numerical analysis of a fin-tube plate heat exchanger with winglets ...

NUMERICAL ANALYSIS OF A FIN-TUBE PLATE

HEAT EXCHANGER WITH WINGLETS

LAFTA FLAYIH MUHSIN AL-JUHAISHI

UNIVERSITI TUN HUSSEIN ONN MALAYSIA

v

ABSTRACT

In this presented work, numerical analysis of heat transfer and flow characteristic using

longitudinal vortex generators (LVGS) in fin and flat tube heat exchanger has been

presented. Conjugate heat transfer 3D numerical model has been developed and

successfully carried out. Rectangular winglets were set in pairs, with downstream

orientation. The effects of impact angles of (20⁰ , 30⁰, and 40⁰ ) as well as tubes and

winglets were placed in one row lined arrangement and air flow by forward

arrangement and backward arrangement. Reynolds number is ranged from 500 to 5000.

The numerical results showed that in the range of the present study, the variation of

these parameters can result in the increase of heat transfer. The study focuses on the

Influence of the different parameters of VGs on heat transfer and fluid flow

characteristics of one row lined circular-tube banks. The characteristics of average Nu

number and skin friction coefficient are studied numerically by the aid of the

computational fluid dynamics (CFD) commercial code of FLUENT ANSYS 14. The

results showed increasing in the heat transfer and skin friction coefficient with the

increasing of Re number. It has been observed that the overall Nuav number of one

circular tubes increases by 23-31% ,by 23-43% and by 23-47% with angles of (20⁰,

30°, and 40⁰) respectively, in forward arrangement and the overall Nuav number of one

circular tubes increases by 23-42%, by 23-46% and 23-52%with angles of (20⁰, 30°,

and 40⁰) respectively, in backward arrangement, with increasing in the overall average

of skin friction coefficient. Also the results showed that the rectangular winglet pairs

(RWPs) can significantly improve the heat transfer performance of the fin and-tube

heat exchangers with a moderate pressure loss penalty.

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ABSTRAK

Dalam ini dibentangkan kerja, analisis berangka pemindahan haba dan mengalir ciri

menggunakan penjana pusaran membujur (LVGS) di sirip dan rata tiub penukar haba

telah dibentangkan. Pemindahan haba konjugat 3D model berangka telah

dibangunkan dan berjaya dilaksanakan. Sayap lawi segi empat tepat telah ditetapkan

secara berpasangan, dengan orientasi hiliran. Kesan sudut kesan (20⁰, 30⁰, 40⁰ dan)

serta tiub dan sayap lawi diletakkan dalam satu baris berbaris susunan dan aliran

udara melalui perkiraan ke hadapan dan ke belakang perkiraan. Nombor Reynolds

adalah di antara 500 hingga 5000. Keputusan berangka menunjukkan bahawa dalam

julat kajian ini, perubahan parameter ini boleh menyebabkan peningkatan

pemindahan haba. Kajian ini memberi tumpuan kepada Pengaruh parameter yang

berbeza VGS pada pemindahan haba dan ciri-ciri aliran cecair daripada satu baris

berbaris bank bulat-tiub. Ciri-ciri purata bilangan dan geseran kulit Nu pekali dikaji

secara berangka dengan bantuan dinamik bendalir pengiraan (CFD) kod komersial

FLUENT ANSYS 14. Keputusan menunjukkan peningkatan dalam pemindahan haba

dan pekali geseran kulit dengan peningkatan jumlah Re. Ia telah diperhatikan bahawa

jumlah keseluruhan Nuav satu tiub bulat meningkat sebanyak 23-31%, oleh 23-43%

dan 23-47% dengan dengan sudut (20⁰, 30 °, dan 40⁰) masing-masing, dalam

susunan ke hadapan dan jumlah Nuav keseluruhan satu tiub bulat meningkat

sebanyak 23-42%, oleh 23-46% dan 23-52% dengan sudut (20⁰, 30 °, dan 40⁰)

masing-masing, dalam susunan ke belakang, dengan peningkatan dalam purata

keseluruhan kulit pekali geseran. Juga keputusan menunjukkan bahawa pasangan

sayap lawi segi empat tepat (RWPs) dengan ketara boleh meningkatkan prestasi

pemindahan haba daripada sirip dan-tiub penukar haba dengan kehilangan tekanan

penalti sederhana.

vii

CONTENTS

TITLE i

DECLARATION ii

DEDICATION iii

ACKNOWLEDGEMENT iv

ABSTRACT v

CONTENTS vii

LIST OF FIGURE x

LIST OF ABBREVIATION AND SYMBOLS xvi

CHAPTER 1 INTRODUCTION

1.1 Heat Transfer Enhancement 1

1.2 Classification of Enhancement Techniques 3

1.3 Fin-Tube Heat Exchanger with Winglet 4

1.4 Problem statement 7

1.5 Project Objectives 7

1.6 Project Scopes 8

1.7 Research Significance 8

CHAPTER 2 LITERATURE REVIEW

2.1 Introduction 9

2.2 Winglet in Tube-Fin Heat Exchanger 9

2.3 Winglets in Refrigerator Evaporator 12

2.4 Winglet-Type Vortex Generators Insert 14

CHAPTER 3 NUMERICAL METHODOLOGY

3.1 Introduction 28

3.2 Computational fluid dynamics 29

3.3 Problem-Solving with CFD 29

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3.4 The CFD Modelling Process 30

3.5.1 Physical Model 32

3.5.2 Governing Equations 33

3.5.3 Boundary Conditions 35

3.6 Finite Volume Method (FVM) 36

3.7 Initial test for a fin-tube plate heat exchanger with winglets 35

3.8 Fluid Flow Computation by SIMPLE Algorithm 36

3.9 Summary 37

CHAPTER 4 RESULTS AND DISCUSSIONS

4.1 Introduction 38

4.2 Calculation of Code Validation 39

4.3. Results and Discussion 40

4.4. Effect of longitudinal vortex generators LVGs 41

4.5. Reynolds number (Re) effect 43

4.6 The angle of winglets α = (20°, 30° and 40°) forward 43

4.7. The effect of Longitudinal Vortex Generators (LVGs) 47

on Skin Friction coefficient (f)

4.8 The angle of winglet α=20⁰ forward arrangement 52

(Local velocity distribution)

4.9 The angle of winglet Forward arrangements 56

(Distribution Temperature)


(Surface Nusselt Number)

4.11 The angle of winglet Forward arrangement 62

(Local Velocity Distribution)





4.14 The angle of winglet forward arrangement 71



ix




4.17 The angle of winglet (α=20°, 30° and 40°) 80

in backward arrangement

4.18 The angle of winglet α=20⁰ backward arrangement 84

(Local velocity)

4.18 The angle of winglet Backward arrangements 87




4.20 The angle α= 30⁰ Backward arrangement 93






4.23. Influence of the angle of attack 101

4.24 The angle α= Backward arrangement 102






4.27 Discussion 113

4.27.1 Influence Longitudinal Vortex Generators of 113

LVGs Parameters on Heat Transfer (Nu)

CHAPTER 5 CONCLUSION AND RECOMMENDATIONS

5.1 Conclusion 114

5.2 Recommendations 116

REFERENCES 117

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LIST OF FIGURES

NO. TITLE OF FIGURES PAGE

1.1 Fin-tubes heat exchanger 2

1.2 Flat plate fin-tubes for heat exchanger 2

1.3 Different types of flat plate fin-tubes for heat exchanger 5

1.4 Flat plate fin-tubes for heat exchanger without winglets 6

1.5 Flat plate fin-tubes for heat exchanger with vortex generator 7

1.6 Fin-tubes heat exchanger with round and oval tubes 7

3.1 The general modelling process 29

3.2 Shows schematic diagram of core region of a fin-and-tube heat 30

exchanger with RWPs.

3.3 Schematic diagram of a fin-tube heat exchanger with winglets 31

3.3 The computational domain 34

3.4 Figures (3.4, a, b, c, and d) as shown above represent initial test 36

4.1 4.1(a, b, and c): shows baseline, Forward and Backward 39

4.2 Relationship between Re and Nu 40

4.3 Distribution of the velocity through tubes bank in tube bank 41

without winglets at (V=0.464 m/s).

4.4 Shows Relation between Reynolds numbers (Re) and Nusselt numbers 42

(Nu) in baseline case

4.5 Shows Relation between Reynolds numbers (Re) and friction factor 42

(f) in baseline case

4.6 shows relation between Reynolds number (Re) and Colburn factor (j) 43

in baseline case


(Nu) of baseline and α=20° forward.





4.10 Shows Relation between Reynolds numbers (Re) and Nusselt 47

numbers (Nu) of baseline and α=20°, 30°, 40° forward arrangement.

xi

4.11 Shows relationship between Reynolds numbers (Re) with friction 48

factor (f) in cases baseline and α=20°, 30°, 40° forward arrangement.

4.12 Shows relationship between Reynolds number (Re) and surface heat 49

coefficient h (w/ in cases baseline and α=20°, 30°, 40°forward

arrangement.

4.13 Shows relation between Reynolds numbers (Re) with Colbrun 50

factor (f) in forward arrangement.

