1 Governing Equations
To focus on the effect of the geometrical
parameters on the heat sink performance, the
following assumptions are made:
1) The governing equations based on Navier-
Stokes can be used to describe the physical
processes.
2) The process is steady and fluid is
incompressible.
3) The flow is laminar.
4) The left and right sides and the top of the
channel are assumed to be adiabatic.
5) The thermal properties of solid and water are
constants.
6) The effect of viscous heating is negligible.
7) The effect of buoyancy is negligible.
8) The effect of radiation heat transfer is
negligible.
The continuity, momentum and energy equations
for the current problem can be written as Li et al.
[6]:
Continuity:
Momentum:
Energy:
(a.) (b.)
Figure. 1: (a.) Schematic of microchannel (Li et al.
[6]). (b.) Cross section of modified channel model.
Nomenclature
cross sectional area
heated surface area
constant pressure specific heat of the flow
Dh , hydraulic diameter
DP pressure difference
apparent friction factor
height
effective heat transfer coefficient
k surface roughness
thermal conductivity
channel length
Normalized
, Nu Nusselt number
pressure
pumping power
q heat flux
combined convective and radiative heat
flux
Reynolds number
thermal resistance
temperature
reference temperature
wall temperature
velocity
x,y,z cartesian coordinates
mean fluid velocity
bottom width
top width
Greek Letters
pressure difference
mean temperature difference
thermal conductivity
viscosity
density
angle in degrees
Subscripts
fluid
glass
inlet
maximum
minimum
outlet
solid
The 5th IMAT, November 12 – 13th
2012
142
3.2 Boundary Conditions
From the model used in Li et al.[6], the inlet and
outlet velocity is given as:
(7)
The temperature for inlet and outlet is given as:
Some modifications were made to Li et al. [6]
model in order to simplify it. The thermal boundary
conditions are given as follows. The boundary
conditions of the left and right sides of the
computational domain are adiabatic:
At the bottom position, the heat flux is a given
value whereas the top is adiabatic:
A constant heat flux of is applied
to all the channels bottom used in this study. The
properties of water used are ,
, and
. The properties of silicon
used are , and
. The properties of pyrex glass
used are , and
.
The 5th IMAT, November 12 – 13th
2012
143
Table 1: Dimensions of channels used in the study.
4. METHODOLOGY
In order to investigate the heat transfer capability, a
unit cell of the complete heat sink is considered.
This is to simplify the investigation so as to find
out solely the effects of geometrical parameters on
the heat transfer of microchannel heat sink. The
analysis is confined to the domain where the
coolant enters the channel and leaves the channel.
The channel is created using GAMBIT. The heat
sink itself is not included in the analysis because
the effect of heat sink material and heat sink
conduction is not within the scope of this study.
The numerical computation was carried out using
commercial CFD package FLUENT by solving the
governing conservation equations and the boundary
conditions. The discretization of governing
equations in the fluid and solid regions was done
using finite-volume method (FVM) with second
order upwind method. The flow field was solved
using the SIMPLE algorithm.
Grid independence test is conducted to determine
optimum grid meshing size for the fluid region
inside the channel to minimize computation time
and also memory required for the modeling.
Computational cells with 51440 grids, 77160 grids
and 102880 grids were compared. Results shown
that 51440 grids with 20 edge interval count and 2
face interval size has almost the same result as
77160 grids. Thus, 20 edge interval count and 2
face interval size were used throughout the whole
study.
5. RESULTS AND DISCUSSION
5.1 Validation
Using optimum grid system, the present numerical
model is validated with available experimental
result by Li et al. [6]. In Fig. 2., there is an
appreciable difference between the numerical
simulation and experimental data. This deviation
may come from the surface roughness effect. The
effect of surface roughness on the flow and heat
transfer in tube becomes apparent as channel size
decreases to the order of microns. In the case of
numerical simulation, FLUENT does not have the
capability to account for roughness effect.
Figure 2: Nusselt number validation
Channel
Wt
(micron)
Wb
(micron) H (micron) Wb/Wt H/Wt k/Dh L/Dh
Surface
material
100micron
45degrees 241.6 41.6 100 0.17218543 0.413907 0 300 silicon
100micron
55degrees 192.2 52.158 100 0.271373569 0.520291 0 300 silicon
100micron
65degrees 157 63.738 100 0.405974522 0.636943 0 300 silicon
100micron
75degrees 130.4 76.81 100 0.589033742 0.766871 0 300 silicon
200micron
45degrees 483 83 200 0.17184265 0.414079 0 150 silicon
200micron
55degrees 385 104.916 200 0.272509091 0.519481 0 150 silicon
200micron
65degrees 314 127.476 200 0.405974522 0.636943 0 150 silicon
200micron
75degrees 260.8 153.62 200 0.589033742 0.766871 0 150 silicon
300micron
45degrees 725 125 300 0.172413793 0.413793 0 100 silicon
300micron
55degrees 577 156.86 300 0.271854419 0.519931 0 100 silicon
300micron
65degrees 471 191.214 300 0.405974522 0.636943 0 100 silicon
300micron
75degrees 391 230.23 300 0.588823529 0.767263 0 100 silicon
The 5th IMAT, November 12 – 13th
2012
144
Table 2: Dimension of channels used in validation
Data
types
Wt
(micr
on)
Wb
(micr
on)
H
(micr
on)
Wb/
Wt H/Wt L/Dh
Surfa
ce
mater
ial
Heat
Flux
Exper
iment
al Li
et al. 770.5 672.6 56.34
0.872
94
0.073
12
299.0
2
silico
n
1.00E+0
6
Prese
nt
work
(2012
) 770.5 672.6 56.34
0.872
94
0.073
12
294.1
17
silico
n
1.00E+0
6
5.2 Nusselt Number
A comparison with literature was made using
almost the same channel dimension. From Fig. 3.,
the present numerical model underestimates at
Reynolds number lower than 240 and
overestimates at Reynolds number higher than 240.
This may due to the constant viscosity model of the
current study. The equation for Nusselt number
taken from (Mahdi, [11]):
Viscosity, μ, has some appreciable effect on the
Nu. But, as an overall, constant properties model
can be used as the trend of Nusselt number against
Reynolds number correlates well with literature.
Gunnasegaran et al. [12] employed constant
property model in their numerical studies to study
the effects of geometrical parameters as well. From
Fig. 3., it is shown that at higher Reynolds number,
the Nusselt number tends to approach a constant.
This correlates well with literature as Mala and Li
[13] show that transition flow regime started at
Re=650.
Figure 3: Nusselt number comparison.
Table 3: Dimension of channels used in
comparison
Data
types
Wt
(micr
on)
Wb
(micr
on)
H
(micr
on)
W
b/
Wt
H/
Wt
L/
Dh
Surfa
ce
mate
rial
Heat
Flux
Experi
mental
Li et al. 770.5 672.6 56.34
0.8
72
0.0
73
29
9.0
silico
n
1.00E
+06
Experi
mental
Wu et
al.
770.4
8
672.6
3 56.34
0.8
73
0.0
73
29
8.6
silico
n
1.00E
+06
Numeri
cal Li et
al. 770.5 672.6 56.34
0.8
72
0.0
73
29
9.0
silico
n
1.00E
+06
Present
work
(2012) 770.5 672.6 56.34
0.8
72
0.0
73
29
4.1
silico
n
1.00E
+06
In order to find out the effects of geometrical
parameters, three sets of channel sizes were used,
that is: 100micron, 200micron and 300micron
hydraulic diameter. Next, with each hydraulic
diameter, 4 types of channel each of different
angles were used, that is: 45degrees, 55degrees,
65degrees and 75degrees. In order to easily
compare with data available in literature, the values
of Wt/Wb, H/Wt, k/Dh and L/Dh were given for
easy reference. These values are important
parameters with which the effects of geometrical
parameters are to be investigated. Surface heat
transfer coefficient, , is given by the following
equation:
Surface Nusselt number is calculated using the
following equation:
From Fig. 4., Fig. 5. and Fig. 6., it is shown that
channels with 45degrees exhibit the highest Nusselt
number for the same given hydraulic diameter. It is
clear that as and increased, Nusselt number
decreased for channels of the same sizes. This is
consistent with the study made by Wong [14] who
found out that reduced distance between side walls
in channel will increase the velocity gradient at the
wall boundaries. This in turn lowers thermal
resistance and increases Nusselt number. In the
case of trapezoidal channel, the smaller heated
surface at channel bottom meant smaller distance
between channel walls. Velocity gradient between
channel wall boundaries at channel bottom
increased due to the smaller distance. This
phenomenon likewise caused the increment of
Nusselt number for 45degrees channel, followed by
55degrees channel, 65degrees channel and finally
75degrees channel. When the value increased,
channel top area decreased in proportion to channel
The 5th IMAT, November 12 – 13th
2012
145
height, H. The decrement of channel top area may
cause the fluid convection between channel top and
channel bottom to decrease due to fact that fluid at
the upper volume remains unheated or not fully
heated compared to the fluid at the channel bottom.
Larger top area means higher volume of unheated
fluid can be carried and effective fluid convection
can occur.
As channel height, H, increased, the Nusselt
number increased as well. Larger channel height
means higher mass flow rate and heat removing
capacity. The present result correlates well with
literature as shown by Cheong [15] who found out
that for channels of same hydraulic diameter, the
size of heated surface does not necessarily increase
heat transfer capabilities of microchannels. The
study showed that channel height also affects heat
transfer performance of a channel.
The result from present study is compared with
available literature. In the experiment conducted by
Wu et al. [7], it was found that as decreased,
Nusselt number decreased.
Figure 4: Nu against Re for 100miron
Figure 5: Nu against Re for 200miron
Figure 6: Nu against Re for 300miron
5.3 Pressure Difference
The pressure difference of present numerical study
was compared with experimental results from
literature. From Fig. 7., it was shown that
appreciable difference occurred between numerical
model and experimental data. This is due to the
different dimension of channels used in both cases.
The trend of the pressure difference correlates well
with literature as shown in Fig.7.
Pressure difference is one of the major issues in
optimization of channel. Higher pressure difference
increases pumping power requirements. The
equation for pressure difference taken from Wu et
al. [7] is given as:
Figure 7: Pressure difference comparison
The 5th IMAT, November 12 – 13th
2012
146
Table 4: Dimension of channel used in PD
comparison
Data types
Wt
(micron)
Wb
(micron)
H
(micron) Dh Wb/Wt H/Wt L/Dh
Surface
material
Heat
Flux
Experimental
Qu et al. 237.01 66.11 109.77 115 0.27893 0.463 262 silicon zero
Present study
(2012) 770.5 672.6 56.34 102 0.87294 0.073 294 silicon 1.00E+06
Fig. 8., Fig. 9. and Fig. 10. showed that pressure
difference is the lowest for 45degrees channel,
followed by 55degrees, 65degrees and finally
75degrees. This may due to the bigger channel
bottom surface area as the trapezoidal angles
changes from 45dgrees to 75degrees. The pressure
difference decreases as hydraulic diameter
increases. Since the same channel depth is being
tested for the same hydraulic diameter, ie:
100miron height for 100micron hydraulic diameter
and so forth, the results also show that pressure
difference is inversely proportional to channel
depth. The result correlates well with literature as
Harms et al. [16] also indicated that pressure drop
is inversely proportional to depth of channel.
Upon close inspection, the magnitudes of the
velocity at channel bottom for both channels are
found to be relatively the same. This means that,
although the velocity contour is different when the
fluid is developing, it does not affect the velocity at
the channel bottom for both channels whilst
developing and after fully developed. Therefore,
the effect of velocity can be ruled out in explaining
the difference of pressure drop between the
45degrees channel and 75degrees channel. The
author proposed that the difference in pressure drop
may be attributed to the difference in surface area
of channel bottom.
Figure 8: PD against Re for 100micron
Figure 9: PD against Re for 200micron
Figure 10: PD against Re for 300micron
The 5th IMAT, November 12 – 13th
2012
147
5.4 Reynolds Number
Analysis was done to find out the best Reynolds in
which channel exhibits the most optimal heat
transfer performance. Both normalized Nusselt
number and normalized pressure difference were
calculated in order to find the best range. Equation
of normalized Nusselt number and normalized
pressure difference are as follows:
The best Reynolds number can be found by finding
the intersecting point between the normalized
Nusselt number and the normalized pressure
difference. This intersecting point corresponds to
the highest Nusselt number and lowest pressure
difference possible. From Fig. 11., Fig. 12. and Fig.
13., it is shown that the best Reynolds ranges from
140-180 for 100micron, 140-170 for 200micron
and 141-167 for 300micron. These ranges of
Reynolds were then used as a reference in the
calculations of channel size optimization in the
next section.
Figure 11: Normalized Nu against Normalized
PD 100micron
Figure 12: Normalized Nu against Normalized
PD 200micron
Figure 13: Normalized Nu against Normalized
PD 300micron
5.5 Channel Size Optimization
or pumping power is the power required to
drive the flow inside a microchannel. Pumping
power is equal to the product of volumetric flow
rate and pressure drop as shown by the equation:
Thermal resistance is given by the following
equation:
Optimization of channel sized was done by first
calculating the normalized channel thermal
resistance and normalized pumping power. Then
these two parameters were plotted against channel
diameter in order to find the optimal size for heat
transfer. The calculation for normalized channel
thermal resistance and normalized pumping power
were given as follows:
Lower pumping power and thermal resistance is
desirable in order to optimize the channel size. Fig.
14. showed that the optimal size is 300micron. This
may due to the fact that for larger hydraulic
diameter, pressure difference between inlet and
outlet is lower. The thermal resistance for larger
hydraulic diameter is also lower. Hence, 300micron
channel exhibits the best heat transfer. The result of
this channel optimization is in agreement with Ong
[17] and Chan [18]. Both Ong [17] and Chan [18]
The 5th IMAT, November 12 – 13th
2012
148
also found out that higher hydraulic diameter
performs better in heat transfer capability.
Figure 14: Normalized Thermal Resistance
against Normalized Pumping
Power
6. CONCLUSION
Numerical simulation on the fluid flow and heat
transfer characteristics in trapezoidal microchannel
was conducted in this study. The effects of
geometrical parameters (hydraulic diameter, height,
width) were extensively studied. Based on the
results, the following conclusions can be made:
1) The Nusselt number in the fully developed
region increases as the Reynolds number
increases. This is in disagreement with
conventional theory of flow in ducts where the
fully developed Nusselt number is a constant.
2) The bottom-to-top width ratio, the height-to-
top width ratio and the hydraulic diameter of
trapezoidal channel were found to have great
effect on the laminar Nusselt number.
3) For channels of the same hydraulic diameter,
lower bottom-to-top width ratio yields higher
Nusselt number. This is in disagreement with
the study made by Wu et al. [7].
4) For channels of the same hydraulic diameter,
lower bottom-to-top width yields lower
pressure difference.
5) For channels of the same hydraulic diameter,
lower height-to-top width ratio yields higher
heat transfer performance.
6) For each of the hydraulic diameters, channels
with 0.17 bottom-to-top width ratio and 0.41
height-to-top width ratio are found to be the
best in terms of Nusselt number and pressure
difference.
7) By evaluating the Nusselt number and pressure
difference, the best Reynolds ranges from 140-
180 for 100micron channel, 140-170 for
200micron channel and 141-167 for
300micron channel.
8) By evaluating the thermal resistance and
channel pumping power, the optimal channel
size is 300micron.
References
[1] Tuckerman, D.B. and Pease, R.F. (1981).
High Performance Heat Sinking for VLSI.
IEEE Electron Device Letters. Vol. EDL-2,
No.5
[2] Pfalher, J., Harley, J., Bau, H.H., and Zemel,
J.N. (1990). Liquid transport in micron and
submicron channels. Sensors Actuators. A21–
A23 431–434.
[3] Peng, X.F. and Peterson, G.P. (1996).
Convective heat transfer and flow friction for
water flow in microchannel structures. Int. J.
Heat Mass Transfer. 39 (12) 2599-2608.
[4] Qu, W., Mala, G.M., and Li, D. (2000).
Pressure-driven water flows in trapezoidal
silicon microchannels. Internat. J. Heat Mass
Transfer. 43 353– 364.
[5] Qu, W. and Mudawar, I. (2002). Experimental
and numerical study of pressure drop and heat
transfer in a single-phase micro-channel heat
sink. Internat. J. Heat Mass Transfer. 45
2549–2565.
[6] Li, Z., Tao, W.Q., and He, Y.L. (2006). A
numerical study of laminar convective heat
transfer in microchannel with non-circular
cross-section. International Journal of
Thermal Sciences. 45 1140–1148.
[7] Wu, H.Y. and Cheng, P. (2003). An
experimental study of convective heat transfer
in silicon microchannels with different surface
conditions. Internat. J. Heat Mass Transfer.
46 2547–2556.
[8] Chen, C.W., Lee, J.J., and Kou, H.S. (2008).
Optimum thermal design of microchannel
heat sinks by the simulated annealing method.
International Communications in Heat and
Mass Transfer. 35 980–984.
[9] Tonomura, O., Kano, M., and Hasebe, S.
(2010). Shape Optimization of Microchannels
Using CFD and Adjoint Method. 20th
European Symposium on Computer Aided
Process Engineering – ESCAPE20. 37-42.
[10] Wang, Z.H., Wang, X.D., Yan, W.M., Duan,
Y.Y., Lee, D.J., and Xu, J.L. (2011). Multi-
parameters optimization for microchannel
heat sink using inverse problem method.
International Journal of Heat and Mass
Transfer. 54 2811–2819.
[11] Mahdi Zhaleh Rafati (2010). Numerical
simulation of fluid flow and heat transfer in a
trapezoidal microchannel. Master of
Engineering, Universiti Teknologi Malaysia,
Skudai.
The 5th IMAT, November 12 – 13th
2012
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[12] Gunnasegaran, P., Mohammed, H.A., Shuaib,
N.H. and Saidur, R. (2010). The effect of
geometrical parameters on heat transfer
characteristics of microchannels heat sink
with different shapes. International
Communications in Heat and Mass Transfer.
37 1078–1086.
[13] Mala, G.M. and Li, D. (1999). Flow
characteristics of water in microtubes.
Internat. J. Heat Fluid Flow. 20 142–148.
[14] Wong Wai Hing (2005). Numerical
simulation of a microchannel. Bachelor of
Engineering, Universiti Teknologi Malaysia,
Skudai.
[15] Cheong Tuck Meng (2006). A micro-channel
heat transfer with a heat source. Bachelor of
Engineering, Universiti Teknologi Malaysia,
Skudai.
[16] Harms, T.M., Kazmierczak, M.J., Gerner,
F.M., Holke, A., Henderson, H.T.,
Pilchowski, J., and Baker, K. (1997).
Experimental investigation of heat transfer
and pressure drop through deep
microchannels in a 110 silicon substrate.
Proceedings of ASME Heat Transfer Division,
in: ASME HTD. vol. 351-1, pp. 347–357.
[17] Ong Jiun Shyong (2012). The effect of Prandtl
number to the performance of microchannel
heat sink. Bachelor of Engineering Thesis,
Universiti Teknologi Malaysia, Skudai.
[18] C
han Zaolon (2012). Second law analysis of
microchannel. Bachelor of Engineering
Thesis, Universiti Teknologi Malaysia, Skuda
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2012
150
Investigation of the Velocity Profiles in a Ninety-Degree
Curved Standing Wave Resonator with PIV
Normah M. G.a, Irfan Abd. R.
b, Quenet T.
c, and Zaki Ab.M.
b
aFaculty of Mechanical Engineering
University Teknologi Malaysia (UTM) Skudai, Johor
Tel : (021) 7270011 ext 51. Fax : (021) 7270077
E-mail : [email protected] bSchool of Mechanical Engineering
University Malaysia Perlis (UniMAP), Kangar, Perlis
Tel : (04) 9885035 Fax : (04) 9885034
E-mail : [email protected] cUniversite de Rennes 1, IUT de Saint Malo, France
ABSTRACT
Travelling wave thermoacoustic heat engines have
been reported to have a higher efficiency than the
standing wave ones. The former are generally large
systems which consist of toroidal shape resonators.
While standing wave heat engines are inherently
smaller, a reduction in size could be considered
which may involve curvatures as compared to the
straight tube conventional systems. However, as with
the streaming losses in the travelling wave
resonators, losses due to the curvature may be
generated. This study involves preliminary
experimental measurements using the Particle Image
Velocimetry (PIV) method to analyze the velocity
profiles in a standing wave resonator before and after
a ninety degree curvature. This design can reduce the
space generally occupied by the straight standing
wave resonator. The overall length of the resonator
fits a quarter wavelength wave based on the straight
closed-end tube type. The working gas is air at 1
atmospheric pressure. Results have shown that the
velocity profiles after the stack but before the
curvature exhibit clear straight paths up just as
reported elsewhere. Signs of disordered motion could
be observed just before the bend and the pattern
continues until after the curvature. The results are
obtained before one periodic cycle and before the
acoustic wave front hit the tube end. The trend is
expected to affect the overall thermoacoustic
performance of the engine as returning gas particles
interact with the oncoming particles that pass by the
curvature.
Keywords : Standing wave resonator,
thermoacoustic heat engines,
Particle Image Velocimetry,
Curvature
1. INTRODUCTION
Thermoacoustic heat engines and refrigerators are
devices that utilize acoustic waves to drive a
thermodynamic process. A thermoacoustic
refrigerator generates cooling from solid-fluid
interactions. Acoustic waves establish a temperature
gradient, transferring heat from one to another, as
fluid particles oscillate over solid boundaries.
Connected to the ambient heat exchanger and a
cooling load, a thermoacoustic refrigerator poses an
attractive alternative to the conventional system.
However, thermoacoustic systems have yet to hit the
commercial market due to the high cost associated
and the selected research community that fully
understand the concept. Studies are being done to
explore possibilities of potential practical applications
as well as towards the better understanding of the
related theories.
There are two types of thermoacoustic systems; the
travelling wave and the standing wave. The former
generally involves a large curved tube with a
regenerator placed inside for the solid-fluid
thermoacoustic effects to take place whilst the
simplest standing wave system consists of a long
straight resonator with a stack located somewhere
between the maximum pressure and displacement of
the oscillating fluid particles. Inspired by the
curvatures associated with the generally large
travelling wave systems, this study looked into the
possibilities of reducing the long tube of a standing
wave resonator by having it bent at a ninety degree
angle. Flows around bend may initiate disorderly flow
at certain velocities and this is being investigated
here. Unlike the common flow around curvatures,
however, the present work involves acoustic waves
that have passed through a thermoacoustic stack
before proceeding by a ninety-degree bend.
IMAT-UI 025
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2012
151
The solid-fluid interactions that generate and
maintain the thermoacoustic effects are of importance
and the heart of thermoacoustic systems.
Experimental research into the thermoacoustic
phenomena include optical methods that are used to
measure the flow activities within such systems
because they are non-intrusive. Visualization
techniques of flow field that have been used in
pulsating/oscillating flow experiments include
Holographic Interferometry [1, 2], Laser Doppler
Anemometry (LDA) [3] and Particle Image
Velocimetry (PIV) [4-7]. The latter gives velocity
data over a large area which is favorable in
thermoacoustic experiments. Fig 1 shows a schematic
of a standing wave thermoacoustic refrigerator
(without the cold and ambient heat exchangers)
consisting of a closed-end long tube with an acoustic
driver on one end.
Figure 1: A schematic of a basic thermoacoustic
refrigerator
2. Experimental Set-Up
The resonator is made from acrylic for transparency,
a necessity in the PIV experiment. It has a square
cross-section of 80x80 mm2 with a center-line length
of 86 meters which follows that of Blanc-Benon et al
[5]. The length was based on a previously completed
design with the same PIV apparatus [8] which was
configured on a half-wavelength straight resonator.
Standing wave resonators are made to fit a quarter or
half wavelength wave within the tube, the shorter one
having lesser losses. However, the present study
followed that of previous reported thermoacoustic
experiments with PIV which utilized a half-
wavelength tube. Fig 2 shows the resonator
fabricated with a ninety-degree curvature.
Figure 2: The ninety-degree acrylic resonator
The experimental set-up is shown in Figure 3. The
resonator was filled with air at atmospheric pressure
with the thermoacoustic stack made of glass plates.
The stack consisted of parallel non-conducting plates
of thickness 2mm and separated by a distance of 1
mm. Each plate is 25 mm long and 79 mm wide.
The stack was positioned 21.5 cm away from the
driver where the amplitude for both pressure and
velocity is high enough for the thermoascoustic
effects to take place within the stack.
The resonator was seeded with smoke particles
located near the acoustic driver. An Nd-YAG laser
and a CCD camera were used to detect and capture
the particles motion. The driver generated acoustics at
200 Hz, which is the theoretical resonant frequency, f,
of this particular resonator based on the simple
relationship,
nL
RTf
(1)
where , R, T, and L are the ratio of specific heats,
gas constant, operating tempearture and resonator
length respectively. Since the tube was designed for a
half-wavelength resonator, n is equal to 2 in this case.
A function generator produced sinusoidal waves
which was amplified by a power amplifier before
passing through a loud-speaker which acted as the
acoustic driver.
Figure 3: The experimental set-up for the PIV
The 5th IMAT, November 12 – 13th
2012
152
3. Results and Discussion The PIV measurements are taken 5 seconds after the
acoustic driver was turned on. Twenty sets of data
were collected each for before and after the curvature
0.5 seconds apart. Since the domain covered by the
PIV is comparatively large, selected regions are
shown here to draw attention to those affected by the
curvature. Fig 4 shows the velocity profiles before
the curvature at locations a, b, c, and d, as described
in Fig 3.
Since the images were captured closed to the ninety-
degree bend, the curved wall effects are clearly seen
here, particularly on the right hand side of (a) and (d),
the top and bottom of the resonator, closer to the
walls. Previous work with PIV on profiles near the
stack showed that profiles are laminar away from the
wall effects with vortices forming close to the stack
walls and significant ones behind think plates, soon
after start-up as well as after steady-state is established [4, 7]. Fig 5 shows the velocity profiles
after the ninety-degree bend.
Figure 4: Velocity profiles after the stack before the curvature at positions (a); (b); (c); (d)
Figure 5: Velocity profiles after the stack after the curvature at positions (a); (b); (c); (d)
The 5th IMAT, November 12 – 13th
2012
153
This time, the curvature effects are observed on the
left hand side of the figures, in all the figures, (a)
through (d). The trend after the bend also exhibits a
tendency towards a disorderly flow as seen
throughout Fig 5. Based on the quarter-wavelength
standing wave resonator calculations, the first wave
front has not hit the tube end yet. This non-linear
behavior is expected to intensify as returning
particles meet with oncoming ones over several
periodic cycles. The consequences are undesirable
since there will be energy lost within the already low
performing standing wave thermoacoustic system.
Thus, a ninety-degree curvature is probably not a
favorable possibility in a standing wave resonator.
7. Conclusion
A preliminary investigation with PIV on a ninety-
degree curved half-wavelength standing wave
thermoacoustic resonator was completed. Although
images were obtained before one periodic cycle was
completed or even before the first wave front hit the
tube end, results showed that a ninety-degree bend is
not a favorable design. This is because the velocity
profiles obtained indicate that the particles did not
maintain its straight paths, moving towards disorderly
behavior as particles approached the bend, worsening
after the bend. The situation is expected to aggravate
as this is a standing wave resonator.
ACKNOWLEDGMENT
The authors wish to thank Ministry of Education
FRGS-KETTHA(9003-00352) for the research grant,
Universiti Teknologi Malaysia (UTM) and Universiti
Malaysia Perlis (UniMAP) for the facilities to do the
research. The authors also appreciate the help from
Mr Johari, the technician at the PIV laboratory who
has assisted in the series of experiments.
REFERENCES
[1] Majid N, and Kamran S. A Critical review on
Advanced Velocity Measurement Technique in
Pulsating Flows. Meas. Sci. Tecnology
2010(21);042002-19pp.
[2] Herman C, Kang E, Wetzel M. Expanding the
Applications of Holographic Interferometry to
the Quantitative Visualization of
Oscillatory Thermofluid Processes Using
Temperature as Tracer. Experimental Fluid1998
(24); 431-446
[3] Bailet H, Lotton P, Bruneau M, Gusev V,
Valiere JC, and Gazembel B. Acoustic
PowerFlow Measurement in Thermoacoustic
Resonator by Means of Lasser Doppler
Anemometry (L.D.A) and Microphonic
Measurement. Appl. Accoustic 2000, 60(1);1-
11
[4] Blance-Benon P, Besnoin E, and Knio O.
Experimental and Computational Visualization
of the Flow Field in a Thermoacoustic Stack. C.
R. Mecanique 2003. 331;17-24
[5] Argenthael B, Marc M, and Blanc-Benon P.
Measurement of Acoustic velocity in the Stack
of a Thermoacoustic refrigerator using
[6] Particle Image Velocimetry. J Heat Mass
Transfer 2007. DOI 10.1007/s00231-007-0316-
x
[7] Shi L, Yu Z, Jaworski AJ, and Abduljalil AS.
Vortex Shedding at the End of Parallel-Plate
Thermoacoustic Stack in the Oscillatory Flow
Condition. World Academy of Science,
Engineering and Technology 2005; 49
[8] Mao X, Marx D, and Jaworski AJ. PIV
Measurement of Coherent Structures and
[9] Turbulence Created by an Oscillating Flow at
the End of Thermoacoustic Stack. Proceeding of
the ITI Conference in Turbulence 2005. 25-28
September. Bad-Zwishenahn, German.
[10] Irwan SA and Mohd-Ghazali N. Stack geometry
Effects on Flow Pattern with Particle Image
Velocimetry (PIV). Jurnal Mekanikal 2011; 33
The 5th IMAT, November 12 – 13th
2012
154
MED+AD Desalination Cycle
Muhammad Wakil Shahzad a, Kim Choon Ng
a,b* , Wai Soong Loh
a, Won Gee Chun
c
a Department of Mechanical Engineering,
National University of Singapore,
9 Engineering Drive 1, Singapore 117576, Singapore b Visiting professor(sabbatical), Water Desalination and Reuse Centre,
King Abdullah University of Science & Technology, Thuwal, 23955-6900, Saudi Arabia. c Department of Nuclear and Energy Engineering,
Cheju National University, 66 Jejudaehakno, Jejusi, South Korea
ABSTRACT
The MED+AD cycle is a hybrid MED+AD. It has
the potential to increase the water production rate of
traditional MED plant by extracting vapor from the
last MED effect and injecting the desorbed vapor into
first stage of MED. It lowers the operational
temperature of MED+AD as compared to traditional
MED. The lower temperature operation typically
ranges from 5-50oC which reduces the corrosion and
fouling chances as well as scavenging the energy
from ambient in last stages of MED. Simulation
results are presented for MED+AD using FORTRAN
linked with IMSL. It can be seen from results that the
top brine temperature can be as low as 50C and the
concentration can be as high as 120,000ppm.
Keywords : Desalination, MEDAD, MED with
thermal compressor, Low
temperature MED.
1. INTRODUCTION
The fresh water demand is increasing with population
increase. The total water on earth covers almost
three-fourth of its surface. However, the seawater is
almost 97% of total water and only a small amount
about 3% is fresh water. The useable water is less
than 1% of fresh water available [1-7]. Global water
consumption is doubling every 20 years, more than
twice the rate of human population growth. If current
trends persist, in 2025, about 67% of population will
be under water stress as shown in Figure1.
1.1- Seawater Desalination
To address these looming crises of fresh water, the
sea is only unlimited source of water that can be use
to fuel the world population in future. The main
bottleneck in direct use of sea water is its salinity
level (≥35,000ppm). The sea water can be desalinated
by desalination methods to reduce the salt
concentration (≈500ppm) to make it useable.
Seawater desalination is being applied at 58% of
installed capacity worldwide, followed by brackish
water desalination accounting for 23% of installed
capacity [2]. The Figure 2 shows the worldwide
installations on the basis of feed water.
In thermal desalination systems, the MED process
has great potential because of high recovery ratio and
high thermal efficiency. Many researchers like; A. E.
Al-Nashar et al. [8], M.Al Shammiri et al. [9], A.O.
Bin Amer [10], J. Blanco et al. [11] and H.K.
Sadhukhan [12] provided the extensive detail on
MED design and operation. The top brine
temperature (TBT) of traditional MED is ranges from
100~ 120C. This high TBT accelerate the corrosion
and fouling of evaporator at high salt concentration.
High corrosion and fouling rate increase the capital
IMAT-UI 026
Figure 1: World population under water stress in
future
(Source: http://experience.sika.com/innovations)
Figure 2: Worldwide desalination
installations on the basis of
feed water [2].
The 5th IMAT, November 12 – 13th
2012
155
and operational cost of plant and also reduce the plant
life.
In conclusion, if above mentioned problem (high
TBT) can be solved; MED can be an attractive
process for desalination in future. To solve this, the
traditional MED is combined with an AD beds and
this novel system called as MED+AD desalination
cycle. The AD cycle shift the whole desalination
cycle toward lower operating temperature (5~50C)
and also increase the performance of plant.
MED+AD cycle simulation is completed by using
FORTRAN linked with IMSL. It is found that by combining the MED with AD, the production rate
increases by 35-40%.
2. MED+AD MODELING
MED+AD is the combination of traditional MED and
adsorption cycle AD. The vapors from last stage from
MED are adsorbed by AD beds, and the desorbed
vapors from AD bed injected back between steam
generator and the first stage of MED. To overcome
the pressure losses due to low pressure desorbed
vapor mixing, steam jet ejector is used. This jet pump
helps to recover the pressure loss due to mixing of
low pressure desorbed vapors. The detailed
model of three stage MED coupled with AD is shown
in Figure 3.
The modeling of the system is completed by using
1) mass conversation, 2) energy conservation and,
3) salt conservation equations. The overall heat
transfer coefficient is calculated by using falling
film correlation developed by M.W. Shahzad et al.
[13] for low pressure evaporation. This novel
correlation is given in equation 1. For the ejector,
in addition to these three equations, momentum
conservation is applied to calculate the throat
velocity and throat diameter. The throat diameter is
designed to make sure that the primary steam
pressure must drop below the desorption pressure
to pull the desorbed vapors.
47.084.0
89.038.0
85.345.0
16.0
32
2
.65.2
1exp.2
PrRe
...00143.0
ref
g
ref
sat
o
ll
l
nevaporatio
v
v
T
q
T
T
S
S
kg
h
(1)
3. RESULTS AND DISCUSSION
MED and AD modeling equations are written in
separate user defined sub-routine of FORTRAN. The
IMSL is used to solve the equations simultaneously.
The tolerance 1x10-7
is used to converge the solution.
Figure 4 shows the adsorber bed and desorber bed
temperatures profiles. The temperature of bed is
increasing during desorption due to heat supply by
hot water for desorption. On other hand, the
temperature of bed is dropping during absorption due
to heat taken by cold water. The switching time helps
the beds (pre-heating & pre-cooling) to prepare for
next operation.
The Figure 5 shows the distillate production from
traditional MED and an advanced MED+AD plant.
The dotted lines represent the typical MED
production. It can be seen that when MED coupled
with AD machine the production increases from 35-
40%. The desorbed vapour pulled by primary vapour
produced in SG in steam jet ejector and mixture of
these two streams introduced into the 1st stage of
MED. There is sharp increase in distillate production
Figure 3: MED+AD detailed
model
Figure 4: AD beds temperatures
profile
The 5th IMAT, November 12 – 13th
2012
156
and then drop gradually due to desorption rate
decrease. It can also be seen that during switching
when no desorption vapour are available the
production is just by traditional MED plant and it
catching the dotted lines of respective effects.
The Figure 6 shows the temperature profiles of AD+
MED stages. The SG and two stages have feed pre-
heater loop inside the chamber to recover the heat
from brine. The dotted line shows the feed
temperature increase from ambient (30C) in feed pre-
heaters. The pressure drop due to mixing of desorbed
vapour is recovered by jet pump. It is observed that
temperature difference between stages varies from 2-
3C.
The salt concentration increase in each effect is
shown in Figure 7. The brine is cascaded to get the
flash affect in next lower pressure stages.
4. CONCLUSION
To overcome the limitations of conventional MED
plants, an advanced desalination cycle is proposed.
The proposed cycle is combination of traditional
MED and AD cycle. In MED+AD cycle, top brine
temperature (TBT) reduced to 50C as compared to
120C in practical traditional MED plants. The low
grade waste heat or solar energy can operate the
system. The distillate water production rate is 35-
40% higher than traditional MED plants as well as
ambient energy is scavenging in last stages of MED.
It is also found that the highest concentration
(120,000ppm) of brine exposed to lowest
temperatures (5C) that reduces the chances of
corrosion and fouling of MED evaporators.
NOMENCLUTURE
l = Liquid viscosity (kg/m-sec)
l = Liquid density (kg/m3)
lk= Liquid conductivity (W/m-K)
Re= Film Reynolds number
Pr = Prandtl number
S = Feed water salinity (ppm)
oS= Reference sea water salinity (30000ppm)
satT=Evaporator saturation temperature (K)
refT= Reference saturation temperature (K) ( refT
=
322.15K )
q = input heat flux (W/m2)
gv= vapor specific volume
ΔT= Tch,out – Tevap
Figure 5: MED+AD production
profile temperatures profile
Figure 6: Temperature profiles of MED+AD
Figure 7: MED+AD salt concentration
profile
The 5th IMAT, November 12 – 13th
2012
157
ABBREVIATION
MED = Multi effect desalination
AD = Adsorption desalination
IMSL =International math and state library
TBT = Top brine temperature
SG = Steam generator
ACKNOWLEDGEMENT
The authors wish to thanks to Dr. Aung Myat for help
and guidance during simulation programming.
REFERENCES [1]. S. A. Kalogirou, Seawater desalination using
renewable energy sources, Progress in Energy
and Combustion Science 31 (2005) 242–281.
[2]. M. A. Eltawil, Z. Zhengming, L. Yuan, A review
of renewable energy technologies integrated
with desalination systems, Renewable and
Sustainable Energy Reviews 13 (2009) 2245–
2262
[3]. J. Buff, P. Re, R. Glynn, Benfield, Silvio
Fischer, Desalination Plants – Technological
development, Risks affecting Engineering
Insurers and Claims Experience, IMIA - WGP
57 (08) Conference Gleneagles, 16th Sep. 2008.
[4]. P. Dickie, Desalination: option or distraction for
a thirsty world? , WWF‘s Global Freshwater
Programme, June 2007,
(www.melaleucamedia.com).
[5]. A. D. Khawaji, I. K. Kutubkhanah, J. m. Wie,
Advances in seawater desalination technologies ,
Desalination 221 (2008) 47–69
[6]. http://environment.nationalgeographic.com/envir
onment/freshwater/freshwater-crisis/
[7]. http://www.globalchange.umich.edu/globalchang
e2/current/lectures/freshwater_supply/freshwater
.html.
[8]. A. M. El-Nashar, A. A. Qamhiyeh, Simulation of
the steady state operation of a multi effect stack
seawater distillation plant, Desalination 101
(1995) 231-243.
[9]. M. A1-Shammiri, M. Safar, Multi-effect
distillation plants: state of the art, Desalination
126 (1999) 45-59.
[10].A.O. Bin Amer, Development and optimization
of ME-TVC desalination system, Desalination
249 (2009) 1315–1331.
[11].J. Blanco, E. Zarza, D. Alarcón, S. Malato, J.
León, Advanced Multi-Effect Solar Desalination
Technology: The PSA Experience , CIEMAT -
PSA, P.O. Box 22, 04200 Tabernas (Almería),
Spain.
[12]. H. K. Sadhukhan and P. K. Tewari, Small
Desalination Plants (SDPS), Thermal
Desalination Processes – Vol. II - Small
Desalination Plants (SDPs), Bhabha Atomic
Research Centre, Mumbai 400085, India.
[13]. M. W. Shahzad, A. Myat, C. W. Gee and K. C.
Ng, Bubble-assisted film evaporation correlation
for saline water at sub-atmospheric pressures in
horizontal-tube evaporator, Applied Thermal
Engineering 50 (2013) 670-676.
Figure 7: MED+AD salt concentration profile
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2012
158
Kinetics of Propane Adsorption on Maxsorb III Activated Carbon
Azhar Bin Ismaila, Loh Wai Soong
a, Ng Kim Choon
a*
aDepartment of Mechanical Engineering,
National University of Singapore,
9 Engineering Drive 1, Singapore 117576
*E-mail : [email protected]
ABSTRACT
Experimental kinetics results of propane in Maxsorb
III activated carbon is obtained at temperatures of
10°C and 30°C, and pressures up to 800kPa using a
magnetic suspension balance. A multi-gradient linear
driving force (LDF) approximation is used for
adsorbate uptake as a function of time. The LDF
mass-transfer-rate coefficients were thus determined.
Using this approach, the experimentally derived LDF
coefficients based on independently measured kinetic
parameters for propane in the activated-carbon bed
agree very well with experimental results. The
computational efficiency is gained by adopting this
extended LDF model.
Keywords : Adsorption, Adsorption Chiller,
Adsorption Kinetics
1. INTRODUCTION
Interest in adsorption refrigeration (AD) has grown
due to its advantages related to its direct utilization of
thermal energy sources such as low grade waste heat
from various industrial sources, solar hot water as
well as geothermal sources. The study of adsorption
in the National University of Singapore is a long
standing and continuous project aimed to achieve
higher refrigeration capacity, better Coefficient of
Performance (COP) and an exploration of diverse
applications of the thermal heat pump system [1-2].
