Thermosyphon flow for cooling of compact devices

THERMOSYPHON FLOW FOR COOLING OF COMPACT DEVICES FILIAN ARBIYANI (B.Eng (Hons), Gadjah Mada University-Indonesia) A THESIS SUBMITTED FOR THE DEGREE OF DOCTOR OF PHILOSOPHY DEPARTMENT OF MECHANICAL ENGINEERING NATIONAL UNIVERSITY OF SINGAPORE 2011 Acknowledgements First of all, I would like to express my infinite gratitude to my Creator, Almighty God, for always being there for me, for granting my wish to get a scholarship to pursue a higher degree, for strengthening me during my PhD, and enabling me to finish my PhD meaningfully. I am also deeply grateful to Yohanes Pembaptis Budi Bayuadi, for encouraging me to apply for doctoral study in NUS, for fertilizing my faith in being able to pursue the higher degree, for supporting me in unpredictable ways, and for his infinite sharing, love and understanding. I would like to express my sincere love and gratitude to my parents, for allowing and believing that I would be able to fulfill my dream of a scholarship, for their love, prayer and nurture that enabled me to develop a strong and independent personality. I would like to express gratitude to my sisters, brother, and family who have continually supported me, and for the many ways in which we could gladly learn from each other. I would like to express my great gratitude to my supervisors. I am sincerely grateful to Prof. Christopher R. Yap, for his kindness, patience, support and enlightenment throughout all the years. I would like to express my greatest appreciation and gratitude to Prof. Nihal E. Wijeysundera, for his patience and guidance, for strengthening my basic knowledge in the early stages of my doctoral study, for always encouraging me in starting my research and for always enlightening and supporting me until the end of my doctoral study. I am most grateful to Prof. Kim Choon Ng, for giving me a lot of knowledge and new skills to improve my research capability, enabling me to complete my doctoral study meaningfully. It is an honour for me to become one of his students. I would also like to extend my gratitude to NUS and AUN/Seed-Net for their financial support that enabled my dreams to come true. i Not to mention, my speechless gratitude is due to my best friends, Wynanda, Christine Windayani, and Albertus Indratno, for always giving me mental support in pursuing my dreams, even though we are on different islands and in different time zones. Our differences make our lives more colorful, and I am glad and grateful that I could know and learn from a good pharmacist, a healthy dentist and a passionate journalist in you. My gratitude is also due to my colleagues, Dr. Mark Aaron Chan, Dr. Kum Ja, Dr. Kyaw Thu, Dr. Loh W. Soong, Dr. Jayaprakash Sathasivam, Mr. Kazi A. Rahman, Mr. Aung Myat, Li Ang, M. Wakil, and Azhar B. Ismail, for sharing good and bad times. I am grateful to the Lab. officers, Mr. Sacadevan, Mr. Yeo, Mr. Chew, Mr. Tan, Mrs. Hung, and Mrs. Roslina, for their kind support. I am sincerely grateful to officers of the ME Fabrication Support Centre, Mr. Lam Kim Song and Mr. T. Rajah, for assisting and teaching me a lot about fabrication work. Finally, I would like to express my gratitude to my fellow NUS Indonesian students, Dr. Patria Kusumaningrum, Dr. Agus Pulung Sasmito, Jundika Candra Kurnia, Tobias Bestari Tjandra, and many others, for their sharing, kindness and support. Singapore, 2011 ii Table of Contents ACKNOWLEDGEMENTS . I TABLE OF CONTENTS III SUMMARY………………………………………………………………………… .VI PREFACE………………………………………………………………………… VIII LIST OF TABLES . IX LIST OF FIGURES . X NOMENCLATURE . XV CHAPTER 1: INTRODUCTION 1.1 General description of two-phase cooling systems 1.2 Research aims 1.3 Scope of study CHAPTER 2: LITERATURE REVIEW 2.1 Cooling systems 2.1.1 Single-phase systems 12 2.1.2 Two-phase systems . 13 2.2 Thermosyphon cooling systems 14 2.3 Heat pipe cooling systems . 