ÇUKUROVA UNIVERSITY INSTITUTE OF NATURAL AND APPLIED SCIENCES

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Serkan MEZARCIÖZ PhD THESIS

IMPROVMENT OF A COACH AIR-CONDITIONING SYSTEM BY APPLYING A DISTINCT AIR CHANNEL

DEPARTMENT OF MECHANICAL ENGINEERING

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APPLYING A DISTINCT AIR CHANNEL Serkan MEZARCIÖZ

PhD THESIS

DEPARTMENT OF MECHANICAL ENGINEERING

We certify that the thesis titled above was reviewed and approved for the award of degree of the Doctor of Philosophy by the board of jury on 05/01/2015.

………... ……….. ……...

Prof. Dr. Hüseyin AKILLI Prof. Dr. Beşir ŞAHİN Assoc. Prof. Dr. Sami AKÖZ

SUPERVISOR MEMBER MEMBER

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Prof. Dr. Ahmet PINARBAŞI Assist. Prof. Dr. Arif ÖZBEK

MEMBER MEMBER

This PhD Thesis was prepared in Mechanical Engineering Department of the Institute of Natural And Applied Sciences of Çukurova University.

Registration Number:

Prof. Dr. Mustafa GÖK Director

Institute of Natural and Applied Sciences Not: The usage of the presented specific declerations, tables, figures, and photographs either in this

thesis or in any other reference without citiation is subject to "The law of Arts and Intellectual Products" number of 5846 of Turkish Republic.

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IMPROVMENT OF A COACH AIR-CONDITIONING SYSTEM BY APPLYING A DISTINCT AIR CHANNEL

Serkan MEZARCIÖZ ÇUKUROVA UNIVERSITY

INSTITUTE OF NATURAL AND APPLIED SCIENCES DEPARTMENT OF MECHANICAL ENGINEERING

Supervisor : Prof. Dr. Hüseyin AKILLI Year: 2015, Pages: 215 Jury : Prof. Dr. Beşir ŞAHİN

: Assoc. Prof. Dr. Sami AKÖZ : Prof. Dr. Ahmet PINARBAŞI : Assist. Prof. Dr. Arif ÖZBEK

The aim of this study is to design an A/C air channel providing uniform air distribution along the coach, which has 12-m length and a capacity of 50 passengers.

The air channel was designed by means of Computational Fluid Dynamics (CFD). In order to determine the most suitable turbulence model for the air channel properties, a turbulence model determination study was conducted with 5 different turbulence models, which are k-ε Standard, k-ε Realizable, k-ε RNG, k-Ω Standard, Spallart- Almaras, and the numerical results were compared with the experimental results measured from the experimental setup designed and produced. A grid independence study was also conducted to be sure that the solution is independent of grid. At the beginning of the study, cooling load of the coach calculated with a detailed study.

After the cooling load calculations, a proper A/C unit selected, fitted to the vehicle and detail road tests, which includes measurements of underbody temperature distribution, A/C nozzle exit velocities, temperature distribution inside the passenger cabin and internal noise level, was performed to compare the current situation and final application. As a result of this study, passenger comfort was improved with the application of distinct air channel from uniform air and temperature distribution along the vehicle and internal noise level points of views.

Key Words: CFD, Turbulence models, Thermal Comfort, Cooling Load, Coach.

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AYRI BİR HAVA KANALI UYGULAMASI İLE BİR YOLCU OTOBÜSÜNÜN KLİMA SİSTEMİNİN İYİLEŞTİRİLMESİ

Serkan MEZARCIÖZ ÇUKUROVA ÜNİVERSİTESİ FEN BİLİMLERİ ENSTİTÜSÜ

MAKİNA MÜHENDİSLİĞİ ANABİLİM DALI Danışman :Prof. Dr. Hüseyin AKILLI

Yıl: 2015, Sayfa: 215 Jüri :Prof. Dr. Beşir ŞAHİN

:Doç. Dr. Sami AKÖZ

:Prof. Dr. Ahmet PINARBAŞI :Yrd. Doç. Dr. Arif ÖZBEK

Bu çalışmanın amacı, 12 metre uzunluğunda ve 50 yolcu kapasitesine sahip bir yolcu otobüsü boyunca düzgün hava dağılımı sağlayacak olan bir klima hava kanalı tasarlanmasıdır. Kanal, Hesaplamalı Akışkanlar Dinamiği (HAD) aracılığı ile tasarlanmıştır. Hava kanlı özelliklerine en uygun türbülans modelinin belirlenmesi için, k-ε Standard, k-ε Realizable, k-ε RNG, k-Ω Standard ve Spallart-Almaras gibi 5 değişik türbülans modeli ile bir türbülans modeli belirleme çalışması yürütülmüş ve nümerik sonuçlar, tasarlanıp imal edilen deney düzeneğinden ölçülen sonuçlar ile karşılaştırılmıştır. Sonuçların, çözüm ağından bağımsız olduğundan emin olunması için, ayrıca bir çözüm ağı bağımsızlığı çalışması da yürütülmüştür. Çalışmanın başında, otobüsün soğutma yükü detaylı bir çalışma ile hesaplanmıştır. Soğutma yükü hesaplamalırından sonra, uygun bir klima ünitesi seçilmiş, araca takılmış ve araç altı sıcaklık dağılımı, klima üfleme menfezi çıkış hızları, yolcu kabini içi sıcaklık dağılımı ve iç gürültü seviyelerini içeren detaylı yol testleri mevcut durum ve yeni uygulamayı karşılaştırmak için gerçekleştirilmiştir. Bu çalışma sonucunda, ayrı bir klima hava kanalı uygulaması ile yolcu konforu, araç boyunca düzgün hava ve sıcaklık dağılımı, ve iç gürültü seviyesi açısından iyileştirilmiştir.

Anahtar Kelimeler: HAD, Türbülans modelleri, Termal konfor, Soğutma yükü, Yolcu otobüsü,

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Hüseyin AKILLI, who gave me the opportunity, support, and freedom that I needed to conduct this research. He has been a great source of motivation, both personal and professional.

I would also like to thank Prof. Dr. Beşir ŞAHİN, for their very useful comments and interest in my research.

I would like to thank Assoc. Prof. Dr. Sami AKÖZ who shares his experiences and comments on this subject with me.

I would like to thank to Prof. Dr. Ahmet PINARBAŞI and Assist. Prof. Dr.

Arif ÖZBEK for their valuable comments as jury members.

I would like to thank the referee of TEMSA GLOBAL and TUBİTAK supported TEYDEB project conducted in the scope of this thesis, Assoc. Prof. Dr.

Arif Emre ÖZGÜR who share his experiences and valuable comments on this project with me.

I would like to thank to all the staff of R&D Center and Test Center of TEMSA GLOBAL because of their help to perform, manufacturing, applications and measurements. Especially thanks to Plant Director of TEMSA GLOBAL, Mr.

İbrahim ESERCE, providing me buses to perform tests and applications and opportunity to use the resources of TEMSA GLOBAL.

I would like to thank to R&D Engineers of TEMSA GLOBAL Mr. Hakan AKGÜN and Bedri ŞAŞMAZ for helping in my study and their friendship, moral support.

Last but not least, I am greatly indebted to my father Yusuf MEZARCIÖZ, my mother Emel MEZARCIÖZ, and my brother Özkan MEZARIÖZ for their encouragement and support.