4.14. a Baseline Re=500 52

4.14. b =20⁰ Forward arrangement and Re=500 52

4.14. c =20⁰ Forward arrangement and Re=1000 53

4.14. d =20⁰ Forward arrangement and Re=2000 53

4.14. e: =20⁰ Forward arrangement and Re=3000 53

4.14.f: =20⁰ Forward arrangement and Re=4000 54

4.14.g: Forward arrangement and Re=5000 54

4.15.a: Baseline Re=500 56

4.15.b: =20⁰ Forward arrangement and Re=500 56

4.15. c =20⁰ Forward arrangement and Re=1000 57

4.15. d: =20⁰ Forward arrangement and Re=2000 57


4.15.f: =20⁰ Forward arrangement and Re=4000 58

4.15.g: Forward arrangement and Re=5000 59

4.16. a: Baseline Re=500 60

4.16. b: =20⁰ Forward arrangement and Re=500 60

4.16. c: =20⁰ Forward arrangement and Re=1000 61

4.16. d: =20⁰ Forward arrangement and Re=2000 61


4.16. f: =20⁰ Forward arrangement and Re=4000 62

4.16. g: Forward arrangement and Re=5000 62

4.17. a: Baseline Re=500 63

4.17. b: Forward arrangement and Re=500 ( velocity) 64

4.17. c: Forward arrangement and Re=1000 (velocity) 64

4.17. d: Forward arrangement and Re=2000 (velocity) 65

4.17. e: Forward arrangement and Re=3000 (velocity) 65

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4.17. f: Forward arrangement and Re=4000 (velocity) 65

4.17. g: Forward arrangement and Re=5000 (velocity) 66

4.18.a: Baseline Re=500 ( Temperature) 66

4.18.b: Forward arrangement and Re=500 (Temperature) 67

4.18. c: Forward arrangement and Re=1000 (Temperature) 67

4.18. d: Forward arrangement and Re=2000 (Temperature) 67

4.18. e: Forward arrangement and Re=3000 (Temperature) 68

4.18. f: Forward arrangement and Re=4000 (Temperature) 68

4.18. g: Forward arrangement and Re= 5000 (Temperature) 69

4.19. a: Baseline Re=500 (Nusselt number) 69

4.19. b: Forward arrangement and Re=500 (Nusselt number) 70

4.19. c: Forward arrangement and Re=1000 (Nusselt number) 70

4.19. d: Forward arrangement and Re=2000 (Nusselt number) 70

4.19. e: Forward arrangement and Re=3000 (Nusselt number) 70

4.19. f: Forward arrangement and Re=4000 (Nusselt number) 71

4.19.g: Forward arrangement and Re= 5000 (Nusselt number) 71

4.20. a: Baseline Re=500 (Velocity) 73

4.20. b: =40⁰ Forward arrangement and Re=500 (Velocity) 73

4.20. c: =40⁰ Forward arrangement and Re=1000 (Velocity) 74

4.20. d: =40⁰ Forward arrangement and Re=2000 (Velocity) 74

4.20. e: =40⁰ Forward arrangement and Re=3000 (Velocity) 74

4.20. f: =40⁰ Forward arrangement and Re=4000 (Velocity) 75

4.20. g: Forward arrangement and Re=5000 (Velocity) 75

4.21. a: Baseline Re=500 (Temperature) 76

4.21. b: Forward arrangement and Re=500 (Temperature) 76

4.21. c: Forward arrangement and Re=1000 (Temperature) 77

4.21. d: Forward arrangement and Re=2000 (Temperature) 77

4.21. e: Forward arrangement and Re=3000 (Temperature) 78

4.21. f: Forward arrangement and Re=4000 (Temperature) 78

4.21. g: Forward arrangement and Re= 5000 (Temperature) 79


4.22. b: Forward arrangement and Re=500 (Nusselt number) 80

4.22.c: Forward arrangement and Re=1000 (Nusselt number) 80

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4.22. d : Forward arrangement and Re=2000 (Nusselt number) 80

4.22. e: Forward arrangement and Re=3000 (Nusselt number) 81

4.22. f: Forward arrangement and Re=4000 (Nusselt number) 81

4.22. g: Forward arrangement and Re= 5000 (Nusselt number) 82

4.23: Relation between Reynolds numbers (Re) and Nusselt 83

numbers (Nu) in case baseline and α=20° backward.

4.24: Shows Relation between Reynolds numbers (Re) and Nusselt 84

Numbers (Nu) in case baseline and α=30° backward.

4.25: Relation between Reynolds numbers (Re) and Nusselt 84

numbers (Nu) in case baseline and α=40° backward.

4.26: Relation between Reynolds number (Re) Nusselt number (Nu) 85

in cases baseline and α=20°, 30°, 40° backward arrangement.

4.27: Relationship between Reynolds number (Re) and surface heat 86

coefficient h(w/ ). in cases baseline and α=20°, 30°, 40°

backward arrangement


4.28. b: =20⁰ Backward arrangement and Re=500 (Velocity) 87

4.28. c: =20⁰ Backward arrangement and Re=1000 (Velocity) 88

4.28. d: =20⁰ Backward arrangement and Re=2000 (Velocity) 88

4.28. e: =20⁰ Backward arrangement and Re=3000 (Velocity) 88

4.28. f: =20⁰ Backward arrangement and Re=4000 (Velocity) 89

4.28. g : Backward arrangement and Re= 5000 (Velocity) 89


4.29. b: =20⁰ Backward arrangement and Re=500 (Temperature) 90

4.29. c: =20⁰ Backward arrangement and Re=1000 (Temperature) 91

4.29. d: =20⁰ Backward arrangement and Re=2000 (Temperature) 91

4.29. e: =20⁰ Backward arrangement and Re=3000 (Temperature) 91

4.29. f: =20⁰ Backward arrangement and Re=4000 (Temperature) 92

4.29. g: Backward arrangement and Re= 5000 (Temperature) 92


4.30. b : =20⁰ Backward arrangement and Re=500 (Nusselt number) 93

4.30. c: =20⁰ Backward arrangement and Re=1000 (Nusselt number) 94

4.30. d: =20⁰ Backward arrangement and Re=2000 (Nusselt number) 94

xiv

4.30. e: =20⁰ Backward arrangement and Re=3000 (Nusselt number) 94

4.30. f: =20⁰ Backward arrangement and Re=4000 (Nusselt number) 95

4.30. g: Backward arrangement and Re= 5000 (Nusselt number) 95







4.31. g: Backward arrangement and Re= 5000 (Velocity) 98

4.32. a Baseline Re=500 (Temperature) 99

4.32.b: Backward arrangement and Re=500 (Temperature) 99

4.32. c: Backward arrangement and Re=1000 (Temperature) 100

4.32. d: Backward arrangement and Re=2000 (Temperature) 100

4.32. e: Backward arrangement and Re=3000 (Temperature) 100

4.32. f: Backward arrangement and Re=4000 (Temperature) 101

4.32. g: Backward arrangement and Re= 5000 (Temperature) 101


4.33. b: Backward arrangement and Re= 500 (Nusselt number) 102

4.33. c: Backward arrangement and Re= 1000 (Nusselt number) 103

4.33. d: Backward arrangement and Re= 2000 (Nusselt number) 103

4.33. e: Backward arrangement and Re= 3000 (Nusselt number) 103

4.33. f: Backward arrangement and Re= 4000 (Nusselt number) 104








4.34. g: Backward arrangement and Re=5000 (Velocity) 108


4.35. b: Backward arrangement and Re=500 (Temperature) 109

xv

4.35. c: Backward arrangement and Re=1000 (Temperature) 110

4.35. d: Backward arrangement and Re=2000 (Temperature) 110

4.35. e: Backward arrangement and Re=3000 (Temperature) 111

4.35. f: Backward arrangement and Re=4000 (Temperature) 111

4.35. g: Backward arrangement and Re=5000 (Temperature) 112


4.36. b: Backward arrangement and Re=500 (Nusselt number) 113

4.36. c: Backward arrangement and Re=1000 (Nusselt number) 113

4.36. d: Backward arrangement and Re=2000 (Nusselt number) 114

4.36. e: Backward arrangement and Re=3000 (Nusselt number) 114

4.36. e: Backward arrangement and Re=4000 (Nusselt number) 114


4.37: Shows relationship between Reynolds number (Re)and Colburn 115

factor (j) and comparison between backward arrangement

for angles ( α=20⁰, 30⁰, and 40⁰ ).

4.38: Shows relationship between Reynolds number(Re) and friction 116

factor and comparison between backward arrangement for

angles ( α=20⁰, 30⁰, and 40⁰ ).

4.39: Shows relationship between Reynolds number (Re) Nusselt 116

number (Nu) and comparison between forward arrangement

forangles (α=20⁰, 30⁰, and 40⁰) and backward arrangement

for angles ( α=20⁰, 30⁰, and 40⁰ ).