17 2.3.1 Grooved-heat pipe 18 2.3.2 Sintered-heat pipe . 19 CHAPTER 3: 3.1 A THERMOSYPHON COOLING SYSTEM 20 Design of thermosyphon water-cooled condenser system 21 3.1.1 Evaporator 23 3.1.2 Riser 23 3.1.3 Condenser . 23 3.1.4 Downcomer 24 3.1.5 Water coolant coil . 24 3.1.6 Working fluid . 24 3.1.7 Heater . 25 3.1.8 Solution bath . 25 3.1.9 Thermocouple . 25 3.1.10 Flowmeter . 26 3.1.11 Insulator 26 iii 3.2 Experimental study of thermosyphon processes . 26 3.2.1 Pre-experiment procedure . 26 3.2.1.1 Thermocouple calibration . 26 3.2.1.2 Leakage investigation . 27 3.2.2 Experiment procedure . 28 3.3 Numerical methodology of study of thermosyphon processes . 28 3.3.1 Numerical method 28 3.3.1.1 Governing equation and Boundary conditions . 28 3.3.2 Simulation program 30 3.4 Characteristics of thermosyphon water-cooled condenser system 34 . . 3.4.1 Effect of mw and Twi on heat dissipation ( Qd ) . 34 3.4.2 Effects of mw and Twi on chip temperature ( Tchip ) 38 . . 3.4.2.1 Effects of mw on chip temperature ( Tchip ) at constant Twi 38 . 3.4.2.2 Effect of Twi on chip temperature ( Tchip ) at constant mw 42 3.4.3 Typical boiling curve 44 3.4.4 Correlation between chip and coolant inlet temperature difference and heat flux 46 3.4.5 Heat-rejection ratio . 50 3.5 Film condensation study in thermosyphon water-cooled condenser system 51 3.5.1 Validation . 51 . 3.5.2 Effect of the water mass flow rate ( mw ) and inlet water temperature ( Twi ) on film condensation rate (  ) 53 . 3.5.3 Effect of the water mass flow rate ( mw ) and inlet water temperature ( Twi ) on Nusselt number film condensation (Nu) 57 . 3.5.4 Effect of the water mass flow rate ( mw ) and inlet water temperature ( Twi ) on Reynold number film condensation (Re ) . 61 3.6 Pressure drop . 65 3.6.1 Pressure drop in riser 65 3.6.2 Pressure drop in downcomer 68 3.6.3 Net pressure drop in thermosyphon system 72 CHAPTER 4: A HEAT PIPE COOLING SYSTEM 76 4.1 Experimental study of the thermal performance of a Helical-grooved heat pipe 78 4.1.1 Experimental study . 78 4.1.2 Experimental procedure 80 4.1.3 Results and discussion 80 4.1.3.1 Uncertainty analysis . 80 4.1.3.2 Temperature distribution and thermal resistance 83 4.1.3.3 The effect of inclination on helical-grooved heat pipe performance 85 4.1.3.3.1 The effect of inclination on thermal boiling resistance of helicalgrooved heat pipe performance 86 4.1.3.3.2 The effect of inclination on overall-system resistance of helicalgrooved heat pipe performance 88 4.1.3.4 The performance of commercial axial-grooved heat pipe 89 4.2 Experimental Study on Thermal Performance of Sintered Heat Pipe 93 4.2.1 Fabrication of Sintered heat pipe 93 iv 4.2.2 Experimental set-up 97 4.2.3 Experimental procedure 99 4.2.4 Results and discussion 100 4.2.4.1 Scanning Electron Microscopy (SEM) . 100 4.2.4.2 Temperature distribution of sintered heat pipe . 100 4.2.4.3 Thermal resistance of sintered heat pipe 102 4.2.4.4 Comparison between helical-grooved heat pipe and sintered heat pipe . 103 CHAPTER 5: CONCLUSIONS . 105 BIBLIOGRAPHY… 108 APPENDIX…………………………………………………………………………115 v Summary The challenge of power dissipation in compact devices requires advanced cooling systems. This research aims to address the two-phase cooling system issues related to various technological problems such as limitation of maximum allowable of chip temperature, design and fabrication of integrated two-phase cooling system, pressure drop in compact cooling system and high condensation rate to achieve high heat dissipation. New designs of three compact cooling systems were explored. A two-phase thermosyphon water-cooled condenser system was built first for an experiment and numerical investigation of the effect of heat load on chip temperature. The effect of heat load on the temperature difference between chip and inlet cooling water, and the heat