In this work, the kinetics of adsorption of propane at
various temperatures and pressures are presented as
an ongoing study of utilizing alternative refrigerants
as adsorbate in an AD system. The Linear Driving
Force (LDF) approach has been adopted to represent
the uptake curves as a function of time. However, the
constant k in this model has been shown to be
subjected to the effects of both temperature and
pressure differences. Due to the sudden compression
effects during charging as well as the isosteric heat
released, the kinetics experiment is non-isothermal as
stipulated by the LDF model. As such, Loh et al
(2012) [3] introduced a model to take into account
the non-isothermal effects on the adsorption kinetics
of assorted adsorbates on Maxsorb III activated
carbon during a Constant volume variable pressure
(CVVP) experiment. He et al [4] on the other hand
identified the effects of pressure differences. In this
paper, a simplified LDF model is implemented to
fully describe the non-isothermal adsorption kinetics
of propane on Maxsorb III.
2. EXPERIMENTAL
2.3 Materials
Figure 1: Schematics Diagram for the magnetic
Suspension Balance unit (Rubotherm)
A magnetic suspension balance (Rubotherm) is used
to measure the instantaneous uptake of the adsorbent
as shown in Figure 1. This balance measures the
weight of the sample with a reproducibility of ±0.03
mg. The advantages of this balance are the high
accuracy and long-time stability due to the balance
being outside the measuring cell and having no
contact to any solvent vapor. Furthermore,
continuous data readout via PC is possible. Maxsorb
III (by Kansai Coke Company, Japan) is utilized with
pure propane, purity 99.5% is utilized. The values of
derived quantities are extracted from NIST
(REFPROP).
2.4 Adsorption Experiments
2.4.1 Buoyancy Correction
Buoyancy forces are taken into account to correct the
influence of gas density on the measured apparent
IMAT-UI 027
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159
weight of the sample. The displacements of gas by
the sample holder, solid adsorbent, and adsorbed
phase are taken into consideration. The correction
due to the sample holder is obtained with blank
experiments performed at different pressures with the
empty holder. The buoyancy due to the solid matrix
of the adsorbent, which results in an apparent weight
loss, is estimated as the product of the skeletal
volume of the adsorbent and the gas density. Finally
the buoyancy effect exerted on the adsorbed phase is
corrected to obtain the absolute adsorption isotherm
(P,T) .
The weight, m, displayed by the balance results from
the net force exerted on the sample:
m = mh (1-ρg⁄ρh ) + ms [1- ρg ⁄ρs +q (1-ρg ⁄ρa)] (1)
Here, mh and ρh are the mass and density of the
sample holder respectively, ms and ρs are the mass
and density of the adsorbent sample, respectively, ρg
and ρa is the density of the bulk gas and adsorbed
phase respectively at the equilibrium pressure and
temperature.
The blank experiments with an empty holder give the
mass and density of the holder from the intercept and
slope of the linear decrease of apparent weight with
gas density:
m = mh – (mh/ρh) × ρg (2)
The density (ρg) of Nitrogen gas is obtained using
values from REFPROP. Adsorption experiments
using a non-adsorbing gas such as helium at a high
temperature of 120°C, provide the mass (ms) and
density (ρs) of the carbon sample:
m - mh (1-ρg ⁄ρh ) = ms - (ms⁄ρs) × ρg (3)
Here, it is assumed that helium acts as an inert gas
that penetrates into all the accessible pore volume of
the carbon without being adsorbed.
Finally, the experiments with the carbon sample
provide the uptake q:
msq(1-ρg⁄ρa) = m - mh(1-ρg⁄ρh) - ms(1-ρg⁄ρs ) (4)
The value of ρa is estimated using the approximation
by Ogawa (1975) [5], given by
ρa = ρa* ⁄ exp [αe (T-Tb ) ] (5)
where ρa indicates the density of the liquid at the
normal boiling point Tb, and αe indicates the thermal
expansion of the superheated liquid. The pressure
dependencies of ρa, ρa*
and αe is negligibly small in
the pressure range of the present work and are thus
neglected. The normal boiling point and the density
of the liquid at the normal boiling point are taken
from Miyamato and Watanabe (2000) [6]. Further,
the value of the thermal expansion was assumed to be
independent of the species of the adsorbate, and the
mean value of the thermal expansion of liquefied
gases (αe = 2.5×10-3
K-1
) was used in the numerical
calculation.
2.4.2 Measurement Procedure
The pressure controller is set to the desired pressure,
and water from the water bath enters the jacket to
maintain the desired temperature. When equilibrium
is reached, the valve is opened to allow the propane
gas to enter the chamber, and the pressure,
temperature and weight changes are logged in the
data logger.
3. SUPPORTING THEORY
3.1 Linear Driving Force (LDF) Model
The LDF model or Lumped Parameter Model (LPM)
[7] has been used to describe adsorption kinetics at
isothermal conditions. It is advantageous in that it is
easy to incorporate in simulation programs given the
computational ease. The LDF model describes the
kinetics of adsorption well due to averaging of the
kinetic properties at the particle, the column, and the
overall cyclic steady state levels [8]. The
characteristics of the models describing the local
rates of adsorption at the particle level are also lost
during these integration processes. In this model, the
mass transfer equation is described by a driving
force, defined as the difference between the
equilibrium uptake (q*) and the instantaneous uptake
(q) [9].
dq/dt = k [q* - q(t)] (6)
where k is the effective particle-phase transfer
coefficient as a function of adsorbate concentration.
The heat transfer equation is then given by
cp· dT/dt = QST· dq/dt – ha(T-To) (7)
The initial and final conditions are
at t = 0, q=qo, T=To (8)
at t = ∞, q=q∞, T=To (9)
qo=qo*(po,To), q∞=q∞
*(p∞,T∞), q=q
*(p∞,T) (10)
q is the adsorbate loading per kg of adsorbent at time
t while qo and q∞ are the equilibrium loading at the
initial and final conditions respectively. T is the
adsorbate temperature at time t while To is the initial
and final temperature of the adsorbate. cp is the heat
capacity of the adsorbent, while h is the external heat
transfer coefficient while a is the external heat
transfer area. QST on the other hand is the isosteric
heat of adsorption. For a differential test where the
changes in the adsorbate loading and temperature are
small, q* can be written as
(q∞- q*)= (∂q
*/∂T)q=q∞, T=To (To-T) (11)
Equation (6), (7) and (11) may be solved
simultaneously to give [10]
The 5th IMAT, November 12 – 13th
2012
160
(q-qo)/( q∞- qo) = (βα2/(1- βα
2))exp(rt)
- exp[-k(1-αβ)t]/ βα2 (12)
where
α = k/(k+r) (13)
β = (QST/cp) (dq*/dT) q=q∞, T=To (14)
r = (-k/2){(1-β+λ)-[(1-β+λ)2-4λ]
1/2} (15)
λ = (ha)/(cpk) (16)
4. RESULTS AND DISCUSSION
4.1 Blank measurements and Buoyancy
Corrections
The blank measurements with Nitrogen gas at
different densities gave a good straight line fit as
shown in Figure 2 to obtain an empty cylinder mass
of 4.4989g and density of 8282kg/m3. The buoyancy
measurements on the other hand were carried out
with inert Helium gas at a high temperature of 120°C
with different pressures to achieve the desired
densities. These buoyancy measurements as shown in
Figure 3 gave a density of 2.2g/cm3 for the Maxsorb
III activated carbon and a mass of 0.1547g of solid
adsorbent in the testing chamber. These results are
tabulated and summarized in the following Table 1.
Table 1: Mass and Densities of Empty Cell and Adsorbent
Mass (g) Density (g/cm3)
Empty Cell 4.4989 8.3
Maxsorb III Adsorbent 0.1547 2.2
Figure 2: Blank Measurements of the empty cylinder
with Nitrogen Gas.
Figure 3: Buoyancy Measurements of the empty
cylinder with Helium gas at high
temperatures of 120°C.
4.2 Adsorption Kinetics of Propane on
Maxsorb III at 10°C and 30°C
The Adsorption kinetics of Propane on Maxsorb III at
temperatures of approximately 10°C and 30°C are
measured and evaluated for the first 300s as
presented in the following Figures 4 and 5
respectively. The LDF model for non-isothermal
adsorption presented earlier are utilized to curve-fit
the data.
Figures 3 and 4 show the experimental temperature
profiles of the adsorbate during the experiment. The
temperature of the adsorbent, which is Maxsorb III
follows the profile when the adsorbent-adsorbate
system is at thermal equilibrium at the beginning of
the experiment and as t approaches infinity,
Figure 3: Temperature Curves of Maxsorb III-
Propane: experimental data at ▲-
To=10.96°C, P∞=497 kPa,
experimental data at ●- To=9.15°C,
P∞=192kPa
time (s)
tem
pe
ratu
re (
°C)
density (kg/m3)
mas
s (k
g)
density (kg/m3)
mas
s (k
g)
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2012
161
Figure 4: Temperature Curves of Maxsorb III-
Propane: experimental data at ▲-
To=28.60°C, P∞=700 kPa,
experimental data at ●- To=28.89°C,
P∞=497 kPa, ■- To=29.27°C,
P∞=195kPa
From equation (12), it may be shown that as time
approaches infinity,
(q-qo)/( q∞- qo) = 1 + (βα2/(1- βα
2))exp(rt) (12)
Hence, using the data from the buoyancy
measurements and the mass recorded by the
MessproTM
software, the uptake of the Masxorb III –
Propane at time t is calculated and curve 1-(q-qo)/(q∞-
qo) can be plotted, and the gradient obtained to give
the value of r and the cut at the y axis gives the value
of -ln(-βα2/(1-βα
2)). The 1-(q-qo)/(q∞-qo) profiles for
To=28.89°C, P∞=497 kPa and To=28.60°C, P∞=700
kPa are presented in Figure 5 as examples.
Figure 5: 1-F(t) profiles for ∆-To=28.89°C,
P∞=497 kPa, experimental data at
○-To=28.60°C, P∞=700 kPa, ---
straight lines to obtain the gradient
which gives the value of r and
intersection at y axis giving the
value of -ln(-βα2/(1-βα
2)).
The parameters q∞, k and the respective errors were
evaluated from 10s onwards and presented in the
following Table 2. The reason for this is in the first
10s, the pressure has not stabilized, and the zero error
correction cannot be determined. β was evaluated to
be 2.0 during these experiments.
Table 2: Parameters of Kinetics Model
T
(°C)
P
(kPa)
q∞
(g/cm3)
k
(1/s)
RMS
error
(%)
9.15 192 0.876 0.00304 1.25
10.96 497 0.729 0.00784 0.65
29.27 195 0.592 0.00495 1.57
28.89 497 0.792 0.00542 1.19
28.60 700 0.835 0.00841 0.75
The fitted models were plotted alongside the
experimental data as shown in Figure 6 and 7.
Figure 6: Uptake kinetics of Maxsorb III-
Propane: experimental data at ○-
To=9.15°C, P∞=192kPa,
experimental data at ∆-To=10.77°C,
P∞=300 kPa, □-To=10.96°C,
P∞=497 kPa --- fitted curves from
the non-isothermal adsorption
kinetics model.
Ln[1
-F(t
)]
time (s)
time (s)
tem
pe
ratu
re (
°C)
up
take
(kg
/kg)
time (s)
The 5th IMAT, November 12 – 13th
2012
162
Figure 7: Uptake kinetics of Maxsorb III-
Propane: experimental data at □-
To=29.27°C, P∞=195kPa,
experimental data at ∆-To=28.89°C,
P∞=497 kPa, experimental data at ○-
To=28.60°C, P∞=700 kPa, --- fitted
curves from the non-isothermal
adsorption kinetics model.
The value of k as expected increases with pressure at
a given temperature, since the value of q* will be
higher at every point in the kinetics curve resulting in
a higher driving force resulting in larger k value.
7. CONCLUSION
The regressed value fit the experimental data very
well for use of simulation in determining the
performance of an AD chiller that runs with propane.
ACKNOWLEDGMENT
The researcher, Azhar Bin Ismail is supported by the
National Research Foundation Singapore under its
National Research Foundation (NRF) Environmental
and Water Technologies (EWT) PhD Scholarship
Programme and administered by the Environment
and Water Industry Programme Office (EWI).
REFERENCES
[1] K.C. Ng, X. Wang, Y.S. Lima, B.B. Saha, A.
Chakarborty, S. Koyama, A. Akisawa and T.
Kashiwagi ―Experimental study on performance
improvement of a four-bed adsorption chiller by
using heat and mass recovery‖ International
Journal of Heat and Mass Transfer, vol. 49, no.
19-20, pp. 3343–3348, September 2006.
[2] B.B. Saha, A. Chakarborty, S. Koyama, K.
Srinivasan, K.C. Ng, T. Kashiwagi, P. Dutta,
―Thermodynamic formalism of minimum heat
source temperature for driving advanced
adsorption cooling device‖ Applied Physics
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[3] W.S. Loh, A. Chakraborty, B. B. Saha , and K.
C. Ng ―Experimental and Theoretical Insight of
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[4] I.I. El-Sharkawy ., J.M. He, K.C. Ng, C. Yap and
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[5] S. Ozawa, S. Kusumi and Y. Ogino ―Physical
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adsorption equation. Journal of Colloid and
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[6] H. Miyamoto and K. Watanabe ―A
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[7] D.D. Do, Adsorption analysis: equilibria and
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[8] S. Sircar and J.R. Hufton ―Why Does the Linear
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[9] J. I. Coates and E. Glueckauf ―Theory of
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vol. 79, pp. 785-796, 198
up
take
(kg
/kg)
time (s)
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ABSTRACT
This paper presents an investigation on the effects
of using solar chimney, gable vent, and the
combination of the two natural ventilations on the
average air temperature and air-flow condition
inside a double-storey house in Malaysia, using
computational fluid dynamics (CFD) method. The
representative model of the house comprises of a
main hall, a kitchen and an upper hall. Both
temperature and air velocity boundary conditions
were prescribed on the model. Results of the
simulation indicates that the average temperature of
the air in the house at 1 pm closely matched the
measured values. It was found that the average
temperature of the air in the house is not so
significantly affected by the types of natural
ventilation used. Opening the kitchen door
causesthe air to flow from the main and upper
halls towards the kitchen and causing a bottle neck
at the pathways. A more uniform air flow is
obtained when solar chimneys are used. When
gable vents are used, high intensity air flow occurs
in the main hall and it spreads uniformly towards
the kitchen and upper hall. The air-flow intensity
becomes even higher in the main and upper halls
when a combination of solar chimney and gable
vents are incorporated into the CFD model.
Keywords : Natural ventilations, Average air
temperature, Air-flow conditions,
CFD simulation, Solar chimney,
Gable vents.
1. INTRODUCTION
Natural ventilation has attracted a strong growing
interest in building sectors because of its potential
advantages over mechanical ventilation systems, in
terms of energy requirement, economic and
environmental benefits. Mechanical ventilation
systems have undesirable energy implication since
they require more electricity to run [1]. Earlier
work on natural ventilation mainly concerned with
aerodynamic loading [2] and they were carried out
in wind tunnel. But with the advancement of
Effects of Natural Ventilations on Indoor Air
of a Double-Storey Residential House in Malaysia
Haslinda Mohamed Kamara, Nazri Kamsah
b & Kam Jia Liq
aFaculty of Mechanical Engineering
Universiti Teknologi Malaysia, Skudai, Johor
Tel : (+607) 5534748. Fax : (+607) 5566159
E-mail : [email protected]
bFaculty of Mechanical Engineering
Universiti Teknologi Malaysia, Skudai, Johor
Tel : (+607) 5534749. Fax : (+607) 5566159
E-mail : [email protected]
IMAT-UI 028
IMAT-UI 028
The 5th IMAT, November 12 – 13th
2012
164
computing technology, more complicated studies
have been conducted using computational fluid
dynamics (CFD) techniques. Nikas et al. [3]
showed that it is possible to get information about
induced velocity and pressure fields for natural
cross ventilation using CFD modelling which
otherwise are quite difficult to extrapolate from
experimental methods.
Various strategies have been proposed in the
literature to enhance buoyancy effect so that
adequate air flow rate and a desired level of
thermal comfort can be achieved inside a building.
One good example is a solar chimney, which is
designed to maximize ventilation effect by
maximizing solar gain [4]. This creates a sufficient
temperature difference between the inside and
outside of the building to drives an adequate air
flow rate. Solar chimney is a thermo-syphoning air
channel in which the principal driving mechanism
of air flow is through thermal buoyancy [5]. One
can find different variations in solar chimney
design, which is affected by a number of factors
such as the location, climate, orientation, size of the
space to be ventilated and the internal heat gains
[6].
Computational methods based on CFD technique
have been used by many to predict flow pattern
inside the chimney as well as in the space (room)
adjoining the solar chimney. The existing CFD
models are able to predict velocity and temperature
profiles along with other flow characteristics
accurately. However, they usually do not consider
the thermal energy storage in the walls of the
building [7]. Nevertheless, the use of CFD
modelling in solar chimney study has been
increasing. These studies have greatly contributed
to the present understanding of the solar chimney.
In this study, we used the CFD method to
investigate the effect of using natural ventilations
in a double-storey residential terrace house in
Malaysia. The natural ventilations considered are
solar chimney, gable vents and the combination of
the two. The focus of this study is not on the types
of ventilation. The main goal of this study is to find
out the effect of using these ventilations on the
thermal and flow conditions of the air inside the
house. For that purpose, the solar chimneys are
represented only as simple square openings located
on the roof of several sections of the house. The
gable vents on the other hand are represented by
long rectangular openings on the upper part of
several walls of the house. During the simulation,
both air velocity and temperature are prescribed on
these openings to model the outward air flow from
the house.
2. METHODOLOGY
2.1 Computational Domain
Figure 1 shows a representative model of the
house and a computational domain for the CFD
simulations. It consists of three sections namely the
main hall, the upper hall and the kitchen. There are
The 5th IMAT, November 12 – 13th
2012
165
only four walls that are considered to be exposed to
solar radiation. These are the eastern and southern
walls of the main hall, northern wall of the kitchen
wall, and the eastern wall of the upper halls. Other
walls are considered to be insulated and are at the
same temperature as the air in the house, which is
at 29C (302K).
2.2 Actual Average Air Temperature &
Humidity
The actual average dry-bulb temperature, wet-bulb
temperature and relative humidity of the interior air
were determined in the three sections of the house:
the main hall, the upper hall and the kitchen, using
a sling psychrometer, for every hour beginning
from 9 am until 4 pm. In each section, all the data
were measured at several locations and then the
average values were computed, for every hour. The
complete hourly data are shown in Table 1. It is
observed that the average dry-bulb temperature of
the air is about 30C and relative humidity is
around 73%.
Figure 1: A representative model of the house
considered for the CFD computational domain
(rear view).
Table 1: Hourly data for the air inside the house.
2.3 Validation of CFD Simulation Procedure
To validate the numerical simulation procedure, a
CFD simulation was performed on the model of the
house to represent a condition when there are no
ventilations. However, a door on the rear wall of
the kitchen was left fully opened to generate some
air flow within the house. We call this as a ―base
case‖ condition. The goal of this simulation is to
estimate the average temperature of the air in the
various sections of the house and compare them
with the actual temperatures measured at 1 pm,
when the kitchen door was opened.
Both temperature and air velocity boundary
conditions were used. A uniform temperature of
47C (320K) was prescribed on the wall of the
main hall facing south and the wall of the kitchen
facing north. A uniform temperature of 29C
(302K) was prescribed on the wall of the main hall
KITCHEN MAIN HALL
SOLAR CHIMNEY
UPPER HALL
GABLE VENT
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2012
166
facing east and the wall of the upper hall facing
east. A constant inlet air velocity of 0.1 m/s, at a
temperature of 29C (302K), was prescribed on all
the door seams. These door seams represent the
gaps between the doors and the slabs and the
clearance between the doors and the walls. The air
velocity boundary conditions allow a turbulent
analysis to be performed on the computational
domain. Turbulent flow analysis using a k- model
with 10% turbulent intensity was performed on the
CFD model until an acceptable convergence was
attained.
2.4 Modeling of the Natural Ventilations
Three natural ventilation systems were considered
in this study, namely a solar chimney, gable vents
and the combination of the two. The solar chimney
is a natural-draft device that is used in many
passive cooling applications for residential houses.
Density of air decreases with increasing
temperature. It means that air with higher
temperature than ambient air is driven upwards by
the buoyancy force. A solar chimney exploits this
physical phenomenon and uses solar energy to heat
air up. Gable vents are usually placed at the top of
the gable on the end of the house. This is to create a
draft through the space by having both intake and
exhaust vents. While gable vents do increase
ventilation they do not offer uniform air flow
within the space and their ability to move large
amounts of air are limited.
The solar chimneys were incorporated into the
CFD model by adding square-shaped openings on
the roof (at the middle) of the main hall, kitchen
and the upper hall. A constant air outlet velocity of
0.3 m/s, at 29C (302K) was prescribed on all the
solar chimneys while a constant inlet air velocity of
similar magnitude was prescribed on all the door
seams as the boundary conditions. The same
temperature boundary conditions as in the base
case were employed in this CFD simulation.
The gable vents were incorporated into the CFD
model by introducing thin rectangular-shaped
openings on the walls of the house. The width of
these openings was made nearly the same as the
width of the walls. An inlet gable vent was placed
on the eastern wall of the main hall, while outlet
gable vents were placed on the southern wall of the
main hall, the northern wall of the kitchen and the
eastern wall of the upper hall. A constant air outlet
velocity of 0.3 m/s, at 29C (302K) was prescribed
on all the solar chimneys while a constant inlet air
velocity of similar magnitude was prescribed on all
the door seams as the boundary conditions. The
same temperature boundary conditions as in the
base case were used in this CFD simulation.
3. RESULTS AND DISCUSSION
3.1 Base Case Conditions
Results of the CFD simulation for the ―base case‖
condition give an average air temperature of about
The 5th IMAT, November 12 – 13th
2012
167
30.4C in both the main hall and the kitchen, and
about 29.7C in the upper hall. These values are
superimposed on the plots of measured temperature
vs. time (within a circle) for the three sections on
the house, shown in Figure 2. It can be seen that the
average temperatures obtained from the CFD
simulation fall within the acceptable range of the
measured temperature range. Thus it is safe to say
that the CFD model, the boundary conditions used
and the turbulent analysis model employed in the
simulation are valid and can be further used in the
proceeding simulations. The air flow distribution in
the house obtained from the CFD simulation for the
―base case‖ conditions is shown in Figure 3. It can
be seen that, with the door on the eastern door of
the kitchen left opened, the air tends to flow from
the main and upper halls towards the kitchen,
producing a bottle neck at the pathway connecting
the main hall and the kitchen. The air flow is seen
fairly uniform in both halls.
Figure 2: Comparison between the average air
temperature obtained from the CFD simulation
and the measured values.
Figure 3: Air-flow distribution (m/s) inside the
house when the door on eastern wall
of the kitchen is left opened.
Figure 4: Air-flow distribution (m/s) inside the
house when solar chimney
ventilation is used.
3.2 The Effect of Solar Chimney Ventilation
Results of the CFD simulation when solar
chimneys are incorporated into the model give an
average air temperature of 302.8 K in the main hall,
302.9 K in the kitchen, and 302.3 K in the upper
Kitchen
Main Hall
Upper Hall
CFD
N
N
S
S
E
E
W
W
The 5th IMAT, November 12 – 13th
2012
168
hall. These are slightly lower than the average air
temperature for the base case condition [303.4 K
(hall); 303.3 K (kitchen); 302.7 K (upper hall)]. On
average, the CFD simulation results indicate that
the average air temperature in the house is reduced
by about 0.6C when three solar chimneys were
incorporated into the model. This is considered as
an insignificant improvement on the average
temperature of the air inside the house. The air
flow condition in the house when solar chimneys
are used is shown in Figure 4. It can be seen that
the air flow is fairly uniform in all three sections of
the house. Slightly higher air velocity occurs at all
the door seams (inward flow) and the solar
chimneys (outward flow). A swirling air flow
condition can be seen near the northern wall of the
kitchen and the southern wall of the main hall.
3.3 The Effect of Gable Vents
Results of the CFD simulation when gable vents
are incorporated into the model give the average air
temperature of 302.4 K in the main hall and in the
kitchen, and 302.2 K in the upper hall. These are
slightly lower than the average air temperature for
the base case condition [303.4 K (hall); 303.3 K
(kitchen); 302.7 K (upper hall)]. On average, the
CFD simulation results indicates that the average
air temperature in the house is reduced by about
0.8C when the gable vents were incorporated into
the CFD model. This can also be considered as an
insignificant improvement on the average
temperature of the air inside the house.
The air flow condition in the house when gable
vents are used is shown in Figure 5. It can be seen
that the air flow is fairly uniform in the main hall
and the kitchen but it is less intense in the upper
hall section. Higher air velocity condition can be
seen at the vicinity of all the gable vents, especially
at the inward flow gable vent on the eastern wall of
the main hall. No swirling air flow condition can be
seen in the figure.
Figure 5: Air-flow distribution (m/s) inside the
house when gable vents are used.
3.4 The Effect of Combined Solar Chimney
Ventilation & Gable Vents
Results of the CFD simulation when a combination
of solar chimneys and gable vents are incorporated
into the model give an average air temperature of
302.2 K in the main hall, 302.3 K in the kitchen
and 302.1 K in the upper hall. When compared to
the base case conditions, it is found that the
average air temperature is dropped by about 1.1C,
1.0C and 0.6C in the main hall, kitchen and the
N
S
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The 5th IMAT, November 12 – 13th
2012
169
upper hall, respectively. These are considered a
mild reduction in the average air temperature inside
the house.
Figure 6: Air-flow distribution (m/s) inside the
house when a combination of solar
chimneys and gable vents are used.
Figure 6 shows the air flow distribution inside the
house when a combination of solar chimneys and
gable vents are used. It is seen that high intensity
air flow occurs in the main hall and the air appears
to move towards the kitchen. The air flow in the
upper hall and the kitchen appears to be less
intense. Higher air velocity is seen at both the inlet
and outlet gable vents. No swirling air flow can be
seen from the figure. Also, the air tends to flow
toward the solar chimneys located on the ceiling of
each section of the house.
4. CONCLUSION
A CFD simulation method has been used to
investigate the effects of several natural
ventilations, namely solar chimney, gable vent and
the combination of both, on the conditions of the
air inside a double-storey residential house in
Malaysia. It was found that the average
temperatures of the air at various sections of the
house, obtained from the CFD simulation for the
base case condition, agree quite well with the
measured values at 1 pm. The average air
temperature drops by about 0.6C when solar
chimneys are used and about 0.8C when gable
vents are incorporated into the CFD analysis. The
temperature drops by about 1C when the
combination of both ventilations are included in the
analysis. When the kitchen door is left opened, the
air tend to flow from the main hall and upper hall
towards the kitchen. Using solar chimney
ventilation results in a more uniform air-flow inside
the house. High intensity air flow occurs in the
main hall and it spreads uniformly towards the
kitchen and upper hall when inlet and outlet gable
vents are used. The air-flow intensity becomes
even higher in the main and upper halls when a
combination of solar chimney and gable vents are
incorporated into the computational model.
REFERENCES
[1] Rakesh Khanal & Chengwang Lei, Solar
chimney - A passive strategy for natural
ventilation, Energy and Buildings 43 (2011)
1811–1819.
[2] P.F. Linden, The fluid mechanics of natural
ventilation, Annual Review on Fluid
Mechanics 31 (1999) 201–238.
N
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E
W
The 5th IMAT, November 12 – 13th
2012
170
[3] K.-S. Nikas, N. Nikolopoulos, & A.
Nikolopoulos, Numerical study of a naturally
cross-ventilated building, Energy and
Buildings 42 (2010) 422–434.
[4] N.K. Bansal, R. Mathur, M.S. Bhandari, Solar
chimney for enhanced stack ventilation,
Building and Environment 28 (3) (1993) 373–
377.
[5] G. Gan, Simulation of buoyancy-induced flow
in open cavities for natural ventilation, Energy
and Buildings 38 (5) (2006) 410–420.
[6] D.J. Harris, N. Helwig, Solar chimney and
building ventilation, Applied Energy 84 (2)
(2007) 135–146.
[7] A. Dimoudi, Solar chimneys in buildings – the
state of the art, Advances in Building Energy
Research 3 (2009) 21–44.
The 5th IMAT, November 12 – 13th
2012
171
Deposit Forming Tendency of Biodiesel and Diesel Fuel due to
High Pressure Exposure
Muhamad Adlan Abdullah, Arshad Salema and Farid Nasir Ani
Faculty of Mechanical Engineering, Universiti Teknologi Malaysia, Skudai, Johor D.T., Malaysia
ABSTRACT
Fuel deposit issues in common rail diesel fuel
system require attention as it will affect the
operation of the closely designed equipment
resulting in increased emissions and poor
performance. However, the fuel propensity in
forming deposit in high pressure common rail
system is inadequately understood. This paper
recounts a test program designed to investigate
the effects of the exposure to high pressure and
temperatures on the diesel and biodiesel fuel
tendency to form deposit. A test rig was built
allowing the fuel to be pre-stressed under
conditions of common rail system and the deposit
forming tendency was determined by using
modified Jet Fuel Thermal Oxidation Tester
(JFTOT) procedure as well as deposition on hot
surface. The result showed that the exposure to
high pressure and temperature such as in common
rail system increases the tendency of deposit
formation.
Keywords : diesel, biodiesel, fuel deposit,
common rail, high pressure diesel
1. INTRODUCTION
Typically, deposit in diesel fuel injection system
lies on the injector nozzles. This deposit blocks
the flow of the fuel, which affects the delivery.
Recently, new type of injector deposits in
common rail diesel fuel systems were reported [1]
to be on the internal of injectors. The new type of
deposit was hypothesized to be a result of
exposure to the high pressure in the common rail
system [2]. Fuels with known storage stability
may be instable after being exposed to operation
in the common rail system. Presence of some fuel
additive components was also demonstrated to
contribute to this deposit [1, 3].
Efforts to quantify the diesel fuel system deposit
covers both on engine tests [4, 5] and laboratory
rigs [6, 7]. Engine tests are often very lengthy and
costly. Hence, some research efforts focus on
using test rigs for the purpose of screening tests
for the deposit forming tendency.
This paper describes a test program to measure
the deposit forming tendency of diesel and
biodiesel after exposure to high pressure of a
commonrail engine using in-house bench test. It
attempts to investigate the tendency of diesel and
biodiesel to form both types of injector deposits –
IMAT-UI 029
The 5th IMAT, November 12 – 13th
2012
172
the injector tip (nozzle deposit) as well as the
internal injector deposit – after the exposure to
conditions experienced in common rail system.
2. EXPERIMENTAL SETUP AND
PROCEDURES
The experiments were conducted as per the
following sequence:
1. The diesel fuel is pre-stressed under high
pressure common rail system
2. The pre-stressed fuel is then subjected to the
deposit forming test using modified JFTOT as
well as the hot surface deposition test
The common rail test rig was built using
BOSCH CP3 common rail pump system as shown
in Figure 1. The fuel pressure was measured by
the pressure sensor located at the fuel rail.
Temperatures of the fuel in the tank, Ti, and at the
return line, To, were measured by means of
thermocouples. The fuel is circulated through the
system simulating the conditions in an engine
operation, through high pressure and returned to
tank at atmospheric pressure. In order to maintain
the fuel temperature, a water cooled heat
exchanger was installed at the fuel return line
from the rail. The test fuel is pre-stressed under
these conditions for a specified duration.
The pre-stressed fuel is then subjected to the Jet
Fuel Thermal Oxidation JFTOT III tester for the
deposit forming study as in Figure 2. Detail of the
procedure was described elsewhere [8].
In principle, the JFTOT exposes the fuel to high
temperatures that represents the conditions found
inside injector nozzles where deposits typically
occur. Thus, this test program attempts to emulate
the conditions where fuel is first exposed to high
pressure and temperature in the common rail
system, and subsequently delivered to the
injectors where it is heated and deposit is formed.
In this case, no fuel evaporation occurs.
The hot surface deposition test was adopted from
the works of Yusmadi [9] as shown in Figure 3.
The heater was used to supply heat to the
aluminium cylinder block until a desired
temperature was achieved i.e. 200°C, 225°C and
250°C. Fuel was dropped onto the surface of the
alumimium block during the deposition test. K-
Type thermocouple was used and was located at
the heater in the cylinder block. In the deposition
experiment, the fuel was dropped from a burette.
Fuel was filled in burette and then released as
droplets to the surface of the aluminium block at
specific rate and the resultant deposit was
observed.
The hot surface deposition test emulates the
deposit formation in the combustion chamber of
diesel engine such as on the injector tips. In this
case, fuel evaporation occurs and the deposit
formation is different from the JFTOT test.
Both the deposit formed under JFTOT and hot
surface deposition tests were quantified by using
JPI Varnish rating as in Figure 4.
The 5th IMAT, November 12 – 13th
2012
173
The test fuel used were typical diesel fuel
obtainable from the market and a biodiesel
produced from waste cooking oil, supplied by
Xtrac Tech Sdn. Bhd.
Figure 1: The common rail test rig
(a)
(b)
Figure 2: (a) The Jet Fuel Thermal Oxidation
Tester and (b) the heater tube
Figure 3: The hot surface deposition test
Figure 4: The JFTOT rating and JPI varnish rating
3. RESULTS AND DISCUSSION
3.1 Effects of exposure to high pressure on
internal deposit
Initial work [8] had demonstrated that diesel fuel
exposure to high pressure and temperature of
common rail system can affect the deposit
formation. Figure 5 shows the JFTOT heater tube
rating for fuels after 4 hours of pre-stressing in the
common rail rig at different pressures and
temperatures. It is shown that for similar inlet
JFTOT rating
JPI Varnish rating
The 5th IMAT, November 12 – 13th
2012
174
temperatures Ti (fuels marked as B, D and E),
higher pre-stressing pressure tends to produce
more deposit (i.e. lower deposit rating). It was
also shown that the temperature at which the fuel
is exposed has significant effects on the deposit
formation.
It was also demonstrated that exposure to high
pressure and temperatures for as short a duration
as 30 minutes may affect the deposit forming
tendency. This is shown in Figure 6 which shows
the deposit forms at lower temperatures as the
pre-stressing time is increased (lower JPI rating
indicates higher deposit).
However, due to some constraints, only diesel
fuel was tested in this study. Biodiesel fuel was
not evaluated for its tendency to form deposit in
the condition of this test
0
10
20
30
40
50
60
70
80
0
2
4
6
8
10
12
unaged 910bar 95C
910bar 82C
700bar 83C
500bar 77C
Inle
t Te
mp
era
ture
He
ate
r tu
be
ra
tin
g
JPI rating inlet temp, °C
A
B
C
D
E
JFTOT temperature :200C
Figure 5: The effects of pressure and
temperatures
0
2
4
6
8
10
12
210 220 230 240 250 260
He
ate
r tu
be
rat
ing
JFTOT temperatures, degC
fresh fuel
30 minutes
4 hours
pressure=900bar
To= 80°C
Figure 6: The effects of pre-stressing
duration
3.2 Effects of exposure to high pressure on
deposit on hot surface
Diesel and biodiesel fuel were tested on the hot
surface tests after exposure to 10 minutes of high
pressure common rail system. The hot surface
temperature was set at 200°C, 225°C and 250°C.
As expected, at higher surface temperature, more
deposit is formed. The amount of fuel dropped (as
given by the longer test duration) also increases
the deposit formation. Figure 7 shows the
photograph of typical deposit of biodiesel formed
at different surface temperature and duration.
Figure 8 to 10 shows the deposit for diesel and
biodiesel for hot surface temperature of 200°C,
225°C and 250°C respectively. The pre-stressing
pressure in the common rail was done at pressures
of 200 bar, 400 bar, 600 bar and 800 bar.
It was shown that for diesel, the pre-stressed fuel
has significantly more deposit than the fresh fuel.
However, increasing the pre-stressing pressure
from 200-800 bar in the common rail system did
The 5th IMAT, November 12 – 13th
2012
175
not increase the deposit formation further. This is
true for all hot surface temperatures tested.
In contrast, for biodiesel, there is some difference
in the quantity and the shape of the deposit
formed as the pre-stressing pressure is increased.
It seems that as the fuel is pre-stressed at higher
pressure, the deposit formation concentrated more
on the fringes. This is probably due to changes in
the fuel‘s surface tension when it has undergone
high pressure and high shear operation in the
common rail.
Note also that at surface temperature of 200C, the
deposit formation was shown to increase with
higher pre-stressing pressure. This is, however,
not evident at other surface temperatures.
4. CONCLUSIONS
From this study, the following conclusions can be
drawn.
1. The deposit forming tendency of diesel and
biodiesel when exposed to high pressure and
temperature operation of a common rail fuel
system was studied.
2. It was demonstrated that fuel exposed to
high pressure and temperatures seen in
common rail system significantly increased
the deposit forming tendency even after a
short duration.
3. Diesel fuel was shown to increase its deposit
formation in JFTOT tests with higher pre-
stressing pressures. However, it did not
show the same trend in hot surface
deposition test. On the other hand, biodiesel
showed increased tendency to form deposit
in hot surface deposition test with increasing
pre-stressing pressure.
ACKNOWLEDGEMENT
The authors are grateful to the Research
University Grant, Universiti Teknologi Malaysia,
Vot 01H03 for the financial support and Research
Management Centre, UTM for the management
support.
REFERENCES
[1] Ullmann, J. Geduldig M., Stutzenberger H.,
Caprotti R., Balfour G., (2008), Investigation
into the Formation and Prevention of Internal
Diesel Injector Deposits, SAE 2008-01-0926
[2] Steve Cook and Paul Richards (2009),
Possible Influence of High Injection Pressure
on Diesel Fuel Stability: A Review and
Preliminary Study, SAE2009-01-1878
[3] Leedham A, Caprotti R, Graupner O, Klaua
T, (2004), Impact of Fuel Additives on Diesel
Injector Deposits, SAE2004-01-2935
[4] Rod Williams (2002), Development of a
Nozzle Fouling Test for Additive Rating in
Heavy Duty DI Diesel Engines, SAE 2002-
01-2721
The 5th IMAT, November 12 – 13th
2012
176
[5] Graupner O, Klaua T, Caprotti R, Breakspear
A, Schik A, Rouff C, (2005), Injector
Deposit Test For Modern Diesel Engines,
www.infineum.com/Documents/.../TAE/Esslin
gen%202005.pdf , accessed on 19th
October
2010.
[6] Chintoo Sudhiesh Kumar (2009),Modelling
Deposit Formation in Diesel Injector Nozzle,
MSc. Thesis, Massachusets Institute of
Technology, June 2009
[7] Stavinoha LL, Barbie JG, Yost DM, (1986),
Thermal Oxidation Stability of Diesel Fuels,
Interim Report BFLRF No 25,. 1986,
Southwest Research Institute
[8] Yusmadi Mohd Arifin (2009), Diesel and
Biodiesel Fuel Deposit on a Hot Wall
Surface, PhD Thesis, Gunma University,
Japan, August 2009
[9] Muhamad Adlan Abdullah and Farid Nasir
Ani, (2012), The Effects of Diesel Fuel
Exposure to High Pressure Common Rail
System on its Deposit Forming Tendency,
SEATUC Conference, Bangkok, March 2012
.
The 5th IMAT, November 12 – 13th
2012
177
Time
(min)
200 ◦C 225 ◦C 250 ◦C
0
5
10
15
20
25
30
Figure 7: The biodiesel deposit formation on the hot surface at different temperature and duration
(fuel without pre-stressing)
The 5th IMAT, November 12 – 13th
2012
178
Figure 8: The effect of pre-stressing on deposit formation for diesel and biodiesel at surface temperature of
200°C
Figure 9: The effect of pre-stressing on deposit formation for diesel and biodiesel at surface temperature of
225°C
Figure 10: The effect of pre-stressing on deposit formation for diesel and biodiesel at surface temperature of
250°C
The 5th IMAT, November 12 – 13th
2012
179
Numerical Analysis of Elastohydrodynamic Lubrication with
Non-Newtonian Lubricant
Dedi Rosa Putra Cupu1, Adli Bahari
2, Kahar Osman
3, Jamaluddin Md Sheriff
3
1Mechanical Engineering Department of Engineering Faculty,
University of Riau, Pekanbaru, Riau, Indonesia
Email1 : [email protected]
2Automation & Mechatronics Section, Industrial Electronics Department,
German Malaysian Institute, Malaysia
3Faculty of Mechanical Engineering,
Universiti Teknologi Malaysia, 81310 UTM Skudai, Malaysia
ABSTRACT
Elastohydrodynamic lubrication is a form of
hydrodynamic lubrication involving physical
interaction between two contacting surfaces and
liquid where elastic deformation of the contacting
surfaces due to heavily loading applied will affect the
elastohydrodynamic pressure and fluid film thickness
significantly. In this paper, a line contact EHL is
modeled through the cylinder contact to a flat surface
to represent the application of roller bearing. This
solution is limited to two dimensional line contact
problem only, an infinite length of cylinder will be
used as physical modeling. The behavior of non-
Newtonian fluid also was investigated using power
law fluid model. Bearing speed is to be assumed in
steady state and temperature is assumed constant. The
bearing performance parameters such as pressure,
film thickness and friction coefficient of lubricated
contacts are calculated using Newton-Raphson
method.
The results show that the peak pressure increases as
the parameters such as velocity, load, material
parameter and power law index were increases and
the spike was found to shift to the center of roller.
The film was almost flat at contact region and formed
a dimple shape near the outlet flow. The coefficient
of friction is reduced as the power law index and
slide to roll ratio were decreased. The value of
pressure spike and minimum film thickness were
smaller at lower speed and were increased during
raising speed then the peak point was found to be
shifted to center of roller.