4.40: Shows relationship between Reynolds number (Re) and friction 117

factor(f) and comparison between forward arrangement

for angles (α=20⁰, 30⁰, and 40⁰) and backward arrangement

for angles ( α=20⁰, 30⁰, and 40⁰ ).

xvi

LIST OF ABBREVIATION AND SYMBOLS

Re Reynolds number

Density of air ρ

Inlet velocity of air (m/s)

Hydraulic diameter (m)

µ Viscosity of air

L Length of heat exchanger (m)

t Fin thickness (m)

Colburn factor

Friction factor

Nu Nusselt number

Average Nusselt number

h Convection heat transfer coefficient

k Thermal Conductivity

α Angle of winglet ( )

Heat flux on the heating surface

Pr Prandtl number

P Static Pressure (pa)

D Tube diameter (m)

w width of heat exchanger (m)

H Fin pitch (m)

xvii

WL Winglet length (m)

Specific heat capacity (J/kg k)

q Heat transfer rate (w)

Central angle (m)

Distance between tube and winglet t (m)

T Temperature (k)

LVs Longitudinal vortices

LVGs Longitudinal vortex generators

RWPs Rectangular winglet pairs

VG Vortex generator

1

CHAPTER 1

INTRODUCTION

1.1 Heat Transfer Enhancement

The conversion, utilization, and recovery of energy in every industrial, commercial,

and domestic application involve a heat exchange process. Some common examples

are steam generation and condensation in power and cogeneration plants; sensible

heating and cooling of viscous media in thermal processing of chemical,

pharmaceutical, and agricultural products; refrigerant evaporation and condensation

in air-conditioning and refrigeration; gas flow heating in manufacturing and waste-

heat recovery; air and liquid cooling of engine and turbo machinery systems; and

cooling of electrical machines and electronic devices. Improved heat exchange, over

and above that in the usual or standard practice, can significantly improve the

thermal efficiency in such applications as well as the economics of their design and

operation.

The engineering cognizance of the need to increase the thermal performance

of heat exchangers, thereby effecting energy, material, and cost savings as well as a

consequential mitigation of environmental degradation had led to the development

and use of many heat transfer enhancement techniques. In the past, these methods

have been variously referred to as augmentation and intensification, among other

terms [1].

2

Enhancement techniques essentially reduce the thermal resistance in a

conventional heat exchanger by promoting higher convective heat transfer coefficient

with or without surface area increases (as represented by fins or extended surfaces).

As a result, the size of a heat exchanger can be reduced, or the heat duty of a heat

exchanger can be increased, or the pumping power requirements can be reduced, or

the exchanger’s operating approach temperature difference can be decreased [2].

Figure (1.1) fin-tube heat exchanger

Figure (1.2) flat plate fin-tubes for heat exchanger

The latter is particularly useful in thermal processing of biochemical, food,

plastic, and pharmaceutical media to avoid the thermal degradation of the end

product. On the other hand, heat exchange systems in spacecraft, electronic devices,

3

and medical applications, for example, may rely primarily on the enhanced thermal

performance for their successful operation.

1.2 Classification of Enhancement Techniques

Sixteen different enhancement techniques have been identified by Bergles [2],

which can be classified broadly as passive and active techniques.

A list of the various methods or devices under each of these two categories is

given in Table (1-1(. The primary distinguishing feature is that, unlike active

methods, the passive techniques do not require direct input of external power.

Table (1-1) Classification of various heat transfer enhancement techniques [2].

Passive Techniques Active Techniques

Treated surface Mechanical aids

Roughness surface Surface vibration

Extended surface Fluid vibration

Displaced enhancement devices Electrostatic fields

Swirl flow devices Injection

Coiled tubes Suction

Surface tension devices Jet impingement

Additives for liquids

Additives for gases

Compound Enhancement

Two or more passive and/or active techniques

that are employed together

Generally, they use surface or geometrical modifications to the flow channel, or

incorporate an insert, material, or additional device. Except for extended surfaces

which increase the effective heat transfer surface area, these passive schemes

promote higher heat transfer coefficients by disturbing or altering the existing flow

behaviour. This, however, is accompanied by an increase in the pressure drop.

In the case of active techniques, the addition of external power essentially

4

facilitates the desired flow modification and the concomitant improvement in the rate

of heat transfer. The use of two or more techniques (passive and/or active) in

conjunction constitutes a compound enhancement.

The effectiveness of any of these methods is strongly dependent on the mode of

heat transfer (single-phase free or forced convection, pool boiling, forced convection

boiling or condensation, and convective mass transfer) and type and process

application of the heat exchanger. In considering their specific applications, a

descriptive characterization of each of the 16 techniques is useful in assessing their

potential.

1.3 Fin-Tube Heat Exchanger with Winglet

Tube-fin heat exchangers are mostly employed as gas-to-liquid exchangers.

Nowadays, many applications such as chemical, petroleum industries, thermal

processing systems in automotive, refrigeration and air conditioning are found for

tube-fine exchangers. Depending on the fin type, these exchangers are referred to as

finned tube exchangers (having normal fins on individual tube) and tube-fin

exchangers (having flat continuous fins). In these exchangers, the fins can be plain,

wavy, or interrupted, while round, flat and oval tubes may be used. A tube-fin

exchanger with flat fins is referred to as a plate finned tube, Figure (1.3).

5

Figure (1.3) different types of flat plate fin-tube for heat exchanger

Apart from the flow structure, geometrical parameters such as tube form (round or

flat tube). Arrangement (in line or staggered), and tube and fin spacing are effective

in the performance of these exchangers. For example, with an inline arrangement, the

horseshoe vortices may not be generated in front of the tubes of the second and the

other rows, while in the staggered arrangement, the horseshoe vortices appear in

front of each tube, which can influence the flow structure on the large area of the fin.

In recent years, vortex generators such as fins, ribs, wings etc. have been successfully

used for heat transfer enhancement of the modern thermal systems. Vortex

generators form secondary flow by swirl and destabilize the flow. They generate the

6

longitudinal vortices and create rotating and secondary flow in the main flow which

can raise turbulent intensity, mix the warm and cold fluid near and in the centre of

channel and increase the heat transfer in the heat exchangers. A lot of investigations

were mainly focused on circular-tube-fin and oval-tube-fin with mounted and

punched longitudinal vortex generators to enhance the heat transfer [3–14].

Figure (1.4) flat plate fin-tubes for heat exchanger without winglets

Figure (1.5) flat plate fin-tubes for heat exchanger with vortex generator

The energy transfer outside tube of economizers in coal-fired boilers needs more

attention and usually faces two significant problems: low heat transfer efficiency of

the gas, and serious tube erosion caused by ash and particles from coal combustion.

Recently, longitudinal vortex generators (LVGs) and dimples have been proved to

7

have a high performance in shell-side gas heat transfer enhancement. Zhang et al.

[15], [16] A typical finned flat and round tube bank with vortex generators is

presented in Fig(1.6) .

Figure (1.6) Fin-tubes heat exchanger with Round tubes and oval tubes

1.4 Problem statement

During my work 23 years experience in cement plant, we have found many problems

related to the subjects .There are many departments in the cement plant, such as

(Cement Mill, Raw mill, Rotary Kiln, and Air Compressors).

For this study we have chosen a problem in a cooling system for the Journal bearing

found in Cement Mill, symmetro gear box and main motor .This problem needs high

efficiency cooling systems, the same work principal with heat exchangers.

Therefore the principal of work in this research will be to increase the efficiency of

heat exchanger by using winglets.

1.5 Project Objectives

This numerical study includes

a) Study the effect of winglets in fin-tube heat exchanger.

b) Exam the effect of different Reynolds number on thermal and flow felid. The

ranges of Reynolds number are (500, 1000, 2000, 3000, 4000, and 5000).

8

c) Investigating the Nusselt number, friction factor for fin-tube heat exchanger with

winglet and without winglet.

d) Study the effect of forward and backward arrangement of winglet in fin-tube heat

exchanger

1.6 Project Scopes

Scope of this research is numerical studies of enhance heat transfer in fin- tube heat

exchanger with and without winglet, and this work can be summarized as follows :

a) Investigation numerically of enhance heat transfer in fin- tube heat exchanger.

b) Investigation numerically of effects of different Reynolds number on fin- tube

heat exchanger, heat transfer coefficient, Nusselt number and distribution of

temperature in tubes.

c) Study the effect of forward and back ward arrangement of winglets

1.7 Research Significance

The heat transfer enhancement can reduce the size and cost for heat exchanger by

using vortex generator element in the fin. Air-cooled condensers (ACCs) with

obvious water conservation benefit can be an important alternative for power plant

near coal mines where water source is of shortage. In direct air-cooled condensers of

power generating units, the ambient air replaces water as cooling medium. The

previous studies show that longitudinal vortex generators generate higher heat

transfer enhancement for the same pressure penalty than transverse vortex generators

in detail a base configuration with embedded rectangular vortex generators in

internal flow.

9

CHAPTER 2

LITERATURE SURVEY

2.1 Introduction

The purpose of this literature review is to go through the main topics of interest.

The enhancement of heat transfer by using a winglet in fin tube heat exchanger is the

subject of growing importance in myriad industrial applications; hence the fin- tube

heat exchanger has been the subject of several studies. This literature review will

address a number of experimental and theoretical studies which focused on the air

side performance of fin- tube heat exchanger. The experimental and numerical

studies carried out in order to cover the enhancement in heat transfer by using

winglet in a fin- tube heat exchanger in all applications and the effect of

configurations for heat exchangers.