load vs heat dissipation were also studied. The results show satisfactory cooling performance of the new design in terms of maintaining chip temperature below the maximum value and maintaining chip temperature at high loads. The boiling process in this thermosyphon water-cooled condenser system was found in the nucleate boiling regime. It was also found that there is a reasonable agreement between numerical predictions and experimental data. A newly designed cooling system with 150 mm and 250 mm helical-grooved heat pipes were also fabricated for an experimental study of the operating temperature as well as temperature distribution along the heat pipe axis for varying conditions, and the thermal resistance and overall heat transfer coefficient over a range of temperatures. It was found that the operating temperature range in the helical-grooved heat pipe was between 27 and 40 ºC with the T of to ºC. A 250 mm helical-grooved heat pipe has better performance owing to lower temperature distribution. The highest overall heat transfer coefficient (76.35 W/(m2.K)) and the vi lowest thermal resistance (1.31 K/W) were found for the heat load of 34 W. The heat load of 2.76 W resulted the lowest overall heat transfer coefficient (33.29 W/(m2.K)) and the highest thermal resistance (3.01 K/W). The present 250 mm heat pipe in this study can dissipate a thermal load of up to 39.05 W with the overall heat transfer coefficient of 76.15 W/(m2.K) and the thermal resistance of 1.32 K/W. Furthermore, a new design of a sintered heat pipe was fabricated for experiments to obtain the operating temperature as well as temperature distribution along the heat pipe axis for varying coolant conditions, the thermal resistances over a range of temperatures, and a comparison with previous studies of helical-grooved heat pipe. The results show that the operating temperature in the current sintered heat pipe was 27 - 67 °C for heat load range of 0.79 – 1.52 W. The overall thermal resistance of sintered heat pipe in this study was 6.85 – 25.24 K/W, and the minimum heat load supplied to a sintered heat pipe operating with 25 °C ambient was 1.52 W performing lower thermal resistances in the higher heat loads. In comparison between sintered heat pipe and helical-grooved designs, it was found that the sintered heat pipe has better performance in the lower . range of heat load ( Qa ct < W). Meanwhile, for good performance in helical-grooved . heat pipe, minimum W of Qa ct must be supplied to this heat pipe. The heat load of W was an onset point for both heat pipes to show better performance. Finally, it is noted that the thesis provides guidelines for thermal management engineers to design and fabricate compact cooling systems. vii Preface This thesis presents the study on the thermosyphon flow for cooling of compact devices. The following publications and achievement are based on research carried out for this doctoral thesis: Achievement: Best Paper Award in International Meeting of Advances in Thermofluids 2010. Conference Papers:  Arbiyani, F., Yap, C.R., Ng, K.C., Sintered Heat Pipe Study, International Meeting of Advances in Thermofluids, 2011, accepted.  Arbiyani, F., Yap, C.R., Ng, K.C., Experimental Study on Thermal Performance of Helical-grooved Heat Pipe, International Meeting of Advances in Thermofluids, 2010.  Arbiyani, F., Wijeysundera, N.E., Yap, C.R., Experimental and Numerical Study of a Thermosyphon Cooling System, Regional Conference in Mechanical and Aeronautical Engineering and Inauguration of JSME Indonesian Chapter, 2008. viii List of Tables Table 3.1 Heat flux range for laminar and wavy-transition regime 62 Table 4.1 Typical properties and applications of some copper-base powder metallurgy materials [56] 93 Table 4.2 General specifications and uses of different composition of copper 94 Table 4.3 Typical properties of copper powder by various methods [56] 94 ix Electronic Equipment, pages 417-434. Hemisphere Publishing Corporation, New York. [26] G.S.H. Lock. 