Lastly, I would like to specially thank to my wife Assist. Prof. Dr. Serin MEZARCIÖZ and my dear son Demir MEZARCIÖZ for their patience, trust and support.

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ÖZ ... II ACKNOWLEDGEMENT ... III CONTENTS ... IV LIST OF TABLES ... VIII LIST OF FIGURES ... XII NOMENCLATURE ... XX

1. INTRODUCTION ... 1

1.1. Significance Of Thermal Comfort ... 1

1.2. Air Conditioning (A/C) Systems For Buses ... 4

1.3. Bus Body Structure ... 9

1.4. Current Air Channel ... 12

1.5. Computational Fluid Dynamics (CFD) ... 13

1.6. Cooling Load Calculation ... 15

2. LITERATURE SURVEY ... 21

2.1. Studies On Thermal Comfort, Cooling Load Calculation and Underhood Thermal Management ... 21

2.2. Studies on Automotive HVAC Systems ... 25

2.3. CFD Studies ... 26

3. MATERIAL AND METHODS ... 33

3.1. Cooling Load Calculation ... 33

3.1.1. Regions Of The Passenger Compartment And Their Properties ... 35

3.1.2. Determination Of Outside Temperature ... 40

3.1.3. Determination Of Total Heat Transfer Coefficients... 45

3.1.4. Heat Transfer by Radiation ... 47

3.1.4.1. Radiation Time Series Method (RTS) ... 48

3.1.5. Cooling Load by Internal Sources ... 53

3.1.6. Cooling Load by Ventilation ... 53

3.1.7. Cooling Load by Infiltration ... 53

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3.2.3. Test Results ... 64

3.3. Design Of Air Channel By Means Of CFD ... 71

3.3.1. Determination Of Channel Geometry ... 71

3.3.2. CFD Analysis Studies ... 74

3.3.2.1. Production Of Turbulence Model Determination Experimental Setup ... 75

3.3.2.2. 3D Flow Domain Design Of Experimental Setup ... 84

3.3.2.3. Grid Generation and Independence Study ... 85

3.3.2.4. Analyses For Determining Suitable Turbulence Model ... 95

3.3.2.4.a. Governing Equations of Fluid Flow and Energy ... 95

3.3.2.4.b. Overview Of Turbulence Models ... 96

3.3.3. Channel Design Studies ... 102

4. RESULTS AND DISCUSSION ... 107

4.1. Cooling Load Calculations ... 107

4.1.1. Temperature Distribution Under The Passenger Compartment ... 107

4.1.2. Cooling Load By Conduction And Convection ... 113

4.1.3. Heat Transfer By Radiation ... 115

4.1.4. Cooling Load By Internal Sources ... 115

4.1.5. Cooling Load By Ventilation ... 116

4.1.6. Cooling Load By Infiltration ... 116

4.2. CFD Analyses ... 118

4.2.1. Determination Of Turbulence Model ... 118

4.3. Channel Design Studies ... 131

4.3.1. Effect Of Main And Exit Channels’ Diameters on Exit Velocity Fluctuations ... 132

4.3.2. Effect Of Zoning ... 142

4.3.3. Effect Of Employing Exit Channels Having Different Diameters ... 151

4.3.4. Final Air Channel Geometry Evaluations ... 164

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4.4.3. Installation Of Air Channel To The Test Vehicle ... 186

4.4.4. Design Verification Tests And Results ... 192

5. CONCLUSSIONS AND RECOMMENDATIONS ... 197

5.1. Conclusions About the Test to Determine the Temperature Distribution Underneath the Passenger Compartment ... 200

5.2. Conclusions About Cooling Load Calculations ... 201

5.3. Conclusions About Determination Of Turbulence Model, Grid and CFD Analyses ... 204

5.4. Recommendations for Future ... 206

REFERENCES ... 207

CIRRICULUM VITAE ... 213

APPENDIX ... 215

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Table 3.2. Detail region’s names and codes ... 35

Table 3.3. Material properties table for region FL1 ... 36

Table 3.14. Material properties table for region RS1 ... 38

Table 3.20. Material properties table for region LS1 ... 39

Table 3.24. Material properties table for region FR1 ... 39

Table 3.25. Material properties table for region FR2 ... 40

Table 3.26. Material properties table for region RR1 ... 40

Table 3.27. Material properties table for region RR2 ... 40

Table 3.28. Material properties table for region RO1 ... 40

Table 3.29. Material properties table for region RO2 ... 40

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cabin during test ... 65

Table 3.33. Vertical temperature difference along the vehicle ... 67

Table 3.34. Temperature differences along the vehicle ... 68

Table 3.35. Internal noise values in different locations with different fan speed levels... 70

Table 3.36. Properties of grids ... 88

Table 3.37. Number of element and equisize skewness of grids ... 88

Table 3.38. Measured and average exit velocity values taken from experimental setup ... 91

Table 3.39. Grid Independency study analysis conditions ... 92

Table 3.40. Resultant average exit velocity magnitudes of each of the analyses ... 93

Table 3.41. Values of deviations between the grids... 94

Table 3.42. Results of residuals. ... 101

Table 3.43. Default Under-Relaxation Factors. ... 102

Table 3.44. Measurement results of evaporator outlet velocities... 104

Table 3.45. Analyses conditions. ... 106

Table 4.1. Average temperatures of the regions under the vehicle floor. ... 113

Table 4.2. Amount of heat transferred from vehicle floor region (FL)... 113

Table 4.3. Amount of heat transferred from right sidewall region (RS) ... 114

Table 4.4. Amount of heat transferred from left sidewall region (LS) ... 114

Table 4.5. Amount of heat transferred from front surface region (FR) ... 114

Table 4.6. Amount of heat transferred from rear surface region (RR) ... 114

Table 4.7. Amount of heat transferred from roof region (RO) ... 114

Table 4.8. Total convection and conduction heat transfer to the passenger compartment ... 115

Table 4.9. Cooling load by ventilation calculation table ... 116

Table 4.10. Details of infiltration cooling load ... 117

Table 4.11. Total cooling load table ... 117

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Table 4.14. Standard deviations of Turbulence models ... 124

Table 4.15. Experimental test results for single fan application ... 125

Table 4.16. Results obtained from experiment and analyses with single fan application ... 126

Table 4.17. Standard deviations of Turbulence models ... 131

Table 4.18. Results obtained form analysis of 150/50 channel... 132

Table 4.22. Results obtained form analysis of 150/50 F2/R4 zoned channel ... 143

Table 4.23. Results obtained form analysis of 150/35/50 F2/R4 zoned channel ... 145

Table 4.24. Results obtained form analysis of 150/50 F3/R3 zoned channel ... 147

Table 4.25. Table of exit diameters of the channel having variable exit diameters . 152 Table 4.26. Analysis results of the channel having variable exit diameters ... 153

Table 4.27. Table of exit diameters of the channel having variable exit diameters . 155 Table 4.28. Analysis results of the channel having variable exit diameters ... 155

Table 4.29. Table of exit diameters of the modified channel having variable exit diameters ... 157

Table 4.30. Analysis results of the modified channel having variable exit diameters ... 158

Table 4.31. Table of exit diameters of the modified channel having variable exit diameters ... 160

Table 4.32. Analysis results of the modified channel having variable exit diameters ... 160

Table 4.33. Analysis results of the channel having variable exit diameters ... 162

Table 4.34. Time-averaged patterns of streamlines and corresponding velocity contours in main channel center plane ... 171

