Accepted Manuscript

Experimental and Numerical Study of the Compact Heat Exchanger with DifferentMicrochannel Shapes

Vladimir Glazar, Bernard Frankovic, Anica Trp

PII: S0140-7007(14)00167-4

DOI: 10.1016/j.ijrefrig.2014.06.017

Reference: JIJR 2823

To appear in: International Journal of Refrigeration

Received Date: 18 February 2014

Revised Date: 12 June 2014

Accepted Date: 29 June 2014

Please cite this article as: Glazar, V., Frankovic, B., Trp, A., Experimental and Numerical Study of theCompact Heat Exchanger with Different Microchannel Shapes, International Journal of Refrigeration(2014), doi: 10.1016/j.ijrefrig.2014.06.017.

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EXPERIMENTAL AND NUMERICAL STUDY OF THE COMPACT HEAT

EXCHANGER WITH DIFFERENT MICROCHANNEL SHAPES

VLADIMIR GLAZAR

Faculty of Engineering, University of Rijeka, HR-51000 Rijeka, Croatia

BERNARD FRANKOVIC

Faculty of Engineering, University of Rijeka, HR-51000 Rijeka, Croatia

ANICA TRP

Faculty of Engineering, University of Rijeka, HR-51000 Rijeka, Croatia

Address correspondence to Vladimir Glazar, Faculty of Engineering, University of Rijeka,

Vukovarska 58, HR-51000 Rijeka, Croatia, tel. +385 51 651 536; fax: +385 51 675 801, E-mail:

vladimir.glazar@riteh.hr

ABSTRACT

Experimental and numerical analysis of heat transfer and fluid flow in the compact heat exchanger

has been done in this paper. In an open circuit wind tunnel, developed on purpose for this

investigation, the measurement of working media temperatures and mass flow rates for heat

exchanger with microchannel coil has been accomplished. In accordance with the heat exchangers

used for experiments, numerical 3D simulation of adequate geometry shapes has been done. With

utilization of air/water side numerical simulation, more detailed results have been achieved in

relation to the simulation that assumes constant temperature or constant heat flux on the pipe wall.

Good agreement between experimentally measured and numerically calculated results has been

accomplished. The influence of different microchannel shapes on heat transfer effectiveness and

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pressure drop has been studied numerically. Comparison of results has been made accompanied by

the discussion and final conclusions.

Key words: Compact heat exchangers; Experimental and numerical analysis; Microchannel shape;

Air/water numerical simulation

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EXPERIMENTAL AND NUMERICAL STUDY OF THE COMPACT HEAT

EXCHANGER WITH DIFFERENT MICROCHANNEL SHAPES

ABSTRACT

Experimental and numerical analysis of heat transfer and fluid flow in the compact heat exchanger

has been done in this paper. In an open circuit wind tunnel, developed on purpose for this

investigation, the measurement of working media temperatures and mass flow rates for heat

exchanger with microchannel coil has been accomplished. In accordance with the heat exchangers

used for experiments, numerical 3D simulation of adequate geometry shapes has been done. With

utilization of air/water side numerical simulation, more detailed results have been achieved in

relation to the simulation that assumes constant temperature or constant heat flux on the pipe wall.

Good agreement between experimentally measured and numerically calculated results has been

accomplished. The influence of different microchannel shapes on heat transfer effectiveness and

pressure drop has been studied numerically. Comparison of results has been made accompanied by

the discussion and final conclusions.

Key words: Compact heat exchangers; Experimental and numerical analysis; Microchannel shape;

Air/water numerical simulation

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NOMENCLATURE

A surface area, m2

B computational block

cp specific heat capacity, J kg-1 K-1

dh hydraulic diameter, m

Fp fin pitch, m

Ft fin thickness, m

h small channel height, m

k thermal conductivity, W m-1 K-1

m& mass flow rate, kg s-1

n normal direction coordinate, m

N total number

p pressure, Pa

∆p pressure drop, Pa

Pt transversal MCHX tube row pitch, m

Q average heat transfer rate, W

Re Reynolds number (uin ⋅ dh / ν )