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Afgan, editors, Heat and Mass Transfer in Rotating Machinery, pages 633644. Hemisphere, Washington, DC, 1984. [33] W. Nakayama, Y. Othsuka, H. Itoh, and T. Yoshikawa. The Effects of Fine Surface Structures on the Performance of Horizontal Rotating Heat Pipes. Proc. 5th Int. Heat Pipe Conf., Tsukuba, 2:121-125, 1984. [34] J. Schneller, B. Pokorny, and F. Polasek. Heat Transfer in Rotating Co-axial and Parallel Heat Pipes and Their Applications in Machinery. In D.E. Metzger 111 and N.H. Afgan, editors, Heat and Mass Transfer in Rotating Machinery, pages 669-688. Hemisphere, Washington DC, 1984. [35] T. Ma, and Z. Hou. Heat Pipe Research and Developments in China. Heat Recovery Syst. CHP, 9(6):499-512, 1989. [36] A. Pal, Y. Joshi, M.H. Beitelmal, C.D. Patel, and Todd m. Wnger. Design and Performance Evaluation of a Compact Thermosyphon. IEEE, 25(4), December 2002. [37] L. Yuan, Y.K. Joshi, and W. Nakayama. Effect of Condenser Location and Imposed Circulation on the Performance of A Compact Two-Phase Thermosyphon. Taylor & Francis, 2003. [38] C. Ramasamy, Y. Yoshi, and W. Nakayama. Combined Effects of SubCooling and Operating Pressure on the Performance of a Two-Chamber Thermosyphon. IEEE, 1998. [39] R. Gaugler. Heat Transfer Device. U.S. Patent 2350348, 1944. [40] G. Grover. Evaporation-Condensation Heat Transfer Device. U.S. Patent 3229759, Application filed Dec 1963, Approved 18 January, 1966. [41] D.K. Anand. Heat Pipe Application to a Gravity-gradient Satellite (Explorer XXXVI). ASME-Aviation and Space-Progress and Prospects, Annual Aviation and Space Conference, Beverley Hills, Calif., pages 634-638, 16-19 June 1968. [42] N. Ghorbani, H. Taherian, M. Gorji, and H. Mirgolbabaei. Experimental Study of Mixed Convection Heat Transfer in Vertical Helically Coiled Tube Heat Exchangers. Experimental Thermal and Fluid Science, 34:900-905, 2010. 112 [43] P. Zamankhan. Heat Transfer in Counterflow Heat Exchangers with Helical Turbulators. Commun Nonlinear Sci Numer Simulat, 15:2894-2907, 2010. [44] D.A. Pruzan, L.K. Klingensmith, K.E. Torrance, and C.T. Avedisian. Design of High-performance Sintered-wick Heat Pipes. International Journal Heat Mass Transfer, 34(6):1417-1427, 1991. [45] S.C. Wong, J.H. Liou, and C.H. Chang. Evaporation Resistance Measurement with Visualization for Sintered Copper-Powder Evaporator in Operating FlatPlate Heat Pipes. 4th International Microsystems, Packaging, Assembly and Circuits Technology Conference, pages 336-339, 2009. [46] N. Popova, Ch. Schaeffer, Y. Avenas, and G. Kapelski. Fabrication and Thermal Performance of a Thin Flat Heat Pipe with Innovative Sintered Copper Wick Structure. IEEE, pages 791-796, 2006. [47] K. D. Hagen. Heat Transfer: with applications. Upper Saddle River, NJ: Prentice Hall, 1998. [48] W.F. Stoecker and J.W. Jones. Refrigeration & Air Conditioning. McGrawHill Inc., 2nd edition, 1982. [49] R. Gregorig, J. Kern, and K. Turek. Wärme Stoffübertrag. 1974. [50] Rahmatollah Khodabandeh. Pressure Drop in Riser and Evaporator in an Advanced Two-phase Thermosyphon Loop. International Journal of Refrigeration, 28:725-734, 2005. [51] T. Yong, C.Ping, and W. Xiaowu. Experimental Investigation into The Performance of Heat Pipe with Micro Grooves Fabricated by ExtrusionPloughing Process. Energy Conversion Management, 51(10):1849-1854, 2010. 