Keywords : Elastohydrodynamic lubrication,
Newton-Raphson, Pressure profile,
film thickness.
1. INTRODUCTION
The whole idea of this study is to write a
programming code that can provide a solution for
fluid film lubrication problem related to
elastohydrodynamic lubrication. This can further be
applied to investigate the effect of parameters on
bearing design and performance such as pressure and
film thickness.
Detailed analysis of gaseous or liquid films is usually
termed hydrodynamics lubrication (HD), while
lubrication by solid is termed solid lubrication. A
specialized form of hydrodynamics lubrication
involving physical interaction between the contacting
bodies and the liquid lubricant is termed
elastohydrodynamics (EHD) lubrication (EHL) and is
considerable practical significance. Another form of
lubrication involves the chemical interactions
between contacting bodies and the liquid lubricant is
termed boundary and extreme pressure lubrication. A
form of lubrication that operates involving the
external force is termed hydrostatic lubrication where
liquid or gaseous lubricant is forced into the space
between contacting bodies.
The most commonly encountered forms of contacts,
commonly known as conjunction, are point and line
contacts. When a sphere comes into contact with a
flat surface, it initially forms a point contact with a
circular shape and the size of conjunction grows as a
function of load. When a cylinder comes into contact
with a flat surface, it forms a line contact and it
grows into a rectangular conjunction as the load is
increased. Incidentally point contact between a ball
and raceway develops into an elliptical conjunction.
In elastohydrodynamic lubrication contact, the
deformed surfaces in lubricated contact are almost
similar to Hertzian contacts with an interposed
lubrication film. A minimum film thickness occurs
near to the outlet region. The film thickness is
IMAT-UI 030
The 5th IMAT, November 12 – 13th
2012
180
important because it is the key parameter to ensure
the protection of mating surfaces for bearing
components.
Dien and Elrod [1] derived the generalized steady
state Reynolds Equation for non-Newtonian fluids
using power law fluid with application to journal
bearing. A. Elsharkawy [2] then applied the formula
into magnetic head-rigid disk interface
hydrodynamically lubricated. The result was shown
that the power law exponent has a significant effect
on the hydrodynamic pressure profile. An early study
of soft elastohydrodynamic in a rolling contact for a
power law fluid by Lim et al. [3] was used in the
exploration of a printing application in which near
pure rolling takes place. The importance of both
power law coefficient and exponent was quantified.
This showed significant impacts for both parameters,
with increases in each resulting in increased film
thickness and maximum pressure. Bohan et al. [4],
explore the application of numerical simulation to
coating applications that involve combined sliding
and rolling mechanisms. This was extend previous
work done by Carvalho and Scriven, [5] and Lim et
al. [6], through the incorporation of actual fluid
properties that exhibit shear thinning. The effect of
power law coefficient, power law exponent and
sliding on the nip performance in terms of pressure
distribution, film thickness profile, strain rate and
viscosity variation through the nip section and flow
rate was investigated.
H.M Chu et al. [7] derived a one-dimensional
modified Reynolds equation for power law fluid from
the viscous adsorption theory for thin film
elastohydrodynamics lubrication (TFEHL). The
lubricating film between solid surfaces was modeled
as three fixed layers, which are two absorption layers
on each surface and a middle layer between them.
The comparison between classical non-Newtonian
EHL and non-Newtonian TFEHL was done. The
result was showing that the TFEHL model can
reasonably calculate the pressure distribution, the
film thickness, the velocity distribution and the
average viscosity. Another result showing that the
greater the thickness and viscosity of the adsorption
layer and the flow index, the greater the deviation in
central film thickness versus speed between EHL
model and TFEHL model produced in the very thin
film regime.
A thermal and non-Newtonian fluid model under
thermal elastohydrodynamics conditions was
proposed by A. Campos et al. [8]. The concept of
apparent viscosity was used to introduce the non-
Newtonian behavior of the lubricant and the thermal
behavior of the contact. The Newton-Raphson
technique was used to obtain the lubricant film
geometry and the pressure distribution inside the
elastohydrodynamics contact. The model was applied
to the analysis of experimental traction curves of a
traction fluid measured in a twin disc machine,
obtained for a significant ranges of the operating
condition. The comparison between numerical and
experimental traction curves showed a very good
correlation.
In 1986, Houpert and Hamrock [9] presented a fast
method to solve the problem facing by Hamrock and
Jacobson (1983). The Reynolds equation is solved
using Newton Raphson iterative procedure and can
solve problem related to higher dimensionless load in
shorter computing time. However this method is not
suitable to be used for point contact problem (three
dimensional) due to memory storage problem. Lin
and Lin in 1990, using the method proposed by
Houpert and Hamrock (1986), derived the Reynolds
equation embedded with power law fluid. Assuming
the flow were compressible and viscous, Lin and Lin
plotted the graph showing the effects of power law
index on pressure profile, film thickness and
coefficient of friction. However their work was only
focusing on Barus equation to represent the pressure-
viscosity model.
2. PHYSICAL PROBLEM DEFINE
In this work, a roller bearing is adapted to the
physical problem for the numerical calculation to be
done. A roller bearing is a device used to support a
rotating shaft to the bearing housing and at the same
time used to reduce the friction between contacting
surfaces. Figure 1 shows a typical roller bearing
commonly used in industrial machine. Since the
scope of this study is limited to two dimensional line
contact problem only, an infinite length of cylinder is
model to be contacted with a flat surface in order to
represent the roller bearing application.
Figure 1: A typical roller bearing
An infinite length of steel roller with diameter, R =
11.4 mm is used for analysis. The cylinder is loaded
against the flat surface. The roller and the flat surface
are then been driven independently to create a mixed
The 5th IMAT, November 12 – 13th
2012
181
rolling/ sliding contact. The temperature is assumed
to be constant. Figure 2 shows some physical
parameters that are applied to the system of roller
bearing. A load, w is applied to the cylinder, which
both the roller and the flat surface are driven with the
different speed, ua and ub. Some of the parameters are
listed in the Table 1 that has been used in the
programming code throughout the study.
Figure 2: An infinite length of cylinder contact
with flat surface
Table 1: Some of the physical parameters involve in
the calculation.
Material Steel: AISI 52100
Elastic Modulus Eb = Ed = 210 GPa
Poisson‘s ratio b = d = 0.3
Ball radius R = 11.4 mm
Dimensionless Applied Load W = 2.0452x10-5
Lubricant Absolute Viscosity 0 = 6.6 Pa.s
Dimensionless Speed U = 1x10-11
Temperature T = 35C (constant)
3. NUMERICAL SOLUTION
The fluid film between two solid surfaces shown in
Figure 3 is considered. Reynolds equation is an
equation to obtain the pressure generated in a fluids
film when two such surfaces undergo relative motion.
However, the fluid film must be sufficiently thin so
that Reynolds‘ assumptions described below will
hold. For simplicity, the lower surface is assumed to
be a plane. The velocity of the fluid in the directions
x, y, and z are denoted by u, v, and w, respectively,
and the velocity of the lower surface is similarly
described by u1, v1, and w1 and that the upper surface
by u2, v2, and w2. In many practical cases, the lower
surface and the upper surface perform a straight
translational motion relative to each other. In this
case, if the x axis is in the translational direction, then
we have w1 = w2 = 0 and so the equations can be
simplified. The gap between the two surfaces of the
liquids film, be donated by h(x,z,t), with t being time.
The coefficient of viscosity of the fluid is donated
by. In deriving Reynolds‘ equation with Newtonian
fluid, the following assumptions are made as follows
[5]:
(i) The flow is laminar.
(ii) The gravity and inertia forces acting on the
fluid can be ignored compared with the
viscous force.
(iii) Compressibility of the fluid is negligible.
(iv) The fluid is Newtonian and the coefficient
of viscosity is constant.
(v) Fluids pressure does not change across the
film thickness.
(vi) The rate of change of the velocity u and w in
the direction and z direction is negligible
compared with the rate of change in the y
direction.
(vii) There is no slip between the fluid and the
solid surface.
Figure 3: Fluid film between two solid surfaces.
Figure 4: A small element of fluid.
The balance of forces acting on a small volume
element in the fluid is considered as shown in Figure
4.
Neglecting the gravity and inertia forces, we obtain
the following equation:
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2012
182
0
dxdydxdz
dxdydxdydzz
dxdzy
dydzdxx
zxyx
x
zx
zx
yx
yx
x
x
(1)
Where x is the normal stress acting on the plane
normal to the x axis and yx and zx are the shear
stress acting on the plane normal to the y axis and z
axis, respectively, in the direction of the x axis.
Equation (1) can be rearranged as follows:
0zyx
zxyxx
(2)
Let the fluid pressure be p, laminar flow of
Newtonian fluid is considered here, and the ―side
leakage pressure‖ zp is neglected, so equation
(2) can be written as follows:
y
u
yx
p
(3)
On the assumption that the rate of change of the flow
velocity u in the z direction is sufficiently small
compared with that in the y direction, the second term
of the right-hand side of the above equation can be
disregarded compared with the first term and
assumption that is constant, the equation of the
balance of forces in the x direction is finally obtained
as follows:
2
2
y
u
x
p
(4)
Integrating equation (4) twice gives the flow velocity
u and w respectively. The boundary conditions for
the velocities are (from the Reynolds assumption: no
slip condition) as follow:
0yatww,uu 11
Then the fluid velocities will be as follows:
211
2
1u
h
yu
h
yyhy
x
pu
(5)
The continuity equation for a small volume element
in an incompressible fluid can be written as follow:
0z
w
y
v
x
u
(6)
Equation (6) can be written as follows in terms of the
surface velocities from the boundary conditions:
0122
200
vv
z
hw
x
huwdy
zudy
x
hh
(7)
21
3h
0uu
2
h
x
p
12
hudy
(8)
21
3h
0ww
2
h
z
p
12
hwdy
(9)
Where and p are assumed to be constant in the y
direction. Substituting these integral (8) and (9) into
Equation (7) gives the following equation:
2121
21
2121
33
2
6
vvwwx
hh
x
hww
uux
hh
x
huu
z
ph
zx
ph
x
(10)
In many practical cases, the x axis can be taken as the
direction of the relative motion of the two surfaces,
so 0wwv 211 . And if the motion of the body
is only in one-dimensional, 0v2 . Assuming that
the viscosity, and the velocity, u1 and u2 are
constant, thus Equation (10) can be reduced to:
x
huu
z
ph
zx
ph
x
21
33
6
(11)
For two-dimensional flow, the ―side leakage
pressure‖ is neglected. Thus Equation (11) finally can
be written as:
x
hu12
x
ph
x
3
(12)
where the entrainment speed, 2
uuu 21
The 5th IMAT, November 12 – 13th
2012
183
Using these dimensionless
parameters,
00
0
2
;;p
pP;
b
xX
;'EG;R'E
uU;
b
hRH;
R'E
wW
H
Reynolds equation (12) can be written as:
X
H
W
U
dX
dPH
X
2
23
4
3
(13)
4. SOLUTION OF REYNOLDS EQUATION
Newton-Raphson formula was used for root finding
of non-linear equation of function, f(x). To find a root
where f(x) = 0, let xi is initial guess for of root.
1ii
i
xx
0xftanx'f
(14)
This equation could be simplified to:
old
i
new
i
old
i x'fxxf
(15)
The Reynolds equation (13) could be written as:
i
eeii
3
ii
HHK
dX
dPHf
(16)
The boundary condition is:
0X
PP,XXat
0P,XXat
out
in
Reynolds Equation (16) is solved by Newton-
Raphson iterative method as describe by Houpert and
Hamrock [21]. The unknowns eeH , jP and 0H are
achieved by two successive iterations represent the
new and old value after iteration.
new
ee
old
ee
new
ee HHH (17)
new
j
old
j
new
j PPP
(18)
new
0
old
0
new
0 HHH
(19)
From the Newton-Raphson definition in Equation
(16), one can write:
old
i
N
jnew
old
j
inew
j
old
ee
inew
ee
old
i
H
f
H
P
fP
H
fHf
02
0
(20)
Equation (20) can be treated as solution of
simultaneous linear equation. Since the number of
equations and the number of unknowns must be
equal, therefore N+1 number of equations is required
to solve the unknowns N+1. An additional load
condition is required, such as:
N
2j
newnew
jj WPC
(21)
A convergence criterion to stop the iteration of
Equation (17), (18) and (19) is chosen to be:
0.00001<P
PPnew
j
old
j
new
j
(22)
A linear system of N+1 equation is therefore to be
solved as below matrix:
old
N
new
N
ee
old
N
N
N
NN
ee
N
Nee
Nee
W
f
f
f
H
P
P
H
CC
H
f
P
f
P
f
H
f
H
f
P
f
P
f
H
f
H
f
P
f
P
f
H
f
2
1
0
2
2
02
0
22
2
22
0
11
2
11
00
(23)
In this simulation, the film thickness is calculated
based on fast approach of published paper by
Houpert and Hamrock [9], as this following formula:
i
2
i0i
2
XHH
(24)
The 5th IMAT, November 12 – 13th
2012
184
The elastic deformation of contacted surface is:
WRln
'dX'XXln'XX'dX
dPend
in
X
Xi
8
4
1
22
1
2
2
(25)
Applied load in dimensionless form can be calculated
from:
2PdXout
in
X
X
(26)
Viscosity of lubricant is able to be calculated from
viscosity-pressure relationship. Barus and Roelands
have proposed their equation to obtain the viscosity.
In this simulation, Barus expression will be used to
calculate the viscosity. p
0e (27)
And the density distribution of lubricant is obtained
from Dowson and Higginson.
iH
9
iH
9
iPp10x7.11
Pp10x6.01
(28)
5. RESULT AND DISCUSSION
Few graphs have been plotted to show the result from
the programming code in Figure 5 to 9. Mineral Oil
of PAO 800 was used for simulation with the
operational viscosity was 6.60 Pa.s and pressure-
viscosity coefficient was810x276.2 m
2/N. The
dimensionless parameters were used in this
simulation, U = 1x10-11
, W = 2.0452x10-5
, and G =
4500.
Figure 5 shows the plot of EHD pressure using 321
nodes and Barus pressure-viscosity model. The inlet
boundary was set at X = 4. The outlet boundary
was calculated at X = 1.17051. Initial condition of
pressure used Hertzian pressure. Using personal
computer, it took about 20 seconds to converge after
19 iterations with absolute relative error, = 0.00001.
The EHD pressure was increasing from the inlet flow
until reach a higher value at center. The pressure was
then drops a bit and increased until reach to the peak
value. The tremendous dropping of pressure is found
after the peak point. The ratio of peak pressure to
maximum Hertzian pressure was 1.377.
Figure 6 shows the relevant elastohydrodynamic film
thickness. The initial guess for dimensionless central
film thickness was chosen to be H0 = 0.6. The shape
is almost flat at contact region and form a dimple
shape near the outlet flow.
The effects of velocity on Elastohydrodynamic
lubrication were shown in Figure 7 and Figure 8.
Figure 7 shows the effect of velocity on EHD
Pressure. At lower the speed, the pressure spike value
is smaller. As the speed is increased the value of
spike also increased and the peak point is found to be
shifted to center of roller. From figure 8, it can be
seen that the effect of velocity on film thickness. At
lower speed, the value of minimum film thickness is
smaller. As the speed is increased the minimum film
thickness also increased and the film dimple shifted
to the center of roller.
Figure 9 and Figure 10 show the effect of load on
EHD Pressure. At higher load (W = 3x10-5
) the value
of pressure spike is smaller and nearly same with the
value of pressure at center of roller, and as the load is
decreased (W = 0.8x10-5
) the value of spike is
increased and the peak point is found to be shifted
near to roller center (Figure 9). Figure 10 shows the
effect of load on film thickness. At lower load (W =
0.8x10-5
) the value of minimum film thickness is
bigger. As the load is increased the value of
minimum film thickness became decreased and the
film shape is more flat along contact region. The film
dimple is found to be shifted from roller center to
outlet region.
Figure 5: EHD Pressure and Hertzian Pressure
The 5th IMAT, November 12 – 13th
2012
185
Figure 6: Elastohydrodynamic Film Thickness
Figure 7: Effect of Velocity on EHD Pressure
Figure 8: Effect of velocity on film thickness
Figure 9: Effect of load on EHD Pressure
Figure 10: Effect of load on film thickness
6. CONCLUSION
The simulation shows that the peak pressure
increases with the parameter such as velocity, and
applied load. In case of velocity, the pressure
increases and approaching its load bearing capacity.
The film thickness is also increases as the velocity
increases and the film dimple approaching to the
center of roller.
REFERENCES
[1] K Dien and H.G Elrod, A Generalized Steady-
State Reynolds Equation for Non-Newtonian
Fluids, With Application to Journal Bearings,
Trans. ASME Journal of Lubrication
Technology vol.105, page 385-390, 1983.
[2] Abdallah A. Elsharkawy, Magnetic head-rigid
disk interface hydrodynamically lubricated with
The 5th IMAT, November 12 – 13th
2012
186
a power-law fluid, Journal of Wear vol.213 page
47-53, 1997.
[3] Lim, C.H., Bohan M.F.J., Claypole, T.C.,
Gethin, D.T. and Roylance, B.J., A finite
element investigation into a soft rolling contact
supplied by a non-newtonian ink, Journal of
Appl. Phsy, vol.29, page 1894-1903, 1996.
[4] M.F.J. Bohan, I.J Fox, T.C Claypole and D.T
Gethin, Numerical Modelling of
Elastohydrodynamic Lubrication in Soft
Contacts using non-Newtonian Fluids.
International Journal of Numerical Methods for
Heat & Fluid Flow, vol. 12, no. 4, 2003.
[5] Carvalho, M.S. and Scriven, L.E., Deformable
roller coating flows: steady state and linear
pertubation analysis. Journal of Fluid
Mechanics, vol.339, pp. 143-172, 1997.
[6] Lim, C.H., Bohan M.F.J., Claypole, T.C.,
Gethin, D.T. and Roylance, B.J., A finite
element investigation into a soft rolling contact
supplied by a non-newtonian ink, Journal of
Appl. Phsy, vol.29, page 1894-1903, 1996.
[7] H.M. Chu, W.L. Li and Y.P. Cheng. Thin film
elastohydrodynamics – a power law fluid
model. Tribology International, vol. 39, page
1474-1481, 2006.
[8] A.Campos, A.Sottomayor and J.Seabra, Non-
Newtonian and Thermal Elastohydrodynamics.
Journal of Mechanica Experimental, vol. 13,
page 81-93, 2006.
[9] Houpert, L.G. and Hamrock, B.J., Fast approach
for calculating thickness and pressures in
elastohydrodynamically lubricated at high loads.
ASME Journal of Tribology, vol.108, page 411-
420, 1986.
The 5th IMAT, November 12 – 13th
2012
187
Latest System Simulation Models in Field of Heating, Refrigeration,
and Air-conditioning, and Development of System Simulator
KiyoshiSa, JongsooJ
b
a,bSchool of Fundamental Science and Engineering
Waseda University, 3-4-1, Okubo, Sinjuku, Tokyo 169-8555, Japan
Tel : 81 (3)52863259. Fax : 81 (3)52863259 aE-mail : [email protected]
bE-mail : [email protected]
ABSTRACT
The energy consumption of heating, refrigeration,
and air-conditioning systems is steadily increasing. It
is, however, not easy to reduce the energy
consumption of such thermal systems because they
have already been greatly improved to save energy.
To meet the demands of the global energy saving
policy, we need to determine the best combination
and total energy management scheme for heating,
refrigeration, and air-conditioning systems.
Simulation is a promising technology for such
investigations because it is not feasible to carry out
experiments with large-scale energy systems. High-
precision simulation models are used for these
investigations, and we are developing such models of
the heat pump, room air-conditioner, variable
refrigerant flow (VRF) system, desiccant
dehumidifier, indirect evaporative cooler, fuel cell,
solar panel, solar collector, etc This paper introduces
highly accurate models of a VRF system and an
absorption heat transformer. The simulator that we
are presently developing is also introduced. Named
‗Energy Flow +M‘, the simulator is very easy to
handle because of its user-friendly graphical user
interface (GUI). It has already been unveiled to the
world through the Internet and is expected to be used
for energy saving in heating, refrigeration, and air-
conditioning systems.
Keywords: Total energy management, Energy Flow
+M, VRF system, Heat transformer,
energy saving
Nomenclature A area m2
COP coefficient of performance -
D mass diffusivity m2s-1
d diameter m
G mass flow rate kgs-1
g gravitational acceleration ms-2
gm mass flow rate per unit length kgm-1s-1
h specific enthalpy Jkg-1
j mass flux kgm-2s1-
K overall heat transfer coefficient kWm-2K-1
L tube length m
l distance between droplets m
p pressure Pa
Q heat transfer rate kW
q heat flux kWm-2
r radius m
T temperature K
t time s
u Specific internal energy kJ kg-1
v velocity ms-1
W compressor power kW
X concentration -
x refrigerant flow direction axis m
Greek symbols
α heat transfer coefficient kWm-2K-1
β mass transfer coefficient ms-1
δ liquid film thickness m
Г mass flow rate per unit length kgm-1s-1
θ angle rad
λ thermal conductivity kWm-1K-1
μ viscosity Pas
ρ density kgm-3
Subscripts
A air
AH high temperature absorber
b bulk
COM compressor
c concentration boundary
EL low temperature evaporator
EVA evaporator
f film
GL low temperature generator
I inlet
in interface, inside
j perpendicular direction axis
O outlet
out outside
R refrigerant
S solution
sh superheat
V refrigerant vapour
W water
1. INTRODUCTION
Recently, the governments of many countries
required industries to further reduce their energy
IMAT-UI 031
The 5th IMAT, November 12 – 13th
2012
188
consumption. In the field of heating, refrigeration and
air-conditioning, the amount of energy consumed is
very high that further reduction is imperative. Hence,
investigations of optimum system combinations and
their total energy management schemes are
important. However, the electrical, air-conditioning,
and water heating loads involved in this field change
greatly as a result of environmental conditions and
human lifestyles. This makes it more difficult to use
only experiments to optimise system compositions
and operation methods. Simulation has thus been a
useful and powerful tool in investigating the
optimization and energy management of combined
systems. We have discussed the efficient and detailed
simulation of a complicated system [1] and also
developed high-precision simulation models of a
compression-type heat pump, dehumidification
system, solar panel, fuel cell, indirect evaporator,
thermal transportation system, etc. As examples, this
paper introduces our latest simulation models of a
multi-type compression air conditioner (variable
refrigerant flow, or VRF system), a multi-stage
absorption heat transformer, and a general-purpose
energy system simulator that can calculate the
characteristics of each of these systems. The
simulator, called ‗Energy flow +M‘, has been
unveiled to the public through the Internet.
2. SIMULATION OF UNSTEADY STATE
OF VRF SYSTEM
A VRF system is an air conditioning system that has
many indoor units connected to a single outdoor unit.
Examples are multi-type air conditioners used in
buildings. Compared to the single-type air
conditioner, the system performs better and saves
space and energy, even though there are several
indoor units. But because there are several indoor
units each placed in different rooms, driving
conditions for each unit differs by its usage
conditions. Thus, when one indoor unit is operated at
maximum load and another at a light load, or
switched off, an imbalance occurs in the system.
Therefore, to improve the performance and efficiency
of the system without reducing its reliability and
usability, accurate simulation is vital. This simulation
must include the various operation conditions of the
indoor units and the unsteady state of the system
during mode change. In this research, we developed a
simulation model of a system with four indoor units
and validated it by running performance tests.
2.5 System Description
The refrigerator used in our research was a
compression-type VRF system composed of one
outdoor unit and four indoor units, as shown in Fig.
1. The outdoor unit incorporated two compressors
and a heat exchanger for subcooling. The indoor
units were cassette-type, set in the ceiling. The
cooling performance of each indoor unit was
controlled by expansion valves, which were set in the
indoor units. The rated cooling capacity was 28.0
kW, and the rated input power of the compressor was
7.19 kW. The refrigerant was R410A. Fig. 1 shows
the system flow.
2.6 Mathematical Model
Fig. 2 shows the heat exchanger model adopted in the
evaporator and condenser. The heat exchanger model
equations are given below. The refrigerant-side
continuity, energy, and pressure drop were calculated
using the following equations:
R RR
GA
t x
(2.1)
_R R R R
R M in
u G hA d q
t x
(2.2)
Figure 1: Schematic flow of VRF system
Expansion valve
Compressor
Accumulator
Condenser
Evaporator
Expansion
valveHeat exchanger Expansion
valve
(a) Schematic flow of VRF system
(b) Outdoor unit
(c)
Indoor unit
The 5th IMAT, November 12 – 13th
2012
189
0RP
x
(2.3)
The air-side energy equation is as follows:
_ _M M
M M out M M in
uA d q d q
t
(2.4)
The tube-side energy equation is as below:
_m O I M out Mg h h d q (2.5)
The following equations were also used to calculate
the heat transfer rates:
M Rq T T (2.6)
M M A Mq T T (2.7)
/EVA COMCOP Q W (2.8)
2.7 Simulation and Experimental Results
The outdoor and the indoor temperatures of the
simulation and experiment were 30 °C and 27 °C,
respectively. In the PH diagram of Fig. 3, the
operation processes of the experiment and the
simulation at the rated capacity are indicated by the
dots and the line, respectively. The two results are in
good agreement. As shown in Fig. 4, we considered
the unsteady condition of the system when the
number of operating indoor units changed from one
to four. We also considered the reverse situation
when the number of operating units changed from
four to one. Further details about the latter are
available in [2]. In the former case, the operation of
the system changed when the cooling capacities and
compressor input power were in good agreement at
the rated capacity, as shown in Fig. 4. Despite the
fact that the heat transfer coefficient was considered
constant and the pressure drop was not considered,
we concluded that it was possible to accurately
predict the performances. A simulation model that
takes the detailed heat transfer into consideration can
also be developed to investigate the characteristics of
the system. However, to ensure an easy evaluation of
the very complicated unsteady state of the VRF
system, a simple simulation model with a good
accuracy is required. This is because it takes a very
long time to obtain results with complicated
simulation models. We are searching for a simulation
method that will optimally combine accurate
simulation results with calculation speed.
200 250 300 350 400 450 5000.4
1
2
4
10
Pre
ssu
re
MP
a
Enthalpy kJ/kg
T =
20
oC 40
60
80
100
120 140
s =
1.0
kJ/
kg
K
1.1
1.2
1.3
1.4
1.5
1.6
1.7
1.8
1.9
2.0
2.1
Figure 3: PH diagram of experiment and
simulation
_
_
_
R I
R I
R I
G
P
h
_
_
_
R O
R O
R O
G
P
h
_ _ _ _A I A I A I A IG P h X
_ _ _ _A O A O A O A OG P h X
x
Air inlet
Air outlet
Refrigerant
outlet
Refrigerant
inlet
(b) Control volume
Figure 2: Heat exchanger model
(a) Heat exchanger of evaporator and condenser
The 5th IMAT, November 12 – 13th
2012
190
3. CHARACTERISTICS OF DOUBLE-
STAGE ABSORPTION HEAT
TRANSFORMER
A considerable amount of steam is used in industry,
most of which is produced by boilers that burn fossil
fuels. To respond to increasing energy saving
demands, a more efficient method of steam
production is vital. There is the need for heat pump
technology to develop a waste heat–driven steam
generator. However, it is difficult for compression-
type heat pumps to generate steam of temperatures
above 120 °C because of the problem it poses to the
stability of the refrigerant and lubricant. In recent
years, there has been more focus on absorption heat
transformers because of their ability to produce high-
temperature steam from virtually only low-grade
waste heat.
20
40
60
80
100
120
Rota
tional
spee
dn
CO
M01 r
ps
0
200
400
600
Val
ve1
open
ing
puls
e
0
200
400
600
Val
ve2
open
ing
puls
e
0
200
400
600
Val
ve3
open
ing
puls
e
0
200
400
600
Val
ve4
open
ing
puls
e
0
5
10
Cooling
capac
ity
QE
VA
1 k
W
0
5
10
Cooling
capac
ity
QE
VA
2 k
W
0
5
10
Cooling
capac
ity
QE
VA
3 k
W
0
5
10
Cooling
capac
ity
QE
VA
4 k
W
0
2
4
6
8
Com
pre
ssor
input
kW
0
10
20
30
Cooling
capac
ity
kW
-200 0 200 400 600 800 10000
5
10
CO
P
Time t s
2000
2500
3000
Pre
ssure
PC
OM
O k
Pa
-200 0 200 400 600 800 1000500
1000
1500
Time t s
Pre
ssure
PC
OM
I k
Pa
Figure 4: Unsteady Simulation Results; operated
units changed from one to four
Absorber
GeneratorEvaporator
Condenser
Steam
Hot
Water
80-90oC
Cooling
Tower
40oC
Tem
pera
ture
Refrigerant
vapor
Refrigerant
lift
temperature
Refrigerant
vapor
Strong
Solution
Solution Heat
Exchanger
Weak
Solution
Feed
Water
temperature
difference
120oC
Figure 5: Concept of absorption heat transformer
3.1 System Description
Fig. 5 shows the concept of a basic absorption heat
transformer. The system comprises five main
components: absorber, generator, evaporator,
condenser, and solution heat exchanger. The lithium
bromide–water pair is used as the working fluid.
Lithium bromide brine is condensed in the generator
by using the temperature difference between the
generator and the condenser. High-temperature steam
is produced by elevating the boiling point of lithium
bromide brine while it is in the absorber. Figs. 6(a)
and 6(b) show the commercialised system, which can
produce 180 °C steam from 80–90 °C hot water. The
system consists of a generator, condenser, evaporator,
absorption evaporator, refrigerant separator, high-
temperature absorber, and steam separator. The
Duhring diagram of the commercialised system is
shown in Fig. 6(c). This commercialised system is
driven by a series-flow double-lift cycle, a test type
of which we fabricated. In this paper, we discuss the
characteristics of the objective heat transformer. A
comparison of the test data and the results of the
analysis is also used to validate the simulation model.
3.2 Mathematical Model
The mathematical model of each component mainly
consisted of equations of continuity, pressure drop,
and energy. We will explain the model of the
absorber as an example. Fig. 7 shows the absorber
model, which is separated into a falling liquid film
and inside tube flow model, liquid drop formation
model, distributoer model, refrigerant vapour flow
model, bifurcation model, mixture model, and pool
model. The falling liquid film and droplet formation
models are representatively shown in this paper.
The 5th IMAT, November 12 – 13th
2012
191
3.2.1 Falling film regime
Fig. 8 shows the falling film regime model. The
mathematical models formulated by Jeong and
Garimella [3] are as follows:
1 S S
f
v j
r
(3.1)
10S S Sv X
r
(3.2)
1 jS S S
f
jh qv h
r
(3.3)
0
, 0
V
j
sh in in
h jh
f X T j
(3.4)
The Nusselt liquid film theory respectively gives the
liquid film thickness and velocity as follows: 1 3
2
3
sin
S S
fS
S g
(3.5)
2
sin2
SS f
S
g xv x
(3.6)
The parameters of the heat transfer are given by the
following equations:
S Wq K T T (3.7)
1 1 1 1ln out
in in W pipe in o S
rK
r r r r
(3.8)
0S inT T
(3.9)
,S T S ST f X h (3.10)
,W T W WT f P h (3.11)
Assuming that the temperature distribution in the
direction of the liquid film thickness is linear, the heat
transfer coefficient is given by the following equation:
8
5
S
S
f
(3.12)
AAHH
AALL
EELL
GGLL
CC
EEHH SS
1100
2000
(a) Objective heat transformer system
(unit: mm)
Condenser
Evaporator
Refrigerant Separator HT Absorber
Generator
Absorbing Evaporator
Steam Separator
C
EL
EH
GL
AL
AH
S
Cooling
Water
Feed Water
Hot Water
Hot Water
Steam
SPRP
SV
T
P
G
Gρ
T
T
T
T
T
T
T
Gρ
T
T
T
T G
TTGρ
P
P
GT
T
T
T
G
T
T
G
GT
TT
T
T
G
P
(b) Schematic flow of system
Q Q
C
EL
EH
AL
AH
GL
Hot Water SteamCooling Water
Saturatedtemperature
Solution temperature
Intermediate steam
(c) Duhring Diagram
Figure 6: Objective absorption heat transformer
Figure 7: Absorber model
The 5th IMAT, November 12 – 13th
2012
192
The mass transfer flow rate is also calculated from
the following equation:
( )film S b inj C X X (3.13)
, = 2.0film
c
mDm
(3.14)
3,
2
b in c
S b
f
X XX X
m
(3.15)
0b SX X
(3.16)
C is the correction factor for the effect of the
surfactant. The heat transfer coefficient inside the
tube adopts the Dittus-Boelter correlation [4] for the
single-phase flow, and the Thome correlation [5] for
the two-phase flow.
3.2.2 Droplet formation regime
Fig. 9 shows the droplet formation regime
model. The mathematical models formulated by
Jeong and Garimella [3] are as below.
_ _ 0form formS I S O VG G G (3.17)
_ _
0form form
V jS I S OGh Gh G h (3.18)
_ _
0form formS I S O
GX GX (3.19)
_ _ 0form formS I S OP P (3.20)
The mass transfer flow rate is determined with the
following equation:
2
_2 formV S form a S I in
nLG d X X
l
(3.21)
1 2
24
7form
form
D
t
(3.22)
1 2
24
7form
form
D
t
(3.23)
tform and da are the formation time and the diameter of
the droplet, respectively. The details of the droplet
formation regime model are presented in [3]. The
system performance is defined as follows:
/( )AH GL ELCOP Q Q Q (3.24)
Figure 8: Falling liquid film regime model
Figure 9: Droplet formation regime model
0 5 10 15 20 251
510
50100
500
Pre
ssu
re P
kP
a
AH EL C
Solution mass flow rate GS kg/min
0.1
0.2
0.3
0.4
CO
P
0 5 10 15 20 25
56
58
60
62
64
Co
nce
ntr
atio
n X
%
Strong solution Middle solution Weak solution
Solution mass flow rate GS kg/min
0
10
20
30
40
Ste
am G
ener
atio
n R
ate
Gst
eam
kg
/h
(b) The effect of solution mass flow rate
Figure 10: Characteristics of the double-stage
absorption heat transformer
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193
3.3 Simulation and Experimental Results
Table 1 shows the simulation and experimental
conditions. The simulation and experimental results
are shown in Figs. 10(a) and 10(b). Figs. 10(a) and
10(b) respectively show the effects of the hot water
inlet temperature and the solution mass flow rate on
the characteristics of the system. As can be seen, the
experimental results agreed with those of the
simulation, thereby validating the simulation models.
We also investigated the characteristics of the
double-stage absorption heat transformer and
confirmed that the system could efficiently generate
high-temperature steam from low-grade heat without
losing stability.
4. DEVELOPMENT OF SIMULATOR
The detailed models of the sample thermal systems
were shown above. The simulation with these models
was not easy owing to the complexity of the models.
We therefore developed a general-purpose analysis
simulator by adding a graphical user interface (GUI).
Fig. 11(a) shows the main screen of the simulator,
called ‗Energy flow +M‘. By pushing the ‗Start
button‘ at the upper right, the pallet is opened on the
display, as shown in Fig. 11(b). The modules are
selected in the pallet using the appropriate icons. One
feature of this simulator is its ability to carry out a
simulation by just connecting modules together. Even
when the flow of some elements changes, the entire
system analysis can easily be carried out again.
Moreover, the energy of a large-scale system and the
simulation of a complicated unsteady state can also
be easily analysed as described in sections 2 and 3.
Furthermore, the calculation results can be obtained
as an Excel document in Fig. 11(c). Fig. 11(d) shows
the simulator that can be also expressed on pallet
with EXCEL interface. This simulator by EXCEL
interface has good features that make simulation to
be easily carried out. As shown in Fig. 11(b) and Fig.
11(d), we could develop two types of simulator by
adopting the modular analysis method into thermal
system. We are also making various calculation
modules including the transient characteristics of
thermal systems for simulator with GUI and EXCEL
interface. The simulator we have developed greatly
reduces the burden on the user in developing
simulation codes, and the simulator by GUI interface
is available on the Internet and can be accessed
anytime.
Figure 12: System flow on EF+M (comparison of
experimental and simulation results)
75 80 85 90 951
510
50100
500
Pre
ssu
re P
kP
a
AH EL C
Hot water inlet temperature THW oC
0.1
0.2
0.3
0.4
CO
P
75 80 85 90 9558
59
60
61
62
63
64
Co
nce
ntr
atio
n X
%
Strong solution Middle solution Weak solution
Hot water inlet temperature THW oC
0
10
20
30
40S
team
Gen
erat
ion
Rat
eG
steam
kg
/h
(a) Effect of hot water inlet temperature
(d) Elements on pallet of EXCEL interface
Figure 11: Simulators for VRF system
The 5th IMAT, November 12 – 13th
2012
194
5. CONCLUSION
In our study, we developed mathematical models of
the multi-type compression air conditioner (VRF
system) and the multi-stage absorption heat
transformer, which were used as examples of heating,
refrigeration, and air-conditioning systems.
Considering the complexity of these models, we also
discussed a simulator that we developed. The
mathematical models were fully validated by the
agreement of the experimental and simulation results.
Anyone in the world can easily use the simulator on
the Internet. We are expanding it for use with other
systems, which will be discussed in a future paper.
REFERENCES [1] K. Saito and J.S. Jeong, ―Latest system simulation
models for heating, refrigeration, and air-conditioning
systems, and Their applications,‖ IJACR , vol. 20, no.
1, pp.13, 2012.
[2] K. Ohno, K. Saito, H. Nakamura, H. Murata, Y. Jinno,
K. Konishi, and Y. Nakaso, ―Unsteady State
Simulation of VRF Systems,‖ 10thIEA Heat Pump
Conference 2011, Tokyo, Japan, 3.33, 2011.
[3] S. Jeong and S. Garimella, ―Falling-film and droplet
mode heat and mass transfer in a horizontal tube
LiBr/water absorber,‖ Int. J. Heat Mass Transfer, vol.
45, pp. 1445–1458, 2002.
[4] F. W. Dittus and L. M. K. Boelter, ―Heat transfer in
automobile radiators of the Tubular Type,‖
Publications in Engineering, 2, 443, Univ. of
California, Berkeley, 1930.
[5] N. Kattan, J. R. Thome, and D. Favrat, ―Flow boiling
in horizontal tubes: part 3–development of a new heat
transfer model based on flow pattern,‖ Trans. ASME,
vol. 120, pp. 156–165, 1998.
(a) Main screen of pallet
(b) Elements on pallet of GUI interface
(c) Simulation results on pallet by EXCEL file
The 5th IMAT, November 12 – 13th
2012
195
Drag Reduction of Bamboo and Abaca Fiber Suspensions
in Circular Pipe
Gunawana, M. Baqi
b, S. Fathernas
c and Yanuar
d
aDepartment of Mechanical Engineering, Faculty of Engineering
University of Indonesia, Depok 16424
Tel : (021) 7270032. Fax : (021) 7270033
E-mail : [email protected]
bDepartment of Mechanical Engineering, Faculty of Engineering
University of Indonesia, Depok 16424
Tel : (021) 7270032. Fax : (021) 7270033
E-mail : [email protected]
cUnder Graduate Student, Department of Mechanical Engineering
University of Indonesia, Depok 16424
Tel : (021) 7270032. Fax : (021) 7270033
dDepartment of Mechanical Engineering, Faculty of Engineering
University of Indonesia, Depok 16424
Tel : (021) 7270032. Fax : (021) 7270033
E-mail : [email protected]
ABSTRACT
The drag reduction of dispersions of fibers in aqueous
solutions of was studied as a function of concentration
with a circular pipe apparatus. Experiments were
carried out by measuring the pressure drop. The
purpose of this research is to investigate the reduction
of pressure drop in a circular pipe with the addition
fiber in aqueous solution. Circular pipe with 4 mm of
diameter is used in this study. Concentration of
bamboo and abaca fibers solutions are 200 ppm and
300 ppm. It was found that fibers solutions give rise to
drag reduction in turbulent flow range. Experimental
was conducted from low to high Reynolds number up
to 55,000. We observed a maximum drag reduction
ratio of 7 % at Reynolds number about 35,000 and
found that increased by increasing a concentration of
fiber solution.
Keywords: drag reduction, bamboo and abaca fibers
solution, pressure drop, turbulent flow.
1. INTRODUCTION
Environmental issues are a major topic of interest
studied mainly in energy efficiency. One topic of
particular interest is drag reduction in fluid transport
systems. The addition of a small amount of additives
suspension such as polymers, surfactans and fibers to
a turbulent Newtonian fluid flow can result in a drag
reduction, which appears in a number of flow fields,
and has received considerable attention. This
phenomena is reach to investigate sice initial
publication of Toms [1]. Using surfactans [2,3] to
obtain drag reduction in turbulent flow is very
effective and low mehanical degradation. However,
surfactans are contain as syntetic chemical so very
dangerous in environment. Although polymers are
safe in environment, they are not practical due to their
significant mechanical degradation. Yanuar et al [4,5]
also investigated the influence of biopolymer solutions
for drag reduction in internal and external flow. His
research show that biopolymer can reduce frictional
drag up to 30% but the mechanical degradaton
occured fastly.
Ogata, Numakawa and Kubo [6] reported that fiber
solutions from bacterial cellulose undergoing a
turbulent flow in a pipe thereby require a lower
pressure drop to maintain the same volumetric flow
rate. The addition of small amounts of fiber to the
flowing fluids can show significant effects on a lot of
flow types. It was found that bacterial cellulose
suspensions give rise to drag reduction in the turbulent
flow range. The maximum drag reduction ratio of 11%
and found that it increased with the concentration of
the fibre suspensions from bacterial cellulose.