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exits ... 176

Table 4.37. Time-averaged patterns of streamlines and corresponding velocity contours in each of exit center planes ... 178

Table 4.38. Time-averaged patterns of streamlines and corresponding velocity contours in each of inlet center planes ... 180

Table 4.39. Table of temperature distribution ... 194

Table 4.40. Vertical temperature difference along the vehicle ... 195

Table 4.41. Temperature differences along the vehicle ... 196

Table 5.1. A/C power requirements in relation to the climatic region and the physiological perceptions of the passengers. (Boltz, et. Al., 2011) ... 202

Table 5.2. Amount of heat transferred from vehicle floor region and heat fluxes .. 203

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registered ... 2

Figure 1.2. Schematic diagram of a conventional bus AC system... 5

Figure 1.3. A sample of integrated system ... 6

Figure 1.4. A sample of rooftop unit ... 6

Figure 1.5. A bus equipped with a rooftop A/C unit and belt driven compressor located in the engine room ... 7

Figure 1.6. A section view demonstrating A/C system integration to a coach ... 8

Figure 1.7. Body structure of the test bus ... 9

Figure 1.8. Outer appearance of a coach ... 10

Figure 1.9. Insulating materials employed in buses in different locations... 11

Figure 1.10. Codes of the regions ... 12

Figure 1.11. A view from current air channel inside ... 12

Figure 1.12. Cooling load components of a bus ... 18

Figure 3.1. Thermo-couple installation plan ... 42

Figure 3.2. Position of sensors (8,9, and 10) to measure the temperature above the front axle ... 42

Figure 3.3. Position of sensor number 20 and 23 to measure the temperature in rear left luggage compartment ... 43

Figure 3.4. Position of sensor number 22 and 25 to measure the temperature in battery case ... 43

Figure 3.5. Position of sensor number 26 and 29 to measure the temperature above the radiator... 44

Figure 3.6. Position of sensor number 28 and 31 to measure the temperature above the exhaust ... 44

Figure 3.7. Position of sensors number 29, 30, and 31 ... 45

Figure 3.8. A screen shot of the calculation algorithm of the program ... 48

Figure 3.9. Calculation options when vehicle is driven in the same direction continuously ... 50

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Figure 3.13. Rear of the vehicle follows the Sun ... 52

Figure 3.14. Photo of the selected A/C unit ... 55

Figure 3.15. Installation area on the roof of the vehicle ... 55

Figure 3.16. A view from the test vehicle ... 55

Figure 3.17. Current cross-section of the air channel and roof top A/C unit ... 56

Figure 3.18. Seats equipped with thermo-couples ... 57

Figure 3.19. Thermo-Couple installation points of each seat ... 57

Figure 3.20. A photo of a sample seat thermo-couple installation ... 58

Figure 3.21. Seats to be equipped with thermo-couples ... 58

Figure 3.22. Installation photos of the Sun light radiation-measuring device ... 59

Figure 3.23. Data logger and the computer ... 59

Figure 3.24. Hourly cooling load distribution at different typical day times around year ... 60

Figure 3.25. Photos of the test vehicle during heat up stage ... 61

Figure 3.26. Measurement of surface temperature of the dashboard of the vehicle with laser thermometer ... 62

Figure 3.27. Measurement of surface temperature of the steering wheel of the vehicle with laser thermometer ... 62

Figure 3.28. Measurement of surface temperature of the inside roof panels of the vehicle with laser thermometer ... 63

Figure 3.29. Measurement of exit velocity of the nozzles ... 64

Figure 3.30. Air velocities at the nozzles measured from the current vehicle ... 64

Figure 3.31. Temperature distribution at the 60^th minute of the test ... 66

Figure 3.32. Internal noise measurement point, front ... 69

Figure 3.33. Internal noise measurement point, middle ... 69

Figure 3.34. Internal noise measurement point, rear ... 70

Figure 3.35. Section view of designed air channel in current luggage rack ... 72

Figure 3.36. Position of the air channel on the vehicle ... 72

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Figure 3.40. Isometric 3D view of the base channel ... 73

Figure 3.41. Evaporator fans used in the selected A/C roof top unit. ... 75

Figure 3.42. 3D CAD model of the fan unit used in the experimental setup. ... 76

Figure 3.43. 3D CAD view of the experimental setup. ... 76

Figure 3.44. Drilling operation and some of the exit holes on the main channel. ... 77

Figure 3.45. Sample of exit channels and reference rings ... 78

Figure 3.46. A sample view from the sealer mastic application and bonded exit channel ... 78

Figure 3.47. General view of the experimental setup at the production stage ... 79

Figure 3.48. Views from CNC machining stage and the resultant machined base mold ... 80

Figure 3.49. Mold and production stages of an inlet channel ... 80

Figure 3.50. Inlet channels ... 81

Figure 3.51. Stages of bonding procedure (a): Sanding, (b): Activator, (c): Primer, (d): Sealer mastic... 81

Figure 3.52. Inlet region of experimental setup ... 82

Figure 3.53. Painting operation of inlet region and a sample of painted exit channel ... 82

Figure 3.54. Installation of fans and harness... 83

Figure 3.55. Installation of control switches and relays... 83

Figure 3.56. General appearance of experimental setup ... 83

Figure 3.57. A sample of inlet channel ... 84

Figure 3.58. A sample of exit channel ... 84

Figure 3.59. 3D model of flow domain ... 85

Figure 3.60. Grid generated for main channel (a): Section view, (b): A portion of side view ... 86

Figure 3.61. Side view of grid generated for inlet channel ... 87

Figure 3.62. Grid generated for exit channels (a): Section view, (b): Side view ... 87

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Figure 3.66. Continuum definitions of flow domain... 90

Figure 3.67. Exit velocity measurement ... 91

Figure 3.68. Graph of grid independency analyses result ... 93

Figure 3.69. Graphics of residuals converged to 10^-5 ... 101

Figure 3.70. Graphics of area weighted average exit velocity of Exit 1 ... 101

Figure 3.71. Graphics of area weighted average exit velocity of Exit 14 ... 102

Figure 3.72. Air nozzle employed in the test vehicle ... 103

Figure 3.73. Evaporator outlet velocities measurement points a: Left, b: Center, c: Right, d: Exits of evaporator fan ... 104

Figure 3.74. Front (a) and side (b) views of the channel having 150 mm main, 50 mm exit diameters and single part geometry ... 105

Figure 4.1. Graphic of the temperatures above the radiator (FL1) ... 107

Figure 4.2. Graphic of the temperatures above the engine (FL2) ... 108

Figure 4.3. Graphic of the temperatures above the exhaust (FL3) ... 108

Figure 4.4. Graphic of the temperatures above the rear left luggage (FL4) ... 109

Figure 4.5. Graphic of the temperatures above the transmission (FL5) ... 109

Figure 4.6. Graphic of the temperatures above the battery (FL6) ... 110

Figure 4.7. Graphic of the temperatures above the rear axle (FL7) ... 110

Figure 4.8. Graphic of the temperatures above the luggage room (FL8) ... 111

Figure 4.9. Graphic of the temperatures above the front axle (FL9) ... 111

Figure 4.10. Graphic of the temperatures above the fuel tank (FL10) ... 112

Figure 4.11. Graphic of the temperatures below the driver platform (FL11) ... 112