113 [52] G.P. Celata, M. Cumo, and M. Furrer. Experimental Tests of a Stainless Steel Loop Heat Pipe with Flat Evaporator. Experimental Thermal and Fluid Science, 34(7):866-878, 2010. [53] M.A. Chan, C.R. Yap, and K.C. Ng. Pool boiling heat transfer of water on finned surface at near vacuum pressures. Journal of Heat TransferTransactions of the ASME, 132(3), 2010. [54] Hugh W. Coleman and W. Glenn Steele. Experimentation, Validation, and Uncertainty Analysis for Engineers. John Wiley & Sons, Inc., Third edition, 2009. [55] Suk-Joong L. Kang. Sintering: densification, grain growth, and microstructure. Oxford: Elsevier Butterworth-Heinemann, 2005. [56] G.S. Upadhyaya. Sintered Metallic and Ceramic Materials. John Wiley & Sons Ltd., England, 2000. [57] M.M. El-Wakil. Fundamental Concepts of Nuclear Power. American Nuclear Society. [58] I.L. Pioro. Experimental evaluation of constants for the Rohsenow pool boiling correlation. International Journal of Heat and Mass Transfer, 42(11):2003- 2013, 1998. [59] S.G. Kandlikar, M. Shoji, and V.K. Dhir. Handbook of Phase Change: Boiling and Condensation, page 528. Taylor&Francis, USA, 1999. 114 Appendix Basic Characteristics of a Thermosyphon System There are three basic characteristics of heat transfer which can explain the process which occurs in thermosyphon system. There are pressure drop, pool boiling and film condensation characteristics. 1. Pressure drop characteristics Condenser Downcomer Riser Evaporato r Figure A.1: Schematic drawing of thermosyphon system. Assumptions: 1. Steady state. 2. Closed system. 3. Only R-113 vapor enters riser tube (single phase). 4. R-113 vapor is fully condensed, thus only R-113 liquid enters downcomer tube (single phase). 115  Pressure drop between evaporator section (1) and riser inlet (2): Pressure drop between and is a pressure drop due to sudden contraction. According to the revised Bernoulli equation [48]: P1   V12 P2 V22 Ploss      2  V2  with Ploss    1  Cc  (A-1) (A-2) This yields the pressure drop between and 2:  l   2        P2  P1  V  V2 1    1       Cc       (A-3) Pressure drop between riser inlet (2) and elbow inlet (3): Pressure drop between and is a pressure drop of R-113 vapor flowing through a straight duct of circular cross section. The fundamental equation is L V2  P  f D (A-4) Thus the pressure drop between and becomes P2  ( P3  H v g )  f R LR V22 v Di , R   LR V22   P2  P3   v  f R  Hg     Di , R      With f R  (A-5) 64 0.316 if Re < 2300 and f R  when Re > 2300. Re Re 0.25 116  Pressure drop between elbow inlet (3) and elbow outlet/condenser inlet (4): The pressure drop between these two inlets, Ploss  V2 ( geometry factor) . Using the revised Bernoulli equation, the pressure drop between and may be determined:   v (V32  V42 )   V32  v     ( geometry factor)  P4  P3       (A-6) And since the ratio of radius of curvature to diameter of riser is very small (mitered), thus the geometry factor is 1.3 [48]. Therefore, the pressure drop between and is: P4  P3    v 0.3V  V42  (A-7) Pressure drop between condenser inlet (4) and condenser section (5): Pressure drop between and is a pressure drop due to sudden enlargement V42  v  A  1   . Therefore, by implementing the with pressure loss, Ploss  A5   revised Bernoulli equation and pressure loss, the pressure drop between and becomes:  v     A 2  1     V  P5  P4  V        A5       (A-8) 117  Pressure drop between condenser section (5) and downcomer inlet (6): Pressure drop between and is same as pressure drop between and 2, which is a pressure drop due to sudden contraction. Therefore, pressure drop between and is:  l   2        P6  P5  V  V6 1    1       Cc       (A-9) Pressure drop between downcomer inlet (6) and downcomer outlet (7): Pressure drop between and is same as pressure drop between and which is a pressure drop of R-113 liquid flowing through a straight duct of circular cross section Therefore, the pressure drop between and is: P6  P7  H l g  f D LD VD2 l Di , D   L V2  P6  P7   l   f D D   Hg    Di , D       (A-10) Pressure drop between downcomer outlet (7) and evaporator section (1): Pressure drop between and is same as pressure drop between and which is a pressure drop due to sudden enlargement. Therefore, the pressure drop between and becomes:  l     A 2  1     V  P1  P7  V        A1        (A-11) 118  The net driving head caused by the different in density between the liquid in downcomer and the vapor/liquid mixture in riser: Pr iser  Pdowncomer ( P2  P3 )  ( P6  P7 )  r2 m f L     fr Lr  d5 d   H g (  l   v )    Di ,r  v Di ,d  l  Thus, mass of refrigerant flowing in thermosyphon system can be determined as: r  m 2.  