T temperature, K

u, v, w velocity components, m s-1

ru velocity vector, m s-1

V& volumetric flow rate, m3 s-1

w small channel width, m

x, y, z Cartesian coordinates, m

Greek

ε heat exchanger effectiveness

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µ dynamic viscosity, Pa s

ν kinematic viscosity, m2 s-1

ρ density, kg m-3

σ standard deviation, K

Subscripts

a air

f fin

in inlet

n computational block number

out outlet

sc small channels

t tube

tot total cross section area

w water

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1. INTRODUCTION

Compact heat exchangers are widely used in many ways in residential, commercial and industrial

HVAC systems. According to Kuppan (2000), compact heat exchangers are those that have high

heat transfer surface area to volume ratio (>700 m2/m3). Serious development and their rapid

growth of usage have began with demands of (mostly) automotive industries for heat exchangers

that use reduced space, weight, support, energy requirement and cost for desired thermal

performance. Fin-and-tube heat exchangers are representatives of compact heat exchangers with

high compactness ratio. Further size reduction of heat exchanger passages, briefly described by

Kandlikar (2007), has led to heat exchangers with flat tubes, commercially known as microchannel

coil. In this paper numerical and experimental analysis of such heat exchanger has been performed.

In last decade of twentieth century respectable number of experimental data for friction factor and

Nusselt number in microchannels that disagree with the conventional theory has been published.

Lots of them appeared to be inconsistent with one another. Various reasons have been proposed to

account for those differences and Morini (2004) made critical analysis in order to highlight the main

results. When dealing with micro scale, heat transfer can be suitably described by standard theory

and correlations, but scaling effects precisely described by Morini (2006) and Rosa et al. (2009),

which are often negligible in macro scale channels, should not be neglected. One of the

experimental investigations that proved their theory has been done by Park and Punch (2008).

Their work was focused on laminar flow (69 < Re < 800) within rectangular microchannel with

hydraulic diameter from 106 µm to 307 µm for single-phase liquid flow. They proposed an

empirical correlation in terms of Nu, Re, Pr and Brinkman number confined to the experimental

range. The friction factor obtained by experiments showed excellent agreement with conventional

hydraulic theory and also supported that the flow inside the microchannels was fully developed

laminar in the range of experiments.

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Current work in the field of compact heat exchangers where single-phase flow occurs, especially for

fin-and-tube type, is usually based on air flow models with assumptions of constant temperature of

heat transfer surface. Strip fin heat exchangers, as one type of compact heat exchangers with high

heat transfer areas, have been used for both single and two-phase flow applications in industrial

processes. Saad et al. (2012) made experimental and numerical investigation of single-phase

pressure drop and two-phase distribution in an offset strip fin compact heat exchanger. For the

purpose of performed investigation, first step of their paper is of particular interest where they have

analysed pressure drop experimentally and numerically via CFD simulations. The experimentally

determined friction factor for laminar flow conditions was plotted versus the Reynolds number and

compared to CFD results and to the Manglik and Bergles (1995) correlations. Numerical

simulations using FLUENT showed good agreement with experimental results in the laminar flow

regime, except at very low Re numbers. At low Re number (Re < 120), all the numerical

calculations for f were at most 15% different from the experimental results.

The use of microchannel heat exchangers ascends due to rapid increase in power density and

miniaturization of electronic devices what made more traditional types of heat exchangers

unpractical, or unable to meet heat power needs of emerging electronic devices. Sui et al. (2010)

studied numerically laminar water flow and heat transfer in three-dimensional wavy microchannels

with rectangular cross section. To simplify analysis they made several assumptions in numerical

modelling of the heat transfer: steady state, incompressible fluid, laminar flow, constant fluid

properties, negligible viscous dissipation and negligible radiative and natural convective heat

transfer. They found that heat transfer performance of the wavy microchannels is much better than

that of straight microchannels with the same cross section area, with the relatively small pressure

drop compared to the heat transfer enhancement. Further enhancement of numerical procedure, of

special interest for this paper, came from air/water side model proposed by Borrajo-Pelaez et al.

(2010) that allows implementation of some scaling effects (entrance effects, conjugate heat transfer

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and viscous heating). Therefore, to achieve higher accuracy of numerically acquired results and to

have better predictions on the exchanger performance, especially when dealing with micro scale,

air/water model should be used.

As opposed to previously mentioned microchannel geometries with rectangular channels, Agarwal

et al. (2010) experimentally tested several circular and non-circular flat tube geometry shapes. They

examined flat tubes with square, barrel, triangular, rectangular channels, channels with W and

channels with N shaped inserts during condensation of refrigerant. Some settings, and geometrical

characteristics (different microchannel geometries), were used as starting points for purpose of, in

this paper conducted, investigation. Although numerical approach gives inexpensive prediction

method, compared to expensive testing of numerous prototypes, experimental research should not

be neglected.