2.2. Winglet in Tube-Fin Heat Exchanger

He et al [17] investigated numerically the effect of rectangular winglet pairs (RWPs)

on heat transfer enhancement and pressure loss penalty for fin-and-tube heat

exchangers. The rectangular winglet pairs RWPs were placed with a special

orientation for the purpose of enhancement of heat transfer with low Reynolds

number flow. The numerical study involved three-dimensional flow and conjugate

heat transfer in the computational domain, which was set up to model the entire flow

channel in the air flow direction. The effects of attack angle of rectangular winglet

10

pairs RWPs, row-number of RWPs and placement of RWPs on the heat transfer

characteristics and flow structure were examined in detail. It was observed that the

longitudinal vortices caused by RWPs and the impingement of RWPs-directed flow

on the downstream tube were important reasons of heat transfer enhancement for fin-

and-tube heat exchangers with RWPs. It was interesting to find that the pressure loss

penalty of the fin-and-tube heat exchangers with RWPs can be reduced by altering

the placement of the same number of RWPs from inline array to staggered array

without reducing the heat transfer enhancement. The results showed that the

rectangular winglet pairs (RWPs) were significantly improved the heat transfer

performance of the fin and-tube heat exchangers with a moderate pressure loss

penalty.

Huisseune et al [18] studied the influence of delta winglet vortex generators

on heat transfer enhancement on air-side of a round-tube heat exchanger with

louvered fins. The contribution of five important design parameters to the thermal

hydraulic performance of the compound heat exchanger was numerically

investigated. Knowing which parameters have the biggest influence is important for

the optimization. At high inlet velocities the performance was mainly determined by

the louvers, while at lower inlet velocities also the delta winglet geometry had a

significant contribution. To validate the simulations, an aluminum compound heat

exchanger was made and tested in a wind tunnel. The experimental result showed

that the friction factor decreased with increased Reynolds number.

Huisseune et al [19] examined numerically the effects of punching delta

winglet vortex generators on Performance enhancement of a louvered fin heat

exchanger. The delta winglet vortex generators into the louvered fin surface in the

near wake region of each tube were numerically investigated using computational

fluid dynamics (CFD). The delta winglet vortex generators cause three important

mechanisms of heat transfer enhancement.

First, due to the swirling motion of the generated vortices, hot air is removed from

the tube wake to the mainstream regions and vice versa. Second, the induced wall-

normal flow locally thins the boundary layer, which also enhances the heat transfer.

Third, the size of the wake zones is reduced because the flow separation from the

tube surface is delayed.

11

Lei et al [20] examined the effects of vortex generators on heat transfer and

pressure drop of an oval heat exchanger by using computational fluid dynamics

(CFD) method. The Reynolds numbers based on fin collar outside diameter varied

from 600 to 2600.From the results, it was found that the delta-winglet vortex

generator with an attack angle of 20° and an aspect ratio of 2 provided the best

integrated performance over the range of Reynolds number computed.

Tian et al [21] investigated numerically the effects of delta winglets on air-

side heat transfer and fluid flow characteristics of wavy fin-and-tube heat exchanger.

The three-dimensional simulations were performed with renormalization-group

(RNG) k 3 model to lay the foundation for the design of the high-performance heat

exchanger. The wavy fin-and-tube heat exchangers which had three-row round tubes

in staggered or inline arrangements were studied. The numerical results showed that

each delta winglet generates a downstream main vortex and a corner vortex. For the

in-line array, the longitudinal vortices enhance the heat transfer not only on the fin

surface in the tube wake region but also on the tube surface downstream of the delta

winglet; for the staggered array, longitudinal vortices were disrupted at the first wavy

trough downstream from the delta winglet and only develop a short distance along

the main-flow direction, and the vortices mainly enhance the heat transfer of the fin

surface in the tube wake region. The longitudinal vortices generated by delta winglet

caused considerable augmentation of heat transfer performance for wavy fin-and-

tube heat exchanger with modest pressure drop penalty.

Tiwari et al [22] studied Heat transfer enhancement in cross-flow heat

exchangers using oval tubes and multiple delta winglets. A three-dimensional study

of laminar flow and heat transfer in a channel with built-in oval tube and delta

winglets was carried out through the solution of the complete Navier–Stokes and

energy equations using a body-fitted grid and a finite-volume method. Oval tubes

were used in place of circular tubes, and delta-winglet type vortex generators in

various configurations were mounted on the fin-surface. An evaluation of the

strategy is attempted in this investigation. The investigation is carried out for

different angles of attack of the winglets to the incoming flow for the case of two

winglet pairs. The variation of axial location of the winglets is also considered for

one pair of winglets mounted in common-flow-down configuration. The results

12

indicate that vortex generators in conjunction with the oval tube showed definite

promise for the improvement of fin–tube heat exchangers.

2.3. Winglets in Refrigerator Evaporator

Sommers et al [23] studied the influence of longitudinal vortex generation

heat transfer enhancement on the fin surface of a heat exchanger. The single row of

delta-wing vortex generators was applied to a refrigerator evaporator with a fin

spacing of 8.5 mm both along the leading edge and at a location halfway along the

flow length for a total of 108 vortex generators. Heat transfer and pressure drop

performance were measured before and after to determine the effectiveness of the

vortex generator under frosting conditions. The heat transfer enhancement

monotonically increased with air velocity and results in a small pressure drop penalty

that is incommensurate with the achieved enhancement. The heat transfer coefficient

is observed to lie between 26-51 W/m2-K for the enhanced evaporator and between

16-26 W/m2-K for the baseline evaporator. Two different performance evaluation

criteria are calculated and both show that the enhanced evaporator outperforms the

baseline specimen for Reynolds numbers greater than approximately 700-750.

2.4 Winglet-Type Vortex Generators Insert

Wua and Tao [24] studied numerically the impact of delta winglet vortex

generators on the heat transfer enhancement and lower pressure loss of a novel fin-

tube surface with two rows of tubes in different diameters. Numerical simulation

results showed that the fin-tube surface with first row tube in smaller size and second

row tube in larger size can lead to an increase of heat transfer and decrease of

pressure drop in comparison with the traditional fin-tube surface with two rows of

13

tubes in the same size. It was clearly noted that the delta winglet pairs could bring

about a further heat transfer enhancement and pressure drop decrease through the

careful arrangement of the location, size and attack angle of delta winglet pairs either

in ‘‘common flow up’’ or ‘‘common flow down’’ configurations. It was also

observed that the utilization of delta winglet vortex generators led to develop more

compact, higher heat transfer efficiency, lower fan power and quieter heat exchanger

of refrigeration and air condition system.

Habchi et al. [25] carried out three dimensional numerical study of heat

transfer in vortex generator-type multifunctional heat exchangers. The first category

were rows of trapezoidal vortex generators in different arrangements; in the second

category the vortex generators were fixed at certain distance from the tube wall, and

the third category had vortex generator rows between which a row of small

protrusions were inserted on the tube wall. The study was conducted for turbulent

vortical flow with Reynolds numbers were varied between 7500 and 15,000 and

constant wall temperature Tw = 360 K. It was also noted that the counter-rotating

vortex pair start to develop from the leading edge of the vortex generator and reaches

maximum intensity near the trailing edge. The total turbulence kinetic energy reaches

its maximum near the trailing edge of the vortex generator due to the decrease in the

shear layer intensity by viscous forces. The longitudinal variation of local Nusselt

number computed on different HEV cross sections showed the advantage of using

protrusions between two successive vortex generator rows, since they increased the

local heat transfer by increasing the temperature gradients and vortices very close to

the heated wall.

Lawson and Thole [26] studied experimentally the effect of a louvered fin

heat exchanger with used practical delta winglets on heat transfer augmentation

along the tube wall. From the results, it was clearly noted that the utilization of delta

winglets with louvered fins provided augmentations in heat transfer along the tube

wall as high as 47% with a corresponding increase of 19% in pressure losses.

Comparisons of measured heat transfer coefficients with and without piercings

indicate that piercings reduce average heat transfer augmentations, but it significantly

increases with respect to no winglets present. It was also observed that the negative

effects of piercings only occurred before the turnover louver, where the winglets

were angled towards the tube wall. Piercings interfered with the winglet influence

14

causing the larger Nusselt number augmentations after the turnover louver for the

pierced louvered fin configurations. Flow passing through the piercings. Friction

factor tests indicated that piercings decreased the friction factors through the test

section relative to the solid louvered fin configuration. Flow through the piercings

caused less shear at the interface between the louver-directed flow and channel-

directed flow regimes.

Ling et al. [27] carried out numerical study on heat transfer and pressure drop

for fin-and-tube heat exchangers with rectangular winglet-type pairs (RWPs) vortex

generators under low Reynolds number flow. The rectangular winglet-type pairs

RWPs were placed with a special orientation for the purpose of enhancement of heat

transfer. The effects of attack-angle of RWPs, row-number of RWPs and placement

of RWPs on the heat transfer characteristics and flow structure were examined. It

was found that the pressure loss penalty of the fin-and-tube heat exchangers with

RWPs could be reduced by altering the placement of the same number of RWPs

from inline array to staggered array without reducing the heat transfer enhancement.