Figure 4.13. Graphics of comparison of area weighted average exit velocities of experimental and analysis with k- Standard turbulence model ... 119

Figure 4.14. Graphics of comparison of area weighted average exit velocities of experimental and analysis with k- RNG turbulence model ... 120

Figure 4.15. Graphics of comparison of area weighted average exit velocities of experimental and analysis with k- Realizable turbulence model ... 120

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experimental and analysis with k-Ω turbulence model ... 121 Figure 4.18. Correlation graphics with calculated correlation coefficients (R²)... 123 Figure 4.19. Single fan configuration ... 125 Figure 4.20. Graphics of comparison of area weighted average exit velocities of

single fan experimental and analysis with k- Standard turbulence model ... 127 Figure 4.21. Graphics of comparison of area weighted average exit velocities of

single fan experimental and analysis with k- RNG turbulence model127 Figure 4.22. Graphics of comparison of area weighted average exit velocities of

single fan experimental and analysis with k- Realizable turbulence model ... 128 Figure 4.23. Graphics of comparison of area weighted average exit velocities of

single fan experimental and analysis with Spallart-Almaras

turbulence model ... 128 Figure 4.24. Graphics of comparison of area weighted average exit velocities of

single fan experimental and analysis with k-Ω Standard turbulence model ... 129 Figure 4.25. Correlation graphics with calculated correlation coefficients (R²)... 130 Figure 4.26. Graphic of area weighted average exit velocities of 150/50 channel .. 133 Figure 4.27. Graphic of mass flow rate of 150/50 channel ... 133 Figure 4.28. Graphic of area weighted average exit velocities of 130/50 channel .. 134 Figure 4.29. Graphic of mass flow rate of 130/50 channel ... 135 Figure 4.30. Graphic of area weighted average exit velocities of 130/30 channel .. 136 Figure 4.31. Graphic of mass flow rate of 130/30 channel ... 137 Figure 4.32. Graphic of area weighted average exit velocities of the channels ... 137 Figure 4.33. Graphic of mass flow rates of the channels ... 138 Figure 4.34. Graphic of area weighted average exit velocities of 150/30 channel .. 140 Figure 4.35. Graphic of mass flow rate of 150/30 channel ... 140

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Figure 4.39. Detail of inlet region of the zoned channel ... 142

Figure 4.40. Graphic of area weighted average exit velocities of 150/50 F2/R4 zoned channel ... 143

Figure 4.41. Graphic of mass flow rates of 150/50 F2/R4 zoned channel ... 144

Figure 4.42. General design of the zoned channel ... 144

Figure 4.43. Detail of inlet region of the zoned channel ... 144

Figure 4.44. Graphic of area weighted average exit velocities of 150/35/50 F2/R4 zoned channel ... 145

Figure 4.45. Graphic of mass flow rates of 150/35/50 F2/R4 zoned channel ... 146

Figure 4.46. General design of the new channel ... 146

Figure 4.47. Detail of inlet region of the new channel... 147

Figure 4.48. Graphic of area weighted average exit velocities of 150/50 F3/R3 zoned channel ... 148

Figure 4.49. Graphic of mass flow rates of 150/50 F3/R3 zoned channel ... 148

Figure 4.50. Comparison graphic of the effect of zoning ... 149

Figure 4.51. Comparison graphic of mass flow rates of all zoning applications ... 150

Figure 4.52. General view of the channel having various exit diameters ... 152

Figure 4.53. Graphic of area weighted average exit velocities of the channel having variable exit diameters ... 153

Figure 4.54. Graphic of mass flow rates of the channel having variable exit diameters ... 154

Figure 4.54. Graphic of area weighted average exit velocities of the channel having variable exit diameters ... 156

Figure 4.55. Graphic of mass flow rates of the channel having variable exit diameters ... 156

Figure 4.56. Graphic of area weighted average exit velocities of the modified channel having variable exit diameters ... 158

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channel having variable exit diameters ... 161

Figure 4.59. Graphic of mass flow rates of the modified channel having variable exit diameters ... 161

Figure 4.60. Graphic of area weighted average exit velocities of the channel 150/40 ... 163

Figure 4.61. Graphic of mass flow rates of the channel 150/40 ... 163

Figure 4.62. General view and sections of the final channel ... 164

Figure 4.63. Front view of the final channel ... 165

Figure 4.64. Side (a) and top (b) views of the final channel ... 165

Figure 4.65. Technical drawing the final channel ... 166

Figure 4.66. Description of inlet’s center planes ... 167

Figure 4.67. Description of main channel and Exit channels’ center planes ... 167

Figure 4.68. Area weighted average contours of velocity magnitude in main channel center plane of final channel ... 168

Figure 4.69. Area weighted average contours of velocity magnitude inlet region where the fans blow air inside to the channel ... 168

Figure 4.70. Contours of turbulent intensity in main channel center ... 169

Figure 4.71. Contours of turbulent intensity in inlet region ... 169

Figure 4.72. Contours of Turbulent Reynolds Number in main channel center plane of the channel ... 170

Figure 4.73. Contours of Turbulent Reynolds Number in main channel center plane near the inlet region ... 170

Figure 4.74. Contours of static pressure on the center plane of main channel ... 175

Figure 4.75. Contours of static pressure on the center plane of main channel near inlet region ... 175

Figure 4.76. General view of air channel in current luggage rack ... 181

Figure 4.77. 3D design model view of distribution box ... 182

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Figure 4.80. General appearance of overhead module and distribution box ... 183

Figure 4.81. Mold to be used to produce the inlet channels ... 184

Figure 4.82. A sample of inlet channel ... 184

Figure 4.83. Photos of exit channel from different point of views ... 185

Figure 4.84. Produced and designed air distribution box inner surface views ... 185

Figure 4.85. Photos of air distribution box assembled to the overhead module ... 186

Figure 4.86. A view from dissembled luggage racks ... 187

Figure 4.87. Detail of inner side of the current air channel ... 187

Figure 4.88. Upper surface of current air channel with hot water hoses ... 188

Figure 4.89. Upper surface of current air channel with A/C gas hoses ... 188

Figure 4.90. Photo of inlet region of the channel ... 189

Figure 4.91. General appearance of the air channel inside the luggage rack ... 189

Figure 4.92. Connections between air channel and A/C fan outlets via inlet channels ... 190

Figure 4.93. A sample of connection of distribution box ... 190

Figure 4.94. Connections between exit channels and distribution boxes at the rear of the channel ... 191

Figure 4.95. A sample of connection of distribution box with overhead module .... 191

Figure 4.96. Final appearance of the test vehicle ... 192

Figure 4.97. A Photo from the test vehicle, equipped with thermo-couples ... 192

Figure 4.98. Measurement of exit velocities ... 193

Figure 4.99. Air velocities at the nozzles measured from the vehicle ... 193

Figure 4.100. Air velocities at the nozzles measured from the vehicle ... 195

Figure 5.1. Comparison of current situation and new application results ... 199

Figure 5.2. Comparison of numerical and experimental results ... 205

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α : Absorption coefficient of the surface Cp : Specific heat of air (W/kgK)

ΔR : Difference between total radiation and reflected radiation (W/m²)

t : Temperature difference between outside and inside (⁰C) Et :Total radiation in unit surface (W/m²)

ED : Amount of direct radiation (W/ m²) Ed : Radiation by diffusion (W/ m²) Er : Radiation by reflection (W/ m²) ε : Surface reflection coefficient F : Heat transfer surface area (m²) Ful : Using factor