Heat transfer flow rate uniformity is another crucial value for microchannel heat exchangers

characterisation. Saleh et al. (2013) proposed approximation assisted optimization of headers for

new generation of air-cooled heat exchangers. They developed a new CFD model for headers used

in next generation of air cooled heat exchangers based on mini and micro tubes. Huang and Wang

(2013) made numerical study with goal to improve the system uniformity flow rate for U-type

compact heat exchangers. They compared flow ratio in small tubes among the various designs with

modified headers: multi-step blocker, trapezoidal blocker, baffle plate, baffle tube modified headers

and optimal module estimated by Levenberg–Marquardt Method. Their results revealed that the

estimated optimal module of U-type heat exchanger yields better system uniformity flow rate when

comparing with the original design of the heat exchanger. Their investigation also showed that flow

ratio significantly changes only in several tubes at the beginning of the header and in last two tubes.

Moreover, they concluded that there are several different methods to modify the inlet header or

adjust the pressure in tubes with obstructs to improve the system uniformity flow rate. One of more

recent papers that deals with flow maldistribution in microchannel heat exchangers was made by

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Zou and Hrnjak (2013). They have made experimental investigation with lower mass fluxes of

water followed with microchannel vertical header visualization. They concluded that distribution is

better at high mass flux in the header due to the higher momentum.

In the present paper thermodynamical and hydrodynamical analysis of compact heat exchangers

with different microchannel shapes has been performed in relation to the heat transfer effectiveness

and pressure drop. The experimental analysis has been accomplished for the heat exchanger with

microchannel coil that belongs to state of the art technologies in heating, ventilating and air

conditioning industry. Its header and internal baffle plate are made to ensure flow uniformity in

most of the flat tubes. Results accomplished by numerical and experimental analysis of heat

exchanger with flat tubes have been compared in order to obtain physical validity of the numerical

simulation. For the purpose of numerical analysis and in accordance with the heat exchangers used

for experiments, 3D models of adequate geometry in direction of both air and water flow have been

created. With utilization of air/water side numerical simulation, more accurate results have been

achieved in relation to the simulation that assumes constant temperature or constant heat flux on the

pipe wall. Numerical 3D simulations of heat exchangers have been performed using the finite

volume method. A detailed mathematical approach and detailed description of experimental

apparatus developed on purpose for this investigation can be found in Glazar (2011). Verified and

experimentally validated air/water side numerical simulation has been used to investigate the

influence of different microchannel shapes on heat transfer and pressure drop. Numerical tests have

been made on heat exchanger composed of flat tubes with circular and non-circular channels.

2. EXPERIMENTAL PROCEDURE

2.1 Experimental Set-up An open circuit wind tunnel was developed on purpose for this investigation. Fig. 1 shows the

schematic diagram of the wind tunnel with capability to supply high pressure air at the ambient

temperature to the test unit. Wind tunnel has been used for the measurement of working media

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temperatures and mass flow rates of several heat exchangers with microchannel coil, and results of

measurements for one heat exchanger have been published in this paper. Air and distilled water

were used as working fluids. The main components of the system were: heat exchanger with

microchannel coil, water flow loops, air supply unit, instrumentations and data acquisition systems.

The open circuit wind tunnel system was used to suck air from laboratory or from open air over the

air handling unit with capability of air preheating. The speed of centrifugal fan could be adjusted by

pressure relief damper from approximately 20 to 100% of maximum air flow. The upper part of

tunnel was composed of circular ducts with diameter of φ600 mm and lower part from rectangular

ducts of appropriate size, 550×450 mm. Measuring station with installed test heat exchanger was

insulated with 25 mm thick thermoplastic insulation to reduce unwanted ducts heat losses.

The inlet and outlet temperature across air side of the heat exchanger were measured by two

platinum resistance thermometer meshes (RTD, Pt100 sensors). The inlet measuring mesh consisted

of three RTD sensors while the outlet mesh consisted of nine RTD sensors. These sensors were

connected to data acquisition system in three-wire configuration. This is relatively simple wiring

arrangement that provides accurate readings with reliable auto correction of any problems caused

by any effect of the temperature range on the wiring itself. Used RTD sensors were pre-calibrated

with an accuracy of ±0.15 K. During each measurement they were periodically checked and

calibrated by standard thermometer of accuracy ±0.2 K. These signals were individually recorded

and then averaged. Fig. 2 shows measuring station of wind tunnel with installed data acquisition

system and part of second loop pipings filled with distilled water.