The results showed that the rectangular winglet pairs (RWPs) significantly improved

the heat transfer performance of the fin-and-tube heat exchangers with a moderate

pressure loss penalty. The staggered-RWPs array provided further reduce the

pressure loss penalty due to the asymmetric arrangement of the vortex generators. It

was expected that the staggered-RWPs array could further reduce the pressure loss

penalty with increase of the Reynolds number.

Lemouedda et al. [28] investigated numerically the effect of angle of attack

of delta-winglet vortex generators in a plate-fin-and-tube heat exchanger. The

optimal angles of attack of the delta-winglets were investigated based on the Pareto

optimal strategy. The angle of attack of a pair a delta-winglet-type VGs mounted

behind each tube was varied between β = -90° and +90°. Three circular tube rows

with inline and staggered tube arrangements were investigated for Reynolds numbers

from 200 to 1200. The results showed that increased the performance of plate-fin

and-tube heat exchangers with used the delta-winglets as heat transfer enhancement

elements. It was found that the used of winglet vortex generators led to increase the

heat transfer in regions of poor heat transfer rate such as the wake regions behind the

tubes by directing a part of the fluid of the mainstream towards these regions.

15

Jiong et al. [29] studied numerically the effect of a slit fin-and-tube heat

exchanger with longitudinal vortex generators on the heat transfer and fluid flow

characteristics. The 3-D numerical simulation was performed on laminar flow of air.

They compared the characteristics of heat transfer and fluid flow for the slit fin and-

tube heat exchanger with longitudinal vortex generators with the heat exchanger with

X-shape arrangement slit fin and heat exchanger with rectangular winglet

longitudinal vortex generators. In front of the first tube, the span-average Nusselt

number abruptly increased due to the formation of horseshoe vortices, which brings

about better mixing and enhances heat transfer in this region. For the slit fin, the

enhanced heat transfer effect of strips led the span-average Nusselt number larger

than the plain fin in this region. A little peak occurs at the downstream of the first

strip as the effect of the longitudinal vortices. The heat transfer performance and

friction factor of slit fin-and-tube heat exchanger with longitudinal vortex generator

is 15.8% and 4.2% larger than that of heat exchanger with X-Arrangement slit fin

and heat exchanger with rectangular winglet longitudinal vortex generators.

Eiamsa-ard and Promvonge [30] studied experimentally the behaviors of

turbulent tube flow through a straight tape with double-sided delta wings (T-W) on

the heat transfer rate and friction factor. They studied the effect of the double-sided

delta wings with forward/backward-wing arrangement, delta wings with alternate

axis (T-WA), three wing-width ratios and wing-pitch ratios on Nusselt number (Nu)

and friction factor f. The overall walls of the tested tube were exposed to a uniform

heat-flux at turbulent air flow over a range of 4000 < Re <20000 .Experimental

results showed that the Nusselt number increased with the reduction of pitch ratio

due to higher turbulence intensity. For the B-wing and F-wing, it was found that the

friction factor get higher values for larger wing-width ratio. It was observed that the

friction factor decreased with decreased wing-width ratios due to lower turbulence

intensity and friction near the tube wall. It was also noted that the thermal

performance factor increased as increased the wing-width ratio. As found, the heat

transfer rate and thermal performance factor of the F-wing were found to be higher

than those of the B-wing. The mean Nusselt number values of the delta wings with

alternate axis (T-WA) was higher than that of the T-W delta wings because of the

high strength of the turbulence intensity in the tube and the better mixing of fluid

16

caused by the interruption of wing elements or alternate axis planes .

Wu et al. [31] studied experimentally and numerically the effect of pair of

delta winglet longitudinal vortex generator punched directly from the plates at attack

angles of 15º, 30º, 45º and 60º on the heat transfer in rectangular channels with air as

a fluid. Experimental results showed that the average Nusselt number on the surfaces

of plate increased with increased of the attack angle of delta winglet pair compared

with that of plain plate without delta winglet pair. The average Nusselt number of the

plate with attack angle of 60º was slightly higher than that of plate with attack angle

of 45º. The average Nu for the plate with delta winglet pair at β=15º increased by 8-

11%, those at β=30º, 45º, 60º increased by 15-20%, 21-29%, 21-34%, respectively at

the experimental range. The average Nu increased with the increase of the Re in the

channel for all cases. Punched vortex generators directly from the fins of heat

exchanger are not only convenient to implement, but it is also beneficial to enhance

the integral heat transfer of the both surfaces of the practical fin due to the transverse

flow through the punched holes.

Kotcioglu et al. [32] studied the second law analysis and heat transfer in a

cross-flow heat exchanger with a new winglet-type vortex generator. The

experimental data showed an increase in the heat transfer enhancement from 15% to

30% and also an increase in the pressure-loss penalty from20% to 30%, in a

comparison with and without vortex generators, respectively. The effectiveness

values of the HX with and without the convergent divergent longitudinal vortex

generator (CDLVG) generators containing gaps for the passage of fluid into the

channel was obtained up to 68-80%, compared to a cross-flow heat exchanger HX

with empty tubes in cross-flow. It was also found that the pressure loss reduced as

the velocity increased; nevertheless the reduction is rather small. Due to the

irreversibility, the motion of mixed fluid in the channels, the number of channels and

the vortex generator surface affect the variation on the cross-flow heat exchanger HX

that increased heat transfer area on both exhaust air and supply air sides, leading to a

higher heat transfer and pressure drop and also increasing the entropy generation

number.

Hiravennavar et al. [33] studied numerically the laminar flow over a delta

17

winglet pair in a channel at various Reynolds numbers and winglet thicknesses. They

solved the three-dimensional, unsteady, incompressible Navier– Stokes equations

and the energy equation by numerical simulation. Three types of parametric were

studied the channel without winglet, with one winglet and a winglet pair, variation of

Reynolds number in case of winglet pair and variation of the thickness of winglet for

the case of winglet pair. It was observed that the Nusselt number Nu was higher for

the case with winglet pair as compared to single winglet case. At the end of the

channel the winglet pair case showed increased of 66.6% in Nusselt number as

compared to the plane channel case and 33.1% as compared to the single winglet

case. It was noted that the enhancement in heat transfer by a winglet pair was twice

that of a single winglet. It was also showed that influence of the thickness of winglet

on the Nusselt number as the thickness of the winglets increased the Nusselt number

Nu increased.

Chen and Hsu [34] investigated theoretically and experimentally the average

heat transfer coefficient and fin efficiency on the fin of annular-finned tube heat

exchangers in natural convection for various fin spicing. The radiation and

convection heat transfer coefficients were simultaneously taken into consideration in

this study. The heat transfer coefficient on this annular circular fin was assumed to be

non-uniform. Thus, the whole annular circular fin was divided into several sub-fin

regions in order to predict the average heat transfer coefficient and fin efficiency

from the knowledge of the ambient temperature, tube temperature and fin

temperature recordings at several selected measurement locations. The results

showed that the average heat transfer coefficient value increased with increasing the

fin spacing (S), and the fin efficiency decreased with increasing the fin spacing.

Chen and Hsu [35] studied theoretically and experimentally the average heat

transfer coefficient and fin efficiency on a vertical annular circular fin of finned-tube

heat exchangers for various fin spacing in forced convection. The distribution of the

heat transfer coefficient on the fin was assumed to be non-uniform, thus the whole

annular circular fin was divided into several sub-fin regions in order to predict the

average heat transfer coefficient and fin efficiency values. The results showed that

the effect of the fin spacing (S) on the average heat transfer coefficient value could

be negligible when the S value exceeded about 0.018 m. The average heat transfer

18

coefficient value increased with increasing the air speed Vair for 1 m/s ≤ Vair ≤ 5 m/s

(1550 ≤ Red ≤ 7760) and increasing the fin spacing for 0.005 m ≤ S ≤ 0.018 m. The

fin efficiency value decreased with increasing Vair for 1 m/s ≤ Vair ≤ 5 m/s and

seemed to be not very sensitive to the fin spacing.

Choi et al. [36] investigated experimentally the heat transfer characteristics of

discrete plate finned-tube heat exchangers with large fin pitches. Thirty-four heat

exchangers were tested with variations of fin pitches, the number of tube rows, fin

alignment, and vertical fin space. The Colburn j-factor of the discrete plate finned-

tube exchanger was analyzed as a function of coil geometry and then compared with

that of the continuous plate finned-tube heat exchanger. For fin pitches of 7.5–15

mm, the Colburn j-factors of the discrete plate finned-tube heat exchangers were 6.0–

11.6% higher than those of the continuous plate finned-tube heat exchangers. The j-

factor of the discrete plate finned-tube heat exchangers decreased with the rise of the

number of tube rows. As the number of tube rows increased, the j-factor decreased

more severely in the continuous plate finned-tube heat exchangers than in the

discrete plate finned-tube heat exchangers. The j-factor for the staggered fin

alignment was at least 5.5% higher than that of the inline fin alignment in the

discrete plate finned tube heat exchangers. For the inline fin alignment, the j-factor

increased with the increase of vertical fin space, but the increase of the j-factor

became negligible for vertical fin space greater than 4 mm. For the staggered fin

alignment, the effects of vertical fin space on the j-factor were almost negligible at

all Reynolds numbers.