The other fibres suspensions also investigated by
reserchers [7,8,9] such as asbestos or nylon fibers.
This research obtained that nylon and asbestos fibers
are effective to reduce drag but requires high
concentation and have a disadvantages with regard to
environmental load.
IMAT-UI 032
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Bamboo and abaca fibres are considered to have a low
environmental load and it is naturally derived from
bamboo and abaca plants. The purpose of this research
is to investigate the reduction of pressure drop in a
circular pipe with the addition fiber in aqueous
solution. Circular pipe with 4 mm of diameter is used
in this study. Concentration of bamboo and abaca
fibers solutions are 200 ppm and 300 ppm. It was
found that fibers solutions give rise to drag reduction
in turbulent flow range. Experimental was conducted
from low to high Reynolds number up to 55,000. We
observed a maximum drag reduction ratio of 7 % at
Reynolds number about 35,000 and found that
increased by increasing a concentration of fiber
solution.
2. EXPERIMENTAL SETUP
Figure 1: Experimental setup
The experimental set up is shown in figure 1. Figure 1
shows the test of rheological properties. The fibers
suspensions are circulated by piston pump. The
pressure drop gradient is measured at 1000 mm length
between each pressure tap by pressure transducer. The
diameter of pressure tap is 2 mm. The inner diameter
of test circular pipe d is 4 mm. The shear stress and
the shear rate can be obtained by measuring the
pressure drop gradient and the gradient of velocity,
respectively. The concentrations of fibers solution in
form of aqueous suspensions are 200 ppm and 300
ppm. The temperature is kept at 25 oC.
The bamboo and abaca fibers were taken from
bamboo and abaca plant. The size of this fibers are
homogen and have length about 0.5 mm. Bamboo and
abaca tree was gently wiped across the surface of a
special smooth metal table. After the bamboo and
abaca are wiped, fibers will separate with the other.
Then the fibers are dried and made the cutting process.
3. RHEOLOGICAL MODELS
The shear stress, τ is proportional to the velocity
gradient, (shear rate), can be described by
Newtonian model:
du
dy
(1)
Where is constant for the particular fluid that
is viscosity. The Newtonian viscosity depends on
the temperature and the pressure and is
independent of the shear rate. The viscosity is
defined as the ratio of shear stress to shear rate.
Several rheological models or rheological
equations of state have been proposed in order to
describe the nonlinear flow curves of non-
Newtonian fluids. Non-Newtonian fluids
Bingham, pseudo plastics, and dilatants are those
for which the flow curve is not linear. The
viscosity of a non-Newtonian fluid is not constant
at a given temperature and pressure but depends
on other factors such as the rate of shear in the
fluids.
Thus, the relationship shear stress and shear rate
may be described by measuring the pressure drop
gradient and the volumetric flow rate in circular
pipe flow is given by:
8
4
D P u
L D
(2)
Where: D is the inner pipe diameter, P is
pressure drop, L is the length of pipe (test
section), and u is the avarage velocity.
Coefficient of friction, f, can be obtained by
Darcy Equation:
2
2D gf h
L u
(3)
Where: f is the coefficient of friction, h is the
head gradient over the considered pipe length,
and g is the gravity acceleration.
Drag reduction in pipe can obtain by equation:
The 5th IMAT, November 12 – 13th
2012
197
100%fiberf f
DR xf
(4)
4. RESULT AND DISCUSSION
Figure 2: Flow curve of fibers suspensions
Figure 2 shows the flow curves of the fibers
suspensions. The wall stress τ and sheer rate were
calculated from the experimental data from laminer to
turbulent flow regime. The solid line in figure 2
indicates the value obtained by the viscosity of water.
The data of bamboo fiber and abaca fiber are shown
linear relationship between wall stress and flow rate. It
is incated that the fibers solutions are Newtonia fluid.
The viscosity is seen to increase with concentration.
The value of figure 2 is used to obtain the Reynolds
number and friction factor.
103
104
105
10-1
Pure water
Abaca Fibre 200 ppm
Abaca Fibre 300 ppm
Bamboo Fibre 300 ppm
Bamboo Fibre 200 ppm
T = 25o C
f = 64/Re*
f = 0.3164*Re* (̂1/4)
f
Re
Figure 3: Flow curve of fibers suspensions
Figure 3 shows the relationship between Reynolds
number and friction factor coefficient based on the
measured pressure drop for 2 suspensions with 2
variation of concentration. The data will be compared
with Hagen Pouiselle equation in laminar flow and the
Blasius equation in turbulent flow. The data of water
also shown in this figure. The coefficient of friction of
fibers suspensions fit with the coefficient of friction of
water for circular pipe in laminar flow. In turbulent
flow, up to Reynolds number about 25.000, the
coefficient of friction also fit with coefficient friction
of water and Blasius equation. The data show that
coefficient of friction fibers at Re > 25.000 is lower
that water data and Blasius equation. The drag
reduction is increase with increasing of fiber
concentration. The data of bamboo fiber is lower that
abaca fiber in same concentration and Reynolds
number.
102
103
3
4
5
6
7
Re.f 1/2
f -1
/2
Bamboo Fibre 300 ppm
Abacca Fibre 300 ppm
Bamboo Fibre 200 ppm
Abacca Fibre 200 ppm
C
B
A
Figure 4: Characteristic of drag reduction
Figure 4 shows the relationship between f -1/2
and
Re.f1/2
where f is denote the Fanning friction factor. It
can be seen that in the Laminar flow regime where
Re.f1/2
is small, the data well fitted by a Newtonian
laminar flow curve (C). In contrast the data in
turbulent flow regime is alligned parallel in the curve
(A). Drag reduction occured if the data is higher than
curve A. The figure shows that increasing the
concentration, can increase the data from the curve A.
Data of bamboo and abaca fibers at high Re.f1/2
are
parallel with curve A buat not fitted. The data is
greater that curve A. The drag reduction that occured
based on this graph is indicates a Type B drag
reduction, which can be seen in fiber and polymer
suspensions. Generally, a type B dra- reducing
mechanism is associated with suppression of vortices.
For fiber suspensions the flow fields influence the
fluid resistance, and the fiber suppress vortices when
they are uniformly distributed in the flow direction,
thus resulting the drga reduction. The friction factors
of high-concentration solutions data still far to the
Virk‘s line (B) according to the increase in the
Reynolds number, Re.
The 5th IMAT, November 12 – 13th
2012
198
2,0x104
3,0x104
4,0x104
5,0x104
2
4
6
8
10
12
Bamboo Fibre 300 ppm
Abaca Fibre 300 ppm
Bamboo Fibre 200 ppm
Abaca Fibre 200 ppm
DR
(%
)
Re
Figure 4: Ratio drag reduction
Figure 4 shows the ratio drag reduction of fiber
suspensions. Based on figure, it can be seen from
these results that drag reduction for a given bammbo
and abaca fibers concentration only occurs above a
critical value Reynolds number. The value of critical
Reynolds number is about 25.000. Below this critical
value the fluid exhibits normal Newtonian viscous
behavior, although the flow is turbulent flow.
The maximum drag reduction occured at the Reynolds
number about 35.000. The drag reduction increases
start from Reynolds number 25.00 up to 35.000. After
Reynolds number about 35.000, the data shows
constant. The drag reduction increased slightly with
increasing concentrations. Drag recution of bamboo
fiber is greater then abaca fiber. The reported value for
bamboo fiber suspensions of 300 ppm and 2000 ppm
in the turbulent flow range in circular pipe has
maximum drag reduction about 7% and 5%. For same
concentration and Reynolds number, drag reduction
for abaca fiber is 6% and 4% respectively.
5. CONCLUSION
Pressure drop measurements for bamboo fiber and
abaca fiber suspensions flowing in circular pipe were
performed and the following result is obtained . The
drag-reduction effect of the bamboo fiber and abaca
fiber were verified. The effect occurred only above
some critical Reynolds number which was affected by
the concentration of the fiber suspensions. Drag
reduction is significantly affected by the type of fiber
and the concentration of fiber. For bamboo fiber, the
range drag reduction is about 5% through 7% depend
on concentration. For abaca fiber, the drag reduction
accured about 4% through 6% respectively. The
maximum drag reduction is 7% at Reynolds number
about 35.000. Drag reduction of bamboo fiber and
abaca fiber is type B drag reduction same as polymer
type drag reduction.
ACKNOWLEDGMENT
This work is supported by the Directorate for Research
and Community Service, University of Indonesia
(RUUI).
REFERENCES
[1] Toms. B. A. ―Somle observations onl the flow
of linear polymer solutions through straight tubes
at large Reynolds numbers," International
Congress onl Rhecology,I Holland. 1948.
Amsterdlam. North I lolh.aid, 1949, Part 11, pp.
135-141
[2] F.-C. Li, Y. Kawaguchi, K. Hishida, and M.
Oshima, ―Investigation of turbulent structures in
a drag-reduced turbulrnt channel flow with
syrfactant additive by stereoscopic particle image
velocimetry‖, Experiments in Fluids, vol. 40, no.
2, pp. 218-230, 2006.
[3] H. W. Bewersdorff, ―Rheology of drag reducing
surfactat solutions‖, in Proceedings of the ASME
Fluids Engineering Division Summer Meeting
(FED‘96), vol. 237, pp. 25-29, San Diego, Calif,
USA, 1996.
[4] Yanuar and Watanabe K. ―Tom‘s effect of guar
gum additive for crude oil in flow through square
ducts.‖ The 14th
International symposium on
transport phenomena. Bali Indonesia. Elsevier
2004. P.599 – 603.
[5] Yanuar, Gunawan and M. Baqi, ―Characteristics
of Drag Reduction by Guar Gum in Spiral Pipes‖
Journal Teknologi. Vol.58 2012, pp. 95–99.
[6] Satoshi Ogata, Tetsuya Numakawa and Takuya
Kubo. ―Drag reduction of bacterial cellulose
suspensions. Advanced in Mechanical
Engineering. 2011. Pp 1-6.
[7] P.S. Virk and R.H. Chen, ―Type B drag reduction
by aqueous and saline solutions of two
biopolymers at high Reynolds number‖, in
Preceedings of the 2nd International Symposium
on Seawater Drag Reduction, pp. 545-558,
Busan, Korea, May 2005.
[8] A.A. Robertson and S.G. Mason, ―The
characteristics of dilute fiber suspensions‖,
TAPPI, vol. 40, pp. 326-334, 1957.
The 5th IMAT, November 12 – 13th
2012
199
[9] W. Mih and J. Parker, ―Velocity profile
measurements and phenomenological description
of turbulent fiber suspension pipe flow‖, TAPPI,
vol. 50, pp. 237-246, 1967.
[10] Yanuar and Watanabe K. ― Drag Reduction of
Guar Gum in Crude oil‖. The 13th
International
Symposium on Trannsport Phenomena. Victoria
Canada. Elsevier 2002. P. 833 – 836.
[11] Yanuar, et al. ―Hydraulics conveyances of mud
slurry by a spiral pipe‖ Journal of Mechanical
Science and Technology 23 (2009) 1835 – 1839.
Springer.
The 5th IMAT, November 12 – 13th
2012
200
A Study on the Effect of Exhaust Gases on the Indoor Air Quality
Onboard Ships
Arman Ariffin
a and Hayati Abdullah
b
aPlan Department, Royal Malaysian Navy Headquarters, Ministry of Defence,
Jalan Padang Tembak, 50634 Kuala Lumpur, Malaysia
Email:[email protected]
bFaculty of Mechanical Engineering, Unversiti Teknologi Malaysia
81310 UTM Johor Bahru, Johor, Malaysia
Email:[email protected]
ABSTRACT
The understanding of the exhaust gas behaviour from
ship plume is necessary in order to avoid serious
operational problems onboard modern ships. The
interference of exhaust gas on the air intake for
ventilation system can result in poor indoor air quality
(IAQ) and can adversely affect the performance of
human and equipment onboard ships. This paper
presents the results of an indoor air quality study
carried out onboard ship and focuses on parameters
such as temperature, humidity and major air
pollutants. An initial study of the velocity ratio K
which represents the ratio of exhaust velocity to
relative wind velocity will also be presented to
investigate the effect of the velocity ratio K on the
indoor air quality for three locations of the air intake
for the ventilation system and two different conditions
of the ship namely alongside and cruising at an
economical speed.
Keywords : Indoor Air Quality, Ship’s Exhaust
Gases, Velocity Ratio
1. INTRODUCTION
The understanding of exhaust gas behavior is
important in ship design and smoke nuisance
onboard ship has long been studied since the
evolution of ship construction. The downwash of
exhaust smoke, especially on modern naval ship
with lower stack in order to reduce the infrared
signature, can cause negative effects such as
suction of hot plume to intake of gas turbine or
HVAC onboard, high temperature and
contamination of topside electric equipment and
interference of exhaust smoke with the flight
deck operation. Problem of the exhaust gas
behavior was reported by Nolan [1]. In the study,
various types of wind tunnel testing were
conducted with the cooperation of Maritime
Commision‘s smoke test program and Langley
Memorial Aeronautical Laboratory. It started
with the actual model S.S America with a scale of
1:96. It was selected as the first model in the test
because of smoke trouble. Some modifications
were made by raising the stack 15 feet (4.57m).
Equipment used in the test includes an
anemometer to measure the wind velocity and an
orifice in the air supply to identify the smoke
velocities. The test was also conducted with
smoke temperatures of 300 oF (149
oC) and 500
°F (260 oC). On the first test, it was concluded
that the height of the stack was affecting the flow
of the smoke. Higher stack will create turbulence
away from the aft stern. The downwash does not
take the smoke down far enough to reach the
turbulence zone, so it will float clear of the ship
and it is an advantage for all on board. Nolan
introduced the S/W ratio or velocity ratio K
where S is the smoke velocity and is evaluated as
the stack gas volume per second divided by the
discharge stack area. W is the wind velocity
relative to the ship. He found that with an S/W
ratio or K below 2, if the wind was at an angle of
even 5 degrees, the smoke could travel all the
way to the base of the stack. Other stacks were
tested with this model but satisfactory results
were only obtained if the S/W ratio was at least 2.
Heated smoke was also used in the study. The
temperature of the smoke was previously about
130 oF (54.4
oC). The temperatures were then
increased to 300 oF (149
oC) and 500 °F (260
oC).
It was concluded that heated smoke floats clear of
the ship better than unheated smoke. The
experiments carried out included different
characteristics of the funnel from the velocity
IMAT-UI 033
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201
ratio K, height of stack, stack design, and exhaust
temperature. It was concluded that the K value
shall be at least 2 to keep the smoke above the
turbulence zone. The stack design should also be
as small as possible to allow high speed of smoke
velocity. The heated smoke floats clear of the
ship because the buoyancy of the smoke exerts
considerable influence on the performance of a
ship‘s stack.
A flow visualization study of exhaust smoke-
superstructure interaction onboard naval ships
was studied by P.R. Kulkarni et al [2]. Four
variants of superstructure arrangements with 1:50
model were studied to gain an understanding of
the typical flow field around the topside of naval
ships and the interaction between bluff body air
wake and the ship exhaust. They noted that most
of the problems arise from the fact that the funnel
superstructure creates a low-pressure zone behind
the exhaust stack or lee side and it will naturally
suck in wind into that area. The flows that will be
sucked in are the flows from the funnel top and
side as well as the exhaust gases from the funnel.
As in the study by Nolan, they concluded that a
velocity ratio of at least 2 is a requirement for
satisfactory operation and to prevent the sucking
of exhaust into the gas turbine intakes.
Apart from wind tunnel studies, numerical
investigations also gave similar results on the
importance of the velocity ratio K on the exhaust
gas behavior. S. Ergin et al [3] carried out
numerical studies on the exhaust smoke-
superstructure interaction on a naval ship and
demonstrated that computational fluid dynamics
can be a powerful tool to study the problem of
exhaust smoke-superstructure interaction. They
investigated 3 main elements that affect the
smoke print onboard which are the yaw angles,
velocity ratios and the exhaust smoke
temperature. The values of the velocity ratio K
studied are 1,2, 3 and 4. The yaw angle, ψ from
Port side included 0°, 5°, 10°, 15°, 20° and 30°,
and exhaust temperature studied are for 15°C,
200°C, 300°C and 400°C. The numerical results
concluded that to minimize the effect of
downwash, the yaw angle, ψ should be more than
10° and velocity ratio K equals to the value of 2
should be maintained. They also note that the
effect of buoyancy forces on the plume rise when
compared to momentum is not as significant.
A numerical study on the effect of air quality
using a very large eddy simulation program was
presented by F.Camelli et. al [4]. They studied
the coupling of the ship topside flow to the
thermal transport and diffusion of the exhaust gas
for 0o and 30
o of angle of attack of the inflow.
They computed the temperature, NOx and SO2
levels and their results compared well with
experimental data from wind tunnel testing.
P.R. Kulkarni et al [5](Kulkarni, Singh, &
Seshadri, 2005) also conducted experimental
study of the flow field around a simplified
superstructure with two funnels. Two
configurations were studied in which the first
configuration is when the 2 funnels are aligned in
the centerline and in the second configuration, the
2 funnels are not aligned. They observed that
when two funnels are aligned with the wind,
there is a momentum shielding effect by the
upstream plume on the downward plume but
when the funnels are offset with respect to the
incident wind, there is no shielding effect of the
forward plume on the aft plume.
Huang J. et al [6] carried out a CFD study on the
temperature and NOx concentration levels. Their
results showed that operation with lower smoke
speed and larger head wind speed will result in a
lower exhaust plume that can move close enough
to the superstructure of the ship to be entrapped
in the down flow downstream of the
superstructure and into the flight deck. This poses
a risk to potentially harm equipment and
ventilation intakes.
2. PROBLEM DESCRIPTION
The tests presented in this paper were carried out
in the South China Sea in the range of 50 nautical
miles close to coastal. The cruising condition is
considered a normal activity without any
additional requirements involved.
The ship has complex Heating, Ventilating and
Air Conditioning (HVAC) System onboard. It is
complete with the capability to be operated in
normal environmental condition up to the
Chemical, Biological, Radiological and Nuclear
(CBRN) contaminated area. In the general
arrangement, it has 3 independent Self
Confinement Zone (SCZ). Each SCZ zone has
its own intake, air handling and air conditioning
system with total independence configuration.
The 5th IMAT, November 12 – 13th
2012
202
In the main machinery configuration, the ship is
powered by two marine diesel engines with a
total of 12 Megawatt and four marine diesel
generators with a total of 2 Megawatt. To
increase the stealthy design of this ship in
infrared signature, it was constructed with 6
points of discharge exhaust at sea water level on
both sides of the ship. The points of exhaust
discharge are depicted in Figure 1.
Figure 1: Points of exhaust discharge
It was reported that the indoor air quality was
affected by the exhaust system. Studies were
then conducted to identify the possible cause of
this problem.
3. METHODOLOGY FOR INDOOR
AIR QUALITY TEST
Indoor Air Quality (IAQ) Test was conducted to
evaluate the IAQ onboard ship. It is used to
confirm the level of concentration at several
measuring points and would be the baseline for
future studies. An initial walkthrough of the area
during normal activity provides information on
all four basic factors influencing IAQ (occupants,
HVAC system, pollutant pathway and
contamination sources). All studied areas were
visited for initial investigation and several
measurements were taken. Appropriate
instruments and measurements were selected for
data collection process.
The sampling probes were located between 75
and 120 cm from the floor of the sampling
position and sampling was carried out when the
ship was alongside and cruising. Typical
measurements using direct reading devices were
employed for measurements of temperature,
relative humidity (RH), carbon dioxide, sulfur
dioxide, nitrogen dioxide and nitrogen oxide.
Measurements using air pumps and collection
media were sent for laboratory analysis.
Air test for temperature, RH, carbon dioxide,
sulfur dioxide, nitrogen dioxide and nitrogen
oxide were taken with the portable gas analyzer
Model Testo 350XL. Air tests for diesel dust,
HCI, H2SO4, HN3, HN2, VOC compounds were
taken with GilAir-5 & Low Flow Module
Constant Flow air pumps. The air sampler was
calibrated using the Gilian Gilibrator 2
Calibration System.
The analysis was conducted at the laboratory
facility of Australia Laboratory Services (ALS).
ALS laboratories operate in compliance with ISO
17025 (General requirements for the competence
of testing and calibration laboratories). The
diesel dust, inorganics acid and VOC are
analyzed using methods with reference to NIOSH
5040, NIOSH 7903 and NIOSH 1500.
4. ENVIRONMENTAL CONDITIONS
The IAQ test was conducted in March 2010 with
two different conditions; alongside and cruising.
The daily weather summary recorded is given in
Table 1. Table 1: Daily Weather Summary.
Temperature:
Mean 28
Max 32
Min 25
Moisture:
Dew Point 25
Average RH 80
Max RH 89
Min RH 63
Sea Level Pressure 1007.77 hPa
Wind:
Average Speed 6 km/h
Max speed 13 km/h
Visibility 8.9 kilometers
Event Rain
5. SAMPLING PROCESS
The IAQ test was conducted in two different
conditions. Each case uses the same method of
sampling process:
a. Case 1: Ship alongside.
b. Case 2: Ship cruising.
There are 3 locations identified to be
significant in the effect to the indoor air quality
The 5th IMAT, November 12 – 13th
2012
203
as indicated in Figures 2, 3 and 4. They are the
intake positions for each SCZ and marked with
―star‖ :
a. Location 1: Intake for SCZ 1.
Position at forward upper deck
starboard side.
b. Location 2: Intake for SCZ 2.
Position at amidships flag deck
starboard side
c. Loaction 3: Intake for SCZ 3.
Position at aft boat deck starboard
side.
Figure 2: Location 1.
Figure 3: Location 2.
Figure 4: Location 3.
The direct reading measurements were
taken using the portable gas analyzer Model
Testo 350XL for temperature, RH, carbon
dioxide, sulfur dioxide, nitrogen dioxide and
nitrogen oxide.
For the diesel dust, HCI, H2SO4, HN3,
HN2, and VOC compounds, measurements were
taken with the GilAir-5 and Low Flow Module
Constant Flow air pumps using the sampling
bottles. The duration for each position is 1 hour
and accumulates 12 liters of air. The collected air
in the sampling bottles were placed in a sealed
box and sent to ALS for laboratory analysis.
6. RESULT AND DISCUSSION
For both cases, the result of the concentration on
sulfuric acid is shown in Table 2.
Table 2: Sulfuric acid concentration (mg/m3) at air intake
area.
Area Case 1 Case 2
Intake SCZ 1 3.42 47.96
Intake SCZ 2 8.09 9.3
Intake SCZ 3 2.42 22.28
The trial for case 1 has the K value of 2.5 and the
yaw angle, ψ = 60° relative to portside. It can be
seen that the intake for all SCZ was slightly
affected by smoke with a higher concentration
level at SCZ 2. The trial for case 2 has the K
value of 0.2 and the yaw angle, ψ = 45° relative
to portside. It can be seen that the intake for all
SCZ was significantly affected by the exhaust
gas. The results indicate a similar trend to the
results obtained in the literature where higher K
values is required for satisfactory ship operation
in terms of exhaust gas behavior and the
downwash phenomena. However, further studies
need to be carried out with more sample data.
Wind tunnel testing for the side exhaust
configuration is planned for the future in order to
further understand the interaction of exhaust
smoke and superstructure interaction for this
configuration.
7. SUMMARY
The study on indoor air quality has indicated
similar trend to the published research results. It
is important to ensure that the concentration
levels of species such SO2 and NOx do not rise
above healthy levels onboard ships so as not to
harm the crew and ship. The results of future
studies in this area will lead to a better
understanding of the recirculation zones and how
they affect the concentration levels of air
pollutants and will be able to assist in the
understanding of the actual effect of exhaust
smoke to the internal environment onboard ships.
The 5th IMAT, November 12 – 13th
2012
204
Acknowledgment
The authors wish to thank Royal Malaysian Navy
and Universiti Teknologi Malaysia for the
support in carrying out this project.
REFERENCES
[1] Nolan, R. W. (1946). Design of stacks to
minimise smoke nuisance. Trans
SNAME, 54, 42-82.
[2] Kulkarni, P. R., Singh, S. N., & Seshadri,
V. (2005). Flow visualization studies of
exhaust smoke-superstructure interaction
on naval ships. Naval Engineers Journal,
117(1), 41-56.
[3] Ergin, S., Parah, Y., & Dobrucali, E.
(2012). A numerical investigation of
exhaust smoke-superstructure interaction
on a naval ship
Sustainable Maritime Transportation and
Exploitation of Sea Resources.
[4] Camelli, F., Sandberg, W. C., &
Ramamurti, R. (2004). VLES Study of
Ship Stack Gas Dynamics. The 42nd
AIAA Aerospace Science Meeting and
Exhibition.
[5] Kulkarni, P. R., Singh, S. N., & Seshadri,
V. (2005). Experimental study of the
flow field over simplified superstructure
of a ship. Int J Maritime Eng, IJME Part
A3, 147, 19-42.
[6] Huang, J., Carrica, P. M., & Stern, F.
(2012). A method to compute ship
exhaust plumes with waves and wind.
International Journal for Numerical
Methods in Fluids, 68(2), 160-180.
The 5th IMAT, November 12 – 13th
2012
205
ABSTRACT
This article presents a study to estimate the potential
saving in annual operating cost of a hypothetical
greenhouse used for planting strawberry, in Johor
Bahru, Malaysia. The greenhouse needs to be
maintained at a constant temperature of 20°C at all
time. The goal of this study is to select a suitable TES
system that can save the annual cost of electricity
usage to meet the cooling load requirement of the
greenhouse, based on a 24 hours operating duration
and local electricity tariff. Comparison is made with
the annual cost for running a conventional air-
conditioning (AC) system to meet the cooling
requirement. The cooling load requirement of the
greenhouse dictates the capacity and size of the
potential TES systems, which was estimated based on
the highest total annual cooling load. Three TES
system operating arrangements were considered in this
study: TES full storage combined with AC systems,
TES full storage and TES partial storage. Among
these three arrangements, the TES full storage was
found to have the highest an annual cost saving of
about RM 58,990 compared to the cost of using the
conventional AC system alone. This represents about
68 % of annual operating cost saving, which is
considered very significant.
Keywords : Thermal Energy Storage (TES), Lowland
Farming House, Cooling Load
Estimation, Greenhouse Air-
conditioning System.
1. INTRODUCTION
Thermal energy storage (TES) can be considered as
the temporary storage of energy for later use when
cooling or heating is needed. For cooling applications,
energy is stored at low temperatures while for heating
applications, the energy is stored at high temperatures
[1]. The interest in cool storage for commercial
applications grew significantly especially for countries
in hot and humid regions where a very high on-peak
demand load occurred in the midday but persisted for
a short period of time [2]. TES technology is seen as
one of the primary solutions to the electrical power
imbalance between its production and continuous
demand. The fundamentals, case studies, design and
history of the TES be found in various literatures
[3,4]. TES may also be a potential cost-saving solution
in countries where the electricity rate is on time-based.
This technology can shift cooling energy usage time
from on-peak periods to off-peak periods and hence
avoids peak demand electricity charges.
In non-residential buildings the TES technology may
become an attractive alternative if one or more of the
following conditions exist [5]: short period of HVAC
demand, frequently varying HVAC loads, infrequent
or cyclical loads, HVAC demand and supply do not
match, economic incentives are provided for using off-
peak energy, energy supply is limited by the utility
company, hence making it impossible to satisfy the
maximum load directly, and the capacity of an
existing chiller is too low to meet the peak load
demand. TES technology can be promoted because it
can substantially reduce the total energy consumption,
conserving fossil fuels and reducing costly imports of
oil and other energy resources. With TES, one can
adjust the time-discrepancy or rate variance between
energy supply and energy demand, thereby playing a
vital role in the conservation of energy [6,7].
TES systems are usually operated in two modes: full
storage and partial storage. The partial storage TES
can further be categorized into load leveling and
demand limiting storage systems [8]. The Full storage
TES systems, also known as load shifting systems are
typically designed to shift all building cooling load
demands from the on-peak period to the off-peak
IMAT-UI 034
Application of Thermal Energy Storage System For a
Lowland Greenhouse
Haslinda Mohamed Kamara, Nazri Kamsah
b & Norull Ahmad Norull Azman
aFaculty of Mechanical Engineering
Universiti Teknologi Malaysia, Skudai, Johor
Tel : (+607) 5534748. Fax : (+607) 5566159
E-mail : [email protected]
bFaculty of Mechanical Engineering
Universiti Teknologi Malaysia, Skudai, Johor
Tel : (+607) 5534749. Fax : (+607) 5566159
E-mail : [email protected]
The 5th IMAT, November 12 – 13th
2012
206
period of a day. In this system, the chiller runs at its
full capacity during the off-peak period and night time
when the building cooling load demand is low and
electricity tariff is cheaper. In those periods, the chiller
charges the storage and meets the building cooling
load requirements simultaneously. Since the full
storage TES systems meet all the building cooling
loads during the day time, it will result in larger and
therefore more expensive chillers and storage units
compared to the partial storage systems. The full
storage TES systems are likely to be attractive under
the following conditions [7]: spikes in the peak load
curve are of short duration, time of use energy rates
based on short-duration peak periods, there are short
overlaps between peak loads and peak energy periods,
high peak demand charges apply, and some utility
companies offer incentives for using TES.
The partial storage TES system provides the best
mode for reducing demand charge and saving
electricity cost. Therefore it represents more that 50%
of the thermal storage installations worldwide.
However, these systems are not capable of shifting as
much load on a day as the full storage TES systems. In
the partial storage system, the chiller operates to meet
part of the cooling load demand and the rest is met by
the storage tank during the day time. Usually in this
system, the chiller is sized at a capacity smaller than
the design load. Partial storage TES systems can be
further classified based on the selected operation
strategies, load leveling, or demand limiting
operations. In a load leveling system the chiller
operates at full capacity for 24 hours of the design
day. When the building cooling load demand is less
than the chiller capacity, the excess cooling is stored
in the storage tank until the tank is full. When the load
exceeds the chiller capacity, the additional cooling is
supplied from the storage tank.
This paper presents a study on the use of TES system
in a lowland greenhouse for planting strawberry. In
Malaysia, Cameron Highland is one of the suitable
areas to plant strawberry in the open due to its suitable
ambient temperature. Strawberry grows healthy in a
temperature range of 17ºC - 20°C [9]. However due to
insufficient land area in Cameron Highland, building
greenhouses at a lowland area is seen as one of the
possible solutions to this issue. Most lowland areas in
Malaysia experiences temperature around 32°C all
year. Conventional air conditioning (AC) systems will
be required to maintain the greenhouse at the
temperature needed for planting strawberry. However,
such systems will consume a lot of electricity to
operate continuously. This will results in high
operating cost for the greenhouses. Thermal energy
storage (TES) systems can therefore be considered as
the way to achieve this goal since they are able to shift
cooling energy use to from peak time to non-peak
times. They can chill storage media such as water, ice,
or a phase-change material during the periods of low
cooling demand for use later to meet the air-
conditioning loads.
The goal of this study is to select a suitable TES
system that can help save the cost of electricity needed
to meet the cooling load requirement of the
greenhouse, based on a 24 hours operating duration.
The cooling load requirement of the greenhouse was
estimated to determine the capacity and size of the
TES systems. This was done based on the highest total
annual cooling load. Three TES system operating
arrangements were considered in this study: TES full
storage combined with AC systems, TES full storage
and TES partial storage. The operating cost of the
these TES systems were compared in term of the total
amount of electricity usage during a 24 hours
operation, based on the local electricity tariff.
2. METHODOLOGY
The effects of operation strategy on electricity
consumption of TES systems for strawberry
production were estimated for three different operating
strategies. These are TES full storage combined with a
conventional AC system, TES full storage and TES
partial storage. For the TES full storage combined
with AC systems, the TES system was designed to
meet all on-peak cooling loads, while the AC system
meets all the off-peak cooling loads. The TES full
storage was designed to meet all on-peak cooling
loads from storage and the TES partial storage was
designed to meet part of the cooling load requirement
from the storage and the other part directly from the
chiller during the on-peak period. The major
components of the TES system consists of an
evaporator, a condenser, a cooling tower, storage tank
and water pumps, as shown in Figure 1. The
evaporator is used to generate chilled water and later
stored in the storage tank. The chilled water is used to
cool the hypothetical green house by discharging it
through the secondary chilled water pump. The heat
from the green house is carried by the water to the
storage tank before it is removed in the evaporator.
Primary chilled water pump is used to transport the
water from the storage tank to the evaporator. The
condenser water pump directs cool water from the
cooling tower into the condenser to absorb heat from
the evaporator. The heat absorbed by the condenser is
rejected to the cooling tower by the cooling water and
rejected it to the surroundings. The TES system was
designed based on the specification given in Table 1.
The 5th IMAT, November 12 – 13th
2012
207
Figure 1: Schematic diagram of the TES system.
Table 1: Design specification for the TES system.
Parameter Value
Indoor temperature 20℃
Chilled water supply temperature 5℃
Chilled water return temperature 15℃
Storage water supply temperature 7°C
Storage water return temperature 17°C
Condenser water supply temperature 30℃
Condenser water return temperature 35℃
The hypothetical greenhouse is a single gable type
with a single-peaked roof, measuring 60 x 10 x 4 m
and is covered with polyethylene material for both the
walls and roof. The greenhouse is to be located in
Johor Bahru, Malaysia, in which the location is 1.3°
North and 103.7° East. Figure 2 illustrates the layout
of the strawberry planting arrangement in the
greenhouse.
Figure 2 Layout of strawberry plants (in mm) in
the hypothetical greenhouse.
The electricity cost for operating the TES systems was
estimated by first estimating the cooling load of the
hypothetical greenhouse for strawberry production.
This information is then used to determine the
capacity and size of the TES systems. The total
cooling load for the greenhouse consists of external as
well as internal thermal loads. The external thermal
load is due to heat transfer by conduction through the
walls, roof, floor and doors. The internal thermal loads
are due to the sensible and latent heat transfer from the
occupants, appliances and the strawberry plants. The
cooling load calculation for the hypothetical
greenhouse was performed using a TROPICA
software [10] that was developed based on a weighting
factor method. Figure 3 shows the flow chart of this
software.
Figure 3 Flow chart of the TROPICA software
[10]
The 5th IMAT, November 12 – 13th
2012
208
2.1 TES Full Storage Combined with AC
System
The TES full storage combined with AC systems
consist of a TES system (chiller and storage) and a
conventional AC system. The AC system includes a
cooling tower and water pumps. The chiller system
was selected based on the highest total cooling load
during the on-peak hours. The highest total cooling
load during on-peak occurs between the month of May
and June which is about 822 kW. The chiller operating
schedule was then determined to estimate the hourly
storage balance. It is done based on the total cooling
load during the on-peak hours and the cooling capacity
of the selected chiller system. The combined TES and
AC systems are designed to operate for 24 hours. The
conventional AC system was selected based on the
highest cooling load during the off-peak hours which
is about 47 kW. In this arrangement, the AC system
will operate to meet the cooling load demand of the
greenhouse during off-peak hours. The TES system
will be in a charging mode during this period. During
the on-peak hours, the AC system will be shut off and
the cooling load demand of the greenhouse is met by
the fully charged TES system.
2.2 TES Full Storage System
In this TES system arrangement, the chiller will
operate at its full capacity during the off-peak hours,
i.e. from 10:00 pm until 8:00 am, to charge the system
and at the same time meet the cooling loads demand of
the greenhouse. During the on-peak hours which is
from 9 am until 10 pm, the charged TES system will
be used to meet all cooling requirements by the
greenhouse. The chiller for this TES system was
selected based on the highest total cooling load in a
day, which is about 1056 kW. This highest cooling
load occurs between the month of May and June. The
full storage system is also designed for 24 hours
operation.
2.3 TES Partial Storage System
The TES partial storage system is designed to meet
part of the cooling load from its storage during the on-
peak hours. The other part is supplied directly from its
chiller system. The chiller system will be in the
charging mode when the cooling load requirement is
less than the output of the chiller. The chiller will be in
the discharging mode when the cooling load
requirement of the greenhouse is greater than the
output of the chiller. For this TES arrangement, the
chiller system was selected based on the highest total
cooling load during the on-peak hours, which is about
822 kW. This highest total cooling load occurs
between the month of May and June. This TES system
is also designed for the chiller to operate at full
capacity for 24 hours.
2.4 Operating Cost Analysis The total cost of electricity usage by the TES systems
depends on the power rating of the electric motor of
each equipment in the systems, the operating duration
of the equipments and the local electric tariff, during
both the on-peak and off-peak hours. Table 2 shows
the electric tariff of the on-peak and off-peak hours in
Johor Bahru, Malaysia.
Table 2: Electricity tariff in Johor Bahru, Malaysia
On-peak 8:00 am - 10:00 pm RM 0.312 /kWh
Off-peak 10:00 pm - 8:00 am RM 0.192 /kWh
Table 3 shows the power rating of the electric motor
for all the three TES system arrangements. The
equipment operating schedule during the off- and on-
peak hours is also shown in the table. The electric
motor ratings were obtained from the corresponding
manufacturers.
Table 3: Electric motor rating and operating schedule of the
TES systems Equipment Motor
Rating
(KW)
Off-Peak
Hours
On-
Peak
Hours
TES full storage with
AC systems:
Chiller
Cooling tower
Chilled water pump
Distribution water
pump
Condenser water pump
AC system
31.19
0.7456
0.75
0.75
1.1
19.3
9
0
9 0
9
0
9
10
0
14
0
0
TES full storage
system:
Chiller
Cooling tower
Chilled water pump
Distribution water
pump
Condenser water pump
37.95
1.12
0.75
0.75
1.1
9
0
9 0
9
1
9
0
14
0
TES partial storage
system:
Chiller
Cooling tower
Chilled water pump
Distribution water
pump
Condenser water pump
19.71
0.37
0.55
0.55
0.75
10
8
10 8
10
0
10
8
14
8
The 5th IMAT, November 12 – 13th
2012
209
3. RESULTS AND DISCUSSION
3.1 Cooling Loads of Hypothetical Green
House
Figure 4 shows the hourly cooling loads profile of the
hypothetical green house in a year. The profile of the
cooling load reflects the local weather data which
dictates the amount of sensible heat conduction into
the green house and the magnitude of solar heat gain.
From 10 am to 1 pm, conduction heat gain and solar
load increases, resulting in the increase of cooling
loads. The peak cooling loads occur at about 1 hour
past noon time. As the external heat gains drop past
the noon time, the cabin air temperature exhibits a
similar decreasing trend. The maximum cooling load
occurs between the month of January and February,
which is about 100 kW.
Figure 4: Hourly cooling load profile for the
hypothetical green house in a year.
3.2 Combined TES Full Storage and AC
Systems
Figure 5 shows the hourly cooling loads profile of the
hypothetical green house when the combined TES full
storage and AC systems are employed to meet the
cooling load demand. In this arrangement, the TES
system is designed to meet all the on-peak cooling
loads, while the AC system meets all the off-peak
cooling loads. As seen from the figure, the chiller of
the TES system is charged from 10 pm until 8 am
during the off-peak hours. During this period, the AC
system is used to meet the cooling load demand of the
green house. It can be seen that the chiller charging
capacity is close to 100 kW.
Figure 5: Hourly cooling load profile for
combined TES full storage and AC
systems.
The AC system is operating at a capacity of about 50
kW when the cooling load requirement of the green
house is about 20 kW. The AC system is running with
cooling capacity much higher than the cooling load
requirement. This is obviously not economical.
However, since this occurs during an off-peak period
in which the electric tariff is lower, a potential saving
of cooling cost can still be achieved. The figure also
shows that from 8 am until 10 pm, which is during the
on-peak hours where the electric tariff is higher, the
cooling load requirement of the green house is met by
the energy stored by the TES system. During this time,
the AC system is turned off to further save electricity
consumption. The same cycle will be repeated for the
next 24 hours period.
3.3 TES Full Storage System Only
Figure 6 shows the hourly cooling loads profile of the
green house when only the TES full storage system is
employed. In this case the TES system is designed to
solely meet all the cooling load demand, during both
the off-peak and on-peak period. As before, the TES
system is charged from 10 pm to 8 am, i.e. during the
off-peak period. The chiller capacity of this system is
now about 120 kW. The extra 20 kW of cooling
capacity is used to meet the cooling load requirement
of the green house.
The 5th IMAT, November 12 – 13th
2012
210
Figure 6: Hourly cooling load profile for TES full
storage system.
3.4 TES Partial Storage System
Figure 7 shows the hourly cooling loads profile of the
green house when the TES partial storage system is
used. With this system, the TES is charged from 10
pm to 5 pm. The chiller runs at a capacity of about 61
kW during the charging period. This is the lowest
charging capacity compared to the previous
arrangements. In this arrangement, the chill water
produced during charging is directly used to meet the
cooling load demand of the green house. At same
time, the excess chill water is stored in the storage
tank. When the cooling load exceeds the chiller
capacity, the additional chill water will be discharged
from the storage tank. This happens from the period of
9 am to 4 pm which is the on-peak hours.
Figure 7: Hourly cooling load profiles for TES
partial storage system.