The air flow rate was measured with pipe orifice installed in circular ducts. A fluid passing through

an orifice construction experiences pressure drop. The orifice meter was calibrated by measuring

the air velocity distribution through the tunnel cross section using a hotwire anemometer and at the

same time recording the pressure drop across the orifice. Pressure drop was measured with portable

measuring system Testo 350 M/XL which has integrated differential probe with accuracy ±1 Pa. Air

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flow rate was then calculated from pressure drop achieved on orifice. The hot/cold water loop was

divided in two separate loops. First one with water connected to water/water heat pump with

capacity of 50 kW and second loop separated with plate heat exchanger. Second loop was built due

to possible creation of lime scale in microchannels and was filled with distilled water. Lime scale is

extremely dangerous for channels with hydraulic diameter smaller than 1 mm and could cause

undesired obstructions on water side of heat exchanger. The temperature of the water in piping was

measured by the same type of RTD sensors used on air side of wind tunnel. First one was installed

on water inlet side and other on water outlet side of piping. RTD sensors were installed in water

stream without direct contact by mean of small tube inserts. Water loop also consisted of circulating

pumps, portable ultrasonic water flow measuring system, regulating valves, strainers, expansion

tank, pressure gauge and thermometers. The National Instruments SCXI data acquisition,

automation and control module system was used. Connection to personal computer was

accomplished by National Instruments DAQCard. All virtual instruments were developed in

LabView which was installed on the personal computer.

During the experiments, the mass flow rate of the water in piping was varied from 200 to 2000 kg

h-1 using the flow control valve and high efficiency smart pump. The air velocity in ducts was

varied from 1 to 6 m s-1 by adjusting the damper positions using a lever mechanism. The outlet

water temperatures, inlet and outlet air temperatures as well as the pressure drops on the air and

water sides, were measured for different flow rates. Based on the above data, the performance of the

test unit was estimated and the corresponding results were used in later investigations.

2.2 Test Heat Exchanger Heat exchanger with microchannel coil (MCHX), developed for purpose of this experiment was

made of 68 parallel flat tubes connected with approximately 700 straight fins. Flat tubes comprise

18 rectangular channels with hydraulic diameter of dh = 0.99 mm. Schematic diagram of examined

heat exchanger and appropriate geometrical parameters have been shown on Fig. 3.

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The tested heat exchanger installed in the test system was insulated by a thin layer of silicone caulk

to reduce heat loses that could increase measuring uncertainties and was completely installed in the

air stream. The piping and ducts were completely insulated with a 10 mm thick layer of

thermoplastic insulation.

2.3 Test Conditions and Method

Measurements were performed in the following range of velocities and temperatures. Frontal air

velocities uin were in a range from 0.5 to 2.5 m s-1 with corresponding Reynolds number between 50

and 400. Air side Re number was based on hydraulic diameter dh of fin passages. The air inlet

temperatures Ta,in were in range from 280 to 305 K. Water velocities were in a range from 0.1 to 0.5

m s-1 with corresponding Reynolds numbers between 100 and 800. Water side Re number was

based on hydraulic diameter dh of small channels. Water temperatures Tw,in were in range from 283

to 325 K. Allowed tolerance for air and water inlet temperatures was ±1 K and was maintained by

electronic regulating system. The choice of water flow rate was based on a principle that a

temperature drop on a water side had to be higher than 5 K with constant flow for all measurements.

Obtained water temperature drop was rather high due to high pressure drop that occurs in

microchannels.

Data acquisition system was used to record air and water inlet and outlet temperatures with

frequency of 4 Hz. Two metrics were used to evaluate the quality of recorded data. The first one

was coefficient of variation σ, i.e. the standard deviation. Standard deviation, for all measurement

results given in this paper (for each used RTD sensor), was in all cases lower than 0.3. Second used

metric was based on capacity deviations and values of heat transfer rate that were used to calculate

heat balance error. Accounting for all instrument errors, property uncertainties and geometry

tolerances, there should be at least an 11.5% margin at the time of any testing (Moffat, 1988; Junqy

et al., 2007; Tatara and Lupia, 2011). In this investigation unmeasured losses in manifold section

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had some additional influence on capacity deviations and value of heat balance error. All the

experimental data was obtained on a basis of heat balance error less than 7%.

