Numerical Analysis of a Miniature-Scale Refrigeration System (MSRS) for Electronics Cooling

Transcription

1 Purdue University Purdue e-pubs International Refrigeration and Air Conditioning Conference Scool of Mecanical Engineering 2004 Numerical Analysis of a Miniature-Scale Refrigeration System (MSRS) for Electronics Cooling Suwat Trutassanawin Purdue University Eckard A. Groll Purdue University Folw tis and additional works at: ttp://docs.lib.purdue.edu/iracc Trutassanawin, Suwat and Groll, Eckard A., "Numerical Analysis of a Miniature-Scale Refrigeration System (MSRS) for Electronics Cooling" (2004). International Refrigeration and Air Conditioning Conference. Paper 679. ttp://docs.lib.purdue.edu/iracc/679 Tis document as been made available troug Purdue e-pubs, a service of te Purdue University Libraries. Please contact epubs@purdue.edu for additional information. Complete proceedings may be acquired in print and on CD-ROM directly from te Ray W. Herrick Laboratories at ttps://engineering.purdue.edu/ Herrick/Events/orderlit.tml

2 1 NUMERICAL ANALYSIS OF A MINIATURE-SCALE REFRIGERATION SYSTEM (MSRS) FOR ELECTRONICS COOLING Suwat Trutassanawin 1 and Eckard A. Groll Purdue University Scool of Mecanical Engineering Ray W. Herrick Laboratories West Lafayette, Indiana, USA 1 Corresponding Autor: (765) ; strutas1@purdue.edu ABSTRACT Tis paper presents te numerical analysis of a Miniature-Scale Refrigeration System (MSRS) for electronics cooling. Te system consists of a simulated electronic cip attaced to a microcannel cold plate evaporator, a compressor, a microcannel condenser, and an expansion device. Te system uses R-134a as te refrigerant. A copper bck eater is designed to simulate te eat generation of an electronic cip by using two cartridge eaters of 200 W eac. Te eat from te simulated CPU is transferred to te cold plate evaporator via a copper eat spreader. Te eat spreader is empyed to provide uniform eat dissipation from te copper bck to te evaporator. In order to analyze te system performance and determine te operating conditions, a numerical simulation of te MSRS was conducted. In addition, te Fluent -software was empyed to analyze te eat spreader. Experimental results tat were obtained using a bread board MSRS were used to validate te model results. 1. INTRODUCTION Te advance of te computer tecnogy as lead to iger processor speeds, smaller sizes, and increased power consumption. Tus, te eat dissipation from te cip becomes a critical issue in te design of ig performance semiconductor processors. Te conventional air cooling metods using eat sinks are expected to not be able to maintain acceptable cip core temperatures in te near future. Terefore, alternative cooling approaces, suc as eat pipes, termoelectric cooling, liquid cooling, spray cooling, jet impingement cooling, and refrigeration cooling, are currently being investigated to accommodate te envisioned ig eat dissipation applications. Te study presented ere focuses on refrigeration cooling tecniques. Te advantages of te refrigeration cooling tecnique are: (1) to maintain a w junction temperature and at te same time dissipate ig eat fluxes; (2) to increase te device speed due to a wer operating temperature; and (3) to increase te device reliability and life cycle time because of a wer and constant operating temperature. Te disadvantages of te refrigeration cooling tecnique are: (4) an increased complexity and cost; (5) te need for additional space to fit te components of te refrigeration system; and (6) a decrease of te system reliability as a result of an additional moving component, i.e., te compressor. 2. MINIATURE-SCALE REFRIGERATION SYSTEMS (MSRS) FOR ELECTRONICS COOLING A scematic diagram of a Miniature-Scale Refrigeration System for electronics cooling is illustrated in Figure 1. Te MSRS is composed of six main components: a cold plate microcannel evaporator, a compressor, a microcannel condenser, two expansion devices in parallel (a needle valve and a capillary tube), a eat spreader, as well as a eat source or eater bck, wic simulates te CPU. Based on te components indicated in Figure 1, a bread board MSRS was designed and constructed using a commercially available small-scale ermetic rotary compressor. Te R-134a compressor is driven by a DC motor and as a cooling capacity of 75 to 140 W, a COP of 1.13 to 1.35, and a maximum power consumption of 103 W at typical otel mini-bar refrigerator operating conditions. Te eat source consists of a copper bck wit dimension of mm 3. Two cartridge eaters wit a maximum eat dissipation power of 400 W are inserted into te base bck tat is bew te copper