3.5 Estimation of Electricity Cost
An economic evaluation of the cooling systems
requires estimation of the annual operating cost of the
system. The operating costs are affected by several
factors such as cooling system capacities, the
operating time period and the usage of the
conventional AC system. Higher chiller capacity
obviously results in higher electricity consumption,
which increases the operating cost. Longer period of
time taken to charge the chiller will also increase the
electricity consumption. Therefore, the operating cost
will also increase. Whether the systems operate only
during off-peak or during both the off-peak and the
on-peak period, it will also has an impact on the
operating costs because the electric tariff is higher
during the on-peak period. The use of conventional
AC system will further increase the electricity
consumption and thus the operating cost. Table 4
shows the cost estimation of the electricity usage for
the three TES system operating arrangements as
described above.
Table 4: Cost estimation of various TES system operating
arrangements
Time/Cost
(RM)
Type of System
AC
System
Full
Storage
With AC
System
Full
Storage
Partial
Storage
A Day 235.27 98.66 75.13 96.83
A Month 7,058.10 2,959.80 2,253.90 2,904.90
A Year 84,697.20 35,517.60 27,046.80 34,858.80
As seen from Table 4, if the conventional AC system
is used to meet the cooling load demand of the
greenhouse, it would cost about RM 84,697 to operate
it in a year. It can also be seen that the operating cost
is significantly reduced when TES systems are
employed to meet the cooling load demand of the
greenhouse. Although the TES full storage system has
the highest of chiller charging capacity of 121.5 kW,
the system actually has the lowest annual operating
cost of about RM 27,047 compared with the other two
arrangements. This is because the system used
electricity only for charging and this is done during
the off-peak hours when the electrical tariff is low.
When the system is used to meet the cooling load
demand of the greenhouse, no electricity is consumed
by the system. Although the TES partial storage
system has the lowest chiller charging capacity of 61
kW, the system costs about RM 34,859 to run in a
year, which is much higher than the TES full storage
system. This is because this system does not only
operate during off-peak hour, but also during on-peak
hour, when electricity tariff is high. It is also seen
The 5th IMAT, November 12 – 13th
2012
211
from the table that the combination of TES full storage
system with AC system has a higher operating cost
among all TES systems. In this arrangement, the AC
system is used to meet the cooling load demand of the
greenhouse during off-peak hours, while the chiller is
charging the TES system. However, the AC system is
operating with capacity higher than what is required
by the greenhouse, thus consumes more electricity
than required. Although the off-peak hour electricity
tariff is low, the operating cost is high because the
power consumption of the AC system is generally
high.
Table 5 shows the possible amount annual saving on
the operating cost of the TES systems when compared
to the conventional AC system. It is seen that the TES
full storage system can give the highest cost saving of
about 68 % compared to the conventional AC system.
The TES partial storage system offers the second
highest cost saving of about 59 %, followed by the
TES full storage combined with the AC system, which
offers a cost saving of 58 %.
Table 5: Annual saving on the operating cost of the cooling
systems using TES compared to the conventional AC system.
System Saving (RM) Percentage (%)
TES Full Storage With
AC System 50,518.80 58.06
TES Full Storage 58,989.60 68.07
TES Partial Storage 51,177.60 58.84
4. CONCLUSION
This study investigates potential saving in electricity
cost to operate a hypothetical greenhouse for planting
strawberry in Johor Bahru, Malaysia. It was found that
if a conventional air-conditioning (AC) system is used
to meet the cooling load demand of the greenhouse the
annual operating cost in term of electricity usage
would amount to RM 84,697, based on the local off-
peak and on-peak hour electricity tariff. The use of
thermal energy storage (TES) system has a potential to
help save the operating cost of the greenhouse. Three
operating arrangements have been considered:
combine TES full storage with AC system, TES full
storage system and TES partial storage system.
Among these three arrangements, the TES full storage
was found to have the highest an annual cost saving of
about RM 58,990 compared to the cost of using the
conventional AC system alone. This represents about
68 % of annual operating cost saving, which is
considered very significant.
ACKNOWLEDGEMENT
The authors would like to acknowledge the supports
from Universiti Teknologi Malaysia and fund
provided by the Ministry of Higher Education,
Malaysia throughout this study under the ERGS Vot
No. 4L404.
REFERENCES
[1] Rakesh Khanal & Chengwang Lei, Solar chimney - A
passive strategy for natural ventilation, Energy and
Buildings 43 (2011) 1811–1819.
[2] P.F. Linden, The fluid mechanics of natural ventilation,
Annual Review on Fluid Mechanics 31 (1999) 201–
238.
[3] K.-S. Nikas, N. Nikolopoulos, & A. Nikolopoulos,
Numerical study of a naturally cross-ventilated
building, Energy and Buildings 42 (2010) 422–434.
[4] N.K. Bansal, R. Mathur, M.S. Bhandari, Solar chimney
for enhanced stack ventilation, Building and
Environment 28 (3) (1993) 373–377.
[5] G. Gan, Simulation of buoyancy-induced flow in open
cavities for natural ventilation, Energy and Buildings 38
(5) (2006) 410–420.
[6] D.J. Harris, N. Helwig, Solar chimney and building
ventilation, Applied Energy 84 (2) (2007) 135–146.
[7] A. Dimoudi, Solar chimneys in buildings – the state of
the art, Advances in Building Energy Research 3
(2009) 21–44.
[8] Pacific Northwest Laboratory, Thermal energy storage
for space cooling. Federal Energy Management
Program (FEMP), Federal Technology Alert (FTA),
Richland, Washington, 2000.
[9] Anita Sønsteby and Ola M. Heide (2006). Dormancy
Relations and Flowering of the Strawberry Cultivars
Korona and Elsanta as Influenced by Photoperiod and
Temperature. Scientia Horticulturae. Volume 110, Issue
1: 57-67.
[10] Mohd Yusoff Senawi (2000). Development of a
Building Energy Analysis Package and its Application
to Analysis of Cool Thermal Energy Storage Systems.
Ph.D. Thesis. Universiti Teknologi Malaysia.
The 5th IMAT, November 12 – 13th
2012
212
Transient Modeling of a Lithium Bromide – Water Absorption
Chiller
Ang Lia, Wai Soong Loh
a, Kim Choon Ng
b
aDepartment of Mechanical Engineering,
National University of Singapore,
9 Engineering Drive 1, Singapore 117576,
bProfessor,
Department of Mechanical Engineering,
National University of Singapore,
9 Engineering Drive 1, Singapore 117576,
ABSTRACT
This article presents a thermodynamic framework
for a lithium bromide – water absorption chiller, in
which a transient model is developed to simulate
the operation process. Local energy and mass
balance within the main components like absorber,
regenerator, condenser, evaporator and solution
heat exchanger is respected to investigate the
behavior of the chiller. Experimental correlations
are used to predict heat transfer of the related
working fluids. The cooling water is set to typical
cooling tower conditions of tropical countries such
as Singapore. The coefficient of performance
(COP) is evaluated against a range of heat source
temperatures from 75oC to 100
oC. The results
indicate the operation conditions of the chiller at its
maximum COP is 95oC to 100
oC.
Keywords : Absorption chiller, lithium
bromide, entropy, modeling
1. INTRODUCTION
Low-grade heat driven and maintenance free are
two attractive features of absorption chillers. As
such, they are favored to produce cooling in
cogeneration and solar systems. This article
presents a thermodynamic framework for a lithium
bromide – water absorption chiller. A transient
model is developed to simulate the operation
process of main components like absorber,
regenerator, condenser, evaporator and solution
heat exchanger. The effect of heat source
temperature to the performance of the chiller is
evaluated.
2. MODELING OF ABSORPTION
COOLING CYCLE
2.1. Description of Operation Process
The transient modeling developed in this work is
based on YAZAKI WFC-900S single effect lithium
bromide – water absorption chiller. A schematic
presentation of the chiller is shown in Figure 1. The
chiller has a rated cooling capacity of 3 tons of
refrigeration (rton). It mainly comprises an
absorber, a regenerator, a condenser, an evaporator
and a solution heat exchanger. The first two
components are the reactors where the absorbent
and absorbate interact to perform thermal
compression. In the absorber, the water, also
known as the absorbate which plays a role of
refrigerant, is attracted by the concentrated LiBr
solution that holds a strong affinity to the former.
This absorption process produces heat as the water
changes its state from gaseous phase to absorbed
phase, and the solution becomes diluted. To keep
the absorption ability of the solution, cooling water
is supplied to carry away the heat produced and
maintain the absorber temperature. In the
regenerator, heat source is provided; the diluted
solution that comes from the absorber is warmed
up and saturated. Water molecules are released,
leading to an increase in the concentration of the
lithium bromide solution. This concentrated
solution flows back to the absorber, and the
absorption / desorption process continues. The
solution heat exchanger is located in between the
two reactors to effectively reuse the energy of two
solution streams, whereby the diluted flow is
preheated prior entering the generator and the
concentrated solution is pre-cooled before going
into the absorber. The water vapor released from
the saturation process of the regenerator flows to
the condenser and condenses to liquid phase. The
cooling water that branched out from the absorber
cooling water is used to remove the heat generated
in the condensation process. The cooling effect of
the chiller is provided by the condensate flowing
into the evaporator where the liquid vaporizes. The
evaporation process lowers down the incoming
chilled water which is circulated between the load
and the chiller. A U-tube is applied to connect the
condenser and the evaporator, and maintain the
pressure difference of the two components. The
absorption cooling cycle is completed when the
refrigerant vapor is absorbed by the concentrated
lithium bromide solution in the absorber, and next
IMAT-UI 035
The 5th IMAT, November 12 – 13th
2012
213
cycle of cooling starts. As the water is used as the
refrigerant, the chiller components operate in sub-
atmospheric conditions.
Figure 1: Schematic presentation of a single-effect
LiBr-water absorption chiller
2.2. Transient Modeling
The transient model developed below simulates the
operation process of the absorption chiller. The key
of the model are the energy and mass balance
equations of each major component which are
differential form and to be solved simultaneously
with respect to time. This computation is
performed in FORTRAN 90 Developer Studio
software with the help of DIVPAG subroutine of
IMSL Library. The equations are solved iteratively
by Gear‘s BDF method, employing the starting
conditions of the chiller components to initialize
the computation. The results that converge in the
tolerance of 10-6
are accepted. Properties of the
lithium bromide solution are calculated using Yuan
and Herold [1] correlation functions except the
vapor pressure for which McNeely [2] correlation
is used. The IAPWS Formulations [3-5] are
implemented for the properties of water. Prior to
model the system, some assumptions are made. 1)
Each component of the chiller is properly insulated.
Heat loss by conduction to the insulation and
radiation to the surrounding are ignored. 2) The
heat exchanger material inside each component has
the same temperature as the content of the
component. 3) The content of the two reactors are
well mixed.
2.2.1. Absorber
The thermodynamic properties of the lithium
bromide solution are dependent on temperature,
pressure as well as concentration. Defining as the
mass ratio of the salt to the solution, the
concentration, X, is expressed as follows:
, /
/
/
100%LiBr ds cs
ds cs
ds cs
mX
m (1)
where the subscription ds and cs denote diluted
solution and concentrated solution, respectively.
In the absorber, water vapor is continuously
absorbed by the solution. To achieve this, the
solution must remain sub-cooled throughout the
process and even upon leaving the absorber [6].
Choose a control volume to be the space taken up
by the solution and heat exchanger in the absorber,
the energy and mass conservation of the absorption
process is given as:
,
,
, ,
ab
hx sol ab
cs sol cs ab cs ds sol ab ds
ab
o i ve g e losscw ab
dTMCp MCp
dt
m h T X m h T X
Cp T T m h T Qm
(2)
And,
ds cs vem m m (3)
In Equation (2), the term on the left hand side
represents the sensible heat required to change the
temperature of the heat exchanger material as well
as the solution content in the absorber chamber. On
the other side, the first two terms denotes the
energy that is brought in or taken out by the
concentrated or diluted solution, respectively. The
third term is the heat removal by the cooling water,
whereas vapor energy flowing into the absorber is
given by the last term.
2.2.2. Regenerator
Attributing to the heat source, the content solution
in the regenerator boils in high temperature and
pressure at which the water vapor molecules are
released. As a result, both temperature and pressure
of the vapor are boosted as compared to its state
upon leaving the evaporator. Another use of vapor
is that it helps to pump the solution out of the
generator. This is achieved by configuring the
generator tube inner diameter to be approximately
equal to the size of bubbles. The mixture of the
bubble and liquid forms the slug flow regimen in
which each bubble lifts a small amount of liquid
upwards. A vapor liquid separator is place on top
of the generator tubes to split the two phases. Such
configuration is known as the ‗air bubble pump‘[7].
Similar to the absorber, considering the heat
exchanger and the solution enclosed space in the
regenerator as the control volume, the energy
balance and mass conservation equation can be
expressed as:
The 5th IMAT, November 12 – 13th
2012
214
,, ,
,
rg
hx sol rg
ds sol ds rg ds cs sol rg cs
rg
o i vc g rg c losshw
dTMCp MCp
dt
m h T X m h T X
mCp T T m h T P Q
(4)
And,
cs ds vcm m m (5)
The amount of water vapor desorbed from the
lithium bromide solution can be obtained from the
equation below that is modified from a correlation
given by Chua et al. [8]. The modification takes
into account the influence of regenerator pressure
to the desorption rate.
w w
w rg rg
rg rg
m mdm dT dP
T P
(6)
Where,
And,
Where θrg is 0 if the solution is in sub-cooled state,
and otherwise, it assumes to be 1. Tdp represents the
dew point temperature, and AD and BD denotes
Duhring constants founded in McNeely [2]
correlation.
2.2.3. Condenser
In the condenser, thermal phenomena involved are
phase change of water and heat transfer to the
cooling water. The conservation of energy can be
described as:
,
,c
vc g rg chx f c
c
fc f c o i losscw c
dTMCp MCp m h T P
dt
m h T mCp T T Q
(7)
The terms on the left hand side of the equation
again represent the sensible heat required to change
the temperature of the heat exchanger material as
well as the liquid content remained in the
condenser chamber. The first two terms on the right
hand side, on the other hand, denotes the energy of
vapor incoming to the condenser and energy of
liquid leaving the chamber. The last term of the
equation gives the heat removal by the cooling
water.
2.2.4. Evaporator
Assume the control volume of the evaporator is the
space enclosed by the U-tube that connects the
condenser with the evaporator, the evaporator heat
exchanger and liquid water. Similar to the
condenser, the evaporator can be modeled as:
e
fc f chx f e
e
ve g e i o losschi
dTMCp MCp m h T
dt
m h T Cp T T Qm
(8)
2.2.5. Solution Heat Exchanger
In the solution heat exchanger, the decrease of
thermal energy of high temperature concentrated
solution is the increase of the low temperature
diluted solution. It can be modeling as following.
, ,rg cs ab ds rg abcs dsCp T T Cp T Tm m (9)
The above equation is solved by ε-NTU method to
determine the heat exchanger output, Tcs,ab and
Tds,rg, with Trg and Tab simultaneously solved by
Equation (2) and (4) as inputs.
2.2.6. External water sources
The above energy balance and mass conservation
equations are solved with coupling to the
calculation of external water sources. For hot
water, cooling water and chilled water of respective
component, the outlet temperature is determined
from UA-LMTD method, and is given by:
, / / , / / / / / , / /
/ / /
/ /
1 exp
o hw cw chi i hw cw chi rg ab c e i hw cw chi
rg ab c e
hw cw chi
T T T T
UA
mCp
(10)
The overall heat transfer coefficient of regenerator,
absorber, condenser and evaporator heat exchanger
U is discussed in the following section.
2.2.7. Heat transfer coefficients
The overall heat transfer coefficient encountered in
the modeling of each component, like the absorber,
regenerator, condenser and evaporator are
determined by the following equation:
1
, ,
ln1
2
i o i i o
t i t o
D D D D DU
h k h
(11)
The local heat transfer coefficient of the tube side,
ht,i, where external water sources for each of the
component heat exchanger, such as cooling water
2
100
- 273.15
w LiBr
rg ds
rg
D D
dp rg
ds ds
m m
T X
dA dBT P
dX dX
(6a
)
2
100
273.15
w LiBr D
rgrg ds
dp
rg
D D
dp rg
ds ds
m m A
dPP X
dT
dA dBT P
dX dX
(6b)
The 5th IMAT, November 12 – 13th
2012
215
for the absorber and condenser, hot water for the
regenerator as well as chilled water for the
evaporator is determined by Dittus-Boelter
correlation for pipe water. For the shall side local
heat transfer coefficients, ht,o, the following
correlations are used. 1) The absorber is
constructed in a way that the lithium bromide
solution falls from the top of the tube bundles. Park
[9] correlation is applied in this case. 2) Pooling
boiling of the LiBr solution is taking place in the
regenerator where Charters [10] experimental
results are used to predict the local heat transfer
coefficient. 3) Nusselt condensation correlation for
horizontal tube bundles counts for the phase change
phenomenon in the condenser. 4) Assuming the
salinity to be zero, the local heat transfer
coefficient of refrigerant water in the evaporator is
calculated by extrapolating from Shahzad‘s [11]
correlation for sub-atmospheric pressures. The heat
transfer area and key simulation parameters are
given in Table 1.
Table 1: Key simulation parameters.
Heat transfer area
Aab 2.98 m2 Ae 1.95 m
2
Arg 2.95 m2 Ac 0.96 m
2
(UA)hx 0.734 kW/k
Flow rate of external water sources
mhw
0.68 kg/s ,
mcw c
0.54 kg/s
,
mcw ab
0.79 kg/s mchi
0.5 kg/s
Initial conditions
Trg,ini = Ti,hw Tcw,ab,ini = Ti,cw
Te,ini = Ti,chi Tcw,c,ini = Ti,cw
2.2.8. Evaluation parameters
The absorption chiller is evaluated by its cooling
capacity QC and the Coefficient of Performance
(COP). Both parameters are defined as follows:
C i o chi
Q Cp T Tm (12)
C
H
QCOP
Q
(13)
Where QH is the total heat input to the system and
calculated from the hot water behavior.
H i o hw
Q Cp T Tm (14)
4. RESULTS AND DISCUSSION Figure 2 shows simulated temperature profile of
chiller major components and water sources during
operation at hot water inlet temperature Ti,hw 90oC,
cooling water inlet temperature Ti,cw 29.5oC and
chilled water requirement To,chi of 9oC. The heat
loss to the surrounding encountered in regenerator
and evaporator is assumed to be 10% of heat input
and cooling effect, respectively. The same quantity
for absorber and condenser is set at 2% of their
total heat rejection. The absorption chiller
experiences a starting period and stabilizes
afterwards. In the formar period, regenerator
content is sub-cooled with no water molecules
being desorbed. Cooling takes into effect when
saturation of concentrated LiBr solution is reached.
In addition, the largest temperature difference of
chiller components and respective external water
sources happens in the regenerator. A deviation of
8.6 o
C is found between regenerator temperature
and outlet temperature of heat source. This is well
explained by the inefficient boiling heat transfer of
lithium bromide solution.
Figure 2: Simulated temperature profile of
chiller components and water
sources during operation at hot water
inlet temperature Ti,hw 90oC, cooling
water inlet temperature Ti,cw 29.5oC
and chilled water requirement To,chi
of 9oC
The effect of heat source inlet temperature Ti,hw to
cooling capacity and overall thermal input is
displayed in Figure 3. The results are calculated at
cooling water inlet temperature Ti,cw 29.5oC and
chilled water requirement To,chi of 9oC. Both
computed quantities unveils linear upward trend
with respect to the temperature of hot water input,
while the amount of thermal input increases faster.
The phenomena are due to the fact that the raising
of heat source temperature results in an increase of
temperature of regenerator content and hence the
amount of refrigerant water released. The addition
of quantity in the refrigerant circulation promotes
the need to the thermal energy, and delivers more
cooling at the same time.
Figure 3: Effect of heat source temperature to
cooling capacity and thermal input at
cooling water inlet temperature Ti,cw
29.5oC and chilled water requirement
To,chi of 9oC
The 5th IMAT, November 12 – 13th
2012
216
Despite a slightly faster increase of thermal input
than the cooling effect with respect to heat source
temperature, it is still wealth of operating the
chiller at the higher end. As implied from Figure 4,
in conditions as above, the chiller‘s coefficient of
performance is improved by higher heat source
temperature. The COP reaches its maximum 0.64 at
95oC to 100
oC, while at lower end of 75
oC the
useful effect is only equivalent to 48% of the total
input. The relation, however, is not linear. The
potential of improvement drops as hot water
temperature increases. The results indicate that at
typical cooling tower conditions (29.5oC) of
tropical countries, the higher end heat source
temperature produces cooling most efficiently for
the lithium bromide – water absorption chiller
presented in this work.
Figure 4: Effect of heat source temperature to the
coefficient of performance (COP) at
cooling water inlet temperature Ti,cw
29.5oC and chilled water requirement
To,chi of 9oC.
5. CONCLUSION
In the current work, a transient model was
developed for a lithium bromide – water absorption
chiller to simulate its operation process. The
cooling capacity, heat input and the COP was
evaluated with respect to heat source temperature.
The results shows that at typical cooling tower
conditions of tropical countries, the chiller operates
at its best COP the hot water temperature of 95oC
to 100oC.
NOMENCLATURE Alphabets Description
A Area, m2
AD Duhring constant
BD Duhring constant
COP Coefficient of performance
Cp Specific heat capacity, kJ/kg.K
D Diameter, m
h Specific enthalpy, kJ/kg
ht Local heat transfer coefficient,
kW/m2.K
k Thermal conductivity, kW/m.K
LiBr Lithium bromide solution
M Mass, kg
m Mass flow rate, kg/s
P Pressure, Pa
Qh Total heat input, kW
Qc Cooling capacity, kW
rton Ton of refrigeration
T Temperature, K
t Time, s
U Overall heat transfer coefficient,
kW/m2.K
X Concentration, wt%
θ State indicator in Eqn (6a) and (6b)
Subscripts Description
ab Absorber
c Condenser
chi Chilled water
cs Concentrated solution
cw Cooling water
dp Dew point
ds Diluted solution
e Evaporator
hw Hot water
i Inlet
ini Initial conditions
loss Heat dissipation
o Outlet
rg Regenerator
sol Lithium bromide solution
REFERENCES
[1] Z. Yuan and K. E. Herold, "Thermodynamic
properties of aqueous lithium bromide using a
multiproperty free energy correlation," HVAC and R
Research, vol. 11, pp. 377-393, 2005.
[2] L. A. McNeely, "Thermodynamic properties of
aqueous solutions of lithium bromide," ASHRAE
Trans vol. 85, pp. 413-434, 1979.
[3] W. Wagner, J. R. Cooper, A. Dittmann, J. Kijima,
H. J. Kretzschmar, A. Kruse, R. Mareš, K. Oguchi,
H. Sato, I. Stöcker, O. Šifner, Y. Takaishi, I.
Tanishita, J. Trübenbach, and T. Willkommen, "The
IAPWS industrial formulation 1997 for the
thermodynamic properties of water and steam,"
Journal of Engineering for Gas Turbines and
Power, vol. 122, pp. 150-180, 2000.
[4] IAPWS. Release on the IAPWS Formulation 2008
for the Viscosity of Ordinary Water Substance
[Online]. Available: http://www.iapws.org
[5] IAPWS. Release on the IAPWS Formulation 2011
for the Thermodynamic Conductivity of Ordinar
Water Substance [Online]. Available:
http://www.iapws.org
[6] M. J. Kirby and H. Perez-Blanco, "Design model for
horizontal tube water/lithium bromide absorbers,"
1994, pp. 1-10.
[7] H.I. Abu-Mulaweh, D.W.Mueller, B.Wegmann,
K.Speith, and B. Beohne, "Design of a Bubble
The 5th IMAT, November 12 – 13th
2012
217
Pump Cooling System Demonstration Unit," Int. J.
of Thermal & Environmental Engineering, vol. 2,
pp. 1-8, 2011.
[8] H. T. Chua, H. K. Toh, A. Malek, K. C. Ng, and K.
Srinivasan, "A general thermodynamic framework
for understanding the behaviour of absorption
chillers," International Journal of Refrigeration,
vol. 23, pp. 491-507, 2000.
[9] C. W. Park, S. S. Kim, H. C. Cho, and Y. T. Kang,
"Experimental correlation of falling film absorption
heat transfer in micro-scale hatched tubes,"
International Journal of Refrigeration, vol. 26, pp.
758-763, 2003.
[10] W. W. S. Charters, V. R. Megler, W. D. Chen, and
Y. F. Wang, "Atmospheric and sub-atmospheric
boiling of H2O and LiBr/H2O solutions,"
International Journal of Refrigeration, vol. 5, pp.
107-114, 1982.
[11] M. W. Shahzad, A. Myat, W. G. Chun, and K. C.
Ng, "Bubble-assisted film evaporation correlation
for saline water at sub-atmospheric pressures in
horizontal-tube evaporator," Applied Thermal
Engineering, vol. 50, pp. 670-676, 2013.
The 5th IMAT, November 12 – 13th
2012
218
Characteristics of Sea-water Ice Slurry for Cooling of Fish
A.S. Pamitrana, M. Novviali
b, H.D. Ardiansyah
b
aDepartment of Mechanical Engineering,
Universitas Indonesia, Kampus UI Depok 16424, Indonesia
Tel : (021) 7270032, Fax : (021) 7270033
E-mail : [email protected] bGraduate Student, Department of Mechanical Engineering,
Universitas Indonesia, Kampus UI Depok 16424, Indonesia
ABSTRACT
A more effective of cooling method is necessary
for fish storage to get high quality and long
freshness of fish. Ice block is not sufficient for fish
storage because of its hard-solid surface that can
damage the fish. Moreover for some remote area it
is difficult to find ice block in good time with
reasonable/low price. One solution for this problem
is the using of sea-water ice slurry for fish cooling.
Ice slurry is formed when the sea-water
temperature goes down to its freezing point, when
the early nucleation is formed. At this moment
there is a chemical potential againts the saturation
condition. Crystal ice can be formed when
chemical equilibrium is occured. The purpose of
this present study is to observe the characteristics
of ice slurry generation using scraper blade
evaportor and orbital rod evaporator. The
experiment is done under some experimental
conditions.
Keywords : Ice slurry, cooling, sea-water, salinity,
evaporator
1. INTRODUCTION Indonesia is an archipelago nation and is located at
tropical area where many kind fish species come
and growth. It has total area of Exclusive Economic
Zone of around 3 million km2, coastal line length of
around 104,000 km, number of fisherman of
around 1.7 million people who can supply fish of
around 9 million Ton in 2010. However, in fact this
figure is not linear with the fisherman welfare. This
situation could be caused by limited quality and
quantity of equipments used for fishing and
storaging the fish. Most of Indonesian fishermans
still use block ice for fish storaging. This way is not
sufficient for the fish because the block ice can
damage the fish and then the value of fish become
lower. Moreover for some remote area in Indonesia
it is difficult to find block ice with reasonable
price. Using outboard engine boat, or using no-
engine boat, traditional fisherman can only do
fishing around coatline, and can only effectively
work around 7-9 months a year. This situation
makes low productivity in fishing. One effective
way for cooling is replacing the ice block cooling
by ice slurry. Because sea-water ice slurry content
natural preservative, it is good for fish storaging
after fishing. This study is devoted to observe the
characteristics of sea-water ice slurry under
experimental conditions variation of salinity, room
tempature, shaft rpm, and sea-water volume in two
developed ice slurry generator types of scraper
blades evaporator and orbital rod evaporator.
E. Stamatiou et al., 2005, defines ice slurry consists
of water solution and ice crystal. Another definition
by Peter W Egolf et al., 2003, ice slurry is an ice
particle with average diameter equal as or less than
1 mm. Ice slurry forming consist of three steps of
supersaturation, nucleation, and propagation.
Supersaturation is when the freezing point
temperature of the fluid is reached. Lower than the
freezing temperature, nucleation is then formed.
Propagation is the phase when the ice crystal
formed.
2. EXPERIMENT AND PROCEDURE
The experiment is done in Refrigeration
Laboratory, Faculty of Engineering University of
Indonesia. As well, under Community Social
Responsibility project supported by the University
of Indonesia, one developed ice slurry generator is
implemented in fisherman community, in
Balongan, Indramayu, Indonesia. The experimental
apparatus is illustrated schematically in Figure 1. It
consists of two systems viz. ice slurry generator
system and cooling system. In ice slurry generator,
sea-water rejects some heat to refrigerant in the
evaporator to generate ice slurry. The present study
develops two kinds of evaporator viz. orbital rod
evaporator and scraper blade evaporator. The
cooling system consists of compressor, condenser,
liquid receiver, filter dryer, sight glass, expansion
valve, and accumulator. The developed test
apparatus is shown in Figure 2.
Ice slurry generator with orbital rod evaporator has
a rod scrapper inside the evaporator. The rod
rotates on the shaft line to scrap the formed ice
slurry on the evaporator inner surface. The present
ice slurry generator with orbital rod evaporator has
63.5 mm inner diameter, 76.2 mm outer diameter,
and 1500 mm length. Figure 3 illustrates developed
shaft of the orbital rod evaporator. Many industries
use scraper blade evaporator because it produces
more ice fraction than others system, as mentioned
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Figure 1: Schematic experimental apparatus
Figure 2: Developed ice slurry generators
Figure 3: Shaft of the orbital rod evaporator (mm) Figure 4: Shaft of the scraper blade evaporator (mm)
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by E. Stamatioua et al. 2005 and T. A. Mouneer et
al.2011. The present developed ice slurry generator
with scraper blade evaporator has 2600 mm
evaporator diameter, 2800 mm total length of
evaporator pipe with 9.525 mm evaporator pipe
diameter. Shaft of the scraper blade evaporator is
illustrated in Figure 4.
Temperature of the ice slurry is measured by
thermocouples and recorded by data aquisition.
Some properties of ice, as follow, are calculated by
referring A. Melinder, 2010.
(1)
(2)
(3)
The ice fraction is calculated by referring Jean-
Pierre Be´de´carrats et al., 2009, and Cecilia Hägg,
2005.
(4)
D.G. Thomas, 1965, and Jacques Guilpart et al.,
2006 use the following equation to calculate
viscosity.
(5)
Enthalpy is calculated by referring T. Kousksou et
al., 2010.
(6)
By referring Taret, 1940, the thermal conductivity
of ice slurry is calculated.
(7)
Parameter of COP (Coefficient of Performance) is
obtained using the following equation.
(8)
The experiment was runned under experimental
condition shown in Table 1.
3. RESULTS AND ANALYSIS
Figure 5 shows ice slurry formation for salinity 26
ppt. The temperature decreases in supersaturation
process from point A to point B. Ice fraction is
initially formed at point B. The freezing point of
sea-water is lower than of water due to its higher
salinity. Referring Melinder et al., 2008, point B is
starting point of nucleation. Ice slurry is initially
Table 1: Experimental condition
Salinity 18, 22, 26, 30 ppt
Shaft rpm 70, 90, 110, 130 rpm
Room temperature 22, 26, 30, 34ºC
Volume 3, 3.5, 4, 4.5 liter
Figure 5: Ice slurry formation
Figure 6: Effect of salinity on ice slurry
formation
formed when the temperature reach freezing point.
Water in the solution is partially freezed, forming
ice fraction.
The effect of salinity on ice slurry formation for
evaporative heat of 2.947 kW is depected in Figure
6. The salinity of sea-water in Indonesia is around
30 ppt. The results show that higher salinity needs
longer time and lower temperature for ice fraction
formation. It means that sea-water with higher
salinity takes more energy for ice slurry formation.
As well, it means that sea-water with higher
salinity can absorb more heat from fish at low
temperature. Figure 6 shows freezing temperatures
for salinities of 18, 22, 26 and 30 ppt are -1.29ºC, -
1.588ºC, -1.667ºC and -1.94ºC, respectively.
In order to validate the measurement data of
freezing temperature, the results are compared to a
reference as shown in Figure 7 with error of around
40%.
Diameter of ice fraction is obtained using IMAGEJ
software as shown in Figure 8. The effect of
Time (s)
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salinity on ice fraction diameter is shown in Table
2. At the same experimental condition, higher
Figure 7: Freezing temperature comparison
Figure 8: Diameter of ice fraction
Figure 9: Salinity effect on ice slurry
formation
Table 2: Diameter of ice slurry
Figure 10: Freezing temperature comparison
salinity results in smaller ice fraction diameter. The
effect of salinity on ice slurry formation is depected
in Figure 9, as well.
Figure 10 shows the enthalpy of ice slurry is
summation of ice enthalpy and sea-water enthalpy.
For salinity of 18 ppt, the ice slurry enthalpy is low
because it has more ice fraction than for the higher
salinity. The effect of cooling capacity on enthalpy
is shown in Figure 10, as well. The enthalpy is
higher for the lower cooling capacity due to
existing of ice fraction. Higher ice fraction may
results in higher pressure drop in ice slurry flow,
therefore it should be avoided.
4. CONCLUSION
The freezing point of sea-water is lower than of
water due to its higher salinity. Higher salinity
needs longer time and lower temperature for ice
fraction formation. Higher salinity results in
smaller ice fraction diameter and higher enthalpy.
ACKNOWLEDGMENT
The work described in this paper was supported by
grants of Hibah Madya 2012 from DRPM
Universitas Indonesia and Hibah Pengabdian
Masyarakat (Community Engagement Grant) 2012
from DRPM Universitas Indonesia.
REFERENCES
[1] E. Stamatioua, J.W. Meewisseb, M.
Kawajia.2004. Ice slurry generation involving
moving parts.International Journal of
Refrigeration 28 (2005) 60–72
[2] Peter W. Egolf, Michael Kauffeld. 2004. From
physical properties of ice slurries to industrial
ice slurry applications.International journal of
refrigeration 33 (2010) 1491-1505
[3] A°. Melinder*, Properties and other aspects of
aqueous solutions used for single phase and ice
Salinity (ppt)
QEvap 1.718 kW
QEvap 2.947 kW
Cooling time 1 hour 3 minutes
Salinity (ppt)
Ice
fra
ctio
n (
%)
*Cooling time 48 minutes
Cooling time 1 hour 3 minutes
Salinity (ppt)
En
tha
lpy
(k
J/k
g)
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2012
223
slurry applications, international journal of
refrigeration 33 (2010 ) 1506e1512
[4] T.A. Mouneer *, M.S. El-Morsi, M.A. Nosier,
N.A. Mahmoud, Heat transfer performance of
a newly developed ice slurry generator: A
comparative study
[5] Jean-Pierre, Thermal and hydrodynamic
considerations of ice slurry in heat exchangers,
Be´de´carrats*, Franc¸oise Strub, Christophe
Peuvrel
[6] Cecilia Hägg ,2005,Ice Slurry as Secondary
Fluid in Refrigeration Systems, Fundamentals
and Applications in Supermarkets,School of
Industrial Engineering and Management,KTH
[7] D.G. Thomas, Transport characteristics of
suspension. VIII. A note on the viscosity of
Newtonian suspensions of uniform spherical
particles, Journal of Colloid Science 20 (1965)
267–277.
[8] Jacques Guilpart*,1, Evangelos Stamatiou,
Anthony Delahaye, Laurence Fournaison
Comparison of the performance of different ice
slurry types depending on the application
temperature, International Journal of
Refrigeration 29 (2006) 781–788
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FLUID FLOW CHARACTERISTIC OF ROUNDED-SHAPE FPSO AND
LNG CARRIER DURING OFFLOADING
Mufti F. M., Jaswar, A. Priyanto, and Efi Afrizal
Department of Marine Technology, Faculty of Mechanical Engineering
Universiti Teknologi Malaysia
Skudai, 81310 Johor, E-mail : [email protected]
ABSTRACT
The design concept of FPSO begins with a shape like a ship.
Nowadays some researcher proposed a different shape for
design concept of FPSO called rounded-shape. The rounded-
shape has advantage in the seakeeping and the construction.
One of the common activity performed by the FPSO is
transfer process of the product which is called offloading
process. There are a wide range of phenomena that occur
during offloading one of which is fluid flow. Fluid flow that
occurs in structure will be different depend on the several
factors, such as shape of structure, Reynolds numbers and
Froude numbers. From the characteristic of fluid flow that
occurs in the structure can be determined the effect of the
fluid flow on the structure. Current research discuss the
concept design of fluid flow around a round-shaped FPSO
LNG interacted with a LNG carrier during side-by-side
offloading condition by CFD method based on the Reynolds
Average Navier-Stokes (RANS) equations.
Keywords: Fluid flow; Offloading; RANS; CFD.
1. INTRODUCTION
The development of oil and gas exploration industry,
particularly exploration in the deep ocean, in recent decades
has increased. Correspondingly with the increase in oil and
gas exploration in deep ocean waters, facilities and
infrastructures required to support exploration activities are
also necessary. One of the supporting facilities in the
activities of oil and gas exploration in deep water is the FPSO
LNG (Floating Production Storage Offloading Liquid Natural
Gas) and LNG Carriers. The FPSO LNG is enabled to
receive and process gas products from the field and save the
LNG into Cargo Containment System (CCS) tank before the
LNG transferred to the LNG Carrier to be distributed to the
market or destination.
Development of design concepts to the FPSO start with
a shape like a ship. The FPSO is designed to accommodate
the construction of the module production process to be the
product oil. The new concept design of the FPSO is currently
a cylinder shaped being developed by Sevan Marine [1] and
SSP Offshore [2]. Wang, Zhang, and Liu [3] study new
design concept of FPSO which is propose a non-ship-shape
FPSO called inverted fillet quadrangular frustum pyramid-
shaped FPSO (IQFP). Another design concept of circular
FPSO for Arctic Deepwater proposed by Srinivasan and
Sreedhar [4].
LNG products that have been produced on the FPSO LNG
will be transferred to the LNG Carrier. The transfer process is
called offloading process, which is a common activity
performed by the FPSO LNG and the LNG Carrier. There are
two methods of offloading process for oil and gas transfer
which is tandem and side-by-side. Tandem configuration is
done when a LNG carrier is moored in tandem with the
FPSO LNG. The hoses or hawsers are connected between the
stern off-loading stations on FPSO LNG to the cross over
manifold of the LNG tanker. While, the side-by-side
configuration is done by moored the LNG Carrier parallel
with the FPSO LNG and offloading is carried out via a
flexible hose between the cross over manifold of the FPSO
LNG and LNG Carrier as shown in Figure 1.
Figure 1. Side-by-side configuration [1]
Environmental loads generally occur in the structure at time
of FPSO LNG and a LNG tanker offloading conditions.
These environmental loads are classified into three types
which area wind, wave and currents load. Wind load exerts a
force on the part of the structure exposed to the air. Wave
load and current load exerts force on the part of the structure
exposed to the water. These three loads will determine
the external forces on structures, stability, and motion of
floating structures and patterns of fluid flow.
Fluid flow around a structure can significantly alter the
structure‘s loading characteristics. Waves and currents is a
fluid that has the most impact on the structure. Influence of
fluid flow generated from waves and currents on floating
structure is seen in the form of motion and
pressure distribution. These motion and pressure distribution
that occurs in the structure depend on the characteristic of the
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fluid flow. Characteristic of the fluid flow is affected by
several factors such as shape of the structure and Froude
number.
The study of fluid flow around hull in side-by-side offloading
condition have been performed by Arslan, Pettersen, and
Andersson [5]. The calculations of three dimensional (3D)
unsteady cross flow past a pair of ship sections in close
proximity and behavior of the vortex-shedding around the
two bluff bodies is investigated numerically by using the
software FLUENT.
The studies of fluid flow around hull of rounded-shape FPSO
has been performed by Lamport and Josefsson [6]. The
studies is shows the current induced velocity fields on the
leeward side of the vessel between round-shaped versus a
traditional ship-shaped FPSO.
Current research is discuss on analysis of flow around a
round-shaped FPSO LNG interacted with a LNG carrier
during offloading. The fluid flow characteristic around the
round-shaped FPSO LNG and LNG carrier during offloading
conducted using Computational Fluid Dynamics (CFD) based
on the governing Reynolds Average Navier-Stokes (RANS)
equations.
2. LITERATURE REVIEW
Studies of the fluid flow on the hull of ship have been
performed by many researchers in the last decade. For the
viscous flow around ship hull field, in 2002 Zhang, Zhao, and
Li [7] performed numerical simulation free surface flow
around Wigley hull. Numerical simulation using RANS
method and SST k-w turbulence model was performed by
Zhao, F., Zhang, Z.-R [8] with a complex modern ship model
DTMB 5415. In 2005, Schweighofer et. al. [9] focuses on the
applicability of different RANS methods to full-scale viscous
flow computations. Alexe [10] study about the effects of
dimensional and movement parameters of the ship on the
pressure distributions in surrounding sea water. Zhao, Zhu,
and Zhang [11] study the flow around Wigley and DTMB
5415 hull using RANS, Visonneau [12] predicting the full-
scale viscous flow field around a ship including the
evaluation of the free surface, the wake field, the
hull/propeller interaction, the resistance and the power.
In 2006, Kinnas, Yu, and Vinayan [13] study the unsteady
viscous flow over the bilge keels of an FPSO hull subject to
roll motions. Wang,Zou,and Tian [14] study the viscous
flow field around a KVLCC2 model moving obliquely in
shallow water using a general purpose computational fluid
dynamics (CFD) package FLUENT. Wang et al [15] study
the wake field in viscous flow and resistance prediction of a
full ship-KVLCC2M by using FLUENT. For coupling the 3D
incompressible RANS equations with level set method was
performed by Wan , Shen, and Ma [16] with numerical
simulation. In 2011, Wackers et. al. [17] reviewed the surface
descretisation methods with different code.
For the turbulence flow around ship hull field, many
researchers have been studied. Kim [18] studies the three-
dimensional turbulent flow using RANS equations. In 2002,
Kim, Kim, and Van Suak [19] developed an efficient and
robust numerical method for turbulent flow calculation.