56 valid experimental measurements have been taken with different air and water temperatures and

flows. Time needed to carry out one measurement was approximately 30 minutes after achievement

of appropriate heat balance, totally one hour for each parameter set up.

3 MATHEMATICAL AND NUMERICAL APPROACH

3.1 Physical Model Air/water numerical simulation has been done. Due to limitations on the computer resources, only

portion of the heat exchanger able to describe flows of air and water was taken into account. Two

symmetry planes were assumed in the z-direction that divide flat tubes in two symmetrical parts.

Assumed planes were perpendicular to the fin surface. A schematic view of the computational

domain has been shown in Fig. 4. The upstream and downstream regions have not been presented in

proportional dimensions.

The computational domain consists of six volume groups: air upstream region (1), internal airspace

(2), air downstream region (3), fins (4), water region (5), and two halves of flat tubes (6). The total

length of the computational domain has been extended 9 times from actual internal airspace. The

upstream region has been extended 2.5 times to ensure inlet uniformity. The downstream region has

been extended 6 times in order to prevent flow recirculation.

3.2 Governing Equations and Boundary Conditions The governing equations in the Cartesian coordinate system for forced steady, laminar, three

dimensional, incompressible fluid flow and heat transfer in air, water and fin subdomains are:

3.2.1 Air and water subdomain:

Continuity:

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( )div 0.ρ =ur

(1)

Momentum:

( ) ( )x... div div grad ;p

u ux

ρ µ∂= − +∂

ur

(2)

( ) ( )y... div div grad ;p

v vy

ρ µ∂= − +∂

ur

(3)

( ) ( )z... div div grad .p

w wz

ρ µ∂= − +∂

ur

(4)

Energy:

( )p

div div grad .k

T Tc

ρ

=

ur

(5)

In case of air subdomain, equations (1) to (5) comprise physical properties of air (ρa, µa, ka and cp,a)

whilst in case of water subdomain, they comprise physical properties of water (ρw, µw, kw and cp,w).

3.2.2 Fin and flat tube subdomain:

p

div grad 0.k

Tc

=

(6)

In case of fin subdomain, equation (6) comprises physical properties of fin (kf and cp,f) whilst in

case of flat tube subdomain, it comprises physical properties of flat tube (kt and cp,t).

3.2.3 Boundary conditions are:

At the air inlet boundary:

a a,in a a a a,inconst., = = 0, const.u u v w T T= = = = (7)

At the water inlet boundary:

w w,in w w w w,inconst., = = 0, const.v v u w T T= = = = (8)

At the air outlet boundary:

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a a a a = 0, = 0, = 0, = 0. u v w T

x x x x

∂ ∂ ∂ ∂∂ ∂ ∂ ∂

(9)

At the water outlet boundary:

w w w w = 0, = 0, = 0, = 0.u v w T

y y y y

∂ ∂ ∂ ∂∂ ∂ ∂ ∂

(10)

At the left and right boundaries (air symmetry) in downstream and upstream region:

a a aa = 0, = 0, 0, = 0.

u w Tv

y y y

∂ ∂ ∂=∂ ∂ ∂

(11)

At the upper and lower boundaries (air symmetry) in downstream and upstream region:

a a aa = 0, = 0, 0, = 0.

u w Tw

z z z

∂ ∂ ∂=∂ ∂ ∂

(12)

At the upper and lower region of water and tube:

w w w tw = 0, =0, 0, = 0, = 0.

u v T Tw

z z z z

∂ ∂ ∂ ∂=∂ ∂ ∂ ∂

(13)

Air – fin, air – tube and water – tube interface respectively:

a f a t w ta f a t w t = , = , = .