Nagarani and Mayilsamy [37] investigated and analyzed experimentally the

heat transfer rate and efficiency for circular and elliptical annular fins for different

environmental conditions. It was found that the elliptical fin efficiency was more

than that of the circular fin. The heat transfer coefficient depends upon the space,

time, flow conditions and fluid properties. If there are changes in the environmental

conditions, there are changes in the heat transfer coefficient and efficiency also.

Alam et al. [38] reviewed the use of tabulators in different forms of ribs,

baffles, delta winglets, obstacles, vortex generator, rings and perforated

blocks/baffles is an effective way to improve the performance of heat exchangers and

19

solar air heaters. Investigators studied the effect of these tabulators for heat transfer

and friction characteristics in air ducts. Based on their view it was found that

perforation in ribs/baffles/blocks and combination of combined rib and delta winglet

leads to the better thermo-hydraulic performance. The correlation presented by

various investigators, in terms of non-dimension al parameters for heat transfer and

friction factor in solar air heater sand heat exchanger shave also been presented.

Du et al [39] studied experimentally the heat transfer and resistance

characteristics of two finned oval-tube heat exchangers (HE1: Double rows of tubes,

HE2: Three rows of tubes.) inclined towards the air incoming flow direction. Four air

inlet angles (90, 60, 45 and 30) were investigated separately to acquire the heat

transfer and pressure drop performances for Reynolds number ranging from 1300 to

13,000. The experimental correlations of Nusselt number and resistance coefficient

of the air side were obtained, and the comprehensive comparisons of heat transfer

performance were carried out. The results showed that whether the heat transfer

performance for heat exchangers positioned obliquely was improved was depended

on their inclined angles.

Du et al. [40] studied numerical Punched longitudinal vortex generators

(LVGs) were employed to enhance air-side heat transfer on the wavy fin surface of

flat tube used in direct air-cooled condenser. The heat transfer enhancement of four

types of the longitudinal vortex generators with different attack angles were

compared by numerical simulations. It was found that the delta winglet pair with

attack angle 25 could reach the greatest performance evaluation criteria (PEC) under

the conditions of the inlet air flow velocity varied from 1 m/s to 5 m/s. The

influences of locations on the wavy fin surface and the row number of the

longitudinal vortex generators were also discussed. One delta winglet pairs at the

middle of the wavy fin surface and the minimum row number, n = 1, with the

average PEC is 1.23, had the best heat transfer performance of all conditions. as well

as the numerical simulations verified that the delta winglet pairs can generate

obvious longitudinal vortex pairs at the down-sweep zone, which can enhance the

heat transfer between the cooling air flow and heated wall surface with acceptable

pressure loss.

Gawande et al.[41] studied numerically the performance characteristics of a

20

solar heater and heat exchangers by using artificial roughness in different forms

shapes and sizes. Artificial roughness was provided in the form of different

geometries such as ribs, dimple shape roughness, wire mesh, baffles, delta winglets

etc. To determine the effect of these geometries on thermal performance of solar

heaters and heat exchangers, several experimental and numerical studies had been

carried out by various researchers. An attempt had been made to review various

roughness element geometries employed in solar air heaters and heat exchangers in

terms of heat transfer, friction factor and flow simulation techniques. Correlations

developed for heat transfer and friction factor for different roughness geometries by

various investigators in solar air heaters were present.

Gholami et al [42] investigated numerically the heat transfer enhancement

and pressure loss penalty for fin-and-tube compact heat exchangers with the wavy-up

and wavy-down rectangular winglets in low Reynolds number flow. The rectangular

winglets were used with a particular wavy form for the purpose of enhancement of

air side heat transfer performance of fin-and-tube compact heat exchangers. The

effect of Reynolds numbers from 400 to 800 and angle of attack of 30° of wavy

rectangular winglets were also examined. The effects of using the wavy rectangular

winglet, conventional rectangular winglet configuration and without winglet as

baseline configuration, on the heat transfer characteristics and flow structure were

studied and analysed in detail for the inline tube arrangements. The results showed

that the wavy rectangular winglet can significantly improve the heat transfer

performance of the fin-and-tube compact heat exchangers with a moderate pressure

loss penalty. It was also clearly found that the numerical results of the wavy winglet

had significant effect on the heat transfer performance and also, this augmentation

was more important for the case of the wavy-up rectangular winglet configuration.

Hasanpour et al [43] reviewed that the enhancing of heat transfer mechanisms

were used in many industrial applications like heat exchanger, air conditioning,

chemical reactors and refrigeration systems. Therefore several techniques had been

promoted to enhance heat transfer rate and to decrease the size and cost of equipment

especially the heat exchangers. One of the leading tools used in passive heat transfer

methods mainly at turbulence flow was twisted tape inserts.

21

Khoshvaght et al [44] studied experimentally a comparative evaluation of

seven common configurations of channels used in plate-fin heat exchangers. All the

channels, including plain, perforated, offset strip, louvered, wavy, vortex-generator,

and pin, were fabricated and tested experimentally. The working fluid was water, and

Reynolds number range was varied from 480 to 3770. To evaluate the performance

of these channels and also selected an optimum plate-fin channel, three mostly used

energy-based performance evaluation criteria were employed. The results were

presented as plots of dimensional and non-dimensional parameters. The results of the

studied channels, the vortex-generator channel showed a significant enhancement in

the heat transfer coefficient and a proper reduction in the heat exchanger surface

area. Therefore, it could be applied as a high quality interrupted surface in the plate-

fin heat exchangers. Moreover, the wavy channel displayed an optimal performance

at low Reynolds numbers.

Li et al [45] investigated numerically a radiantly arranged LVGs enhanced

fin-and-tube structure to enhance air side heat transfer. The arrangement of LVGs is

totally different from existing publications. In the proposed structure there exist 12

winglets around each tube. The attack angles are 50, 50, 50, 50, 70 and 110,

respectively. The height ends of the winglets were further away from the tube, while

the closed point ends of winglets were close to the tube wall. Heat transfer and

pressure drop performance was compared with other three structures: an arc-shaped

wavy fin-and-tube surface, a common-flow down LVGs enhanced fin-and-tube

surface and a plain plate fin-and-tube surface. The results showed that the 12

winglets form five local passages which could guide the moving fluid from the main

flow to the tube wall, leading to some impinging effect or reducing the wake region

behind the tube. The performance evaluation of the four structures was conducted by

using the newly proposed ln (Nue/Nuo) vs. ln (fe/fo) plot based on energy saving. It

was found that the proposed radiantly arranged LVGs enhanced fin and- tube surface

was the best. The field synergy principle is adopted to analyse the four structures and

it was found that the domain averaged synergy angle of the proposed radiantly

arranged LVGs enhanced structure was significantly less than that of other three

cases. Finally characteristics of the proposed fin and- tube surface with five tubes

were investigated at five fin pitches and compared with the wavy structure of six

tubes at the same other conditions.

22

Ranut et al [46] studied numerically the Optimization of heat exchangers, as a

consequence of their vital role in several industries and applications, has attracted a

lot of interest in the last years, and in particular the necessity of improving their

performances is well recognized. The coupling of optimization techniques with

Computational Fluid Dynamics (CFD) had demonstrated to be a valid methodology

for easily explore this work; a CFD-based shape optimization of a tube bundle in

cross flow was presented, as a natural extension of the work.

Saha et al [47] investigated numerically the performance of a plate-fin heat

exchanger with an emphasis on acquiring fundamental understanding of the relation

between local flow behaviour and heat transfer augmentation mechanism. Numerical

simulations were performed in a rectangular channel containing built-in longitudinal

vortex generators on the bottom wall arranged periodically both in the stream wise

and span wise directions. Two types of vortex generators, namely, rectangular

winglet pair (RWP) and delta-winglet pair (DWP) with two different flow

arrangements, common-flow-up (CFU) and common-flow-down (CFD) had been

explored to assess the influence of shape and flow arrangements on heat transfer

enhancement. The basic mechanisms of flow structure and heat transfer

characteristics had been examined with the help of secondary velocity vectors,

streamlines, and temperature contours. Additionally, the mechanism of the local heat

transfer augmentation had been explained using a novel concept called the field

synergy principle. The performance of the vortex generators had been compared

based on integral quantities such as Nusselt number, pressure loss, performance

evaluation factor and domain averaged synergy angle. The computations reveal

enhanced mixing of fluid between the wall layer and the core due to strong

secondary flows produced by vortex generators. The performance analysis indicates

that the RWP was more effective in terms of heat transfer enhancement as compared

to DWP. The field synergy analysis showed that the sites with higher Nusselt number

are associated with smaller synergy angle or better coordination between the velocity

vector and the temperature gradient.

Sun et al [48] tested experimentally and numerically the effect of elliptical

tube in finned-tube heat exchangers (FTHE). Most of the previous studies evaluated

the two tube shapes only based on the air-side performance of FTHE, and did not

consider any interaction effect of the axis ratio with other parameters. among seven

23

design factors including number of rows, axis ratio, transversal tube pitch,

longitudinal tube pitch, fin pitch, air velocity, water volumetric flow rate. Response

surface analysis was used to evaluate the axis ratio effect on the overall thermal

hydraulic performance which was quantified by the heat transfer rate per unit power

consumption. The results showed that the axis ratio strongly interacted with air

velocity and water volumetric flow rate. The increase of axis ratio improved the

overall thermal hydraulic performance at higher air velocity or lower water

volumetric flow rate, but the opposite effect was observed at lower air velocity or

higher water volumetric flow rate.