H g (l  v )  f L f L  8 5r r  d5 d   Di ,r  v Di ,d  l  (A-12) Pool boiling characteristics Pool boiling is defined as the process in which the formation of vapor is due to heat added to the liquid by a surface in contact with or submerged within the liquid [57]. Pool nucleate boiling is assumed as boiling phenomenon in the present study, since in nucleate boiling, the bubbles of vapor are formed within the liquid which this vapor bubble plays a key role in this cooling system. Furthermore, saturated boiling is also assumed in the present study rather than subcooled boiling. This assumption is made as in saturated boiling, the bubbles can detach to liquid surface, which then flow to the riser and then condenser. However, some of experimental data are from subcooled boiling. Therefore, the cooling characteristics are not as dominant in this particular case. 119 Heat transfer correlation of nucleate boiling is defined by Rohsenow’s correlation: a. Nusselt number Nu L  hnb Lc Ja ,  kl C NB Pr m Where Jakob number (Ja) is Ja  (A-13) C p ,l Tchip  Tsa t  Prandtl number (Pr) is Pr  h fg C p ,l  l kl The values of the constants C and m depend on the boiling liquid and the nature of the heating surface. In the present study, the boiling liquid is R-113 and the heating surface is made of copper, thus C = 0.0022 and m = 2.25 [58]. b. Heat transfer rate Q nb  Achip hnb Tchip  Tsa t  (A-14) Equation (A-13) and (A-14) follows that:  Achip kl C p ,l   Qnb    Tchip  Tsa t  3 m  Lc C Prl h fg  (A-15) c. Characteristic length Characteristic length is considered to be proportional with the radius of the bubbles at the time it breaks away from the heating surface. The estimation as follows   t Lc  Rb      l   v  g  (A-16) 120 3. Film condensation characteristics Steady-state laminar film condensation on radial system is assumed as a condensation phenomenon between vapor and outer surface of coolant coil. Figure A.2: Schematic of film condensation in radial system [58]. Assumptions 1. The coolant coil surface is maintained at a constant temperature (Ts). 2. The saturation temperature of the quiescent vapor in which the surface is placed is Tsat. 3.3.1 Fluid mechanics consideration Consider the schematic and fluid element shown. Assumption:  Neglect inertia force  Steady flow  Viscous force and buoyancy force is balance. 121 Hence, the equilibrium equation of the fluid element:    xy   Lccxy l g  PLcc  ( P  P ) Lccy   xy Lccx   xy   y  Lccx  (17) y      For the quiescent vapor: P  x v g For a Newtonian fluid:  xy   l u y Thus, by integrating for the velocity, the solution gives the velocity distribution as: u ( y)  g (  l   v ) sin   y2    y  C l   (A-18) 3.3.1.1 Boundary condition At the coolant coil surface: u(0) = 0, thus C = 0. Hence, the velocity distribution becomes: u ( y)  g (  l   v ) sin   y2    y   l 2  (A-19) 3.3.1.2 Mass flow rate of condensate liquid The liquid flow rate at a section of the film at a half side is given by:  1 side   Lcc  l udy (A-20) Substituting u(y) and then integrate, thus: Lcc g (  l   v )  l sin      side    l 3 And thus,   Lcc 11 23  l (A-21)  l (  l   v ) g sin  13 (A-22) 122 3.3.1.3 Rate of condensation Rate of condensation for half side is defined as: c  m 1 x , where x = R (A-23) 3.3.1.4 Film condensation characteristic flow The characteristic velocity to denote the flow of condensate film is: um   A 4(  Lcc ) and Dh  , thus the Reynolds number is  (  l  Lcc ) p given by Re   u m Dh  l l  4 l (A-24) Therefore, the