T T T T T Tk k k k k k

n n n n n n

∂ ∂ ∂ ∂ ∂ ∂∂ ∂ ∂ ∂ ∂ ∂

(14)

3.3 Numerical Treatment The governing differential equations were discretized using the finite volume method, fully

described by Versteeg and Malalasekera (1995), on a hybrid, non-orthogonal grid. The domain has

been divided in 30 computational blocks and meshed using GAMBIT software. The number of

control volumes for MCHX was close to 110 million divided in 30 computational blocks (23 fins

each computational block, total of 690 fins, approximately 4 million control volumes per

computational block). Computational blocks were used due to limitations on the computer resources

(available physical memory) that could severely slow down calculation process. For each control

volume all calculated values (velocities, temperatures, pressures, etc.) on interface regions between

computational blocks were copied to the next block as shown on Fig. 5. Numerical simulation used

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this way saves some time and excludes need for supercomputing on hardware with large amount of

physical memory. Certainly, some tests of numerical procedure validity had to be taken.

In order to validate the solution independency of the grid volumes number, three different grid

systems of MCHX were investigated. Air and water input parameters were: ua,in = 0.29 m s-1, Ta, in =

301.5 K, vw,in = 0.06 m s-1 and Tw, in = 285.9 K. The grid volumes number dependence on final

results has been made for heat exchanger consisting of five fins. The same reason for use of heat

exchanger with only five fins, as before mentioned, was due to the available computing resources,

especially of the available physical memory. The predicted air and water outlet temperatures, as

well as air and water pressure drops for selected grid systems have been shown in Tab. 1. It can be

seen that the solution is grid-independent and that the grid system with smallest number of grid

volumes can be used for further investigations. Further reduction of grid volumes number was not

possible as solver was not able to converge to solution.

Similar validation has been done in order to validate the solution independency of the total number

of used computational blocks. Validation has been done for geometry of tested MCHX with

reduced number of fins (25, compared to total number of 690) and results of this simulations have

been shown in Tab. 2. Therefore, reduced number of fins was chosen in a way that it doesn’t exceed

computer physical memory limits during of calculations for the whole heat exchanger at once (one

computational block). Calculated temperatures and pressure drops have been compared with results

acquired for two other cases (domain consisting of 5 computational blocks with five fins each and

other consisting of 25 computational blocks with only one fin). From Tab. 2 it can be seen that the

solution system of the numerical model of heat exchanger with microchannel coil can be regarded

as computational block size-independent.

Heat transfer and fluid flow simulations were performed using commercial fluid flow and a heat

transfer solver FLUENT. All velocity vectors and temperatures between computational blocks were

linked with internal scripting language employed in CFD software. Fluid was assumed to be

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incompressible with constant property values and the flow was assumed to be laminar. The

SIMPLE algorithm for pressure-velocity coupling was used to ensure mass conservation and to

obtain a pressure field. The convection-diffusion terms have been discretized using the power law

scheme.

3.4 Matching of Computational Domain and Heat Exchanger Used in Experiment For the purpose of numerical analysis and in accordance with the heat exchangers used for

experiments, models with adequate geometry in direction of both air and water flow have been

developed. Fig. 6 shows positioning of RTD sensors set for temperature measurement of air leaving

heat exchanger relative to computational domain. Temperatures for each point, Ta,out, used in later

comparison have been calculated as average values of three measured temperatures for each RTD

sensor column:

3 3 3

a,out1 a,out,n1 a,out2 a,out,n2 a,out3 a,out,n31 1 1

= , = , = . n n n

T T T T T T= = =∑ ∑ ∑ (15)

3.5 Comparison Between Experimental and Numerical Results Comparison between numerical and experimental results has been accomplished in same ranges of

air and water temperatures and velocities used in experimental investigation. Series of numerical

calculations have been done in order to analyse heat transfer and fluid flow. Due to limited paper

length, results of comparisons have been shown only for several parameter setups given in Fig. 7.

and Tab. 1., accompanied with measured and numerically obtained water and air temperatures.

Setups have been chosen with caution and care of operational conditions of similar compact heat

exchangers used in practice. Error bars on Fig 7. have been given with fixed value of ±0.15 K

according to maximal obtained standard deviation of air outlet temperatures.

It can be seen that numerical simulation results coincide well with the experimental data and that

deviations are within an acceptable range. Temperature differences are smaller than ±0.5 K for all

valid measurements, what can be taken as assertion of measurement methodology validity as well

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as proof of used numerical simulation validity. Comparison of experimental and numerical analyses

has been done in order to validate used air/water numerical simulation. Therefore, it can be

concluded that air/water side simulation described in this paper can be used for further

investigations of enhanced and improved surface geometries of heat exchangers with microchannel

coil both on water side and air side.