Tarlet et al [49] studied experimentally a first and second law analysis of the

heat transfer characteristics of a mini shell-and tube heat exchanger equipped with

multi-scale distributor/collector. Experiments of heat exchanger with or without

transverse baffles installed in the shell side were conducted, under laminar flow

conditions (average channel Re number between 8 and 100). The temperature field at

the shell side was obtained by using infrared thermograph. The effects of transverse

baffles on the thermal performance and entropy generation of the heat exchanger

system were quantified and discussed. Experimental results show that the integration

of multi-scale branched distributor and collector guarantees uniform flow distribution

among parallel tubes. The installation of baffles provides a locally cross flow and

globally counter current flow arrangement so that the recirculation, passive zones can

be largely eliminated. Enhancement of heat transfer has been verified by first law

(global heat transfer coefficient) and second law (entropy generation) analyses.

Xuehong et al [50] investigated by numerical study the effect of composite

fin and advantage of longitudinal vortex generator and slit fin, respectively. The

performance of air-side heat transfer and fluid flow was investigated by numerical

simulation for Reynolds number ranging from Re = 304 to 2130. Stepwise

approximation method is applied on the mesh generation for the irregular domains of

delta winglets and slit fins. The mechanism for augmenting heat transfer was also

analysed based on the local fluid field, field synergy principle and entrance

dissipation principle. The numerical results showed that some eddies are developed

behind the X-shaped slit and delta winglet, which produce some disruptions to fluid

flow and enhance heat transfer; compared with plain fin and slit fin, it was found that

24

the composite fin had better heat transfer performance. By applying on the field

synergy principle and entrance dissipation principle to analyse the composite fin, the

results showed that composite fin was improved the synergy of temperature gradient

and velocity fields, and its equivalent thermal resistance was smaller and its

irreversibility of heat transfer is lower.

Yamashita et al [51] studied experimentally a new type of heat exchanger, in

which flue gas flows inside thin tubes and cool water is on the shell side, was

proposed to develop the performance and compactness of shell and tube type heat

exchangers for latent heat recovery from flue gas. The experimental heat transfer

characteristics of single tubes were systematically investigated to determine the

effects of tube diameter (1.0-5.0 mm) and length (7 -100 mm). Furthermore, a

correlation between the non-dimensional bulk mean temperature and the ratio of

effective tube length to the thermal entrance region was proposed and was correlated

well with the measurement data. Prediction of the heat exchanger performance using

this correlation was possible. As a result, it was found that the using mini-tubes is

remarkably effective to reduce the size of heat exchanger due to enhancement of heat

transfer coefficient and enlargement of heat transfer surface. The volume with 1 mm

inner diameter of tubes was approximately 5 present of that with 5 mm in diameter.

Zhao et al [52] numerically studied the heat transfer performance and reduce

erosion of economizers in coal-fired power plants. The heat transfer and erosion

characteristics was tested for the single H-type finned oval tube with enhanced heat

transfer structures including bleeding dimples, longitudinal vortex generators

(LVGs), and compound dimple-LVG. The results showed that the oval tube with

compound LVG-dimple achieved the highest overall heat transfer performance while

the oval tube with LVG works most efficiently in the anti-wear performance. Then

based on the H-type finned oval tube, the LVG structure on the first row of tubes

together with hemisphere protrusions design, while the compound LVG-dimple on

the rest tubes were also simulated. The optimized H-type finned oval tube bank heat

exchanger was demonstrated of high performance on both heat transfer and anti-

wear.

Zhou et al [53] studied experimentally the performance of plane and curved

winglet (rectangular, trapezoidal and delta) vortex generators (VGs) with and without

110

REFERENCES

[1] R. M. Manglik ''Heat transfer Enhancement '', Handbook of Heat Transfer,

(Adrian Bejan and Allan D. Kraus), Wiley & Sons, Inc, Hoboken, New

Jersy, Chapter 14, pp. 1029-1130, 2003.

[2] W.J.Marner and A.E. Bergles. “Augmentation of highly viscous laminar

heat transfer inside tubes with constant wall temperature Experimental

Thermal and Fluid Science, Volume 2, Issue 3, July 1989 pp 252–267

[3] M. Sohal, J. O’Brien, Improving air-cooled condenser performance

usingwinglets and oval tubes in a geothermal power plant, Geotherm. Res.

Counc. Trans. 25 (2001) 1–7.

[4] S. Tiwari, D. Maurya, G. Biswas, V. Eswaran, Heat transfer enhancement in

cross-flow heat exchangers using oval tubes and multiple delta winglets, Int.

J Heat Mass Transfer 46 (15) (2003) 2841–2856.

[5] P. Chu, Y. He, Y. Lei, L. Tian, R. Li, Three-dimensional numerical study on

finand–oval-tube heat exchanger with longitudinal vortex generators, Appl.

Therm. Eng. 29 (2009) 859–876.

[6] A. Joardar, A. Jacobi, A numerical study of flow and heat transfer

enhancement using an array of delta-winglet vortex generators

in a fin-and-tube heat exchanger, J. Heat Transfer 139 (2007)

1156–1168.

[7] C. Lin, Y. Liu, J. Leu, Heat transfer and fluid flow analysis for plate-fin and

oval tube heat exchangers with vortex generators, Heat Transfer Eng. 29

(2008) 588–596.

[8] Y. Chen, M. Fiebig, N. Mitra, Conjugate heat transfer of a finned oval

tubewith a punched longitudinal vortex generator in form of delta winglet–

parametric investigations of the winglet, Int. J. Heat Transfer 41 (1998)

3961–3978.

[9] Y. Chen, M. Fiebig, N. Mitra, Heat transfer enhancement of a finned oval

tube with punched longitudinal vortex generators in line, Int. J. Heat

Transfer 41(1998) 4151–4166.

[10] Y. Chen, M. Fiebig, N. Mitra, Heat transfer enhancement of finned oval

111

tubes with staggered punched longitudinal vortex generators, Int. J. Heat

Transfer 43 (2000) 417–435.

[11] G. Biswas, K. Torii, D. Fujii, K. Nishino, Numerical and experimental

determination of flow structure and heat transfer effects of longitudinal

vortices in a channel flow, Int. J. Heat Mass Transfer 39 (1996) 3441–

3451.

[12] F. Dupont, C. Gabillet, P. Bot, Experimental study of the flow in a compact

heat exchanger channel with embossed-type vortex generators, J. Fluids

Eng. 125(2003) 701–709.

[13] A. Jain, G. Biswas, D. Maurya, Winglet-type vortex generators with

commonflow-up configuration for fin-tube heat exchangers, Numer. Heat

Transfer, Part A 43 (2003) 201–219.

[14] M. Gupta, K. Kasana, R. Vasudevan, A numerical study of the effect on

flow structure and heat transfer of a rectangular winglet pair in a plate fin

heat exchanger, Proc. Inst. Mech. Eng. Part C-J. Mech. Eng. Sci. 223

(2009) 2109–2115.

[15] X.B. Zhao, G.H. Tang, X.W. Ma, Y. Jin, W.Q. Tao “Numerical investigation

of heat transfer and erosion characteristics for H-type finned oval tube

with longitudinal vortex generators and dimples” Applied Energy 127

(2014) 93–104.

[16] Zhang YH, Wu X, Wang LB, Song KW, Dong YX, Liu S. “Comparison of

heat transfer performance of tube bank fin with mounted vortex generators

to tube bank fin with punched vortex generators”.Exp Thermo Fluid .

33(2008)58–66.

[17] Ya-Ling He , Pan Chu, Wen-Quan Tao, Yu-Wen Zhang, Tao Xie “Analysis

of heat transfer and pressure drop for fin-and-tube heat exchangers with

rectangular winglet-type vortex generators”. Applied Thermal Engineering

(2012)p 1-14.

[18] Henk Huisseune , Christophe T’Joen, Peter De Jaeger, Bernd Ameel, Sven

De Schampheleire, Michel De Paepe “Performance analysis of a

compound heat exchanger by screening its design parameters” Applied

Thermal Engineering (2013).p 490-501.

[19] Henk Huisseune , Christophe T’Joen, Peter De Jaeger, Bernd Ameel,Sven

De Schampheleire a, Michel De Paepe “Performance enhancement of a

112

louvered fin heat exchanger by using delta winglet vortex generators”

International Journal of Heat and Mass Transfer (2013) 475–487.

[20] Yong-Gang Lei, Ya-Ling He, Li-Ting Tian, Pan Chu, Wen-Quan Tao

“Hydrodynamics and heat transfer characteristics of a novel heat

exchanger with delta-winglet vortex generators” Chemical Engineering

Science (2010)p1551–1562.