Fsa : Lightening equipment factor g : Acceleration due to gravity h : Sensible enthalpy

hin : Inner surface heat transfer coefficient (W/mK) hout : Outer surface heat transfer coefficient (W/mK) ho : Outer surface total heat transfer coefficient (W/m²K)

I : Unit tensor

IAC : Inside curtain coefficient J : Diffusion flux

k : Thermal conductivity of the material (W/mK) keff : Effective thermal conductivity of the region kt : Turbulent thermal conductivity

Lk : Characteristic Length (m) m : mass flow rate (kg/s) N : Number of people inside

Nu : Nusselt Number

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Pr : Prandtl Number

qb : Heat gain by direct radiation (W) qd : Heat gain by solar reflection (W) qc : Heat gain by solar conduction (W) qsensible : Amount of sensible heat (W) qlatent : Amount of latent heat (W)

Re : Reynolds Number

s : Thickness of the material (m) Sh : Volumetric heat generation term

SGHC( ) : Heat gain coefficient by direct radiation (SHGC)D : Heat gain coefficient by diffusion Sm : Mass source term

ti : Insdie temperature (⁰C) to : Outside temperature (⁰C) ts : Surface temperature (⁰C)

tb : Adjacent volume average temperature (⁰C) U : Total heat transfer coefficient (W/ m²K) W : Total lightening power (W)

: Viscous stress tensor

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1. INTRODUCTION

1.1. Significance of Thermal Comfort

Maintaining thermal comfort for occupants in buildings or other enclosures, even in extreme climatic conditioning requirements and irrespective of the environmental outside conditions, has been the main focus for the Heating Ventilation and Air Conditioning (HVAC) design engineers and systems developers (Alahmer et. al., 2011).

Especially in the vehicles thermal comfort is more important, since the thermal environment and air quality in a passenger car can affect driver's and passengers' health, performance and comfort (Ruzic, 2011).

The early history of transportation systems starts mainly with the horse drawn carriage. This was eventually surpassed by the invention of the automobile. Early automobiles had cabin spaces that were open to the outside environment. This means that the occupants had to adjust their clothing to allow for different weather conditions. Closed cabin spaces were eventually introduced which required heating, cooling and ventilating to meet customer expectations (Daly, 2006).

On the other hand, thermal comfort of vehicular occupants is gaining more importance due to the rising focus on comfortable mobility, in addition to the fact that, the time that people spend in vehicles (private or public transport) has grown substantially (Alahmer et. al., 2011). The following figure shows the percentage of the Air Conditioning (A/C) equipped vehicle to the total number of vehicle registered between 1992 and 2004 years in United Kingdom (Daly, 2006). As can be seen clearly that in early of 90s, the percentage of A/C equipped vehicle is below 10%, in 2004 this rate is above 75%. Now this rate is above 95%. Air conditioning has become a standard option on most vehicles, enhancing comfort and safety (Kaynaklı and Horuz, 2003).

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Figure 1.1. Percentage of the A/C equipped vehicle to the total number of vehicle registered. (Daly, 2006)

Modern man's desire for mobility has been a major factor in technical developments. Technical solutions for conveying people in comfort were sought at an early stage. The bus (from Latin omnibus: "for all") has played a significant role in this development. With in the introduction of enclosed bodywork, passengers were given a certain level of comfort. Today, buses still account for a major potion of road transport with an upward trend. The most convincing argument to the passenger in favor of bus travel, however, is the comfort offered, and the interior climate plays a key role here (Boltz et. al., 2011).

The operation condition of mobile air conditioning systems is considerably different from the residential air conditioning. The ambient temperature, solar radiation, vehicle speed have great influences on the system performance (Qi et. al., 2007). And also, the vehicular in-cabin environments is affected by a large number of parameters that include the different interior trim surface air temperatures, the air velocity and its profiles over the different geometries, the relative humidity, the solar intensity and its reflections over the different material types and surface finishes in the cabin, the angles of incidence, the type of clothes, etc. Also many of these parameters are dependent with unknown relationships (Han and Huang, 2005).

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Thermal environment in passenger cars differs from those in buildings and is often highly non-uniform and asymmetric, from the following reasons:

* Interior volume is small, compared with the number of persons,

* Changing of microclimate parameters could be rapid (vehicle is changing the orientation to the sun, etc),

* The shape of the cabin interior is complex,

* Glazing area is large in comparison to cabin surface,

* Passengers seat near surfaces with temperatures, which could be significantly higher or lower than interior air temperatures,

* Passengers are not able to change position within the cabin, and changes of body posture are limited,

* The air-conditioning is usually not activated when there is nobody in the car or when the engine is not running, resulting in occurrence of extreme microclimate conditions (Ruzic, 2011).

On the other hand, Travel comport in buses presents extra challenges in several respects: large vehicle sizes with an internal volume of about 50 m³ (Compared to 1,5 to 3 m³for a motor car), limited usable waste heat from diesel engine, large window surfaces, special design such as double-decker and articulated buses, vehicle zones subject to different thermal influences, use of buses in different climatic zones around the world, the different perception of comfort of up to 180 passengers and not least the driver as an interface between man and machine. All these factors must be taken into account in the conception of suitable solutions. The declared goal is to ensure the well being of each individual passenger (Boltz, 2011).

Nevertheless, conditions for human thermal comfort in vehicles should be the same as in buildings. Standards for thermal comfort suggest that combination of microclimatic parameters of indoor environment (temperature, air velocity and relative humidity) and individual parameters (clothing and activity) should be within certain limits, spatially homogeneous and steady state. Vehicle HVAC system

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obtains desired interior environment by introducing the cooled/heated/dried air into the cabin air through the system's outlets (Ruzic, 2011).

Vehicle HVAC system delivers conditioned air into the cabin space, and human body is cooled by convection and sweat evaporation. The heat loss from the body will mainly depend on air velocity and temperature over the individual body part. The resulting distribution of skin temperature over the body surface will dictate human thermal sensation and thermal comfort (Ruzic, 2011). Also, the overall accuracy of the thermal comfort predictions will depend on how accurately we can predict the airflow and thermal environment in the passenger compartment (Huang and Han, 2002). So airflow velocity and uniform distribution inside the vehicle play an important role in thermal comfort of the passengers.

1.2. Air Conditioning (A/C) Systems For Buses

In order to satisfy passengers‟ comfort requirements in buses, an Air- Conditioning (A/C) system is employed. Although different solutions exist for air conditioning for buses, general working principle for the system is the same. General working principle of a bus A/C system can be summarized as follows; Figure 1.2 shows the schematic diagram of a conventional bus AC system, which comprises of two cooling coils to provide the conditioned air to both rows of the passenger compartment, one compressor, a receiver, dryer, condenser, and 2 expansion valves.

The saturated steam of refrigerant is injected into a heat exchanger (evaporator) energetically connected to the interior of the bus. The refrigerant has a boiling or evaporating temperature, which is pressure regulated by an expansion valve and is lower than the temperature of the vehicle interior. This causes the liquid portion of the refrigerant to vaporize. The refrigerant absorbs energy from the vehicle interior by the endothermic evaporation process. A compressor draws in the vaporized refrigerant and conveys it at a higher pressure to the condenser, through which ambient air flows. In turn, a phase transition takes place in the condenser:

from a gaseous to a liquid state. Liquefaction is an exothermal process through which

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the energy from the interior and introduced by compression in the compressor is released to the surroundings (Boltz et. al., 2011).