4 THERMODYNAMICAL DEPENDENCE ON MICROCHANNEL SHAPE

4.1 Description of the Microchannel Shapes The influence of different microchannel shapes on single-phase heat transfer and pressure drop has

been studied numerically. Numerical tests have been conducted on flat tubes with circular and non-

circular channels. Agarwal et al. (2010) tested experimentally the effect of tube shape on heat

transfer during condensation. Rectangular, barrel and circular shapes used in their paper were used

as base for development of microchannel shapes examined in this paper. Microchannel shapes used

in this investigation are shown on Fig. 8.

Flat tubes with multiple extruded parallel channels of rectangular (1), barrel (2), circular (3),

hexagonal (4), vertical rectangular (5) and square diamond shapes (6) were examined numerically.

Number of channels per flat tube was in all cases 20. Total cross section area of small channels was

held constant in all cases. It consists of cross section areas of 20 small channels (Atot = 26.4 mm2).

Fin pitch of tested heat exchangers was Fp = 1.6 mm, and fin thickness was Ft = 0.1 mm. Fins were

assumed to be flat and construction material of flat tubes and fins was aluminium.

4.2 Microchannel Shape Effect on Heat Transfer Effectiveness and Pressure Drop

Comparison between different microchannel shapes has been done according to correspondent heat

transfer effectiveness and pressure drop. Both values were obtained by numerical simulations. The

heat transfer effectiveness was defined as the ratio of the actual amount of heat transferred to the

possible amount of heat that could be transferred. According to the Bosnjakovic and Knoche

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(1998), heat transfer effectiveness is equal to the ratio between temperature difference of fluid with

lower heat flux and given inlet temperature differences of both fluids. Heat flux of both fluids was

checked for each case of both numerical and experimental investigation and fluid with lower heat

flux was determined. In this investigation lower heat flux was on air side in all cases and heat

transfer effectiveness was defined by following equation:

a,in a,out

a,in w,in

.T T

T Tε

−=

−

(16)

The selected operating conditions of Ta,in = 273 K, Tw,in = 313 K, ua,in = 0.5 to 2 m s-1 and vw,in = 0.5

to 2 m s-1 were applied in numerical simulation for the air/water heat exchanger. Fig. 9 shows

microchannel shape effect on air side heat transfer effectiveness for different inlet air velocities and

constant water inlet velocity (vw,in = 1 m s-1).

The heat transfer effectiveness has highest values in range of low inlet air velocities. In that range

microchannel shape effect on heat transfer effectiveness is almost negligible. With rise of inlet air

velocity, heat transfer effectiveness falls down and at average inlet air velocity close to 2 m s-1gets

values between 0.6 and 0.7. Best results from heat transfer effectiveness point of view, which are

more noticeable in range of higher air velocities, are accomplished with channels of diamond and

hexagonal shapes. It should be noted that difference between heat transfer effectiveness for different

shapes gets bigger with increase of inlet air velocity. Fig. 10 shows appropriate microchannel shape

effect on water side pressure drop with constant air inlet velocity (ua,in = 1 m s-1).

Undesirable raise of pressure drop comes with increase of water velocity. In range of lower water

velocities microchannel shape effect on pressure drop is lower than in range of higher water

velocities. Best results, meaning lowest values of pressure drop, are obtained with heat exchanger

that consists of flat tubes with rectangular shape of high aspect ratio that is close to square shape.

Second best obtained results, according to microchannel shape effect on pressure drop, is for flat

tubes with circular and barrel shapes. These results coincide well with results of Hasan et al. (2009)

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who investigated influence of channel geometry on the performance of counter flow microchannel

heat exchanger. Extremely high pressure drops on water side of microchannels with hexagonal,

square and diamond shapes discard gains achieved in point of higher heat transfer effectiveness.

In case of rectangular microchannel shapes, aspect ratios (h/w) influence on pressure drop,

shouldn’t be neglected. According to Lee and Garimella (2006), its value should always be higher

than 0.6 to assure best thermodynamical and hydrodynamical characteristics of heat transfer.

Comparison of obtained pressure drops for two numerically investigated rectangular shapes has

been shown on Fig. 11.

At low water velocities, advantages of better selected rectangular microchannel aspect ratios are not

so significant. As opposed, these advantages become more and more significant with the increase of

water velocity, since proper selection of these aspect ratios prevents undesired and expensive

increase of pressure drop. Rectangular microchannel aspect ratios should be selected with extreme

caution.