flow characteristic of film condensation is as follow: 3.3.2  Re < 30 is laminar flow  30 < Re < 180 is wavy flow  Re > 1800 is turbulent flow Heat transfer consideration Assumptions:  Velocity is small, so that convective energy transfer can be neglected  Temperature gradient in the direction is also small Energy balance for the small fluid element:  q  qdx  q  dy dx y   (A-25) Hence, heat flux at the coolant coil surface is: 123 k  Q s   l (Ts  Tsa t ) p2   (A-26) Heat transfer coefficient is defined as: Q s  h c p2 (Ts  Tsa t ) (A-27) From (27) and (28): k  hc   l    (A-28) is the expression for the laminar condensation heat transfer coefficient. The energy balance:  c dx  (  d)h f  Q s dx h f  hg m (A-29) Using relation (23) and from (26): 1  g l (  l   v ) kl (Tsa t  Ts ) R   1.1438   h fg l   (A-30) 3.3.2.1 Average heat transfer coefficient and Nusselt number Hence, for a tube where the condensate flows on both sides, and representing the heat transfer rate by the heat transfer coefficient: Q s  21 h fg  2Rhc (Tsa t  Ts ) , thus by substituting (30), the average film condensation heat transfer coefficient is:  g l (  l   v ) kl h fg  h c  0.728     l (Tsa t  Ts ) D  (A-31) Therefore, the average Nusselt number for a tube of diameter 2R becomes: 124  g (    v ) h fg D h c (2 R) Nu   0.728  l l kl   l kl (Tsa t  Ts )    (A-32) And heat transfer rate:  g l (  l   v ) kl h fg   Qs  0.728     l (Tsa t  Ts ) D  p (Ts  Tsa t ) (A-33) 3.3.2.2 Modified latent heat The work by Rohsenow showed from experiments that the modified latent heat for condensation is given by: h"fg  h fg 1  0.68 Ja  (A-34) Thus, the latent heat which is used in equation (A-31), (A-32) and (A33) should be as modified latent heat, hence average heat transfer coefficient, Nusselt number and heat transfer rate becomes:  g l (  l   v ) kl h "fg  h c  0.728     l (Tsa t  Ts ) D  (A-35)  g (    v ) h fg D h c (2 R) Nu   0.728  l l kl   l kl (Tsa t  Ts ) "  g (    v ) kl h 'fg  Q s  0.728  l l    l (Tsa t  Ts ) D     (A-36) p (Ts  Tsa t ) (A-37) When using the above expression, all liquid properties are evaluated at the mean temperature between coolant coil surface and the saturated vapor ( T f ). Meanwhile, the vapor properties are evaluated at saturation vapor temperature (Tsat). 125 3.3.2.3 Heat transfer at coolant coil Heat rate of water flowing in coolant coil is also meant heat dissipation.  wC p,wdTw dQ d  dQ w  m Two  wC p , w  dT Q d  Q w  m Twi  wC p,w (Two  Twi ) Q d  Q w  m (A-38) Nusselt number for water coolant inside the tube is given by: Nu D  0.023 Re 0D.8 Pr 0.4 , Furthermore, Nu  h w Di , thus heat transfer coefficient for water kw coolant flow inside the tube is: hw  kw (0.023Re 0D.8 Pr 0.4 ) Di (A-39) 3.3.2.4 Overall heat transfer coefficient Heat transfer coefficient from R-113 vapor to water coolant is determined by overall heat transfer coefficient through the following equation: U Di  hw  hc ( Do  h c )  ( Di  hw ) (A-40) Furthermore heat transfer from coolant coil surface to water coolant is determined by: Q w  p1hw (Ts  Tw ) , hence (A-41) It is necessary to define Tw which is mean temperature of water coolant (Tm), thus the heat transfer becomes: Q w  p1hw (Ts  Tm ) (A-42) 126 [...]... Effect of inlet water temperature on Nusselt number of film condensation at constant mw of 0.0067 kg/s 58 Figure 3.37: Effect of mass flow rate of inlet water on Nusselt number of film condensation at constant Twi of 10 °C 59 Figure 3.38: Effect of mass flow rate of inlet water on Nusselt number of film condensation at constant Twi of 15 °C 60 Figure 3.39: Effect of mass flow. .. Reynold number of film condensation at constant of mw 0.0067 kg/s 63 Figure 3.43: Effect of mass flow rate of inlet water temperature on Reynold number of film condensation at constant Twi of 10 °C 64 Figure 3.44: Effect of mass flow rate of inlet water temperature on