Figure 1.2. Schematic diagram of a conventional bus AC system (Boltz et. al., 2011) In some cases instead of 1 condenser, 2 condensers can be used. It should be noted that the two condensers are connected in series to allow the condensation process to occur in two stages whereas the evaporators are connected in parallel to allow equal cooling rate to both passenger rows (Mansour et. Al., 2007).

The general cooling cycle described above can be applied to the buses with different configurations. These may be units integrated into the vehicle interior, which is installed inside luggage rack and called integrated system, or a rooftop unit, which may be installed to the top of the vehicle. In both cases compressor is driven

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by the diesel engine of the bus. A sample of integrated system and a rooftop unit can be seen in Figure 1.3 and 1.4 respectively.

Figure 1.3. A sample of integrated system (Spheros, 2014a)

Figure 1.4. A sample of rooftop unit (Spheros, 2014b)

The use of a rooftop passenger vehicle (buses or trains) air-conditioning (A/C) system has been steadily growing in countries experiencing hot and humid

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tropical climates. It is quite common to see a bus AC compressor driven by an auxiliary engine to ensure a constant speed operation (Mansour et. al., 2007). A general view of a bus equipped with a rooftop A/C unit, which is the component numbered with 2, and a belt driven compressor, which is the component numbered with 3 in Figure 1.5 can be observed.

Figure 1.5. A bus equipped with a rooftop A/C unit and belt driven compressor located in the engine room (Temsa, 2012)

Rooftop units are mounted on the top of the vehicle roof structure. This unit accommodates all the components of the A/C system except the compressor, which is driven by the engine and installed to the engine room of the vehicle. The conditioned air is send to the air channel of the bus, and then it is delivered to the passenger cabin via air nozzles on the luggage rack, which is located on top of the passengers. A section view demonstrating A/C system integration to a coach can be seen in Figure 1.6.

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Figure 1.6. A section view demonstrating A/C system integration to a coach (Spheros, 2014c)

Another aspect of vehicle A/C system is the fuel consumption. An A/C system is the second biggest energy consumer component in either intercity or city buses. Clutch cycling (or on/off cycling) is the most common method for the control of bus A/C systems. If the AC system is driven by the main engine, driver can easily feel the drop in the vehicle power when the electromagnetic clutch is engaged (Mansour et. al., 2007).

AC systems are often over-designed first to ensure a fast response so that the cabin temperature drops quickly when the system is switched on and second to overcome the irregular and rare conditions of extremely high humidity and high atmospheric temperature. Thus, under normal conditions of low cooling load, a lot of the energy is unnecessarily wasted (ASRHAE, 2000) and this also results in higher consumption of fuel. Only in the United States alone approximately 26 billion liters of fuel are consumed annually for cooling vehicle passenger compartments (Rugh and Hovland, 2003).

Global petroleum energy is estimated to be depleted in about 40 years and the recent instability in oil price is much related to the rapid development of the transportation and industrial sectors. This cost can be reduced by improving climate control systems design. For this reason, researchers and AC system manufacturers seek to improve automobile AC systems design and technology to reduce the fuel

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consumption rate without forfeiting passenger thermal comfort (Mansour et. Al., 2007).

1.3. Bus Body Structure

Heat input and loss are greatly influenced by the design (shell) and glazing.

Buses differ from motorcars not only in size and use, but also in their design (Boltz et. al., 2011). Structural elements such as high-strength rectangular tubing and rolled sections, sometimes made of stainless steel, are weld with gusset plates. Views from different perspectives of body structure of the test bus to be evaluated can be seen in Figure 1.7.

Figure 1.7. Body structure of the test bus

This shell is covered with siding panels, e.g. aluminum, steel sheet or Fiber Reinforced Plastic (FRP), which are partially welded and/or glued. And the side glasses are also glued to the structure. The resultant appearance of the coach to be studied at the end of application of external panels, FRP covers, side glasses and luggage lids can be seen in Figure 1.8.

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Figure 1.8. Outer appearance of a coach (Temsa, 2014)

Thermal protection is increasingly important in the development process of passenger cars. Tightly packaged engine compartments and strongly increased engine power demand extensive testing and analysis (Binner et. al., 2006). The influence of environmental changes on underhood and underbody components of a vehicle during various driving cycles is an important issue in the automotive industry (Apolloni, 2006). Underhood component temperatures are sensitive to all three modes of heat transfer, conduction, convection and radiation (Weidmann et Al, 2007).

In order to protect the passenger compartment from heat coming from outside and underneath of the bus. The cavities between the outer and inner lining are filled with noise and heat insulating materials. And side windows of coaches are double- glazed. Some insulating materials employed in buses in different locations can be seen in Figure 1.9 with its section views.

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(a): Engine Room

(b): Roof Structure

(c): Side body

Figure 1.9. Insulating materials employed in buses in different locations

In the present study temperature distribution under the passenger compartment of a coach in different usage conditions will be tested by considering the components that can be assumed as a heat sources under the vehicle like, engine, transmission, exhaust muffler, axles etc.

The bottom of the passenger compartment will be sub-divided into 11 regions by considering the mechanical components in the regions and the section properties.

Codes of the regions can be seen in Figure 1.10.

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Figure 1.10. Codes of the regions

It was decided to perform the temperature determination tests under 3 different conditions. These are high speed, highway test condition, forced performance, uphill climbing test condition and stationary, idle test condition.

1.4. Current Air Channel

In the current situation, an air channel created by the boundaries of roof structure and luggage racks is employed in the vehicle. Also this air channel accommodates luggage rack connection brackets, heating water and refrigerant fluid pipes of A/C unit, electrical harnesses and some small brackets to mount electrical components. A photo showing inside of current air channel is provided in Figure 1.11.

Figure 1.11. A view from current air channel inside

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As a result of this structure, some pressure drops, air leakages to undesired locations, non-uniform air distribution along the air channel, non-uniform temperature distribution along the vehicle and in different levels of the passenger cabin (Head, Lumbar and ankle levels of the passenger) are observed in the vehicle.

In the current study, air channel of the A/C system of a 12-meter coach will be developed by applying a distinct, dedicated air channel providing uniform air distribution and consequently a uniform temperature distribution inside the passenger cabin by means of CFD.

1.5. Computational Fluid Dynamics (CFD)

For the design of distinct air channel, CFD, standing for Computational Fluid Dynamics, technique will be employed, which is the most famous and powerful design tool of last decades. CFD technique can be used in lots of applications from aviation to transportation, from HVAC to textile industry. Especially in Automotive, CDF technique can be used in a wide range of fields. Flows around the vehicles, underhood thermal analyses are some examples only.

On the other hand, since HVAC systems are one of the most important parts of a vehicle, they must be examined in detail. These examinations can be done by theoretically, experimentally or numerically.

In the current study, a series of Computational Fluid Dynamics studies will be conducted to determine the most suitable air channel dimensions and configuration.

CFD represents a vast area of numerical analysis in the field of fluid‟s flow phenomena. Headway in the field of CFD simulations is strongly dependent on the development of computer-related technologies and on the advancement of our understanding and solving ordinary and partial differential equations (ODE and PDE). However CFD is much more than “just” computer and numerical science.

Since direct numerical solving of complex flows in real-like conditions requires an overwhelming amount of computational power success in solving such problems is very much dependent on the physical models applied. These can only be derived by

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having a comprehensive understanding of physical phenomena that are dominant in certain conditions (Sodja, 2007).