5 CONCLUSIONS

The main aim of present analysis was to make comparison of the compact heat exchangers with

different microchannel shapes. Comparison has been made regarding to heat transfer effectiveness

and pressure drops in single-phase heat transfer. After experimental validation of air/water

simulation, numerical analysis has been carried out. Results showed that heat transfer effectiveness

decreases with rise of inlet air velocity. Best achievements regarding same parameters were

accomplished with microchannels of diamond and hexagonal shape. Although higher heat transfer

effectiveness of mentioned shapes, according to water side pressure drop the best performance is

achieved by rectangular shape. This kind of geometry shape has already been implemented in water

chillers of several manufacturers as a condensing unit. High pressure drop and consequently higher

needed pumping power are some of the reasons why this construction type of compact heat

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exchanger hasn’t been implemented in greater number of HVAC units that include single-phase

heat exchangers.

ACKNOWLEDGMENTS

This research has been performed as part of the scientific project Research and Development of

Renewable Energy Components and Systems, supported by the Ministry of Science, Education and

Sports of the Republic of Croatia.

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Figure Captions

Fig. 1 – Schematic diagram of the wind tunnel test apparatus.

Fig. 2 – Measuring station of wind tunnel.

Fig. 3 – Schematic diagram of a tested heat exchanger and used geometrical parameters.

Fig. 4 – Schematic view of computational domain.

Fig. 5 – Schematic view of computational blocks.

Fig. 6 – Positioning of RTD sensors and computational domain

Fig. 7 – Comparison of numerical and experimental results for different air volume flows: 1200 m3

h-1 (a), 1950 m3 h-1 (b) and 2500 m3 h-1 (c).

Fig 8 – Microchannel geometries of rectangular (1), barrel (2), circular (3), hexagonal (4), vertical

rectangular (5) and square diamond shapes (6) used in numerical analysis (all units in mm).

Fig. 9 – Microchannel shape effect on air side heat transfer effectiveness.

Fig. 10 – Microchannel shape effect on water side pressure drop.

Fig. 11 – Water side pressure drops for two different rectangular shape aspect ratios.

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Table Captions

Table 1 – Variation of the predicted temperatures and pressure drops for three grid volume number

systems

Table 2 – Variation of the predicted temperatures and pressure drops for three computational block

size systems

Table 3 – Comparison of numerical and experimental results for water side temperatures.

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Table 1 - Variation of the predicted temperatures and pressure drops for three grid volume

number systems.

Grid volumes number (MCHX – 5 fins)

Tw,out [K]

Ta,out [K]

∆pw [Pa]

∆pa [Pa]

709 920 285.993 285.932 18.0 3.1

1 891 260 285.993 285.932 18.0 3.1

7 386 645 285.993 285.934 18.6 3.1

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Table 2 - Variation of the predicted temperatures and pressure drops for three computational

block size systems.

Size of computational blocks (MCHX - number of fins)

Tw,out [K]

Ta,out [K]

∆pw [Pa]

∆pw,tot [Pa]

∆pa [Pa]

1 286.220 286.005 3.5 87.5 3.1

5 286.210 286.010 18.1 90.5 3.1

25 286.194 286.048 89.1 89.1 3.1

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Table 3 – Comparison of numerical and experimental results for water side temperatures.

Case No.

Tw,in [K]

Experimental Tw,out [K]

Numerical Tw,out [K]

Error /

[K]

Average heat transf. rate

Q [kW]

#1 310.0 306.6 306.8 +0.2 7.10

#2 308.6 305.7 305.6 -0.1 6.50

#3 308.1 304.6 304.5 -0.1 7.33

#4 307.5 304.5 304.3 -0.2 6.40

#5 308.5 304.1 304.3 +0.2 9.22

#6 307.3 303.2 303.0 -0.2 8.99

#7 307.6 302.4 302.5 +0.1 11.31

#8 305.4 300.3 300.5 +0.2 11.25

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ACCEPTED MANUSCRIPT1. Compact heat exchanger with microchannel coil (MCHX) is investigated experimentally. 2. The parameters affecting heat transfer efficiency and pressure drop are clarified. 3. Numerical air/water simulation of MCHX was established and validated. 4. Guide to more efficient use of MCHX with different microchannel shapes has been given.