Ciortan et al [20] investigate the free surface incompressible
turbulent flow around the hull by the numerical solution of
the unsteady Navier-Stokes equations for slightly
compressible flows. Deng, Queutey, and Visonneau [21]
study the simulation of two appended hull configurations
using all hexahedral unstructured grids. The three-
dimensional turbulent flow around a Wigley hull using
slightly compressible flow formulation performed by Ciortan,
Wanderley, and Soares [22]. Lungu [23] presented a
methodology for computing the 3D turbulent free-surface
flow.
Ciortan, Soares, and Wanderley [24] study the turbulent and
laminar free-surface flow around ship hulls using slightly
compressible flow formulation. Ahmed, Fonfach, and Soares
[25] investigated the flow pattern around the DTMB 5415
hull at two speeds. In 2011, Ahmed [26] uses Volume of
Fluid method (VOF) to simulate the flow pattern around the
DTMB 5415 hull at two speeds. Ciortan, Wanderley, and
Soares [27] study the simulation of flow around a Wigley
hull using the slightly compressible flow formulation.
In 2006, Tahara et. al. [28] evaluates the computational fluid
dynamics (CFD) as a tool for hull form design along with
application of state-of-the-art technology in the flow
simulations. Two Reynolds-averaged Navier-Stokes (RANS)
equation solvers were employed, namely CFDShip-Iowa
version 4 and Flowpack version 2004e, for the towing and
self-propulsion cases, respectively. An accurate, efficient
algorithm for solving free surface flows around ship hulls
using a compressive advection discretization which maintains
a sharp free surface interface representation without relying
on a small time step [29].
The application of the Fluent code to the numerical
simulation of the free-surface flow around a model naval
ship; the DTMB 5415. Simulations were performed using
both a structured hexahedral mesh and an unstructured
tetrahedral mesh of lower resolution. The results show that
Fluent is able to accurately simulate the total ship resistance,
near-field wave shapes, and the velocity field in the propeller
plane [30].
Tahara et al [31] conducted research on high-speed multi
hull. Multi hull which is used in that research is catamaran
hull with forward speed. In 2011, Broglia, Zaghi and Di
Mascio [32] study about the simulations of the flow around a
high speed vessel in both catamaran and monohull are carried
out by the numerical solution of the Reynold averaged
Navier–Stokes (RANS) equations.
3. METHODOLOGY
Current research about fluid flow characterization of round-
shaped FPSO LNG and LNG Carrier during offloading. The
steps for the working of this research as follows (Figure 2.):
a) Data Collection
Data collection for environmental condition and dimension of
the rounded-shape FPSO LNG and LNG carrier using a
previous study.
b) Side-by-side Offloading Arrangement
The majority of planned offshore LNG transfer systems are
designed for side-by-side configuration loading procedures,
although offshore transfer operations are limited to
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significant waves heights of Hs = 3 m [33]. Safety distance
between a rounded-shape FPSO LNG and a LNG carrier in
side-by-side offloading conditions was from 5 [34] to 10
[33] meters.
c) Design RANS Code for CFD
The simulation code created in Fortran software, result of the
calculation exported to Visual Basic for visualization. The
code from the Fortran software and the environmental data
will be used to analyze. From the result of software, then
determine the characteristic of fluid flow around hull of
rounded-shape FPSO LNG and LNG carrier during
offloading. Visualize the fluid flow around hull of rounded-
shape FPSO LNG and LNG carrier during offloding by using
Visual Basic software.
START
Data Collection
RANSE Round-shaped FLNG
and LNG Carrier
Dimension
Input Dimension and
Environmental Condition Data
into RANS Code
Running the Software to get
fluid flow of Round-shaped
FLNG and LNG Carrier
Environmental
Condition
Governing Equations
of RANS
Design Code of RANS
Equation
Analysis the Out Put from
Software of Fluid Flow
Determine the
Characteristic of Fluid
Flow
Finish
Conclusion
Figure 2. Flowchart of Methodology
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4. ROUND-SHAPE FPSO
According to Paik and Thayamballi [35], the first
floating, production, storage and offloading vessel
(FPSO), Shell‘s Castellon, was installed in 1977.
Since the industry has seen a large and diverse suite of
different FPSO solutions, from converted tankers to
purpose-built barge shaped vessels. Until recently,
most FPSOs had one thing in common that the design
philosophy was based on classic ship-shaped vessels.
While a ship has beneficial characteristics for
transporting cargo from one location to another, with
great maneuverability and little water resistance, it‘s
slender and non-ax symmetrical shape presents major
disadvantages when permanently moored in one
location.
One such disadvantage is that ship-shape vessels must
be able to align themselves with predominate sea state
to minimize their motions and vessel stresses. The oil
and gas industry mitigated this problem by developing
turrets and swivels, which allowed the ship shaped
vessels to weathervane into the predominant sea state.
Though swivels and turrets allow ship-shaped vessels
to weathervane, they are costly, have long lead times
and are typically available from only few specialized
designers and fabricators. Swivels and turrets also
have associated maintenance requirements and
potential downtime (from leaking seals, for example).
According to Lamport and Josefsson [6], slender ship-
shapes are subjected to significant bending loads due
to hogging and sagging and, as a result, are subject to
fatigue damage. In the case of converted hulls, the
fatigue problem is exasperated when using hulls built
after 1985 where high tensile strength steel was used
extensively to reduce weight. Ship-shapes are also less
efficient in storage volume per plated area than more
compact shapes of the next generation round-shaped
FPSOs.
To overcome short comings associated with using
traditional ship-shaped vessels for FPSOs, the industry
is now developing fit-for-purpose FPSOs. Unlike
traditional ship shaped FPSOs, which must
weathervane into the predominate sea state to
minimize water resistance and motions, the next
generation FPSOs are being designed to have similar
motion characteristics from all directions and to
eliminate yaw excitation. This eliminates the need for
a costly turret and swivels, minimizes the bending
loads and fatigue and increases the storage capacity
per plated area.
Round-shaped FPSOs also have the advantage of
being more easily approachable by service and
installation vessels with minimum collision risk.
According to Lamport and Josefsson [6] Round-shape
have the several advantages, as follow:
a) More efficient storage shape and the smaller
bending load.
b) The motions are similar from all directions with
little to no yaw excitation.
c) More efficient storage shape and the smaller
bending load.
d) The pie-shaped tanks in round-shaped units create
smaller sloshing forces.
e) Providing additional savings in structural
reinforcement
f) Allows for larger freeboard, decreases the risk of
green water on the deck
g) Simple block construction and repeatable
fabrication.
Figure 3. Design Concept Round-Shape FPSO
5. FLOW AROUND THE SHIP SECTION
Computational fluid dynamics, usually abbreviated
as CFD, is a branch of fluid mechanics that uses
numerical methods and algorithms to solve and
analyze problems that involve fluid flows. Computers
are used to perform the calculations required to
simulate the interaction of liquids and gases with
surfaces defined by boundary conditions. With high-
speed supercomputers, better solutions can be
achieved. In this section discuss the theory of the fluid
flow for CFD simulation.
5.1. Incompressible Potential Flow
Incompressible flow is constant density flow, i.e.
. Visualize a fluid element of fixed
mass moving along a streamline in an incompressible
flow. Because its density is constant, then the volume
of the fluid element is also constant. Determine to
the time rate of change of the volume of a fluid
element, per unit volume. Since the volume is constant
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for a fluid element in incompressible flow, the
equation becomes:
Futhermore, if the fluid element does not rotate as it
moves along the streamline, i.e. if its motion is
translational only, then the flow is called irrotational
flow. For such flow, the velocity can be expressed as
the gradient of a scalar function called the velocity
potential, denoted by .
Combining Eqs. (1) and (2),
or,
Equation (3) is Laplace‘s equation – one of the most
famous and extensively studied equations in
mathematical physics. From Eq. (3), can be determine
that inviscid, irrotational, incompressible flow
(sometimes called potential flow) is governed by
Laplace‘s equation.
5.2. Reynolds-Averaged Navier-Stokes (RANS)
Equation
The non-dimensional RANS equations for unsteady,
three-dimensional incompressible flow can be written
in Cartesian tensor notation as,
where
and are the Cartesian
components of mean and fluctuating velocities,
respectively, normalized by the reference velocity U0,
is the dimensionless
coordinates normalized by a characteristic length ,
is the Reynolds number, is the kinematic
viscosity, the barred quantities
Reynolds stresses normalized by , and is the
upressure normalized by are related
to the corresponding mean rate of strain through an
isotropic eddy viscosity, , i.e.
Where is the turbulent kinetic
energy, Equation (4) becomes
where 1/Rø =1/Re+vt, and ø=Ui (i=1,2,3). Equations
(5) and (7) can be solved for Ui and p when a suitable
turbulence model is employed to calculate the eddy-
viscosity distribution.
5.3. Hess-smith Method
A.M.O. Smith at Douglas Aircraft directed an
incredibly productive aerodynamics development
group in the late ‘50s through the early ‘70s. In this
section we describe the implementation of the theory
given above that originated in his group. *Our
derivation follows Moran‘s description6 of the Hess
and Smith method quite closely. The approach is to i)
break up the surface into straight line segments, i i)
assume the source strength is constant over each line
segment (panel) but has a different value for each
panel, and i i i) the vortex strength is constant and
equal over each panel.
Roughly, think of the constant vortices as adding up to
the circulation to satisfy the Kutta condition. The
sources are required to satisfy flow tangency on the
surface (thickness).
Figure 4 illustrates the representation of a smooth
surface by a series of line segments. The numbering
system starts at the lower surface trailing edge and
proceeds forward, around the leading edge and aft to
the upper surface trailing edge. N+1 points define N
panels.
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Figure 4. Representation of a smooth airfoil with
straight line segments.
The computer program based on the Hess-Smith panel
method (HSPM) approximates the body surface by a
collection of panels and expresses the flow field in
terms of velocity potentials based on sources and
vortices in the presence of an onset flow.
(8)
where, is the total potential function and its three
components are the potentials corresponding to the
free stream, the source distribution, and the vortex
distribution. These last two distributions have
potentially locally varying strengths and ,
where is an arc-length coordinate which spans the
complete surface of the airfoil.
The potentials created by the distribution of
sources/sinks and vortices are given by:
Combining Eqs. (8), (9) and (10)
The potential relation given above in Eq. (4-22) can
then be evaluated by breaking the integral up into
segments along each panel:
Since Eq. (12) involves integrations over each discrete
panel on the surface of the airfoil, we must somehow
parameterize the variation of source and vortex
strength within each of the panels. Since the vortex
strength was considered to be a constant, we only need
worry about the source strength distribution within
each panel.
This is the major approximation of the panel method.
However, you can see how the importance of this
approximation should decrease as the number of
panels, (of course this will increase the cost of
the computation considerably, so there are more
efficient alternatives.)
Hess and Smith decided to take the simplest possible
approximation, that is, to take the source strength to be
constant on each of the panels
Therefore, we have unknowns to solve for in
our problem: the panel source strengths qi and the
constant vortex strength . Consequently, we will need
independent equations which can be obtained
by formulating the flow tangency boundary condition
at each of the panels, and by enforcing the Kutta
condition discussed previously. The solution of the
problem will require the inversion of a matrix of size
.
5.4. Cubic Spline
Cubic spline interpolation is a useful technique to
interpolate between known data points due to its stable
and smooth characteristics. The cubic spline has been
utilized within the grid generation procedure to
accurately model curves that may be found in
engineering situations. An example of such a curve is
a hull ship section which, using cubic splines, can be
regenerated using relatively few data points. The cubic
spline fits a cubic polynomial between each set of
defining data points. The cubic spline is equal at the
data points and the spline is thus continuous. If the
gradient and the curvature are also assumed to be
continuous then the spline can be derived.
The fundamental idea behind cubic spline
interpolation is based on the engineer‘s tool used to
draw smooth curves through a number of points. This
spline consists of weights attached to a flat surface at
the points to be connected. A flexible strip is then bent
across each of these weights, resulting in a pleasingly
smooth curve. The mathematical spline is similar in
principle. The points, in this case, are numerical data.
The weights are the coefficients on the cubic
polynomials used to interpolate the data. These
coefficients ‘bend‘ the line so that it passes through
each of the data points without any erratic behavior or
breaks in continuity.
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Cubic Splines Derivation
Consider a collection of known points ,
, ... , , , ...
. To interpolate between these data points
using traditional cubic splines, a third degree
polynomial is constructed between each point. The
equation to the left of point is indicated as
with a value of at point . Similarly, the
equation to the right of point is indicated as
with a value of at point .
If a set of data points is defined a unique cubic
polynomial can be defined between each set of points.
where is a third degree polynomial defined by
for
The first and second derivatives of these
equations are fundamental to this process, and they are
for
6. CONCLUSION
The new concept design of the FPSO is currently a
cylinder shaped or non ship-shape proposed by several
researchers which is providing a better design. LNG
products that have been produced on the FPSO LNG
will be transferred to the LNG Carrier. The transfer
process is called offloading process. There are two
methods of offloading process for oil and gas transfer
which is tandem and side-by-side. Current research
focuses on for side-by-side configuration.
Fluid flow around a structure can significantly alter
the structure‘s loading characteristics. Waves and
currents is a fluid that has the most impact on the
structure. Current research discuss the concept design
of fluid flow around a round-shaped FPSO LNG
interacted with a LNG carrier during side-by-side
offloading condition by CFD method based on the
Reynolds Average Navier-Stokes (RANS) equations.
ACKNOWLEDGMENT
Special thank to Pengajian Tinggi Malaysia (MOHE)
and Universiti Teknologi Malaysia for supporting this
research.
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2012
233
Simulation of Organic Rankine Cycle System Using Turbocharger with
Cycle Tempo and Environmentally Friendly Fluid
Ruli Nutranta a, Idrus Al Hamid
a, Nasruddin
a, Harinaldi
a
aFaculty of Engineering
University of Indonesia, Depok 16424
Tel : (021) 7270011 ext 51. Fax : (021) 7270077
E-mail : [email protected]
Abstract
Organic Rankine cycle (ORC) is a modified rankine cycle
with working fluids, of organic material (Refrigerant).
Refrigeran pentane has low boiling point, therefore ORC can
be used in power plant which uses low temperature resources,
such as solar thermal exhausted gases and geothermal wells.
Organic Rankine Cycle (ORC) is used to convert heat energy
into mechanical energy or electricity generated by a low
temperature of the hot sun. The working fluid used is R12,
R22, R134a and Pentane. Simulations performed with an
organic Rankine cycle temperature and pressure with cycle
tempo program. By programming the simulation cycle tempo
and got the result on the maximum power a turbine to the
conditions of the working fluid Pentane to the input turbine T
= 700C and pressure = 2 bar can generate 2.07 kW.
Turbocharger is one of the alternatives in the energy
conversion of the energy of motion into electrical energy.
Turbocharger rotation will be used to turn a generator and
converts the energy of motion into electrical energy.
Keyword : organic rankine cycle, energy, working fluid,
turbocharger
1. Introduction
Power energy in Indonesia has increase rapidly in national
daily consumsion and has developed renewable energy
include solar energy. Solar energy plays an important role in
the utilization of electrical energy in Indonesia. On average
the sun shines in Indonesia about 12 hours per day. If it takes
three hours on average for the utilization of solar thermal
then it becomes extremely beneficial to the interests of
society, especially in rural areas. To meet the demand for
electricity has not come into the house (by 33.4% in 2010),
then the solar power into one of the renewable alternative
energy that should be developed [1]. In this solar energy
utilization, solar thermal energy is one of Indonesia are very
rarely used. Utilization of solar thermal by finding a suitable
working fluid and meets the latest technology, is one reason
researchers simulate materials such as R12, R22, R134a and
Pentane.
2. Study Literature The majority of electrical generating plants are variations of
vapor power plants where water is the working fluid. The
basic components of a simplified Rankine cycle are shown in
Figure 1
Figure 1. Basic Organic Rankine Cycle
The Rankine cycle is the thermodynamic cycle that models
of Figure 1. In analyzing this cycle, one neglects the stray
heat that takes place between the plant components and their
surroundings. Further, kinetic and potential energy effects are
ignored. Finally, each component is considered to be
operating at steady state
Table 1. Equation of Basic ORC
The Rankine cycle differs from the Carnot cycle in that the
heat trans-fer processes take place at constant pressure
instead of constant tem-perature. On the other hand, much of
the Rankine cycle cooling and heating processes include
phase changes, and thus, they also preserve the isothermal
character while the phase change is in progress. For this
reason the efficiency is very good, but still less than the
Carnot efficiency. For example, the isothermal character is
not preserved for the first part of the heating process, when
the liquid is in the compressed form, and or the last part of
the heating process, when the fluid is in the superheated
vapor region. The steam must be in a superheated state before
entering the turbine so that the liquid state inside the turbine
can be avoided. Condensation in the turbine can cause blade
erosion.
Subsistem Equation
Pump (1) =
Evaporator
(2)
Turbine (3) Condenser
(4)
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The 5th IMAT, November 12 – 13th
2012
234
Many investigation is conducted the simulations and
experiments to find an effective working fluid used in the
ORC. Yamamoto [2] investigated on the estimation of
operating conditions using a ORC working fluid of water and
HCFC-123. Saleh [3] examine 31 working fluid on the
condition of the sub-critical and Supercritical of geothermal
power generation on the ORC and Tchanche et.al.[4] using
organic solar Rankine cycle system to investigated 20
working fluid that has the best work in low-temperature.
In organic Rankine cycle, Mills [5] examines several
technologies including solar-powered ORC is capable of
generating electricity in several countries. Research on a
combination of fuel cell engine, gas turbine and ORC has
been done by Sanchez [6]. This study used the work of some
of the refrigerant fluid in the ORC machines. To R245fa
found to have the best advantage to produce electricity.
Vankeirsblick [7] also found that the ORC is the most
efficient engine to run a small generator instead of steam
power plants. In this study, ORC with simulated regenerator
to produce power 368.2 kw generator with an efficiency of
95%. Research on solar-powered ORC was also carried out
by Nasri [8] in the Sahara desert region in the form of
modeling and simulation. Although many studies examine
both ORC working fluid and the system, but not many
researcher is substituted turbocharger as turbine in Organic
Rankine Cycle
The objective of this study is to investigate the performance
working fluid R12, R22, R134a and pentane to electricity in
solar low-temperature ORC with turbocharger as a turbine
with cycle tempo.
3. Research Methodology The research is to compare the simulation and solar-
powered ORC engine. Activity of research is taken to create
a prototype solar-powered ORC with turbocharger. The ORC
will be connected with solar thermal collectors consisting of
flat and/or parabolic collectors. Research methodology in this
study can be seen in Figure 2 below. Starting from the
previous test, which has a solar thermal flat and parabolic
solar collectors, the research will investigated the
characteristics of water as a medium. The results obtained are
solar radiation and the efficiency of collectors, Heat Removal
Factor and Heat Loss Coefficient for some combination of
flat and parabolic
Figure 2. Reasearch Methodology
Setup experiment was conduct to small ORC with
temperature below 800C.
Figure 3. Organic Rankine Cycle with Turbocharger
The consideration of simulation are listed as follow:
a. The condensation temperature is 350C (308 K)
b. The evaporator temperature was set to 700C (343 K)
c. The reference state temperature was set to 300C (303 K)
d. The efficiency of the pump was set to 0.75
e. The efficiency of the turbocharger was set to 0.75
The 5th IMAT, November 12 – 13th
2012
235
Working fluid were investigated in this study is a working
fluid circulating in Indonesia and environmentally friendly.
Working fluid mixture of from R12, R22, RC134a and
Pentana. For Pentane, Indonesia still imports from abroad.
But working is required characteristics and perfomanya for
further research in the UI. The selected working fluid can be
seen in Table 2 below
Table 2. Physical properties working fluid
Molecular
weight
(g/mol)
(oC)
(MP
a)
Std 34
Safety
group
ODP GWP
(100
year)
R12 120.91 112 4.114 A1 1.000 10,890
R22 86.47 96.2 4.99 A1 0.055 1810
R134a 102.03 101.1 4.06 A1 0 1300
Pentane 72.15 196.6 3.37 A3 0 20
The turbocharger turbine, which consists of a turbine wheel
and a turbine housing, converts the engine exhaust gas into
mechanical energy to drive the compressor.The gas, which is
restricted by the turbine's flow cross-sectional area, results in
a pressure and temperature drop between the inlet and outlet.
This pressure drop is converted by the turbine into kinetic
energy to drive the turbine wheel.
Figure 4. Gasoline Turbocharger 1300 cc
As the radial-flow turbine is the most popular type for
automotive applications, thefollowing description is limited
to the design and function of this turbine type. In the volute
of such radial or centripetal turbines, exhaust gas pressure is
converted into kinetic energy and the exhaust gas at the
wheel circumference is directed at constant velocity to the
turbine wheel. Energy transfer from kinetic energy into shaft
power takes place in the turbine wheel, which is designed so
that nearly all the kinetic energy is converted by the time the
gas reaches the wheel outlet.[9]
4. Result and discussion
Research has done with electricity with condesor temperature
of 313K, 348K and evaporator temperature ambient
temperature of 300K. 0.85 pump efficiency, turbine
efficiency 0.7 and 5 kW of electric power [10]. Thus the
calculation of this system can be seen in Table 3 below.
Figure 5. Efficiency - Power 5 kW
In this research, thermal efficiency of Pentane also has a
number greater than most other working fluid. It is 8.9626 at
348K evaporator temperature. In Figure 4 can be seen five
working fluid thermal efficiency. Working fluid which has
the smallest thermal efficiency is HRC 12 with a value of
6.5233 is almost equal to R134a the value of 6.5975
Table 3. Summary of working fluid with Cycle tempo
No Fluid
System
Subsystem pin
(Bar)
Ppump
(W)
m
(kg/s)
P
(kW)
1 R 12
Pump 9
30 0.121 2.15 Evaporation 10
Turbine 10
Condenser 8.462
2 R 22
Pump 14
20 0.107 0.16 Evaporation 15
Turbine 15
Condenser 13.55
3 R134a
Pump 12
20 0.11 0.52 Evaporation 13
Turbine 13
Condenser 8.87
4 Pentane
Pump 1
20 0.121 2.07 Evaporation 2
Turbine 2
Condenser 0.977
This research, simulation cycle tempo conducted to
determine how the power generated in the turbine. R12, R22,
R134a and pentane are included as a working fluid
temperature at 35 ° C minimum and 80 ° C conditions at its
maximum. Simulation in cycletempo at table 3 shown the
highest value for the power is in R12 (2.15 kW). However, to
generate power, its required pump 30W and pressure up to 10
bar. Comparing with pentane which only requires power 20
W and pressure pump 2 bar, it is clear that for pentane
working fluid lighter than R12
The 5th IMAT, November 12 – 13th
2012
236
Figure 6. Pentane Simulation
The results of pentane is obtained power 2.07 kW at inlet
turbine pressure 2 bar and 0.977 bar at outlet pressure in
figure 6. The data can be explained:
- Pump: pump power 20 W with 75% efficiency entropy and
mechanical efficiency of 65%.
- Evaporator: subtitud by PHE with inlet temperature 35 ° C
and outlet temperature 80°C. A pressure of 2 bar for water
entering the evaporator temperature 80°C entry pressure of
1 bar, the incoming flow 0123 kg / s
- Turbine: from the inlet temperature 70 C to 40°C outlet
temperature. Pressure of 2 bar entry to 0.977 bar 2.07 kW
turbine generates power
- Condenser: the incoming pressure 0.977 bar and
temperature 35°C.
5. Conclusion
Conclusion can be drawn from the cycle tempo simulation by
using a turbocharger and environmentally friendly working
fluid:
a. The turbocharger can be used by decrease flow rate. This
adjustment can be made by used inlet nozzle
turbocharger.
b. Pentane have good conditions for development, especially
at low pressure and temperatures below 100 ° C.
Therefore the use of pentane substance is suitable for use
in Indonesia because of its nature.
References
[1] Sumiarso L.Regulasi dan pengembangan energi baru
terbarukan dalam rangka energi bersih. UMB 2011.
[2] Yamamoto T, Furuhata T, Arai N, Mori K. Design and
testing of the organic Rankine cycle. Energy
2001;26:239-51.
[3] Saleh B, Kogibauer G, Wedland M, Fischer J. Working
for low temperature organic Rankine cycle. Energy
2007;32:1210-21
[4] Tchanche BF, Papandakis G, Lambrios G, Frangoudakis.
A Fluid selection for low temperature solar organic
Rankine cycle. Appl Therm Eng 2009;29:2168-76
[5] D. Mills, Advanced in Solar thermal electricity
technology, Solar Energy (2004), vol 76 0hal 19-31
[6] D Sancez, JM Munoz de Escalona, B Monje, R
Chacartegui, T Sanchez, Preeliminary analysis of
compound system based on high temperatur fuel cell, gas
turbine and organic Rankine Cycle, Journal of Power
Source (2011), vol 196 hal 4355-4363
[7] Vankeirsblick, Vanslambrouck, Gusev, De Paepe,
Organic Rankine cycle as efficient alternative to steam
cycle for small scape power generation, International
Conference of Heat Transfer, Fluid Mechanic and
Thermodynamic, Mauritius (2011)
[8] Faozi Nasri, Chouki Ali, Habib Ben Bacha, Electricity
production system form solar heated rankine cycle :
modelling and simulation, IJRRAS (2011) vol 8 hal 176-
183
[9] Brown S, The Turbomustangs.com : Complete
Turbocharging guide, Underpsi of utahstang.com, 2003
[10] Nutranta. R, AlHamid. MIA, Nasrudin, Harinaldi. Studi
karakteristik fluida kerja hydrokarbon ramah lingkungan
pada siklus rankine (SRO) bertenaga surya. SNTTM X
Universitas Brawijaya Malang, 2011
The 5th IMAT, November 12 – 13th
2012
237
Thermophysical Properties of Novel Zeolite Materials for Sorption Cycles
Kyaw Thua, Young-Deuk Kim
a, Baojuan Xi
b, Azhar Bin Ismail
b, Kim Choon Ng
a,b,*
aWater Desalination and Reuse Center
King Abdullah University of Science and Technology, Thuwal 23955-6900
Saudi Arabia Tel : (966) 2-8084969.
E-mail : [email protected]; [email protected]
bDepartment of Mechanical Engineering
National University of Singapore, Singapore 117576
Tel : (065) 65162214. Fax : (065) 6779-1459
E-mail : [email protected]; [email protected]; [email protected]
*Corresponding Author
ABSTRACT
This article discusses the thermophysical properties of
zeolite-based adsorbents. Three types of zeolite (Z-01, Z-02
and Z-05) with different chemical compositions developed by
Mitsubishi Plastics, Inc. are analyzed for possible
applications in adsorption chillers and desalination cycles.
Static volumetric method is adopted with N2 gas sorption at
77 K. Thermophysical properties such as pore surface area,
micropore volume and pore size distribution are evaluated
using standard multi-point Brunauer-Emmett-Teller (BET)
and Non-Local Density Functional Theory (NLDFT)
methods. It is observed that Aluminosilicate functionalized
Z-02 exhibits the highest surface area with huge micropore
volume.
Keywords : Adsorption, Zeolite, Thermophysical properties
1. INTRODUCTION
Frequent and severe natural disasters such as super storms
and earthquakes are claimed to be attributed to the
environmental issues such as Global warming and
greenhouse gas emissions. The burning of hydrocarbons for
various applications i.e., the industrial processes, the
transportation and the life comfort, inevitably contributes
damages to the environment. Thus, secondary fuels
(photovoltaic, winder turbine and fuel cells) gain much
attention as clean and environmental-friendly alternates to
depleting oil [1-3]. However, the primary systems powered
by secondary fuels conventionally operate at low efficiencies,
typically below 60%. With the introduction of the
cogeneration concept such as Combined Heat and Power
(CHP) systems, the overall system efficiency can be realized
as high as 80% [4]. The combine systems normally include
waste heat-driven absorption and adsorption cycles producing
useful effects such as cooling and potable water [5].
The adsorption cycles employed sorption principles between
the solid adsorbent and the vapor phase adsorbate. Common
adsorbent materials for adsorption chillers are silica gel,
zeolite and activated carbon whilst water, ethanol and
methanol are used as adsorbate. Poor coefficient of
performance (typically less than 0.7) of adsorption cycles
calls for the development of new adsorbent materials and
improved heat and mass recovery schemes [6-11]. The
adsorbent selection depends on the quality of the available
waste heat and type of refrigerant. Silica gel is commonly
employed for low-temperature waste heat application whilst
zeolite and activated carbon are used where the waste-heat is
higher than 100 °C [12-15]. With recent development in the
novel zeolite materials, it is possible to regenerate the
adsorbate from the zeolite using low-temperature heat
sources, typically as low as 60 °C [16-18].
This article presents the thermophysical properties of novel
zeolite materials developed by Mitsubishi Plastics, Inc. These
materials are developed for possible applications where
regeneration temperature is as low as 55 °C with water vapor
adsorption. Three types of powdered-zeolite materials with
different chemical compositions are investigated for their
thermophysical properties such as pore surface area, pore
volume and pore size distribution.
2. MATERIAL AND METHOD
Three types of zeolite materials code names (Z-01, Z-02 and
Z-05) are investigated. The chemical composition and
Scanning Electron Microscope (SEM) pictures of these
samples are given in Table 1.
Table 1: Chemical composition and SEM pictures of the
samples.
Z-01 Z-02 Z-05
Composition FeAlPO2.nH2O SiAlPO2.nH2O AlPO2.nH2O
SEM image
(Particle
description)
SEM image (high
resolution)
Static volumetric method is used with N2 gas adsorption at 77
K. The AutoSorb-1 analyzer manufactured by Quantachrome
Corporation is employed to investigate thermophysical
properties. The minimum relative pressure available by this
IMAT-UI 039
The 5th IMAT, November 12 – 13th
2012
238
analyzer is 1.0 x 10-7
with N2 sorption whilst the applicable
lowest surface area is 0.01 m2/g without upper limit. The
AutoSorb-1 has the capability of measuring adsorbed or
desorbed volumes of nitrogen at relative pressures in the
range 0.001 to slightly lower than 1.0. The supplied software
(ASWIN) facilitates the analyses of the surface
characteristics using various methods such as BET surface
area (single and/or multipoint), Langmuir surface area,
adsorption and/or desorption isotherms, pore size and surface
area distributions, micropore volume and surface area using
an extensive set of built-in data reduction procedures. All the
samples are performed degasification at 140 °C under
vacuum for 12 hours prior to the experiments.
3. RESULTS AND DISCUSSION
2.8 N2 Isotherm
The N2 gas adsorption by three types of selected zeolite
materials is given in Figure 1. It is observed that all the
adsorbent and adsorbate pairs exhibit Type II isotherms and
Aluminosilicate-based Z-02 has the highest uptake followed
by Z-01 and Z-05. The amount of uptake by Z-02 is
significantly higher than those of Z-01 and Z-05 with average
uptake of 170 cm3/g.
Figure 1: N2 gas adsorption by selected adsorbent
2.9 Micropore Analysis
Using the N2 adsorption data, the present of micropore and
the external surface area are evaluated using the t-method of
de-Boer [19]. A t-plot is a plot of the volume of gas adsorbed
versus t, the statistical thickness of an adsorbed film and
Figure 2 gives the t-plot of the zeolite materials whilst Table
2 provides the summary of the micropore analysis by t-
method.
The results show that all the samples present micropore.
However, Z-02 exhibits huge amount of micropore and the
micropore volume and micropore surface area of Z-02 are
found to be one order than those of Z-01 and Z-05.
Figure 2: Micropore analysis of the selected Zeolite
materials using t-plot method
Table 2: The summary of the micropore analysis of the selected
Zeolite materials by t-method.
Parameter Z-01 Z-02 Z-05
Slope 6.386 1.891 4.183
Intercept 18.551 172.00 13.90
Correlation coefficient, r 0.999994 0.998 1.000
Micropore volume (cm3/g) 0.029 0.266 0.021
Micropore area (m2/g) 67.387 737.000 41.032
External surface area (m2/g) 98.781 29.246 64.696
2.10 Surface Area Analysis
Using the Brunauer-Emmett-Teller (BET) method, the
surface area of the adsorbent can be determined as,
1 1 1
0 01
C P
P W C W C Pm mW
P
(1)
where W is the weight of the adsorbed gas at a relative
pressure (P/P0), Wm is the weight of adsorbate at a monolayer
coverage and C is the BET constant. This is related to the
adsorption energy of the first adsorbed layer, indicating the
magnitude of the adsorbent/adsorbate interactions. Table 3
summarizes the surface area analysis results. It is noted due
to the present of huge micropore, Z-02 possesses huge
surface area of 766 m2/g.
Table 3: BET analysis summary of Zeolite materials.
Parameter Z-01 Z-02 Z-05
Slope 20.936 4.544 30.053
Y-intercept, i 2.21E-02 6.96E-06 3.26E-02
Correlation coefficient, r 1 1 0.999995
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239
BET constant, C 946.833 653308.334 922.153
Surface area (m2/g) 166.168 766.374 115.752
2.11 Pore Size Distribution
According to the IUPAC recommendation, the classification
of pores present in an adsorbent is such that micropores (<
2nm), mesopores (between 2nm and 50nm), and macropores
(larger than 50nm) as to their pore widths. Pores in
adsorbents represent higher surface area as well as high
selectivity in reaction and adsorption. Pore Size Distribution
(PSD) analysis is the most essential information of the pore
and it is used to evaluate the population of pores as a function
of the pore width [20]. The pore-size distribution analyses of
the mentioned Zeolite materials are conducted using the Non-
Local Density Functional Theory (NLDFT) method with the
application of the provided software package by the
AutoSorb-1. Here, the Equilibrium Model is adopted to
determine the PSD of the aforesaid adsorbents. Figure 3
shows the cumulated pore volume of the selected adsorbent
materials. It is observed that Z-02 type adsorbent exhibits the
highest total pore volume followed by Z-01 and Z-05.
Figure 3: Cumulated pore volume of the selected Zeolite
materials by NLDFT method
Figure 4 gives the PSD comparison of three zeolite materials
whilst Table 4 shows the analyses summary. The pore width
of all the adsorbent is found to be between 2 and 5 nm. It is
noted that Z-05 exhibits two maxima distribution or bimodal
type.
Dubinin-Astakhov (DA) analysis of the adsorption isotherm
of the Zeolite materials for N2 gas adsorption is depicted in
Figure 5 where the DA equation is given as,
0
0
lnn
RT P PW W Exp
E
(2)
here W is the weight of adsorbed amount at relative pressure,
P/P0 and T, W0 is the total adsorbed weight, E is the
characteristic energy and n is the DA constant. The
significantly higher pore volume of Z-02 is detected here.
Finally, Table 5 gives the analysis summary using DA
method. DA analysis reveals that Z-02 possesses the highest
micropore volume (0.29 cm3/g) as well as the highest
characteristic energy (71 kJ/mol).
Figure 4: Pore Size Distribution summary of different
types of Zeolite materials Table 4: Summary of NLDFT (Equilibrium Model) for PSD analysis
of Zeolite materials.
Pore Volume
(cm3/g)
Pore width
(Å)
Fitting Error
(%)
Z-01 0.108 34.18 1.414
Z-02 0.320 45.7 0.310
Z-05 0.084 11.14 0.83
Figure 5: DA plot of the Zeolite materials for N2 gas
adsorption Table 5: DA analysis summary of the Zeolite materials.
Parameter Z-01 Z-02 Z-05
Characteristic energy, E (kJ/mol) 6.122 70.566 5.988
DA constant, n 1.0 1.0 1.0
DA Micropore Volume (cm3/g) 0.100 0.288 0.069
Pore Radius (Å) 7.10 3.20 7.20
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4. CONCLUSION
Three types of novel zeolite materials with different chemical
compositions provided by Mitsubishi Plastics, Inc. have been
investigated for their possible application in adsorption
refrigeration and desalination cycles. The samples are
analyzed using the static volumetric method with N2 gas as
adsorbate at 77 K. It is observed that Z-02 type adsorbent
with Aluminosilicate composition exhibits superiority in
terms of surface area (766 m2/g) and micropore volume (0.32
by NLDFT method and 0.29 by DA method). It is also noted
that the pore diameter of all the selected adsorbents is
between 2 and 5 nm with Z-05 showing bimodal type or two
maxima distribution. Based on these analyses, it is suggested
that Z-02 type material is suitable for adsorption cycle
application with superior surface characteristics.
ACKNOWLEDGMENT
The authors gratefully acknowledge Mayekawa
Manufacturing Co., Ltd. for the supply of zeolite materials.
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Review Paper: Sea-water Ice Slurry Generator and Its Application on
Indonesian Traditional Fishing
A.S. Pamitrana, H.D. Ardiansyah
b, M. Novviali
b
aDepartment of Mechanical Engineering,
Universitas Indonesia, Kampus UI Depok 16424, Indonesia
Tel : (021) 7270032, Fax : (021) 7270033
E-mail : [email protected] bGraduate Student, Department of Mechanical Engineering,
Universitas Indonesia, Kampus UI Depok 16424, Indonesia
ABSTRACT
Sea-water ice slurry generator is aimed to produce ice-
slurry using sea-water while fishing on boat. Some models
of ice slurry generator have been researched and developed.
Indonesia as an archipelago and maritime nation has big
potential in fishery. Appropriate model for Indonesian
traditional fishing boat is necessary. The most important
part of ice slurry generator is evaporator, including auger
and scrapper. Some studies on ice slurry are presented in
this review paper in order to get larger view on design of
ice slurry generator.
Keywords: Ice slurry, cooling, sea-water, evaporator,
fishing
1. INTRODUCTION
Nowadays international agreement has been applied to
protect environment from using refrigerant which contain
chlorine. Every chlorine refrigerant has effect for
environment destructive where the parameter is said Ozone
Depletion Potential. Refrigerant contains hydrogen was
developed to replace CFCs, such as NH3. However, the
toxicity of NH3 must be considered for a refrigeration
system. The properties of refrigerants are very important to
be considered in a system due to their adhered effects.
Therefore, secondary refrigerant such as water solution can
be largely used to minimize accident and leak effect of
refrigerants.
Ice slurry is useful secondary refrigerant for many
applications. Low temperature of primary refrigerant can
absorb more heat and change phase of fluid from liquid
become mixture of ice-liquid. Fluid for ice slurry can be
pure water or solution has freezing point depressant such as
Sodium Chloride, Ethanol, Ethylene Glycol, Propylene
Glycol (Kauffeld et al., 2005) and sea water (A.S Pamitran
et al., 2012). Figure 1 illustrates a schematic diagram for ice
slurry as a secondary refrigerant published in Meewise,
2004.
The characteristic and advantage of ice slurry has been
researched before 1975. One of interesting topic of ice
slurry study for some researchers are heat transfer and
pressure drop of ice slurry even
Figure 1: Secondary refrigerant scheme (Meewise,
2004)
some topic presented ASHRAE Meeting at June 1998
Toronto, Canada (Kirby P. Nelson et al., 1998). Energy
storage of ice slurry is higher than others secondary
refrigerants because ice slurry contain ice particle which
has high latent heat in solution. Moreover, ice slurry is fast
and effective for cooling because of large contact surface
for heat transfer between ice particle and product. Ice slurry
can reduce dimension of tank, pipe, chiller, and can reach
economic efficiency by reducing more than 70% power of
pump compare with water (Kasza et al., 1988). Nowadays,
some researchers concern on ice slurry because of its large
benefit. For example application in industry (Wang and
Kusumoto,2001; Rivet, 2009), medical and direct cooling
for food or fish (Wang and Goldstein, 2003; Pineiro et al.,
2004).
Some published papers study on thermofluids
characteristics of ice slurry. Gupla dan Frazer (1990)
explained ice slurry in heat exchanger with 6% ethylene
glycol, ice fraction of 0%-20%, flow rates of 1.18 m3/hour
and 2.16 m3/hour, diameters of ice slurry of 0.125 mm and
0.625 mm. The result showed total heat transfer coefficient
proportional to flow rate and opposite to increasing ice
fraction. Pressure drop is constant until ice fraction of 20%
and then rapidly increase more than 20%. Kauffeld (1999)
compared ethanol solution and potassium carbonate
solution as fluids to produce ice slurry. Ethanol solution
result small ice particle and has heat transfer coefficient
increase along with ice fraction increase. The opposite
result was found for potassium carbonate solution with big
ice particle. The heat transfer coefficient decreases along
with ice fraction increase. Knodel (2000) concluded heat
transfer coefficient decrease with ice fraction increase. As
well Gupla and Frazer (1990) showed the same result.
Knodel (2000) explained heat transfer decrease because
IMAT-UI 040
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moving of ice slurry flow change from turbulent to laminar
due to increasing of the fraction ice. Different result was
reported by Bellas J et al. (2002) that measured ice slurry of
5% propylene glycol solution in plate heat exchanger with
0-25 % ice fraction and 1-3.7 m3/hour flow rate. Bellas et
al. (2002) explained that increase of ice fraction 0-20%
make pressure drop about 15% and total heat transfer
coefficient rapidly increase along with increase of flow rate.
Variation of ice fraction has insignificantly effect on heat
transfer coefficient.