Computational fluid dynamics has become a popular tool in performance analysis and design, but one of the major challenges of CFD is turbulent flow. There are many turbulence models available (Neel, 1997). One of the most important parts of a CFD analysis is the determination of the suitable turbulence model.

In the current study, the geometry designed will be analyzed with 5 different turbulence models, which are k-ε Standard, k-ε Realizable, k-ε RNG, k-Ω Standard, and Spallart Almaras. Then the numerical results will be compared with the experimental results performed in the experimental set up, which will be produced.

The determination of a proper grid for the flow over or through a given geometric shape is a serious matter. The way that such a grid is determined is called grid generation. The matter of grid generation is a significant consideration in CFD;

the type of grid you choose for a given problem make or break the numerical solution. Because of this, grid generation has become an entity by self in CFD (Anderson, 1995). In the current study, Grid will be generated by GAMBIT^®package program.

The matter of grid independence is also a serious consideration in CFD. In general when you solve a problem using CFD, you are employing a finite number of grid points (or a finite mesh) distributed over the flow field. You must increase the number of grid points until you reach a solution, which is no longer sensitive to the number of points. When you reach this situation, then you have achieved grid independence (Anderson, 1995). In the current study a grid independence study will be conducted with 4 different grids having different densities and the suitable grid will be determined.

After determining the suitable turbulence model and grid, air channels having different dimensions (inlet/outlet diameters) and configurations (with/without zoning) will be analyzed and by evaluating the results of exit velocities and mass flow rates it will be determined the suitable air channel dimensions.

Another aspect of the numerical solution is that of boundary conditions.

Because, without the physically proper implementation of boundary conditions and

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their numerically proper representation, we have no hope whatsoever in obtaining a proper numerical solution to our flow problem (Anderson, 1995). In order to start the numerical analysis, the inlet boundary condition must be known. So the A/C unit must be selected firstly. And the evaporator exit velocity values must be measured.

1.6. Cooling Load Calculation

In order to select the A/C unit, which will be used in the test vehicle, the cooling load of the vehicle must be calculated.

The first step of an HVAC system design is to calculate the cooling and/or heating load of the media to be conditioned. The number of parameters affecting cooling load is very high. The most important parameter in calculating cooling load, differently from the heating load, is the load coming from the sunlight. Another factor making the cooling load calculation difficult is that the parameters must be taking account with dependently not individually. Because of these reasons, cooling load calculation is a complex and high intensive process.

Cooling load calculation of a bus as a moving media is more difficult and complex than a stationary building. Because the position of the vehicle to the sun is changing continuously and the convective heat transfer coefficient on the outer face of the vehicle is also chancing due to the vehicle speed (Büyükalaca et. al., 2011).

In literature it is not seen a special procedure or a standard to calculate the bus cooling load calculation. But it is seen some special design considerations. For the standard, German DIN, VDI-Regeln and American ASHRAE sources were investigated. (VDI, 1996), (Alahmer et al, 2011), (ASHRAE, 2001).

Since a special procedure and standard is not available for bus cooling load calculation, firstly the procedures and standards to calculate the cooling load of buildings that can be applied to bus are examined. Lots of calculation techniques to determine the cooling load of a building from simple to complex were developed.

“American Society of Heating Refrigeration and Air-Conditioning Inc.” (ASHRAE) has played an important role in development of the most of these techniques. In

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ASHRAE Handbooks different load calculation procedures were published in time (Büyükalaca et. al., 2011).

Earlier ASHRAE heating and cooling load methods include the Total Equivalent Temperature Differential/Time-Averaging method (TETD/TA), the Transfer Function Method (TFM) and the Cooling Load Temperature Differential (CLTD)/Solar Cooling Load (SCL)/Cooling Load Factor (CLF) method.

(ANSI/ASHRAE, 2001), (Steven, 2004), (Mitalas, 1972), (ANSI/ASHRAE, 1979), (McQuiston and Spitler, 1992). It was shown that no significant difference was found between the peak heating loads calculated by the TFM and TETD/TA methods (Rudoy and Duran, 1975), (Niu et al, 2002), (Zhang and Niu, 2003), (Spitler and Fisher, 1999), (Mui and Wong, 2007).

These techniques can be sub-divided into 2 groups. The first group procedures (like TFM and TETD/TA) consist of 2 steps. In the case of determining the cooling load of the media, firstly the heating load is calculated and than by considering the heat storage properties of the building, cooling load is determined. In the second group procedures, (like CLTD/CLF), cooling load is directly calculated (Büyükalaca et. al., 2011).

The 2001 ASHRAE handbook illustrates a new cooling load calculation method—Radiant Time Series (RTS)—to replace the TETD/TA, TFM and CLTD/SCL/CLF methods (Mui and Wong, 2007).

RTS is a new simplified method for improving the calculation accuracy while maintaining the design engineer‟s ability to apply his/her experience and judgment to the process. It allows the characterization of time delay effects due to exterior surface and building mass in a readily understandable and quantitative form; it also allows for individual component contribution to the total cooling load (Rees et. al., 1998), (Spitler and Fisher, 1999a), (Spitler and Fisher, 1999b), (Spitler et. al., 1997). As a cooling load calculation design tool, RTS is „„well behaved‟‟ in that it would generally over-predict rather than under-predict the peak-cooling load (Mui and Wong, 2007).

Heat Balance (HB) and RTS are the newest load calculation techniques released in ASHRAE Handbook 2001 (ASHRAE, 2001). In these techniques firstly

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the heating load of the media is calculated and than the cooling load is determined.

HB and RTS techniques are very similar to each other basically. However, HB technique requires solving lots of equations accordingly, it is only possible using this technique with detailed computer software. In other words RTS technique can be assumed as the practical form of HB technique.

ASHRAE recommends using HB or RTS techniques to calculate the cooling load of medias out of house, instead of the other cooling load calculation procedures.

Thus in this study it is decided using RTS method, which is the most reliable and modern technique for bus cooling load calculation.

The most important parameter that must be considered in calculating the cooling load of a media whose outer surfaces are subjected to sunlight is the radiation-cooling load. In the building applications, the building is fixed and the sun moves. So it is simpler to determine the heat transferred to the media over the outer wall in any time interval of a day. However, because of the movement of both bus and sun, it is very complex to determine the solar radiation from the outer surface (Büyükalaca et. Al., 2011).

Cooling load components of a coach can be seen in Figure 1.12 as:

a. Radiation cooling load from opaque outer surfaces b. Radiation cooling load from non-opaque outer surfaces c. Conduction cooling load from adjacent volumes d. Cooling loads from Internal sources

i. People ii. Lightening iii. Equipments

e. Infiltration cooling load f. Ventilation cooling load

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Figure 1.12. Cooling load components of a bus

In the current study, the newest and the most effective technique, Radiation Timer Series (RTS) method will be employed to calculate the cooling load of the coach by referencing the study conducted Büyükalaca et al., (2011).

The main aim of this work is to design an A/C air channel for a 12-m coach providing uniform air distribution by means of CFD and consequently a uniform temperature distribution along the passenger compartment, satisfying the following comfort conditions (Temsa, 2007).

* Maximum 5 m/s average velocity at the exit of the nozzles located above each of the passengers with a tolerance of +0,5 m/s.