3 2 bck. Te base bck as dimensions of mm 3. A variable AC transformer is used to adjust te power input to te eaters. A eat spreader is empyed in order to dissipate te eat from te eat source to te eat sink. Te eat spreader is made of copper wit te dimensions of mm 3. A termal conductive paste was applied between te copper bck-eat spreader mating surfaces and te eat spreader-eat sink mating surfaces to reduce te termal resistance. Te termal conductive paste as a termal resistance of less tan C-in 2 /W for a inc layer. Te evaporator is a microcannel eat excanger consisting of 41 rectangular cannels. Eac cannel as a cross section area of mm 2. Te microcannel condenser as a eat dissipation capacity of 225 W and dimensions of mm 3. Te expansion devices are composed of a capillary tube wit OD and ID, and a and operated needle valve. Te system was carged wit 100 g of R-134a. A potograp of te entire bread board MSRS is sown in Figure 2. Te target operating conditions of te MSRS are as folws: cooling capacity of 200 W; evaporating temperature range from 10 to 20 C; supereat of te refrigerant at te compressor inlet in te range of 3 to 8 C; condensing temperature in te range of 40 to 60 C; subcooling temperature of te refrigerant at te condenser outlet in te range of 3 to 10 C; and ambient air temperature in te range of 30 to 50 C. Figure 1: Scematic of bread board MSRS. Figure 2: Potograp of bread board MSRS. 3. REFRIGERATION SYSTEM MODEL A refrigeration system simulation model was written in MATLAB. Te model is divided into four parts: an evaporator model, a compressor model, a condenser model, an expansion device model. Te main assumptions of te system model are negligible pressure drop in te evaporator and condenser, as well as negligible eat ss from te connecting pipes. Te system model begins its calculations at te inlet to te compressor wit tree inputs: suction pressure, discarge pressure, and supereat temperature. Te outputs of compressors model are refrigerant mass fw rate, compressor power input, and compressor outlet temperature. In te condenser model, te inlet conditions are assumed equal to te outlet of te compressor. To simplify te condenser model, a lump capacitance metod was empyed and te total eat rejection of te condenser was set equal to te sum of te cooling capacity of te evaporator and te power consumption of te compressor subtracting te eat ss of te compressor. Te condenser outlet pressure and a guess outlet temperature of te condenser are te inputs to te expansion device model. Te expansion process is assumed isentalpic and te outlet pressure equals te compressor inlet pressure. Te refrigerant quality is te output of te expansion device model. Te evaporator model is divided into small segments wit known inlet quality and constant pressure. Te outlet state of eac segment is calculated by assuming a constant eat flux to all segments. Until te segment outlet quality reaces a state of saturated vapor, a omogeneous two-pase analysis is used on eac segment. Afterwards, a single pase analysis is empyed. Te analysis marces troug te evaporator until te total lengt of all segments is equal to te evaporator lengt. Te outputs of te evaporator model are cal and average refrigerant eat transfer coefficients, as well as te evaporator outlet temperature. In te next step, te condenser outlet temperature is determined from te condenser model and is compared to te guess expansion device inlet temperature. If bot values are not equal, ten te condenser outlet