N. Putra et al. (2004) used ice breaker to make ice slurry.
Heat transfer coefficient increases whilst flow rate and ice
fraction increase in plate heat exchanger. They assumed
heat transfer coefficient and pressure drop are functions of
viscosity, Reynolds number, diameter and ice fraction of ice
slurry. Stamatiou dan Kawaji (2005) studied heat transfer
coefficient in vertical rectangular channel with given heat
flux on wall surface. Nusselt number increase along with
increase of ice fraction and heat flux. The fusion of ice
particle formed on pipe result in higher convection heat
transfer than conduction heat transfer along of the pipe. The
profile of velocity on wall surface for ice slurry and liquid
is different. Liquid velocity was not function of heat, but ice
slurry velocity was function of fusion of ice particle. Lee
D.W. et al. (2006) deeply researched heat transfer with
6.5% ethylene glycol in diameter of 13.84 mm and length
of 1500 mm, ice slurry mass flux of 800-3500 kg/m2s and
ice fraction of 0-25%. The result was heat transfer
coefficient increase along with increase of flow rate and ice
fraction, but effect of ice fraction was insignificant in
higher flow rate and in lower flow rate region. Niezgoda-
Zelasko (2006), Niezgoda et al. (2006) and Grozdek (2009)
researched on heat transfer and pressure drop with ice
slurry of 10% ethanol in horizontal pipe. High ice fraction
and high velocity give high heat transfer coefficient and
pressure drop. Heat flux has small effect on heat transfer.
Ice fraction of 10-15 % increases heat transfer coefficient in
laminar flow, but not for turbulent flow under same heat
flux. Beyond the value has high heat transfer coefficient.
J.P. Nedecarrats et al. (2009) used corrugated and smooth
pipe with given heat flux on pipe wall, velocity of 0.3-1.9
m/s and ice fraction of 0-30%. Pressure drop and heat
transfer coefficient increase along with fraction ice and
velocity. They founded critical point of pressure drop and
heat transfer coefficient. Comparison of corrugated and
smooth pipe was reported with result of heat transfer and
pressure drop are 2.5 times higher for corrugated pipe.
Report from some researchers can be concluded that ice
slurry characteristic depend on the solution, flow rate, ice
fraction, and diameter of ice particle. Although many
researchers have offered some methods and correlations
regarding ice slurry characteristic, but their results could
not used largely to predict heat transfer and pressure drop in
heat exchanger (Ayel et al., 2003). Therefore, study on ice
slurry is still opened in investigating of their thermofluid
characteristic. The other interesting topics on ice slurry are
its application, ice slurry generator, melting of ice slurry,
ice slurry formation, microscopic observation, measurement
and control, modeling and simulation.
2. ICE SLURRY DEFINITION
Ice slurry consists of liquid and ice particle (E Stamatiou et
al.,2003). Ice slurry is defined as fine-crystalline ice slurry
or liquid has ice particle with average diameter of equal or
less than 1 mm (Peter W Egolf et al., 2003). Nandy P. et al.,
2006, mentioned general definition for ice slurry:
a. Solution and solid with temperature up to -15 oC.
b. Ice slurry can be produced from brine solution with
freezing temperature lower than the freezing temperature
of water up to -50 oC.
c. Ice slurry has different characteristic and behavior
compared to brine solution.
d. Ice slurry is 2 phase fluid non Newtonian in high ice
fraction.
e. Ice slurry needs different calculation and prediction pipe,
pump, heat exchanger even storage tank.
3. ICE SLURRY FORMATION
Ice slurry formation has been explained by E Stamatiou
(2003). Generally, process of ice slurry formation consists
of supersaturation, nucleation, and growth. Moreover, there
are other processes of attrition, agglomeration and ripening
can happen on ice slurry formation.
Supersaturation occurs when driven force complete.
Therefore, supersaturation needs conditions of instability
and difference of chemical potential between solution and
solid crystal.
(1)
Chemical potential difference occurs because of
temperature or pressure driven force. Rate of crystallization
is influenced by grade of supersaturation solution.
Considering Raoult Law, when a liquid mixture with
methanol, ethylene glycol, propylene glycol, sodium
chloride, magnesium chloride, potassium chloride, etc, has
mixture pressure between vapor partial every component
and freezing temperature is lower than water.
Nucleation is formed when molecule getting stable. There
are two kind of nucleation viz. homogeneous and
heterogeneous. Nucleation can be recognized from existing
ice fraction in solution. Equilibrium of chemical potential
from driven force makes separation of solution. Pure water
separated from solution is partially freezing become ice
fraction.
Growing of crystal occurs in three steps viz. mass transfer
with molecule diffusion in bulk liquid through boundary
layer near nucleus, molecule merger to wall area and heat
transfer simultaneously from crystal to bulk area include
change of phase. Part of growing ice is supported by
rotation of shaft auger to accelerate transfer of mass and
heat. Some points of shaft auger break slug of ice crystal on
evaporator inner wall, and ice crystal is revolved in center
of generator. Interaction between nucleation and growth of
crystal determine crystal characteristic such as diameter,
distribution, and morphology of crystal (Mullin, J. W.
2001).
The 5th IMAT, November 12 – 13th
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243
Figure 2 shows that point t0 to t1 is supercooling or
supersaturation, then t1 to t2 is area where ice fraction
forming, t2 to tf is sensible heat until complete of nucleation
(T.A. Mouneer et al,.2011).
Figure 3 present a similar illustration as Figure 2 but with
different result. Nucleation is marked with increasing bulk
temperature. Nucleation starts from forming ice particle.
Figure 2: Time dependence curve for generator and
forming of volumetric ice concentration
(scraper) (T.A. Mouneer et al,.2011)
Figure 3: Time dependence curve and torque of shaft auger
(scraper) (Frank Qin et al., 2006)
4. ICE SLURRY GENERATOR
Every ice slurry generator has similar process with different
evaporator and scraper system. Scraper of ice slurry
generator consists of scraper itself and shaft auger as
illustrated in Figure 4. The purpose of scraper is to avoid
lump of ice on evaporator inner wall. There is a clearance
between scraper and the wall surface. Thermal resistance
makes ineffective heat transfer when ice attach on the wall.
Many industries use scraper system because it is cheaper
than others and produce high ice fraction (E. Stamatioua et
al., 2005, T.A. Mouneer et al., 2011). Heat transfer
equipment can be made with shell and tube or flooded
evaporator.
M. Miguel Leon (2006) modified scraper model with
helical scraper and flooded evaporator as shown in Figure
5.
There are some systems of ice slurry generator such as
falling liquid film ice slurry generator, orbital rod ice slurry
generator, fluidized bed,
Figure 4: Scraper of Ice Slurry Generator (T A
Mouneer et al,. 2011)
(a) (b)
Figure 5: (a) Flooded Evaporator, (b) Helical Scaper
(M. Miguel Leon et al., 2006)
(a) (b)
Figure 6: (a) Falling liquid system (b) Refrigeration
system
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244
Figure 7: Orbital Rod Ice Slurry Generator (E.
Stamatioua et al., 2005)
Figure 8: Vacuum Freezing (Michel Barth)
vacuum freezing, supercooled slurry ice production,
supercooled water jet, direct contact heat transfer. The
system is called falling film because ice slurry flow with
gravitation. Falling film is depicted in Figure 6. Falling
liquid system works without motor but needs pump for
circulation of ice slurry. Falling film system works similar
as scraper system, shown in Figure 7, but the motor for
auger has less friction. Falling film has higher speed of
rotary than scraper but has less power (E. Stamatioua et al.,
2005).
Vacuum freezing, as shown in Figure 8, use vacuum
pressure in evaporator to get low temperature for ice
crystal. This method needs vacuum pump. Asaoka et al.
2006 researched ice slurry using ethanol solution with
vacuum freezing. Hasegawa et al. 2002 used pure water
with this method (Hasegawa et al., 2002).
Fluidized bed, as shown in Figure 9, used bed heat
exchanger mechanism. Refrigerant flow through evaporator
in small pipes with flooded pipe, then
Figure 9: Fluidized Bed (Michel Barth)
Figure 10: Supercooled Slurry Ice Production
(SlurryICE TM
Manual Book)
Figure 11: Super Cooled Water Jet (T A Mouneer et
al., 2011)
ice slurry will separate on certain diameter as mesh
gravitation system. Ice fraction in small pipe is pushed by
pump (Pronk et al., 2005).
Supercooled Slurry Ice Production, as illustrated in Figure
10, is ice slurry producing method with supercooling.
Water flow with low velocity can be cooled to bellow of
freezing temperature without ice forming. Before leave
evaporator flow of supercooled water is teased with ice
crystal forming. Formed ice fraction depends on
supercooling solution leave evaporator.
Figure 11 illustrates a method develop by T.A. Mouneer et
al. (2011). Supercooled water jet applies water jet method
to improve velocity of solution. Pump is a main component
for circulation. The model works without shaft auger.
Different than others, in Direct Contact Heat Transfer
model refrigerant is blended with solution using coolant
nozzle. In application, the system can produce 40% ice
fraction (N.E. Wijeysundera et al., 2004).
Figure 12: Coolant Nozzle (N.E. Wijeysundera et al.,
2004)
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2012
245
Figure 13: Ice slurry installation for Fishing (M. J Wang et
al., 1999)
5. ICE SLURRY FOR FISH COOLING
There are three aspects for maintenance fish quality, viz.
cooling, handling and cleaning. Handling is effort to avoid
bacteria with cutting broken area. The second is cleaning
bacteria source with hygienic equipment after get fish from
sea, and the last is cooling system. However from all steps,
cooling has dominant effect for fish quality.
Traditional fishing still uses block ice, commonly, rather
than ice flake. Schematic diagram for ice slurry installation
is shown in Figure 13. Today, ice slurry become popular
applied to keep fish quality in developing countries. Ice
slurry can avoid air between fish and ice slurry, so fish
cooling become faster because of larger contact surface,
and slowing growth of bacteria (M.J Wang et al. 1999).
Ice slurry with depressant temperature made active protein
function and protect probiotics good from heat risk
(T.Vajda, 1999). Ice slurry has three times cooling faster
than ice flake to get 2oC (J Paul, 2002). Ice slurry does not
injure the fish because of soft ice texture (Pineiro et al.,
2004). Ice generator works as heat exchanger for water and
refrigerant commonly called harvest tank where ice particle
is formed. Harvest tank consists of scraper and auger shaft.
Sea water is injected to push ice fraction above liquid with
different density. Ice fraction
Table 1. Time for good quality of fish with storage
temperature (Masyamsir, 2001).
Storage temperature Time for good quality
16 oC 1-2 days
11 oC 3 days
5 oC 5 days
0 oC 14-15 days
with low density goes to third tank. Ice fraction in
third tank is added by sea water because of technical reason
while still stirred by scraper. Wang et al., 1999, reported ice
slurry application in fisheries.
Table 1 lists limitation of time when fish kept in a certain
low temperature. Lower temperature is better in getting
longer time for fish quality.
6. ICE SLURRY FOR INDONESIAN FISHIERY
Indonesia as archipelago country has 18,306 islands which
are united by ocean with length of coast line of 81,000 km.
Indonesia is the 4th
largest fishery county has fishery
potential of around 6.4 million/years (Dahuri et al., 2002).
Although Indonesia has huge potential in sea, many
fishermen are still far away from their prosperity. Indonesia
Fishery Minister and BPS in 2010 reported amount 7.87
million destitute people spread in coastal area. It is means
that totality fishermen are poor. Therefore, using
advantages ice slurry for Indonesian fishermen are
supposed can improve their economic income.
In 2010, from 590,352 fishing boats, just 6,370 units (less
than 2%) can be classified as modern boats, and they are
more than 30 GT. Inboard motor boats are 155,922 units
(26%), outboard motor boats are 238,430 units (40%), and
189,630 units (32%) are sail boats (KKP, 2010). Outboard
motor boat with simple equipments for fishing of traditional
fishermen just can catch fish in coastal area during 7 to 9
months/year. It is low productivity and there is reduction of
209 kg/month. They use unfair profit sharing system
between owner and workers. This condition makes number
of new fishermen utilizing fishery source is just 69.68%
(Yonvitner, 2007). New technology such as ice slurry
generator portable suppose can repair fish cold chain of
Indonesian fishermen.
In several survey for community engagement program
supported by grant of Hibah Pengabdian Masyarakat 2012
from DPRM University of Indonesia, authors concluded
some considerations must be taken for ice slurry application
in outboard motor boat, such as:
a. Ship stability
b. Increasing of draft (capacity)
(a)
(b)
The 5th IMAT, November 12 – 13th
2012
246
(c)
Figure 14: (a) Fishing boat 20 GT with inboard motor, in
Muara Angke, North Jakarta, (b) Fishing boat
with outboard motor, in Tidung Island,
Kepulauan Seribu, Jakarta, (c) Fishing boat with
outboard motor, in Balongan, Indramayu
c. Average production of fish
d. Duration and distance for fishing
e. Wide of close area and open area in board to place
equipments and activity area
f. Economic level
g. Local wisdom of fishermen
Figure 14 shows sample of some local boats used by
Indonesia fishermen.
7. SEA-WATER ICE SLURRY
Salinity is main parameter denote salt content in water.
Salinity unit is ppt (part per ton). NaCl (gram) contain 1000
gram sea water (Wibisono, 2004). Salinity may depend on
tidal, rainfall, distillation, and topography. Figure 15 shows
chart
Figure 15: Salinity chart (Calor M Lalli,.2006)
Table 2. Ion in sea water salinity of 35 ppt (Calor M
Lalli,.2006)
Ion
Concentration
(g/kg) Weight (%)
Chloride (Cl-) 18.98 55.04
Sodium (Na+) 10.56 30.61
Sulphate (SO42-) 2.65 7.68
Magnesium (Mg2+) 1.27 3.69
Calcium (Ca2+) 0.4 1.16
Potassium (K+) 0.38 1.1
Bicabonate (HCO3-) 0.16 0.41
Bromide (Br-) 0.07 0.19
Borate (H3BO3) 0.03 0.07
Strontium (Sr2+) 0.01 0.04
Figure 16: Effect of concentration of NaCL with
Freezing Point (K. S. Hilderbrand., 1998)
of salinity for location around the world.
The highest component in sea water is NaCl. Table 2
explain composition of ion in sea water. NaCl has number
of 85.65% of weight gram sea-water. Therefore, calculating
of ice slurry can use properties of NaCl.
K. S. Hilderbrand (1999) presented correlation of NaCl in
solution with freezing temperature, as shown in Figure 16.
The figure is discovered eutectic point at -21.1 oC. Eutectic
temperature is limit temperature for NaCl solution that
direct change from solution to solid salt and ice. The
mixture percentage is called mixture eutectic.
Table 3. Effect of salinity with freezing temperature (Feistel
et al., 2008)
The 5th IMAT, November 12 – 13th
2012
247
Salinity (ppt) Freezing temperature (oC)
5 -0.269
10 -0.536
15 -0.803
20 -1.074
25 -1.348
30 -1.625
35 -1.908
40 -2.195
45 -2.487
50 -2.784
55 -3.087
60 -3.396
65 -3.711
70 -4.033
Table 4. Effect of salinity with freezing temperature (The
Practical Salinity Scale 1978)
Salinity (ppt) Freezing temperature (oC)
0 0
5 -0.274
10 -0.542
15 -0.812
20 -1.083
25 -1.358
30 -1.638
35 -1.922
40 -2.212
Feistel et al. (2008) have been discovered relation salinity
with freezing temperature, as listed in Table 3. The
Practical Salinity Scale 1978 and the International Equation
of State of Seawater 1980, Unesco Technical Papers in
Marine Science No.36 determine relation of salinity with
freezing temperature, as listed in Table 4.
Ice slurry is mixture of ice particle and solution, so it‘s
important to know sea water properties for calculate
parameters of ice slurry thermofluids. Mustafa H et al.
(2010) reported about sea water thermofluid. Figure 17
shows (a) effect of temperature on density, (b) effect of
temperature on dynamic viscosity, (c) effect of temperature
on thermal conductivity, (d) effect of temperature on
specific heat, and (e) effect of temperature on specific
enthalpy.
(a)
(b)
(c)
(d)
(e)
Figure 17: Experimental results of Mustafa H. et al.,
2010
The 5th IMAT, November 12 – 13th
2012
248
A. Melinder (2010) proposed some correlation of ice slurry
properties as function of temperature.
(2)
(3)
(4)
(5)
Ice fraction can be calculated using correlation proposed by
Jean-Pierre Be´de´carrats et al., 2009 and Cecilia Hägg,
2005.
(6)
Total enthalpy can be calculated using correlation proposed
by T. Kousksou et al. (2010) and A. Melinder et al. (2005).
(7)
Thermal conductivity can be calculated using correlation
proposed by Taref (1940).
(8)
D.G. Thomas (1965) dan Jacques Guilpart et al. (2006)
proposed correlation of ice slurry viscosity.
(9)
Moreover, Kasza and Hayashi (2001) have reported micro
scale of shape and surface of ice slurry. Kasza and Hayashi
(2001) and Cecilia Hägg (2005) explained that ice slurry
with big diameter and rough surface had bigger pressure
drop.
8. DISCUSSION
Research on sea-water ice slurry characteristics, including
the formation, flow, heat transfer, and others related topics
is necessary in order to develop ice slurry generator. A
compact and cheap product of ice slurry generator is
supposed can improve fishing productivity, especially for
traditional fishermen in Indonesia. Using appropriate
product, fishermen can work at more far location from coast
line with longer time because they can produce ice slurry
while fishing. Research and implementation of sea-water
ice slurry generator have been doing by Lab. of
Refrigeration Engineering, Department of Mechanical
Engineering, Universitas Indonesia. Optimizing in some
important components and topics, such as evaporator,
auger, scraper, flow, and energy source, is necessary in
developing product.
ACKNOWLEDGMENT
The work described in this paper was supported by grants
of Hibah Madya 2012 from DRPM University of Indonesia
and Hibah Pengabdian Masyarakat (Community
Engagement Grant) 2012 from DRPM University of
Indonesia.
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Improving Hydrogen Storage Capacity on Lithium-Doped Carbon
Nanotubes Using Molecular Dynamics Simulation
Nasruddina, Engkos A. Kosasih
a, Supriyadi
b and Abdul Jabbar
a
aFaculty of Engineering
University of Indonesia, Depok 16424
Tel : (021) 7270011 ext 51. Fax : (021) 7270077
e-mail : [email protected] bFaculty of Engineering
Trisakti University of Indonesia, Jakarta 11440
Tel : (021) 5663232 ext 8431. Fax : (021) 5665840
e-mail : [email protected]
ABSTRACT
The technical challenge for the coming years is to found new
alternative energy sources that are environmentally friendly
and continuously renewable. Hydrogen has been under
investigation for its potential use as an alternative energy,
because of non-polluting and high thermal efficiency.
Hydrogen has a wide range of applications, as both stationary
and mobile use. The use of hydrogen require development of
efficient storage method. Adsorption method is considered as
a very efficient and safety. For this purpose, to improve the
performance of carbon nanotubes have attracted a lot of
interest due to their high volume and pore size distribution.
Experimental research on carbon nanotubes generally still too
expensive, it is necessary to be supported by another method
ie Molecular Dynamics Simulation. The use of metal doping
to increase the capacity of hydrogen storage has been done,
and in this study will be conducted using lithium. The aim of
this work is to build and to make two simulation models
using LAMMPS to study hydrogen storage capacity on
lithium-doped and without lithium-doped carbon nanotubes.
Doping metal presentation will focus on weight, position, and
distribution of elements on the carbon nanotubes substitution
and its influence on adsorption of hydrogen. Result of two
models simulation can be compared, for the simulation with
temperature range 253 to 293 K in the range pressure from 1
to 12 atm, hydrogen storage capacity of li-doped SWNT can
be enhanced significantly.
Keywords : CNT, hydrogen, adsorption, metal doping,
molecular dynamics, LAMMPS
1. INTRODUCTION
The generation and utilization of clean and renewable energy
is one of major worldwide concern. Currently, energy
consumption is intimately linked with CO2 emissions, a
significant human contributor to climate change [1].
Hydrogen is regarded as a pollution-free energy carrier.
Many investigations have been therefore carried out in recent
years for the utilization of hydrogen as an effective and cheap
storage system. Hydrogen adsorption in porous solids is one
of the alternatives that has been investigated [2]. Different
types of carbon nanostructures were investigated for their
suitability as hydrogen storage material. In addition to
compressed hydrogen gas and liquid H2 vessels, hydrides,
chemical hydrides or light element for hydrogen storage [3].
To obtain the maximum storage capacity of carbon nanotubes
is still interested for many researchers. Dilon, 1997,
published on the SWNT hydrogen storage capacity is not
pure, at room temperature and moderate pressure. By using
thermal desorption spectroscopy is obtained 12.01 wt%.
Based on the results obtained Dilon, Hirscher predict if the
SWNT was purified, the hydrogen storage capacity will
increase to about 5-10 wt%. Measurement of hydrogen
storage capacity in this type of material, which has been
published in various journals ranging from 0.1 wt% to 67
wt%. Spectacular results up to 67 wt%, far exceeding the
target set by DOE in the amount of 6.5 wt% or by 62 kgH2.
M-3
, performed by Baker and Rodriguez on the type of
carbon nanofiber at a pressure of 110 atm, at room
temperature and monitored for 24 hours [4].
Callejas and colleagues (2004) conducted a study to improve
the hydrogen storage capacity of SWNT pasa through the
reduction of the sample. SWNT produced by arc-discharge
modified using Ni/Y as a catalyst with a different
percentage. SWNT are synthesized by using electric-arc-
discharge method has a ratio of metal elements and catalyst
2/0, 5 and 4/1 (Ni / Y). Next examined the metal content and
value BETnya and obtained: Ni / Y 2/0, 5 (266), Ni / Y 2/0, 5
+ 3500C / l at (728), Ni / Y 4/1 (207), Ni / Y 4/1 + 3500C / l
at (585) [5].
The effects of thermal treatments and palladium loading on
sorption characteristics of single-walled carbon nanotube
(SWCNT) samples were investigated by Kocabas et al
(2008). The thermal treatment experiments were carried out
in a temperature range of 300–8000C. The sorption
characteristics of nitrogen and hydrogen on the original, heat
treated and the palladium loaded samples were investigated.
The highest surface area samples obtained at 5750C were
loaded by 3.3, 6.3 and 10.1 wt% palladium and hydrogen
adsorption isotherms on these samples were obtained at 77.4
K. The hydrogen sorption capacities of the original and the
10.1 wt% palladium loaded samples were found to be 0.76
and 1.66 wt%, respectively [6].
Boron substitution in carbon nanotubes is investigated by
Sankaran et al (2008), maximum storage capacity of 2 wt% at
80 bar pressure is obtained for BCNT1, whereas pure carbon
nanotubes shows 0.6 wt% [2]. Shevlin and Guo (2008)
performed ab initio density functional theory simulations on
IMAT-UI 041
The 5th IMAT, November 12 – 13th
2012
252
titanium-atom dopants adsorbed on the native defects of an
(8,0) nanotube. Adsorption on a vacancy strongly binds
titanium, preventing nanoparticle coalescence (a major issue
for atomic dopants). The defect-modulated Ti adsorbs five H2
molecules with H2 binding energies in the range from -0.2 to
-0.7 eV/H22 , desirable for practical applications. From their
simulations is obtained indicate that this complex is stable at
room temperature, and simulation of a C112Ti16H160 unit cell
finds that a structure with 7.1 wt % hydrogen storage is stable
[7, 8].
Zubizarreta et al (2009) investigated the effect of nickel
distribution and content in Ni-doped carbon nanospheres on
hydrogen storage capacity under conditions of moderate
temperature and pressure. It was found that the nickel
distribution, obtained by using different doping techniques
and conditions, has a noticeable influence on hydrogen
storage capacity. It was found a higher storage capacity in
samples containing 5 wt.% of Ni. This is due to the greater
interactions between the nickel and the support that produce a
higher activation of the solid through a spillover effect [9].
Hydrogen storage by chemisorption on MWCNTs was
studied by Zuttela et al (2010). By oxidation treatment to
produce defects and subsequent loading with a Pd-Ni catalyst
significantly increased the hydrogen storage capacity up to
6.6 wt%. In the same manner Meiyan et al (2010)
investigated Li-doped charged SWNTs [10].
2. POTENTIAL AND INTERACTION MODELS Molecular dynamics simulation based on statistic mechanics
and statistic thermodynamics to simulate the particles‘
interactions and consists of several processing methods, such
as trajectories, position updates, the cut-off radius, and the
initial condition. Hydrogen-hydrogen and hydrogen-carbon
interactions are both modeled with Lennard-Jones potential.
For a pair of particles i and j separated by the distance r , the
interaction between them is given by [11, 12]:
(1)
where i and j denote hydrogen, carbon or lithium particles,
ε/k and σ are the energy and size potential parameter, are
obtained from the literature., which are 0.3158 K and 2.915 Å
for hydrogen, 0.026 K and 2.27 Å for lithium and 0.2327 K
and 3.4 Å for carbon, respectively. The cross interaction
parameters σij and εij are obtained from Lorentz-Berthelot
mixing rules [12].
(2)
Diameter of nanotubes can be calculated using simple
formula,
(3)
The number of hydrogen molecules, besides of simulation
box volume, its depends on the external temperature and
pressure. Hydrogen is real gaseous, relation between
volume, pressure and temperature can be expressed by
equation real gas:
(4)
Where P is the absolute pressure of gas (atm), V is volume
(m3), n is number of moles (mol), R is the ideal gas constant
equal to 8.314 (J/mol.K), T is temperature gas (K), a and b is
van der waals constant. For example, using this formula,
simulation at 5 atm and 253 K , the number of hydrogen
molecules is 160.
3. SIMULATION METHOD
We investigated two models of simulation. The first model,
simulation in (10,10) armchair SWNT with diameter about
1.375 nm and 1.122 nm length. The SWNT was simulated
using molecular dynamics at 253 – 293 K temperature for
pressure ranging 1 atmosphere to 18 atmospheres. In the
simulation, the nanotubes are placed in a 5.5385 x 6.413 x
12.243 nm simulation box.
In the second model, we build (8,8) armchair SWNT with
diameter about 1.155 nm and 3 nm length. The SWNT was
simulated using molecular dynamics at 253 – 453 K
temperature for pressure ranging 1 atmosphere to 11
atmospheres. Figure 3.1 shows the initial condition of atoms.
In the simulation, the nanotubes are placed in a 6.9625 x
6.9625 x 22.6915 nm simulation box. The lithium atoms are
placed surrounding the nanotubes. Proportion number of
atoms is 415 atoms carbon, 35 atoms lithium and 16 – 352
atoms hydrogen, respectively.
Periodic boundary conditions on the position of the atoms
are use in all directions to eliminate surface effect [6].
Figure 1. Initial position of the simulated system, atoms
C (cyan), Li (purple) and H (green).
In this study molecular dynamics simulations performed on
the with a variety chirality, diameter and length of CNT,
while the doping element used Lithium. The results are then
used to construct models of the CNTs. Interactions between
atoms in the simulation is computed using the Lennard-Jones
potential, the Coulombic and van der Waals forces. Lennard-
Jones potential dominates at the short distances so that the
distance between atoms in excess of the cut-off radius will be
ignored, while for the coulomb force is more dominant at
longer distances, so long as the charge between atoms is still
significant value calculated despite the relatively large
distance between the atoms.
The 5th IMAT, November 12 – 13th
2012
253
Figure 2. Final position of the simulated system, atoms
C (purple), Li (green) and H (cyan).
Purpose molecular dynamics simulation is to update the
position, direction and speed at any time, due to hit each
other, push each other, due to the forces acting on each
particle. The condition is most difficult is determining the
initial conditions that simulated the entire atom. Position,
distance and tolerance should be adjusted to the conditions of
system that exists. In this simulation to determine the initial
conditions through several stages. To specify one or several
groups of molecules made in the program Avogadro, the
coordinates of all atoms are stored in the form file.pdb, while
to build a CNT can use a variety of software including
Generator CNT, CNT Wrapping and VMD. In this simulation
using the existing facilities Nano Builder in VMD.
Coordinates of all atoms obtained is stored in the form of
file.pdb. Furthermore, to determine how many molecules are
desired, the position where and how much tolerance between
molecules, used packmol program.
To update the position of each atom is run by software
LAMMPS (Large-Scale Atomic / Molecular Massively
Parallel Simulator). Input from this program in the form of
the main program and each of the data presented in notepad
++ format. LAMMPS program is only doing calculations to
update the position, whereas longer required to display the
results of visualization programs such as Atom Eyes, Pizza,
and VMD. In this work VMD program was used as a tool for
visualization
4. RESULTS AND DISCUSSIONS
We make two simulation models of hydrogen adsorption in
SWNT. System conditions is emphasized on the effect of
temperature and pressure in the simulation space. For the first
model, system temperatures ranged from 253 K to 293 K
with a pressure of 1 to 18 atm. In this variation to change the
pressure or temperature is done by varying the amount of
hydrogen or can also be done by changing the volume of the
simulation box. For the second model, system temperatures
ranged from 253 K to 453 K with a pressure of 1 to 12 atm.
In this variation to change the pressure or temperature is done
by varying the amount of hydrogen or can also be done by
changing the volume of the simulation box. Hydrogen
storage capacity of SWNTs were observed for 700,000 until
1,000,000 running steps.
Table 1. Storage capacity (% weight) at first model simulation
Storage (% weight)
Pressure Temperature (K)
(atm) 253 273 293
1 0.14 0.05 0.05
2 0.14 0.05 0.18
4 0.32 0.23 0.18
6 0.46 0.28 0.18
8 0.42 0.46 0.28
10 0.60 0.51 0.46
12 0.87 0.78 0.69
14 1.01 0.78 0.55
16 1.05 0.64 0.69
18 1.14 0.87 0.78
Table 1 shows the hydrogen storage in SWNT Li-doped at
pressure in the range between 1 atmosphere to 12
atmospheres and temperature between 253 to 453 K.Table 2
shows the hydrogen storage in Li-doped SWNT at pressure in
the range between 1 atm to 12 atm and temperature between
253 to 453 K.
Table 2. Storage capacity (% weight) at second model simulation.
Storage (% weight)
Pressure Temperature (K)
(atm) 253 273 293 313 333
1 0.35 0.31 0.27 0.27 0.24
2 0.74 0.70 0.66 0.47 0.43
4 0.97 0.82 0.66 0.63 0.59
6 1.43 1.20 1.17 1.17 1.05
8 2.49 2.19 2.08 1.89 1.74
10 2.75 2.42 2.19 2.16 2.16
12 3.38 3.09 3.42 3.16 2.72
Table 2. (continued)
Storage (% weight)
Pressure Temperature (K)
(atm) 353 373 393 413 433
1 0.24 0.20 0.20 0.20 0.16
2 0.43 0.43 0.39 0.39 0.35
4 0.59 0.51 0.51 0.47 0.35
6 0.97 1.05 0.97 0.86 0.86
8 1.63 1.63 1.51 1.36 1.32
10 2.12 2.31 1.89 2.04 1.51
12 2.68 2.34 2.04 1.93 2.16
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254
Figure 3 presents results for hydrogen storage in CNT
without Li-doped in the range from 253 K until 293 K and
pressures in the range from 1 to 18 atm.
Figure 3. The hydrogen adsorption in various temperature
and pressure (model 1)
Figure 4 and Figure 5 presents results for hydrogen storage in
Li-doped CNTs in the range from 253 K until 453 K and
pressures in the range from 1 to 12 atm. From three tables
above, it can be seen that the hydrogen adsorption increased
with a decreasing temperature at constant pressure, and at
constant temperature the greater pressure the greater
hydrogen storage capacity will be obtained. In order to
increase the hydrogen storage at a constant temperature the
pressure should be increased, and similarly, to increase the
hydrogen storage at a constant pressure, the temperature
should be decreased.
Figure 4. The hydrogen adsorption in various
temperature (model 2).
Figure 5. Variation of hydrogen storage at various pressures
(a)
(b)
(c)
Figure 6. Storage capacity (% wt) in Li-/Non-Li-Doped
SWNT. (a). Simulation at temperature 253 K, (b).
Simulation at temperature 273 K and (c).
Simulation at temperature 293 K
Figure 6. (a), (b) and (c), show that hydrogen storage
capacity of Li-doped SWNT can be enhanced significantly. It
can be said that doping effect of the second model is better
than the first model.
Besides that, several parameters , diameter, length and metal
doping are the dominant parameters that influence the
hydrogen storage capacity.
4. CONCLUSION
Generally the second model, Li-doped SWNT, in the range
from 253 K until 453 K and pressures in the range from 1 to
12 atm, has better storage capacity than the model without
Li-Doped SWNT. Diameter, length and metal doping are the
dominant parameters that influence the hydrogen storage
capacity.
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REFERENCES
[1] S.A.Shevlin and Z.X.Guo, High-Capacity Room-
Temperature Hydrogen Storage in Carbon Nanotubes
via defect-Modulated Titanium Doping, J. Phys. Chem. C
2008, 112, 17456-17464.
[2] A. Toprak and T. Kopac, Surface and Hydrogen Sorption
Characteristics of Various Activated Carbons Developed
from Rat Coal Mine (Zonguldak) and Antrachite,
Separation Science and Engineering. Chinese Journal of
Chemical Engineering, 19(6) 931-937 (2011).
[3] M. Hentsche, H.Hermann, D.Lindackers, and G. Seifert,
Microstructure and low-temperature hydrogen storage
capacty of ball-milled graphite, International Journal of
Hydrogen Energy 32 (2007) 1530 - 1536.
[4] Sang-Hun Nam, Seong H. J, Soon-Bo Lee, and Jin-Hyo
Boo, Investigation of hydrogen adsorption on single wall
carbon nanotubes, Physics Procedia 32 (2012) 279 – 284
[5]. Callejas, M.A., et al, Hydrogen adsorption studies on
single wall carbon nanotubes, Carbon 42 (2004) 1243 –
1248
[6]. Kocabas, S., Kopac, T., Dogu G., Effect of thermal
treatments and palladium loading on hydrogen sorption
characteristics of single-walled carbon nanotubes,
International Journals of Hydrogen Energy 33 (2008)
1693 – 1699.
[7]. M. Sankaran, B. Viswanathan, S. Srinivasa Murthy,
Boron substituted Carbon Nanotubes - How appropriate
are they for hydrogen storage?, International Journal of
Hydrogen Energy 33 (2008), 393 – 403.
[8]. S. A. Shevlin and Z. X. Guo, High-Capacity Room-
Temperature Hydrogen Storage in Carbon Nanotubes
via Defect-Modulated Titanium Doping, J. Phys. Chem.
C 2008, 112, 17456 – 17464
[9]. L. Zubizarreta, J.A. Menéndez, J.J. Pis, A. Arenillas,
Improving hydrogen storage in Ni-doped carbon
nanospheres, International Journal of Hydrogen Energy
34 (2009) 3070 – 3076
[10]. A. Zuttela, Ch. Nutzenadela, P. Sudana, Ph. Maurona,
Hydrogen sorption by CNT and other carbon
nanostructures, Journal of Alloys and Compounds 330–
332 (2002) 676–682
[11]. C. Gu, G.H. Gao, Y.X. Yu, and Z.Q. Mao, Simulation study of
hydrogen storage in single walled carbon nanotubes,
International Journal of Hydrogen Energy 26 (2001) 691 -
696.
[12]. Banerjee, S., Molecular Simulation of Nanoscale
Transport Phenomena, Desertasi, Virginia Polytechnic
Institute and State, 2008.
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2012
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Performance of Thermoelectric and Heat Pipe Refrigerator Cooling
System
Ali A. Sungkara, Firman Ikhsana, Saripudinb, M.Afin Faisola, M. Zilvan Beya, Nandy Putraa
aDepartment of Mechanical Engineering, bDepartment of Electrical Engineering
Faculty of Engineering Universitas Indonesia, Depok 16424
Tel : (021) 7270011 ext 51. Fax : (021) 7270077
E-mail : [email protected]
ABSTRACT
Most of the refrigerators commonly use the conventional
refrigeration system known as Vapor Compression
Refrigeration System becoming a big issue lately due to
ozone depleting substance it uses as the refrigerant. This
paper will shows step by step of an experiment with the
objective of constructing a refrigeration system based on
thermoelectric which is reliable and compete able with the
Vapor Compression Refrigeration System. The designing of
this refrigeration system shows attention to the environment
that is combined with the knowledge so the environmental
friendly technology can be applied. The performance of
thermoelectric refrigerator was conducted in variation
input power (40Watt, 72Watt, and 120Watt) and operated
in ambient temperature and cooling load of water 1000mL
to investigate the characteristic of system, the performance,
and also the COP values. The COP values is decrease
increasing of cooling load, QL. The best actual COP is
0.182 reached when the refrigerator operated at input power
40W. The result, it showed that decreasing of temperature
ambient affects the decreasing of cabin temperature.
Thermoelectric and heatpipe refrigerator cooling system
can reach cabin temperature with power 120 Watt (8.73A,
14V) produces temperature of compartment is 10.63˚C
indicates effective performance work-based thermoelectric
applications. Keywords: heat pipes, thermoelectric, thermoelectric
refrigerator, COP, environmental friendly.
1. INTRODUCTION
In engineering and applications, we have known cooling
system using vapor compression refrigeration system and
absorption refrigeration system. Vapor compression
refrigeration system, which is the most common used
refrigerating system has the advantages due to its high
cooling capacities and COP values, but this system has an
environment issues due to the ozone depletion causes by it
refrigerant, and also its unstable working temperature[1].
Absorption system has the advantage of its un noisy
working condition because this system is a-Heat driven
refrigerating system. Absorption system can use an exhaust
heat from other system as the heat generator so it can
increase its efficiency. Absorption system has some
shortages due to its low COP values and bulky design[2].
Thermoelectric system has the advantages that it is a solid
state device which is very practical use and there is no work
fluid applied that makes it environmentally friendly, the
working temperature of this system is stable, easy to
control, and the reliability of system is high so it has a long
life time. Nevertheless, this system shows a low COP
values and expensive[1]. Thermoelectric technology has
been widely used for both cooling and power generation[2].
Thermoelectric cooling have been applied in wide range of
application, from the portable vaccine carrier box, surgical
device, cooling system for electronic equipment, food
processing equipment, military & aerospace instruments[3-
11]. Thermoelectric device has a given operating
temperature range beyond which its operation may cease.
For this reason, all thermoelectric coolers (TECs) require
heat sinks in order to dissipate the energy generated or
absorbed at the two junctions. Design and selection of a
heat sink is crucial to the overall operation of a
thermoelectric system, the heat sink should be design to
minimize the thermal resistance. The heat transfer between
thermoelectric module and heat dissipation device may be
further improved by use of the heat pipe [12]. Heat pipe is a
device with a very high thermal conductivity and typically
consist of a sealed tube with an internal wick. The heat pipe
is charged with refrigerant, such as water, ethanol or
methanol and nano fluids. Heat pipes are widely adopted
for their high efficiency, cooling capability, reliability and
shape flexibility [13-17].
II. PRINCIPAL OF WORK AND
CALCULATION
2.1 Thermoelectric
Each thermoelectric module consists of two or more
semiconductor which is connected electrically in series and
thermally in parallel. All of the thermoelectric elements are
attached to a pair of ceramic substances. This ceramic
substances act as the body that mechanically holds the
structure of the connections and as the insulator electrically.
The semiconductors used are the ―N‖ and ―P‖ type made of
Bismuth Telluride. The arrangement of these
semiconductors causes the heat transfer when a current
flows between upper and lower ceramic trough every
semiconductor element P and N. The ―N‖ type
semiconductor that has been doted becomes surplus of
electrons while the ―P‖ type semiconductor that has been
doted becomes lack of electrons. The surplus of electron at
the ―N‖ type and the hole produced by the lack of electron
at the ―P‖ type semiconductors become the way of
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transferring heat in the thermoelectric materials. Figure 1
shows the heat transfer caused by the electrical current (I)
that
Figure 1. Working scheme of Thermoelectric (Phillip C.
Watts, 2011)
NOMENCLATURE
Qc = Heat pumping capacity (J)
P = Input power (Watt)
N = Number of termocouple
α = Seebeck coefficient (V/K)
G = Geometry factor-area/length of thermoelectric
(cm)
ρ = Resisitivity (ohm cm)
K = Thermal conductivity (W/m.K)
I = Electrical current (A)
Tc = Cold side temperature (°C)
Th = Hot side temperature (°C)
Tm = Average temperature (°C), where
Tm = (°C)
(1)
Z = Figure of merit (K-1
)
∆T = Temperature difference of cold side temperature
and hot side temperature (°C)
applied to the thermoelectric module. Most of
thermoelectric made with an equal number of ―P‖ and ―N‖
type semiconductors in couple.
.
Figure 2. Electric current scheme in thermoelectric
(source: http://www.tec-microsystems.com)
Figure 2 shows the electron flow from the ―P‖ type
semiconductors that lacks of energy, absorbing heat from
cooled space and then electron flows to the ―N‖ type
semiconductors. The ―N‖ type semiconductor became
surplus of energy and dissipates the excess energy to the
environment. The amount of heat flux (heat pumped by
thermoelectric) will equal to DC current flow which is
applied. Controlling the amount of DC current, we can
control the heat flow and temperature as well.
2.2 Heat Pipes
Heat pipes generally consisting of a tube are a sealed tube
at both ends. The tube is made of metal which can absorb
and deliver heat (thermo conductive metal) very well; such
as aluminum or copper. Inside this tube contains a liquid
coolant (such as water, ethanol, or mercury) and a number
of gas from the liquid. On the side of the inner tube there is
an axis with the character of capillary walls that serves to
drain the steam produced by cooling liquid that evaporates
due to receive the amount of heat. Heat pipe consists of
three parts: the evaporator is located at one end, where the
heat is absorbed and the liquid is evaporated, then
condenser at the other end, where the vapor condensed and
the heat is released; final adiabatic section located between
the two. Adiabatic is a state where there is no heat transfer
to or from the surrounding environment. Adiabatic can
occur under two possibilities: the system perfectly isolated,
or the temperature inside and outside the same.