* It is expected that the A/C system of the vehicle cool down the passenger compartment to 25 °C (Average) at the end of a 60-min test.

* Maximum 3°C of temperature difference between the measuring points.

(Vertical and along the passenger cabin).

Test will be performed under the following conditions according to the (Temsa, 2007).

* Minimum 37°C of outside temperature

* Test vehicle must be heated up under a minimum 950± 95 W/m² of solar radiation at least 3 hours.

* Vehicle speed of min. 90 km/h.

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Thesis outline can be summarized as follows;

In the next section, which is literature survey, the studies conducted on thermal comfort, cooling load calculation, underhood thermal management, Automotive HVAC systems and CFD applications in different industries will be mentioned.

There will be 3 main chapters in Material and Method section namely, cooling load calculation, current situation, and design of air channel by means of CFD.

In Cooling load calculation chapter; procedure and formulas to calculate the cooling load components of the coach will be given, passenger compartment of the coach will be sub-divided into small section to investigate in detail and the section properties of all regions will be determined. The tests to be performed to determine the temperature distribution underneath the passenger compartment will be discussed. And calculation procedure of total heat transfer coefficients of the regions will be mentioned. Radiation Time Series (RTS) method will be mentioned briefly.

And the proper A/C unit will be selected from the market.

In current situation chapter, current channel structure, test preparations and conditions to determine the temperature distribution along the passenger compartment and exit air velocities of each of the nozzles over the passengers will be explained with the measurements and results.

CFD Analyses chapter will start with the determination of the general geometry of the air channel to be applied to the vehicle. Section and general dimensions of the base channel will be determined and discussed with the reasons of selected criteria. Design and manufacturing processes of the experimental setup to determine the suitable turbulence model will be explained. 3D design of flow domain, grid generation and grid independence studies will be explained in detail.

After determination of suitable grid, turbulence model and boundary conditions, the way followed in designing the air channel will be explained.

The section, Results and discussion, having 4 different sections, will start with the test results of temperature distribution under the passenger compartment and determined average temperatures. Cooling load calculation results of all load

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components and total cooling load of the vehicle will be given with the selected A/C unit in the first chapter.

The second chapter of this section will start with the results of turbulence model determination studies. The comparison of the numerical and experimental results of 3-fan and single fan tests will be discussed.

The way starting from the basic channel and going to the final channel will be explained and discussed with exit velocities, mass flow rates, velocity contours, pressure contours and streamlines in the third chapter, which is channel design studies.

In design verification studies chapter, 3D design considerations and production of the final air channel, installation of the air channel to the test vehicle, and measurements of velocity and temperature distribution will be explained with the test results.

In conclusion section, a general conclusion will be made with the results and comments.

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2. LITERATURE SURVEY

Studies conducted on thermal comfort, cooling load calculation methods, and underhood thermal management will be given in the first chapter. In the second chapter, studies about the HVAC applications of vehicles will be mentioned briefly.

In the last chapter, general CFD applications from different industries, CFD applications in automotive industry and HVAC studies conducted with CFD will be mentioned and turbulence model studies will be introduced.

2.1. Studies on Thermal Comfort, Cooling Load Calculation and Underhood Thermal Management

Lots of studies can be found in literature about thermal comfort in buildings and different enclosures. An example to this is the study conducted by Atmaca (2006), which investigated the effects of thermal comfort parameters such as temperature, humidity and air velocity on cervical range of motion experimentally.

For this purpose, each of the three thermal comfort parameters was used as a variable while the others are kept constant in a controlled laboratory room, and the effects of variable parameter were investigated. Human responses to thermal environment such as skin temperature; skin wettedness, sensible and latent heat losses were determined from the developed simulation program. Finally, experimental data and simulation results were assessed simultaneously, and the most appropriate thermal environmental conditions, which did not cause cervical mobility limitations, were determined and they were suggested for both academic and industrial users.

Since thermal comfort in vehicles is a very important subject, researchers and academicians pays attention to this subject and lots of studies can be found in literature about vehicular thermal comfort. Here are some examples.

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Körbahtı (1999), analyzed the parameters that affect the heating condition in buses and heat distribution in a bus and on the outer shell of the intercity bus was calculated under definite conditions.

Kaynakli and Kilic (2005), presented a theoretical and experimental analysis of the in-cabin thermal comfort during the heating period by dividing the human body into 16 segments, with the change of temperature measured and calculated in both experimental and theoretical basis. The air temperature, velocity and relative humidity inside the automobile were acquired experimentally through multiple sensors distributed across the passenger compartment.

In the study performed by Sevilgen (2010), Three-dimensional numerical analysis of temperature and airflow distribution in the automobile cabin was conducted by using Computational Fluid Dynamics method. For this purpose, a three dimensional automobile cabin including window and outer surfaces was modeled by using the real dimensions of a car. In order to evaluate the results of numerical analysis according to thermal comfort, a virtual manikin divided into 17 parts with real dimensions and physiological shape was added to the model of the automobile cabin.

Alahmer et. al. (2011), provided a comprehensive review of the different models developed to predict vehicular cabins thermal comfort, in addition to the different experimental techniques used. The study investigated the different challenges that exist in predicting and evaluating the thermal comfort for vehicular cabins when compared with thermal comfort in buildings. These challenges are mainly related to the fast transient behaviors involved especially the cases of cooling the cabin after a hot soak condition, in addition to the non-uniformities in the thermal environment associated with the high localized air velocity, air temperature distribution, solar flux, and radiation heat flux from surrounding interior surfaces; in addition to other variations related to trip durations and passenger clothing levels.

An overview of preferable microclimate conditions in warm indoor ambient is given in the study conducted by Ruzic, (2011). The data presented in the study were based on numerous experiments with human subjects under different ambient conditions. Focus was on combination of air temperature and local velocity of

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airflow in the region of head and upper body, as thermally most sensitive parts of human body. The results showed that values proposed by standards for thermal comfort, generally used for assessment of indoor thermal environment, could be too restrictive. Preferable conditions are shifted towards higher air velocities.

Furthermore, microclimate parameters that will provide thermal balance of the passenger's body with the surrounding (cab interior) and thermal comfort could not be presented by single value, but by the range of values. The chosen combination of the values would be dependent of individual preferences and local and overall microclimate conditions around different parts of the body. This means that the system must allow precise regulation of local air temperature and velocity in several zones around each passenger's body.

Wu and Ahmed (2012), studied a novel mode of air supply, which has the potential to improve ventilation performance, without increasing the fresh air supply rate. Because, Aircraft cabin ventilation is essential during commercial passenger flights and efficient fresh air delivery has become an important research issue in the field of HVAC.

Different cooling load calculation techniques were studied and compared by several researchers.

Spitler et. al. (1997), described the radiant time series method and the generation of the response factors and the radiant time series coefficients and gives a brief comparison to the heat balance method in their study. At the end of the study, RTS method was compared to the other cooling load calculation methods and RTS method found simpler, more accurate than the other techniques.

Mui and Wong (2007), investigated energy performance and consumption, an example weather year and some occupant load profiles of offices. In the study, the usefulness of the existing example weather year and occupant load variations was investigated. The methods established would be useful for an effective design and an accurate cooling load calculation of air-conditioned buildings, in meeting the demand of occupant loads and updated outdoor information. In particular, a new example weather year and a mathematical model of generating the time variant occupant load profiles using Monte Carlo sampling techniques were used as a basis to calculate the