Reynold number of film condensation at constant Twi of 15 °C 64 Figure 3.45: Effect of mass flow rate of inlet water... systems into commercialization 1.3 Scope of study This study focused on two-phase behavior in compact devices The size of compact devices was limited to the dimensions of current commercial laptop Therefore, the results of this study may be accurately compared to the actual performance of existing compact cooling systems Furthermore, in the fundamental research of the present study, film condensation... Although a separate study of cooling system has been carried out to achieve a passive, compact and free-orientation cooling system, an integrated system of passive, compact and free-orientation cooling system has not been achieved Studies of pressure drop have been well established in two-phase flow behavior, but not that on compact cooling systems Studies of condensation in several cooling systems have... 3.11: Effect of Twi on cooling capacity at constant mw of 0.0033 kg/s 37 Figure 3.12: Effect of Twi on cooling capacity at constant mw of 0.005 kg/s 37 Figure 3.13: Effect of Twi on cooling capacity at constant mw of 0.0067 kg/s 38 Figure 3.14: Effect of heat flux on chip temperature at constant Twi of 10 °C 40 Figure 3.15: Effect of heat flux on chip temperature at constant Twi of 15 °C ... However, the mode of condensation in two-phase cooling systems to achieve high rate of condensation in compact devices has not been explored The main objective of this research is therefore to address the two-phase cooling system issues related to various technological problems such as limitation of maximum allowable of chip temperature, design and fabrication of integrated twophase cooling system, pressure... Schematic drawing of thermosyphon system 28 Figure 3.6: Schematic of film condensation in radial system [58] 29 Figure 3.7: Solver algorithm 33 Figure 3.8: Effect of mw on cooling capacity at constant Twi of 10 °C 35 T Figure 3.9: Effect of mw on cooling capacity at constant wi of 15 °C 36 T Figure 3.10: Effect of mw on cooling capacity at constant wi of 20 °C 36 Figure... development of cooling systems, starting from single-phase cooling systems and followed by two-phase cooling systems In the second part, the published thermosyphon studies which motivate the enhancement of current thermosyphon systems are provided In the third part, the study of the enhancement of heat pipes will be further discussed 2.1 Cooling systems This section discusses two key aspects in the study of cooling. .. at constant Twi of 15 °C 56 Figure 3.33: Effect of mass flow rate of water on film condensation rate at constant Twi of 20 °C 56 Figure 3.34: Effect of inlet water temperature on Nusselt number of film condensation at constant mw of 0.0033 kg/s 57 xi Figure 3.35: Effect of inlet water temperature on Nusselt number of film condensation at constant mw of 0.005 kg/s... Helical-grooved heat pipe Sintered heat pipe o Characteristics of thermosyphon water-cooled condenser system Effect of mass flow rate of coolant water ( mw ) and inlet water temperature ( Twi ) on cooling capacity ( Qd ) 4 Effect of mass flow rate of coolant water ( mw ) and inlet water temperature ( Twi ) on chip temperature ( Tchip ) Typical boiling curve of thermosyphon water-cooled condenser system Correlation . THERMOSYPHON FLOW FOR COOLING OF COMPACT DEVICES FILIAN ARBIYANI (B.Eng (Hons), Gadjah Mada University-Indonesia) A THESIS SUBMITTED FOR THE DEGREE OF DOCTOR OF PHILOSOPHY. Effect of mass flow rate of inlet water on Nusselt number of film condensation at constant wi T of 15 °C. 60 Figure 3.39: Effect of mass flow rate of inlet water on Nusselt number of film. Effect of wi T on cooling capacity at constant . w m of 0.0033 kg/s. 37 Figure 3.12: Effect of wi T on cooling capacity at constant . w m of 0.005 kg/s. 37 Figure 3.13: Effect of wi T on cooling