Figure 3. Working scheme of heat pipe
(source : http://heatpipe.nl)
2.3 COP Calculation
COP is ratio of work or useful output to the amount of work
or energy . There are two equation used to determine the
COP values for thermoelectric.
(2)
.....(3)
(4)
III. EXPERIMENTAL SET UP
Determining the characteristic of thermoelectric
refrigeration system, the experiment was conducted to
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collect the data from every important spot from system
Testing was conducted to obtain performance data to
determine the performance of refrigeration systems work
performance. The test procedure is the thermocouples are
placed on each side of the thermoelectric, cabins, and
outside the cabin as a temperature sensor. The whole
thermocouple is connected to the module is used as a
National Instrument data acquisition. Later, it connected to
a computer. Perform set-up and testing of the software
labview each sensor to ensure that the sensors are working
normally. Monitor the temperature and wait until all sensors
temperatures approaching ambient temperature. Connecting
the power cable thermoelectric cooling system with power
supply as the source of power used. After all the above
procedures have been implemented, data collection is done
with the initial conditions the system off for five minutes.
Furthermore, the power supply is turned on so that the
system works. Data is collected for four hours, including
five minutes of the initial conditions. Figure 3 is the scheme
of experiment.
Figure 4. Design of experiment performance of
Thermoelectric and Heat Pipes Refrigerator
Figure5. Construction of Cooling System
From figure 5, can be seen that thermoelectric system
consists of several components that have different
functions. The following will explain the components used
in thermoelectric systems are:
1. Fan brushless DC 12 V
2. Heat Pipes PC Cooler
3. Thermoelectrics-Peltier element
4. Heat sink
Source of six thermoelectric cooling uses the Peltier
elements are arranged in series. Heat removal process is
done by pumping heat in the form of a fan with power
13Watt for 6 fans. Heat flows through the air as
intermediaries and excreted through the bulk head under the
system. The end result made the development of tools that
can be seen in Figure 6.
(a)
(b)
Figure 6. The experiment of thermoelectric and heat pipes
refrigerator. (a) in front of view, (b) behind view.
1. Thermoelectric and heat pipes refrigerator
2. Power supply
3. Thermocouple (sensor of temperature)
4. Module DAQ data acquisition National
Instrument
5. PC (data display)
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IV. RESULT
4.1 Performance of thermoelectric cooling system
temperature cold and hot side
Figure 7. Measurement temperature cooling systems cold
and hot side at variation of input power by Flir i50
thermograph
Figure 7 shows the thermographic (temperature
distribution) of thermoelectric placed on heatsink and
heatpipe with a variety of voltage and current supplied
power supply.
1. Figure 1.1 at 8V, 1.3A shows the temperature, Thotside =
60˚C and Tcoldside = -11.4˚C,
2. Figure 1.2 at 12V, 2.93A shows the temperature, Thotside
= 40.10˚C and Tcoldside = -8.50˚C, and
3. Figure 1.3 at 14V, 3.36A shows the temperature, Thotside
= 48.70 C and Tcoldside = -13.60˚C
It explains that the thermoelectric works by controlling the
temperature difference. In applications as a cooling system
is needed both in terms of performance so that will be
utilized thermoelectric the lower temperature. Higher
temperature will be controlled so that the lower temperature
can reach the minimum temperatur side. Temperature
control is optimized to move higher into the environment
using the principles of forced convection heat transfer by
fans.
4.2 Experimental Result of Thermoelectric and Heat
Pipes Refrigerator
The experiment to determine performance of heat pipes
and thermoelectric refrigerator was conducted in variation
of the input power 40 Watts (8.2V, 4.9A), 72 Watt (11.4V,
6.29A), 120 Watt (14V, 8.73A) with 1000mL water cooling
loads and operated at ambient temperatures range 27-30˚C.
4.2.1 Performance of Refrigerator Operated in Input
Power 40W
Figure 8 shows the cabin temperature conditions and water
temperature of the cooling load 1000mL at ambient
temperature conditions. It can be seen that the refrigerator
can be operated on low power. Within 35 minutes the
temperature decreased significantly from the cabin
temperature ambient. Then, the decrease of temperature
fluctuation is small, ranging from 0.01 to 0.03˚C. After 2
hours of operation, the cabin temperature increase in
fluctuates around 2˚C due to higher ambient temperatures.
Ambient temperature increases due to heat from the cooling
system is wasted to the environment and cause the
temperature around the test increased slightly. In the end,
the cabin temperature reached at T = 16.06˚C. Meanwhile,
the temperature of the water cooling loads tend to decrease
linearly up to 150 minutes of operating time. Then the
temperature tends to fluctuate with the small end
temperature reached 15.3˚C. The end of the cabin
temperature slightly greater differences compared to the
final temperature of the cooling load due to forced
convection in the have ability to maintain than the air
temperature. It means that water have a good stability than
air by fluctuation of temperature.
Figure 8. Performance of Thermoelectric and Heat Pipes
Refrigerator in various input power 40W
4.2.2 Performance of Refrigerator Operated in Input
Power 72W
Figure 9 shows the performance conditions refrigerator
operated at input power 72Watt with 1000mL water cooling
load remains at a stable ambient temperature conditions
ranging from 27-28˚C. It can be seen that. cabin
temperature has decreased dramatically in the first hours of
operation and then remained stable until the temperature
reaches the end of 12.67˚C. In addition, the cooling load
temperature chart also had a significant reduction in cabin
temperature approaching 160 minutes of operation and
produce a final temperature that is slightly above the
temperature of the cabin 13.17˚C. This suggests that the
performance refregerator operated at 72W power has a
higher temperature difference than the power of 40W at
nearly the same ambient temperature conditions.
Figure 9. Performance of Thermoelectric and Heat
Pipes Refrigerator in various input power
72W
1.1 1.2 1.3
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4.2.3 Performance of Refrigerator Operated in Input
Power 120W
Figure 10. Performance of Thermoelectric and Heat
Pipes Refrigerator in various input power
120W
Figure 10 shows the chart performance of heat pipes and
thermoelectric refrigerator with input power of 120 W at
ambient temperature of 27-28˚C. It explains that the cabin
temperature has decreased significantly during the 40
minutes at the start of the operation. Then, tend to be stable
until it reaches the end temperature is 10.63˚C. Decreasing
of cabin temperature is proportional in decrease of water
cooling load temperature. It decreases linearly until reaches
the end of the temperature is 10.96˚C. Decreasing of
temperature in both conditions is the largest compared to
operation on the input power 40W and 72W at the same
ambient temperature conditions. This illustrates that the
value of the temperature difference between the
environment and the cabin refrigerator depends on the
performance of the cooling system works through a given
input power. The input power will affect the inseide of
cabin temperature that happen heat transfer in forced
convection between coldsink and the air of the cabin.
4.3 Performance of Cabin Temperature in Variation of
Input Power
Figure10. Graph testing performances of cabin temperature
the power variation
Figure 10 above shows the temperature conditions in the
test cabin refrigerator with the power variation. Testing was
done by giving the power 40, 72 and 120Watt on the
refrigerator, with environmental temperature 28°C and
loaded by water with a volume of 1000ml. By looking at
these graphs can be concluded that the higher power that is
given then the lower the cabin temperature. As said in
previous discussions, this happens in the cabin heat transfer
by forced convection.
4.4. Result of COP Calculation
COP result calculation is given below show that the
performance of cabin temperature due to variation of
cooling load and ambient temperature.
Table 1. COP Value of thermoelectric and heat pipes
Refrigerator
abased on equation 4 bbased on equation 2 with the Qc values was based on equation 3
Table 1 shows performance comparison of thermoelectric
refrigerator in input power variation of running condition.
The COP values are decrease with increasing the input
power. The decreasing of actual COP values is caused by
the value of Qc that is tend to increase due to increasing of
cooling load, QL while the input power applied stay in small
difference of cooling capacity. It means the highest COP
values (optimum COP and actual COP) reached when the
refrigerator operated at input power 40W in constant
1000mLwater cooling load. The lowest of actual COP when
operated at input power 120W and cabin temperature is
10.63oC.
ACKNOWLEDGMENT
We would appreciate Assistant of Professor Ridho
Irwansyah, S.T., M.T. dan Wayan Nata, S.T., M.T. at the
Applied Heat Transfer Laboratory Faculty of Engineering
Universitas Indonesia for helpful advices and kindly
providing us with thermoelectric and its measurement
technique. We are also thankful to Dikti by the project in
Students Creativity Programme for funding this research
and member of Universitas Indonesia‘s Robotics Team for
technical support.
V. CONCLUSION
1. Decreasing of temperature ambient affects the
decreasing of cabin temperature
2. Increasing of input power applied to system causes the
increasing temperature difference of system so that it
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261
affects the cabin and cooling load temperature of
thermoelectric and heatpipes refrigerator.
3. The measurements of cabin temperature with power 120
Watt (14V, 8.73A) produces temperature of
compartment is 10.63˚C in steady temperature indicates
effective performance work-based thermoelectric and
heat pipe cooling applications.
4. The COP values is decrease caused by the value of Qc
that is tend to increase due to increasing of cooling load,
QL. The highest COP values (optimum COP and actual
COP) reached when the refrigerator operated at input
power 40W in constant 1000mLwater cooling load.
5. As the result, we can see that thermoelectric and heat
pipes refrigerator has potency to replace the
conventional refrigerators. Non-CFC refrigeration
products could be replacement of the box of
conventional cooling or heating due its ability to keep
the temperature steady so that the materials can be
stored in optimal conditions.
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thermoelectric air-conditioners versus vapour compression
and absorption air-conditioners,Applied Thermal
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[2] Kong Hoon Lee, Ook Joong Kim, Analysis on the cooling
performance of the thermoelectric micro-cooler,
International Journal of Heat and Mass Transfer 50 (2007)
1982-1992.
[3] Nandy Putra, Desing, Manufacturing And Testing Of A
Portable Vaccine Carrier Box Employing Thermoelectric
Module And Heat Pipe, Journal of Medical Engineering &
Technology, 33 (2009) 232-237.
[4] Nandy Putra et al, The characterization of cascade
thermoelectric cooler in cryosurgery device, Cryogenics 50
(2010) 729-764
[5] S.B. Riffat, Xiaoli Ma, Thermoelectrics : a review of present
and potential applications, Applied Thermal Engineering, 23
(2003) 913-935
[6] Hsiang-Sheng Huang et. al, Thermoelectric water-cooling
device applied to electronic equipment, International
Communications in Heat and Mass Transfer 37(2010) 140-
146.
[7] Rieyu Chein, G. Huang, Thermoelectric cooler application in
electronic cooling, Applied Thermal Engineering 24 (2004)
2207-2217.
[8] Miguel A. Sanz-Bobi et al, Thermoelectricity applied to the
cryoconcentration of orange juice, 15th International
Conference on Thermoelectric (1996) 259-263
[9] A. Hamilton, J.Hut, An electronic cryopore for cryosurgery
using heat pipes and thermoelectric coolers : a preliminary
report, Journal of Medical Engineering & Technology 12
(1993) 104-109.
[10] Hiroki Takeda et al, Development and estimation of novel
cryoprobe utilizing the peltier effect for precise and safe
surgery, Cryobiology 59 (2009) 272-284.
[11] M.R. Holman, S.J Rowland, Design and development of new
surgical instrument utilizing the peltier thermoelectric effect,
Journal of Medical Engineering & Technology 21 (1997)
106-110.
[12] S.B Riffat et al, A novel thermoelectric refrigeration system
employing heat pipe and phase change material : an
experimental investigation, Renewable Energy 23 (2001)
313-323
[13] Faghri A. Heat pipe science and technology, Taylor &
Francis, 1995.
[14] David Reay, P.A. Kew, Heat pipes theory, design and
applications, Elsevier, 2006.
[15] Y.H Yau, M. Ahmadzadehtalatapeh, A review on the
application of horizontal heat pipe heat exchangers in air
conditioning system in the tropics, Applied Thermal
Engineering 30 (2010) 77-84.
[16] Te-En et al, Dynamic test method for determining the thermal
performances of heat pipes, International Journal of Heat
and Mass Transfer 53 (2010) 4567-4578.
[17] Leonard L. Vasilev, Heat pipe in modern heat exchangers,
Applied Thermal Engineering 25 (2005) 1-19.
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Preliminary Study on Length of Candle Filter Surface on the Flow Pattern
in Freeboard of Fluidised Bed Gasifier
A. Farhan Faudzi1*
, Kahar Osman1, Nor Fadzilah Othman
2,
Mohd Hariffin Bosrooh2
1 Computational Fluid Mechanics Laboratory, Faculty of Mechanical Engineering,
Universiti Teknologi Malaysia,
81310 UTM Skudai, Johor, Malaysia, [email protected]
2 TNB Research Sdn. Bhd.,
No. 1, Lorong Ayer Itam,
Kawasan Institusi Penyelidikan,
43000, Kajang Selangor, Malaysia,
ABSTRACT
This paper presents numerical study on flow pattern in
freeboard of fluidized bed gasifier during syngas filtration
due to different length of candle filter surface which are
0.25m, 0.30m and 0.35m. This numerical study was done
by considering two flows which are air flow and coal flow
by using ANSYS Fluent 14. The air inlet velocity flows
from previous study are 0.11m/s, 0.16m/s and 0.21m/s
whereas coal inlet velocities choose to be 10% of air flow
velocity. From the results, observed that flow from mixture
of air and coal for 0.25m of candle filter surface has
uniform distribute as flow move throughout freeboard. On
the other hand, for 0.30m and 0.35m of candle filter
surface, flow tend to become higher in velocity nearer wall
of freeboard. By increasing the flow inlet velocity, the flow
pattern in freeboard remain same as before but increase in it
magnitude. In comparison for all length of candle filter
surface in this study, 0.25m considered best as flow
produce in freeboard is more uniformly.
Keywords : Example: Freeboard fluidized bed, candle
filter, flow pattern
1. INTRODUCTION
The candle filter design much important in order to enhance
the ability to filter any impurities or contaminations that
exist in freeboard bed fluidized. The higher filtration rate
mean that higher in provide uniform flow distribution in
freeboard of fluidized bed gasifier. This is because the non-
uniformities may lead to insufficient filter cleaning [1].The
previous study shown that velocity profile not achieve
better at inlet vent and distribution of inlet velocity no fully
understood. Apart from that, flow velocity at inlet vent not
uniform distributed lead to lower usage rate of filter media.
However, a more uniform flow inlet can be obtained by
adjustment of length or angle of candle filter surface [2].
From previous study, results indicate that distribution
uniformity of flow could be improved by reducing inlet
velocity, diverging angle and particle diameter [3]. On the
other hand, changes of velocity magnitude during inlet
process also important to see as it is adversely affect
operation of plant and filter durability [4]. This research
was conducted to seek the flow uniform distribution in
freeboard of fluidized bed gasifier due to the different
length of candle filter surface.
2. CFD MODELLING DESCRIPTIONS
In this paper, the flow pattern inside the freeboard of
fluidized bed gasifier study using ANSYS Fluent 14.0 by
dimension as 2D model with viscous model of k-epsilon
model. In the multiphase model, it stated to be mixture with
2 eularian phases which are air and coal. The geometry of
freeboard of fluidized bed gasifier was simplified as in
figure 1 with varies in length of candle filter surface. The
flow of air and coal into freeboard was at bottom part which
was inlet vent. The vertical length of freeboard is 0.3m with
width of 0.05m whereas candle filter surface stated at
opposite of inlet vent with vertical length of 0.03m from
inlet vent [5].
Figure 1: 2D model for freeboard of fluidized bed
gasifier
The air inlet velocity varied at 0.11m/s, 0.16m/s and
0.21m/s meanwhile coal inlet velocity is 10% of air inlet
velocity which were 0.011m/s, 0.016m/s and 0.021m/s.
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Pressure stated as 2000pa at surface of inlet vent. The
simulation study on candle filter neglect it effect in filtering
particle as study much concern about velocity flow pattern
in freeboard of fluidized bed gasifier. Any chemical
reaction and temperature neglected.
3. RESULTS AND DISCUSSION
All the results from simulation by ANSYS Fluent 14.0
presented were converged by running 40 to 80 iterations.
Figure 2 show the mixture flow pattern of air and coal in
fluidized bed gasifier with different in length of candle
filter surface of 0.25m, 0.30m and 0.35m with different
flow inlet velocity.
Figure 2 (a): Flow pattern for 0.11m/s air inlet
Figure 2 (b): Flow pattern for 0.16m/s air inlet
Figure 2 (c): Flow pattern for 0.21m/s air inlet
Figure 2: Flow velocity coloured by contour vector
From Figure 2, left show 0.30m, upper part show 0.35m
and lower part show 0.25m of candle filter surface. Candle
surface length of 0.25m show flow velocity of air and coal
not reach higher value as shown by contour vector.
Moreover, as filter media introduce in this model, the
filtering of particle become easier as flow exist nearer filter
surface. The flow pattern for 0.30m of length of candle
filter surface stated almost same with 0.25m but the flow
form has higher magnitude in velocity. However, at length
of 0.30m the velocity starts to become higher at nearer wall
of freeboard. This show that flow exist caused fluid or solid
particle exist will flow away from candle filter and filtering
process become more inefficient. For 0.35m of candle filter
surface, flow pattern more toward at nearer freeboard wall,
moreover the contour vector stated at high value.
Furthermore, flow exist less at candle filter surface and the
vortex can occur at candle filter surface due to different in
velocity profile exist. The vortex exist prove of circulating
movement which is non uniform flow which lead to low
rate usage of filter surface [6].
Meanwhile, for all length of candle filter surface as
increasing air inlet velocity, the magnitude of flow inside
freeboard bed gasifier increase. Figure 3 show the change in
magnitude of flow inside freeboard at candle filter surface
as air inlet velocity in varies. From the figure 3, green line
represent length of 0.25m, red line for 0.30m and blue line
for 0.35m of candle filter surface. The velocity in freeboard
show steady flow as it passes through candle filter then start
to produce higher flow velocity at nearer wall of freeboard
as length of candle filter surface increase.
Figure 3 (a): Magnitude of flow velocity at
0.11m/s air inlet
The 5th IMAT, November 12 – 13th
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264
Figure 3 (b): Magnitude of flow velocity at 0.16m/s
air inlet
Figure 3 (c): Magnitude of flow velocity at 0.21m/s
air inlet
Figure 3: Magnitude of flow velocity at candle filter
surface with different length of candle filter surface.
4. CONCLUSION
The numerical result shown that the shorter length of candle
filter surface provide more uniform flow hence lead to high
rate of filter surface. Moreover, flow pattern exist at nearer
candle filter surface enhance filter process.
ACKNOWLEDGMENT The authors would like to thank Universiti Teknologi
Malaysia and TNB Research Sdn. Bhd. for supporting this
research activity.
REFERENCES
[1] T.G. Chuah, C.J. Withers and J.P.K. Seville,
―Prediction and measurement of the pressure and
velocity distributions in cylindrical and tapered rigid
ceramic filters‖, Separation and Purification
Technology 40, 2004, 47-60
[2] Chia-Jen Hsu and Shu-San Hsiau, ―Experiment study
of the gas flow behavior in the inlet of a granular bed
filter‖, Advanced Powder Technology 22, 2011, 741-
752
[3] Gong Jinke, Tian Chan and Wu Gang, ―Numerical
simulation on distribution characteristics of particle
distribution uniformly in a radial style diesel
particulate filter‖, Advances in Computer Science and
Engineering, 2012, 795-805
[4] S. Ito, T. Tanaka and S. Kawamura, ―Changes in
pressure loss and face velocity of ceramic candle
filters caused by reverse cleaning in hot coal gas
filtration‖, Powder Technology 100, 1998, 32-40
[5] Andrea Di Carlo and Pier Ugo Foscolo, ―Hot syngas
filtration in the fluidized bed gasifier: Development of
a CFD model‖, Powder Technology 222, 2012, 117-
130
[6] Chia-Jen Hsu, Shu-San Hsiau, Yi-Shun Chen and Jiri
Smid, ―Investigation of the gas inlet velocity
distribution in a fixed granular bed filter‖ Advanced
Powder Technology 21, 2010, 614-622
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2012
265
Preliminary Study on the Effect of Type Distributor Plate on Airflow
Pattern in Bubbling Fluidised Bed
Nofrizalidris Darlis1*
, Kahar Osman1, Ab Malik A. Hamat
1 , Nor Fadzilah Othman
2, Mohd
Hariffin Bosrooh2
1 Computational Fluid Mechanics Laboratory, Faculty of Mechanical Engineering,
Universiti Teknologi Malaysia,
81310 UTM Skudai, Johor, Malaysia,
2 TNB Research Sdn. Bhd.,
No. 1, Lorong Ayer Itam,
Kawasan Institusi Penyelidikan,
43000, Kajang Selangor, Malaysia
ABSTRACT
This paper presents numerical study on airflow pattern in
bubbling fluidized bed due to the different shapes of
distributor plate which are flat, convex and concave. The
numerical study was done at a grace of a Commercial
Computational Fluid Dynamics ANSYS Fluent 14. The air
inlet velocity based on previous studies which are 0.28m/s,
0.33m/s and 0.37m/s. From the results, observed that
airflow for flat plate distributor has uniform upward
movement. Meanwhile, concave and convex has airflow
movement upward and downward resulting circulating flow
which is good for mixing process. By increasing the air
inlet velocity, the airflow inside the bubbling fluidized bed
increase. Compared all the distributor plate in this study,
convex shape considered the best distributor plate in mixing
process.
Keywords Example: Gasification,distributor
plate,airflow
1. INTRODUCTION
The gas distribution plate is the key element in fluidization
technology. There are many types of distributor plate
differentiated by hole type, number of hole and plate shape.
Main requirement for a distributor plate are to promote
uniform distribution of fuel particle to make sure a good
chemical conversion and uniform temperature throughout
the bed [1]. In the same time, mixing pattern will affect the
performance of gasification [2]. This type of plate will
influenced the mixing pattern between air and fuel particles.
In the other hands, distributor plates also govern to the
particles circulation behavior in the gasification chamber
[3]. Although many research regarding distributor plate
have been carried out, the effect of different shape type of
the plate still in minimum study. The airflow and its
distributor primary factor influenced fluidized bed
processing [4]. The higher heating value reached its peak
value at a fluidization velocity of 0.28 m/s but remained
fairly constant at the fluidization velocities of 0.33 and 0.37
m/s [5]. Therefore this research was conducted to seek the
air distribution in bubbling fluidized bed due to the
different shapes of distributor plate which are flat, convex
and concave. The geometry of the bubbling fluidized bed
was referring to laboratory scale fluidized bed gasifier at
TNB Research Sdn. Bhd.
2. CFD MODELLING DESCRIPTIONS
In this paper, the airflow pattern inside the gasification
chamber will study by using Commercial Computational
Fluid Dynamics ANSYS Fluent 14. Standard K-epsilon
model with enhanced wall treatment was used as the model
of this study.
Figure 1: 2D geometry modeling for flat, concave and
convex distributor plate
The geometry of bubbling bed fluidized was simplified as
figure 1 and the air inlet was state at the holes plate. The
inner diameter 250mm and its bed height is 300mm [6].
The air velocity inlet is varied at 0.28m/s, 0.33m/s and
0.37m/s.
Figure 2: Types of distributor plate [1]
Figure 2 show the types of distributor and the angle of
convex and concave plat was 20 degree. The simulation
study was neglected any chemical reaction and temperature.
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3. RESULTS AND DISCUSSION
All simulation results presented throughout this paper were
converged by running 800 to 1500 iterations. Figure 3 show
the air flow pattern in bubbling fluidized bed by using flat,
concave and convex distributor plate with different air inlet
velocity.
Figure 3 (a): Airflow pattern for 0.28m/s air inlet
Figure 3 (b): Airflow pattern for 0.33m/s air inlet
Figure 3 (c): Airflow pattern for 0.37m/s air inlet
Figure 3: Airflow velocity coloured by contour vector
Flat plate distributor show that the airflow uniformly
moving upward without any circulating. From the airflow
pattern, it is reasonable to say that this type of distributor
plate will give low mixing and turnover. A. E. Ghaly and K.
N. MacDonald (2012) reported that localized mixing caused
upward movement of the bubbles by flat distributor plate
was clearly evident but no bed material turnover was
observed. Besides, by using concave plate observed that
upward movement at the center and downward at near the
wall. The airflow velocity near the wall region is less than
the center. Meanwhile, convex plate type show the upward
movement beside the wall and downward at the center. The
velocity of the airflow near the wall is higher than the
center region. Both concave and convex distributors give
upward and downward movement of the air inside the
fluidized bed resulting circulation movement which is good
for the mixing process. The different speed of airflow can
be used to give more time for carbon conversion of the fuel
and also minimize the bed material from leave the reactor.
This is concurrent with past research that to improve
mixing properties of the binary mixture, which has great
tendency for segregation due to density differences, an
angled distributor plate should be used [1].
For all type of distributor, by increasing the air inlet
velocity, the airflow inside the bubbling fluidized bed
increase. Figure 4 show clearly the graph of velocity
changed for each type of distributor plate.
Figure 4 (a): Distribution of airflow velocity at
0.28m/s air inlet
Figure 4 (b): Distribution of airflow velocity at 0.33m/s air
inlet
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267
Figure 4 (a): Distribution of airflow velocity at 0.37m/s air
inlet
Figure 4: Airflow distribution at 0.05m of bed high with
different type of distributor plate
The airflow velocity change for convex distributor was
highest compared with concave and flat plate. The rapidly
drop of airflow at the center region resulting circulating
flow. Increasing circulating flow will increase the particle
turnover, thus increasing mixing efficiency.
4. CONCLUSION
In this paper, the effect of type distributor plate shape has
been studies using ANSYS Fluent 14. These numerical
results give a good agreement with the past study by A. E.
Ghaly and K. N. MacDonald (2012). Convex distributor
plate shows very good airflow movement inside the
bubbling fluidized bed reactor compared with others. It is
reasonable to say that by using convex distributor plate will
give high efficient mixing process during gasification.
ACKNOWLEDGMENT The authors would like to thank Universiti Teknologi
Malaysia and TNB Research Sdn. Bhd. for supporting this
research activity.
REFERENCES
[1] E. Ghaly and K.N. MacDonald, ―Mixing patterns and
residence time determination in a bubling fluidized
bed system‖ American Journal of Engineering and
Applied Sciences, 2012, 5 (2), 170-183
[2] Chih-Jung Chen, Chen-I Hung, and Wei-Hsin,
―Numerical investigation on performance of coal
gasification under various injection pattern in an
entrained flow gasifier,‖ Applied Energy, in press.
[3] Zbigniew Garncarek, Longin Przybylski, John S.M.
Botterill and Christopher J. Broadbent, ―A quantitative
assessment of the effect of distributor type on particle
circulation,‖ Powder Technology 91, 1997, 206-216
[4] Frederic Depypere, Jan G. Pieters and Koen
Dewttinck,―CFD analysis of air distribution in
fluidized bed equipment‖, Powder Technology 145,
2004, 176-189.
[5] Sadaka S.S., A. E. Ghaly and M. A.
Sabbah,―Development of an air-stream fluidized bed
gasifier‖, Misr Journal of Agricultural Engineering,
1998, Vol 15(1), 47-52.
[6] Nor Fadzilah Othman, Mohd Hariffin Bosrooh and
Kamsani Abdul Majid,―Partial gasification of different
types of coals in a fluidized bed gasifier‖, Jurnal
Mekanikal, 2007, 40-49
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2012
268
Natural Convection in A Differentially Heated Cavity
Using Splitting Method
Ubaidullah S.a, Kahar Osman
a
aFaculty of Mechanical Engineering
Universiti Teknologi Malaysia, Skudai, Johor, Malaysia
E-mail : [email protected]
ABSTRACT
The solution of thermally-driven flow of Navier-
Stokes equation, in Boussinesq approximation, is
presented. The results are obtained using splitting
method and in good agreement with available
benchmark numerical solutions. The convection
terms are explicitly integrated using 3-step Range-
Kutta scheme. Buoyancy-driven fluid flows data in a
differentially heated cavity for low Rayleigh
numbers (R=103, R=10
4 and R=10
5) are presented.
Keywords : Natural convection, thermally driven,
splitting method
1. INTRODUCTION
Thermally-driven fluid/gas flow can be found in
many engineering applications such as room
ventilation system, heat exchanger, solar energy
collector, cooling of electronic components, factory
stack emission etc. The fluid velocity is caused by
buoyancy force generated from the change of fluid
density as a result of significantly hot fluid
temperature. This type of flows are normally
investigated to improve performance of engineering
systems, increase efficiency of ventilation systems
that may reduce the cooling energy required, predict
contaminants between buildings and other purposes.
Researches on thermally-driven flows have been
done in many areas of engineering. In solar collector
study, Fan et. al. [1] studied the standby heat loss
effects to thermal performance of a heat system by
analyzing the flow around the hot storage tank. In
construction and design of buildings, Chun. al. [2]
investigated the buoyancy-driven ventilation in a
reduced-scale building, Rakesh [3] studied the flow
in solar chimneys, Guohui [4] studied the impact of
computational domain in buoyancy –driven flow
simulation. In air pollution research, Baik and Kim
[5] investigated the urban street canyon flows with
street-heating in between two buildings. Various
solar radiation conditions in urban street canyon
have been studied by Xiaomine et. al. [6]. Rezwan
et. al. [7] investigated the building aspect ratio and
wind speed effect on temperature distribution in
urban street canyon. A lot of problems related to
thermally-driven flows can be found in the
literature. In brief, natural convection flows are still
a major interest for researchers in heat and fluid
flow studies.
There are numerous approaches to solve Navier-
Stokes equation that governs the thermally-driven
flow problems. One of the most successful
approaches nowadays is the projection methods.
Projection methods lead to easy-to-implement and
efficient algorithms by decoupling the diffusion and
convection terms of the Navier-Stokes equation.
Fractional step method and pressure correction
methods are two kinds of projection methods.
Fractional step method is based on a full splitting of
the diffusion and incompressibility constraint
(pressure) in different sub steps (Karniadakis [8]).
Full splitting suffers from erroneous solutions as a
result of improper boundary condition of pressure.
However, the pressure boundary conditions have
been discussed extensively in the literature [8], [12],
[13], [14]. Meanwhile, pressure correction methods
are based on predictor-corrector procedure between
velocity and pressure fields. Initial approximation of
pressure allows the momentum equation to be
solved without satisfying the divergence-free
constraint and requires additional pressure
correction procedure. In this method, a Poisson
equation for a new defined quantity is solved instead
of pressure Poisson equation. A homogeneous (zero)
Neumann condition need to be used to ensure
divergence-free condition of the velocity which is
not valid for pressure itself [14]. However, the final
velocity fields satisfy the divergence-free condition
for semi-discrete formulation.
A splitting algorithm based on pressure correction
method is chosen in this study for solving the
thermally-driven flow problems in a differentially
heated cavity, the standard problem used for
benchmarking new computer programs.
2. NUMERICAL MODEL
The so called Boussinesq equations to model non-
isothermal flow problems are used
(1)
(2)
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269
(3)
Where
R = (gβ∆Tl3)/(k/ν) (4)
Pr = ν/k (5)
are commonly used Rayleigh and Prandtl numbers
and g = (0, 1)T . Here, u, p and T are fluid velocity,
pressure and temperature respectively, g is
acceleration of gravity, β is thermal expansion
coefficient, l is characteristic length, ∆T is
characteristic temperature difference, k is thermal
diffusivity and v is kinematic viscosity of the fluid.
The equations are non-dimensionalised according to
Paolucci and Chenowith [10]. The equations are
solved using algorithm by Minev [9].
Instead of using spectral element method as in the
pressure correction method by Minev [9], present
study applied finite difference method. The
equations are discretized in time with a second order
backward difference scheme and a second order
central difference scheme for space. The algorithm
leads to two Helmholtz equations for velocity
components, and a Poisson equation for pressure.
All of them are solved implicitly in present study.
The boundary conditions are
Nusselt numbers (Nu) are calculated using simple
finite difference formula at the hot vertical wall
Nu = ∂T/∂x |x1=0 (6)
3. RESULTS
0 20 40 60 80 100 120 140 1600
20
40
60
80
100
120
140
160Streamline
5 10 15 20 25 30 355
10
15
20
25
30
35
(a) (b)
5 10 15 20 25 30 355
10
15
20
25
30
35
5 10 15 20 25 30 355
10
15
20
25
30
35
(c) (d)
Figure 1: (a) Streamlines (b) iso-U velocity
contour (c) iso-V velocity contour (d)
Isotherms for buoyancy driven flow at
Rayleigh number = 103.
0 20 40 60 80 100 120 140 1600
20
40
60
80
100
120
140
160Streamline
5 10 15 20 25 30 355
10
15
20
25
30
35
(a) (b)
5 10 15 20 25 30 355
10
15
20
25
30
35
5 10 15 20 25 30 355
10
15
20
25
30
35
(c) (d)
Figure 2: (a) Streamlines (b) iso-U velocity
contour (c) iso-V velocity contour
(d) Isotherms for buoyancy driven
flow at Rayleigh number = 104.
0 20 40 60 80 100 120 140 1600
20
40
60
80
100
120
140
160Streamline Iso-U
10 20 30 40 50 60 70 80 90 100
10
20
30
40
50
60
70
80
90
100
(a) (b)
Iso-V
10 20 30 40 50 60 70 80 90 100
10
20
30
40
50
60
70
80
90
100
Isotherm
10 20 30 40 50 60 70 80 90 100
10
20
30
40
50
60
70
80
90
100
(c) (d)
Figure 3: (a) Streamlines (b) iso-U velocity
contour (c) iso-V velocity contour
(d) Isotherms for buoyancy driven
flow at Rayleigh number = 105.
u1=0
u2=0
T=1
u1=0, u2=0,
dT/dn=0
u1=0
u2=0
T=0
u1=0, u2=0,
dT/dn=0
Ω
(0,0)
(0,1) (1,1)
(1,0)
x1
x2
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270
Table 1: Buoyancy-driven flow in an enclosed
cavity. Present results (P) compared with benchmark
numerical solution (B) by Minev [9] and the
derivation (D) for R=103, 10
4 and 10
5.
Variable Source R=103
R=104
R=105
U1,max B 3.630 12.627 34.73
P 3.675 14.893 37.12
D (%) +1.2 +12.2 +6.9
X1,max B 0.813 0.823 0.875
P 0.834 0.868 0.914
D (%) +2.6 +3.5 +4.4
U2,max B 3.693 19.617 68.59
P 4.232 19.520 58.72
D (%) +14.6 -8.0 -14.4
X2,max B 0.170 0.125 0.079
P 0.152 0.106 0.066
D (%) -10.6 -5.6 -16.5
Numax B 1.507 3.531 7.717
P 2.652 5.534 9.275
X2(Nu) B 0.08 0.143 0.080
P 0.13 0.140 0.067
Numin B 0.692 0.586 0.726
P 1.405 1.082 0.847
X2(Nu) B 1.0 1.0 1.0
P 0.98 1.0 1.0
0 2 4 6 8 100
1.0
0.75
0.5
0.25
Nusselt Number (Nu)
Y -
Coord
inate
R=104
R=105
R=103
Figure 4: Comparison of local Nusselt number
along the hot wall (X1=0).
0.25 0.5 0.75 1.00-60
-40
-20
0
20
40
60
X1 - Coordinate
U2-V
elo
city
R=105
R=103
R=104
Figure 5: Variation of vertical velocity at
X2=0.5.
-30 -20 -10 0 10 20 30
1.0
0
0.5
0.25
0.75
U1 - Velocity
X2 -
Coord
inate
R=104
R=105
R=103
Figure 6: Variation of horizontal velocity at X1=0.5.
0 1.00.50.25 0.750
0.2
0.4
0.6
0.8
1
X1 - Coordinate
Tem
pera
ture
R=105
R=104
R=103
Figure 7: Variation of temperature at mid-
height (X2=0.5).
4. DISCUSSION
The initial conditions are set as zero velocity, zero
pressure and zero temperature. The temperature on
the left cavity is then set to 1 and the algorithm is
then applied. After some initial transience, the
solutions reach steady-state values and plotted as in
the above figures. The number of grid used is 150 x
150 for all Rayleigh numbers. For transitional and
high Rayleigh number flows (R > 105), they are
unstable with present grid size. They require finer
mesh and are not presented here. The time step is
obtained experimentally and varies a lot with
Rayleigh number. Its value is 0.004 for R=103 and
become as small as 0.0004 for R=105. The algorithm
takes 1.51 s per time step on Intel Core i5 CPU (3.5
GB RAM) for second order central difference
approximation scheme.
As in figure 1, figure 2 and figure 3, the streamlines,
iso-velocity contour for both directions x and y and
the isotherms plots are in very good agreement with
the benchmark numerical solution of a differentially
heated cavity flow. Although consistent contour
The 5th IMAT, November 12 – 13th
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271
plots are obtained, errors can be found in the
position of maximum velocities and its values. The
spatial discretization of the algorithm is expected to
be the main contributing factor for errors as the
benchmark solution used high order spectral element
discretization technique. Besides, Karniadakis [8]
recommended the use of high order pressure
boundary condition especially in splitting algorithm
for low Reynolds number flow which is not being
implemented in the algorithm. In addition, present
study employed full-discrete formulation, discrete
time and space, instead of continuous projection,
discrete time with continuous space, as in Minev
[9].
As for Nusselt number, the values are very sensitive
to the point where it is calculated especially for
R=105. As in figure 4, the Nusselt number values
along the X2 axis has the same trend as Wan et. al.
[11]. For U1, U2 and temperature variation along X1
and X2 axes as in figure 5, figure 6 and figure 7,
they are also in good agreement with Wan et. al.
[11]. However, their values are shifted slightly
higher from the benchmark solution due to the
prescribed reasons mentioned above.
5. CONCLUSION
Present study has successfully investigated the
natural convection in a differentially heated square
cavity using splitting method. Iso-contours of
streamlines, horizontal velocity, vertical velocity
and temperature are the same as available
benchmark numerical solutions. Present algorithm
has slightly over-predicted the Nusselt number
values although they have relatively similar trends
along vertical hot wall as benchmark solutions.
REFERENCES
[1] J. Fan and S. Furbo, ―Buoyancy Driven Flow In
A Hot Water Tank Due To Standby Heat
Loss‖, Solar Energy, vol. 50, 2012, pp 1266-
1274.
[2] P. L. Chun, H. T. Lin, J. H. Chou, ―Evaluation
Of Buoyancy-Driven Ventilation In Atrium
Buildings Using Computational Fluid
Dynamics And Reduced-Scale Air Model‖,
Building and Environment, vol. 44, 2009, pp
1970-1979.
[3] K. Rakesh, L. Chengwang, ―Flow Reversal
Effects On Buoyancy Induced Air Flow In A
Solar Chimney‖, Solar Energy, vol. 86, 2012,
pp2783-2794.
[4] G. Guohui, ―Simulation Of Buoyancy-Driven
Natural Ventilation of Buildings – Impact Of
Computational Domain‖, Energy and
Buildings, vol 42, 2010, pp 1290-1300.
[5] J. J. Kim, J. J. Baik, "Urban street-canyon flows
with bottom heating," Atmospheric
Environment, vol 35, no. 20, 2001,
pp.3395-3404.
[6] X. Xiamoine, H. Zheng, W. Jiasong, Z. Xie,
"Impact of solar radiation and street layout
on pollutant dispersion in street canyon,"
Building and Environment, vol. 40, no. 2,
2005, pp. 201- 212.
[7] A. M. Rezwan, Y.C.L. Dennis, C. H. Liu,
"Effects of building aspect ratio and wind
speed on air temperatures in urban-
like street canyons," Building and
Environment, vol. 45, no. 1, 2010, pp.
176-188.
[8] G. E. Karniadakis, M. Israeli, S. A. Orszag,
―High Order Methods for The Incompressible
Navier-Stokes Equation,‖ International
Journal of Computational Physics, vol. 90,
1991, pp 414-443.
[9] P. D. Minev, F. N. Van De Vosse, L. J. P.
Timmermans and A. A. Van Steenhoven, ‖A
Second Order Splitting Algorithm For
Thermally-Driven Flow Problems‖,
International Journal of Numerical Method
and Fluid Flows, vol 6. No. 2, 1995, pp 51-60.
[10] S. Paolucci and D. R. Chenowith, ―Transition
To Chaos in a Differentially Heated Vertical
Cavity,‖International Journal of Fluid
Mechanic, vol. 201, 1989, pp 379-410.
[11] D. C. Wan, B. S. V. Patnaik, and G. W. Wei,
―A New Benchmark Quality Solution for the
Buoyancy-Driven Cavity by Discrete Singular
Convolution‖, Numerical Heat Transfer, vol.
40, 2001, pp 199-228.
[12] S. A. Orszag, M. Israeli and M. O. Deville,
―Boundary Conditions for Incompressible
Flows‖, Journals of Science Computing, vol 1,
1986, pp 75-111.
[13] P. M. Gresho and R. L. Sani.‖On Pressure
Boundary Conditions for the Incompressible
Navier-Stokes Equation‖, International
Journals of Numerical Methods in Fluids, vol
7, 1987, pp 1111-1145.
[14] L. J. P. Timmermans, P. D. Minev and F. N.
Van De Vosse, ―An Approximate Projection
Scheme for Incompressible Flow Using
Spectral Elements‖, International Journals of
Numerical Methods in Flu