[21] Liting Tian, Yaling He, Yubing Tao, Wenquan Tao “A comparative study

on the air-side performance of wavy fin-and-tube heat exchanger with

punched delta winglets in staggered and in-line arrangements”

International Journal of Thermal Sciences .(2009) p1765–1776.

[22] S. Tiwari, D. Maurya, G. Biswas, V. Eswaran. “Heat transfer enhancement

in cross-flow heat exchangers using oval tubes and multiple delta

winglets”. International Journal of Heat and Mass Transfer. (2003) p

2841–2856.

[23] J.M. Wua, W.Q. Tao,“Impact of delta winglet vortex generators on

the perfor-mance of a novel fin-tube surfaces with two rows of

tubes in different diameters”, Energy Conversion and Management

vol.52 ,2011,pp 2895–2901.

[24] C.Habchi, S.Russeil, D. Bougeard, J. Harion, T. Lemenand, D.Valle, H.

Peerhossaini,“Enhancing heat transfer in vortex generator-type

multifunctional heat exchangers”, Applied Thermal Engineering ,vol.38

,2012,pp 14–25.

[25] M. J.Lawson, K.A. Thole, “Heat transfer augmentation along the tube

wall of a louvered fin heat exchanger using practical delta winglets”, Int.

Comm. in Heat and Mass Transfer ,vol.51 , 2008, pp 2346–2360.

[26] Ya-Ling He, Pan Chu, Wen-Quan Tao, Yu-Wen Zhang, T. Xie,“Analysis

of heat transfer and pressure drop for fin-and-tube heat exchangers with

rectangular winglet-type vortex generators”, Applied Thermal

Engineering, 2012. In Press.

[27] A. Lemouedda ,M. Breuer , E. Franz , T. Botsch , A.Delgado, “Optimization

of the angle of attack of delta-winglet vortex generators in a plate-fin-and-

tube heat exchanger”, Int. J. of Heat and Mass Transfer ,vol.53 ,2010,pp

5386–5399.

[28] J. Li, S. Wanga, J. Chen , Y. Lei, “Numerical study on a slit fin-and-tube

heat exchanger with longitudinal vortex generators”, Int. J. of Heat and

Mass Transfer, vol.54,2011, pp 1743–1751.

[29] S. Eiamsa-ard, P. Promvonge, “Influence of Double-sided Delta-wing

Tape Insert with Alternate-axes on Flow and Heat Transfer Characteristics

113

in a Heat Exchanger Tube”, Fluid Flow and Transport Phenomena, vol.19 ,

2011,pp 410–423.

[30] J. M. Wu, W.Q. Tao, “Effect of longitudinal vortex generator on heat

transfer in rectangular channels”, Applied Thermal Engineering, vol.37,

2012, pp 67–72.

[31] I. Kotcioglu, S. Caliskan, A. Cansiz, S. Baskaya, “Second law analysis and

heat transfer in a cross-flow heat exchanger with a new winglet-type

vortex generator”, Energy ,vol.35 ,2010,pp 3686–3695.

[32] P. Promvonge, S. Eiamsa-ard, “Heat transfer and turbulent flow friction in

a circular tube fitted with conical-nozzle turbulators”, Int. J. of Heat and

Mass Transfer, vol. 34, 2007, pp 72–82.

[33] J.M. Wua, W.Q. Tao,“Impact of delta winglet vortex generators on the

perfor-mance of a novel fin-tube surfaces with two rows of tubes in

different diameters”, Energy Conversion and Management ,vol.52

,2011,pp 2895–2901.

[34] C.Habchi, S.Russeil, D. Bougeard, J. Harion, T. Lemenand, D.Valle, H.

Peerhossaini,“Enhancing heat transfer in vortex generator-type

multifunctional heat exchangers”, Applied Thermal Engineering ,vol.38

,2012,pp 14–25.

[35] Han Taw Chen, Wei-Lun Hsu, Estimation of heat transfer co efficient on

the fin of annular finned tube heat exchangers in natural convection for

various fin spacings, International Journal of Heat & Mass Transfer (2007)

1750-1761.

[36] Choi, Jong Min; Kim, Yonghan; Lee, Mooyeon; Kim, Yongchan. Applied

Thermal Engineering vol. 30 issue 2-3 February, 2010. p. 174-180

[37] N.Nagarani and K. Mayilsamy (2010). "EXPERIMENTAL HEAT

TRANSFER ANALYSIS ON ANNULAR CIRCULAR AND

ELLIPTICAL FINS." International Journal of Engineering Science and

Technology 2(7): 2839-2845

[38] Alam, Tabish; Saini, R.P.; Saini, J.S. Renewable and Sustainable Energy

Reviews vol. 31 March, 2014. p. 289-304

[39] X.P. Du, M. Zeng, Z.Y. Dong “Wang Experimental study of the effect of

air inlet angle on the air-side performance for cross-flow finned oval-tube

heat exchangers” Experimental Thermal and Fluid Science. 52 (2014) pp

146–155

114

[40] Xiaoze Du, Lili Feng, Li Li, Lijun Yang, Yong ping Yang “Heat transfer

enhancement of wavy finned flat tube by punched longitudinal vortex

generators.

[41] Vipin B.Gawande, A.S.Dhoble,D.B.Zodpe “Effect of roughness geometries

on heat transfer enhancement in solar thermal systems – A review”

32(2014) pp347–378

[42] A.A. Gholami, Mazlan A.Wahid, H.A. Mohammed “Heat transfer

enhancement and pressure drop for fin-and-tube compact heat

exchangers with wavy rectangular winglet-type vortex generators”

International Communications in Heat and Mass Transfer 54

(2014) 132–140

[43] A. Hasanpour, M. Farhadi , K. Sedighi “A review study on twisted tape

inserts on turbulent flow heat exchangers: The overall enhancement ratio

criteria” International Communications in Heat and Mass Transfer 55

(2014) 53–62

[44] M. Khoshvaght-Aliabadi, F. Hormozi, A. Zamzamian “Role of channel

shape on performance of plate-fin heat exchangers: Experimental

assessment” International Journal of Thermal Sciences 79 (2014) 183e193,

[45] M.J. Li, W.J. Zhou, J.F. Zhang, J.F. Fan, Y.L. He, W.Q. Tao “Heat transfer

and pressure performance of a plain fin with radiantly arranged winglets

around each tube in fin-and-tube heat transfer surface” International

Journal of Heat and Mass Transfer 70 (2014) 734–744

[46] Paola Ranut , Gábor Janiga , Enrico Nobile , Dominique Thévenin “Multi-

objective shape optimization of a tube bundle in cross-flow” International

Journal of Heat and Mass Transfer 68 (2014) 585–598

[47] Pankaj Saha, Gautam Biswas , Subrata Sarkar “Comparison of winglet-

type vortex generators periodically deployed in a plate-fin heat exchanger

– A synergy based analysis” International Journal of Heat and Mass

Transfer 74 (2014) 292–305

115

[48] Lei Sun, Chun-Lu Zhang “Evaluation of elliptical finned-tube heat

exchangerperformance using CFD and response surface methodology”

International Journal of Thermal Sciences 75 (2014) 45e53

[49] Dominique Tarlet, Yilin Fan, Stéphane Roux, Lingai Luo “Entropy

generation analysis of a mini heat exchanger for heat transfer

intensification” Experimental Thermal and Fluid Science 53 (2014) 119–

126

[50] Wu Xuehong, Zhang Wenhui, Gou Qiuping, Luo Zhiming, Lu Yanli

“Numerical simulation of heat transfer and fluid flow characteristics of

composite fin” International Journal of Heat and Mass Transfer 75 (2014)

414–424

[51] Junpei Yamashita, Yoshio Utaka “Improved performance of secondary heat

exchanger for latent heat recovery from flue gas using mini-tubes” Applied

Thermal Engineering 67 (2014) 230e239

[52] X.B. Zhao, G.H. Tang X.W. Ma, Y. Jin, W.Q. Tao “Numerical

investigation of heat transfer and erosion characteristics for H-type

finned oval tube with longitudinal vortex generators and dimples”

Applied Energy 127 (2014) 93–104

[53] Guobing Zhou*, Zhizheng Feng “Experimental investigations of heat

transfer enhancement by plane and curved winglet type vortex

generators with punched holes” International Journal of Thermal

Sciences 78 (2014) 2635

[54] Anderson, John D. (1995). Computational Fluid Dynamics: The Basics

with Applications. Science/ Engineering/Math. McGraw-Hill Science.

[55] Patankar, Suhas (1980). Numerical Heat Transfer and Fluid

Flow.Hemisphere Series on Computational Methods in Mechanics and

Thermal Science.Taylor & Francis.

[56] M.Venturino, P.Rubini, “Coupled fluid flow and heat transfer analysis of

steel reheat furnaces”, School of Mechanical Cranfield University, 1995.

[57] R. S. Vajjha, D. K. Das, D. P. Kulkarni, “Development of new correlations

for convective heat transfer and friction factor in turbulent regime for

116

nanofluids”,Int. J.of Heat and Mass Transfer,vol. 53, 2010,pp4607.4618

[58] M. Corcione, “Empirical correlating equations for predicting the effective

thermal conductivity and dynamic viscosity of nanofluids”, Energy

Conversion and Management, vol.52, 2011, pp 789–793.

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