4 3 temperature is updated. Te calculations are repeated until bot values are equal. A fw cart of te system model is presented in Figure 3. After system convergence is acieved, te coefficient of performance (COP) and all oter operating conditions are calculated. Te properties of te refrigerant R-134a are calculated using te Reference Fluid Termodynamic and Transport Properties software, REFPROP 7.0, by NIST. 3.1 Compressor model Te folwing assumptions are made witin te compressor model: neglecting canges in kinetic and potential energy, te compressor operates at steady-state condition, and pressure sses in te suction and discarge lines are negligible. A ermetic compressor modeling approac is empyed wit compressor speed, swept volume, compressor inlet and outlet pressures, as well as supereat temperature as input parameters. Te volumetric, isentropic, and mecanical compressor efficiencies are assumed equal to 67, 80, and 45 %, respectively. Te motor efficiency is 80 % and te compressor eat ss factor is 10%. Te refrigerant mass fw rate is computed from known volumetric efficiency, compressor speed, and swept volume by te folwing equation: actual actual η vol (1) teoritical N( Vswept / vin ) Te overall compressor efficiency is te product of te isentropic, mecanical, and motor efficiencies: η η η η (2) o isen mec motor Te isentropic efficiency is te ratio of isentropic compression work to te actual compression work: isen dis, isen suct η isen (3) refrig dis suct Te mecanical efficiency is related to te actual refrigerant compression work and te saft work as folws: Te total compressor power input is: refrig η mec (4) saft m ( ) isen suct comp, e (5) η o Te eat ss from te compressor sell is determined by equation (6): ( η mecη motor ) Wcomp e, 1 (6) Q ss sell fq, In addition, eat transfer occurs from te discarge line back to suction refrigerant. Te eat transfer is calculated by: ( 1 fq )( ηmec ) comp e Q dis _ to_ suct 1, (7) Tus, te suction, discarge, and outlet entalpies of te refrigerant are computed as folws: suct dis outlet Q dis _ to_ suct inlet + (8) Q comp, e ss, sell suct + (9) Q dis _ to _ suct dis (10)

5 4 Te outputs of te compressor model are te refrigerant mass fw rate, compressor outlet temperature, compressor eat ss, and compressor power input. 3.2 Condenser model Te condenser model is used by assuming lump capacitance metod and neglecting pressure drop in te condenser. Te total eat rejection rate equals te sum of te cooling capacity of evaporator and te compression power of refrigerant. Q Q + (11) cond ad Te condenser outlet entalpy is obtained by: cond comp Q cond, o cond, i (12) 3.3 Expansion device model Te expansion model requires te refrigerant inlet and outlet pressures as well as an estimated expansion inlet entalpy or temperature. An isentalpic (constant entalpy) process is assumed for te expansion device: exp, o exp, i (13) Te outlet quality of te refrigerant is computed by: o f x exp, (14) fg Te outlet states of expansion device are assumed to be te inlet states of te evaporator model. 3.4 Evaporator model Te folwing assumptions are made witin te evaporator model: neglecting canges in kinetic and potential energy, steady-state operating condition, pressure drop in te cold plate is negligible, and tere is no eat ss to te surroundings, i.e., te entire eat dissipated by te CPU is transferred to te cold plate evaporator. Te evaporator model is divided into two regions: two-pase region and single pase or supereat region. Two-pase region: Te two-pase eat transfer coefficient is calculated using te Kandlikar (1999) correlations [2]: Were TP, NBD and TP, CBD TP,NBD TP larger of (15) TP,CBD are te two-pase eat transfer coefficients in te nucleate boiling dominant and convective boiling dominant regions, as expressed by te folwing equations: ( Co Fr Bo F fl )(1 x TP, NBD ) (16) ( 1.136Co Fr Bo F fl )(1 x TP, CBD ) Were, te convection number (Co) and boiling number (Bo) are defined as: ρ Co ρ f g 0.5 ( 1 x) x Te fluid-surface parameter for R-134a is F fl (17) q Bo (18) Te Froude number for te wole fw as liquid, Fr, is a parameter for te stratified fw region. For te given application, te Froude number was set equal to 1, since tere is no stratified fw in microcannel eat excangers. G fg

6 5 Te single pase eat transfer coefficient te wole fw as liquid is computed using te Gnielinski correlation [2]: For 10 Re 5 10 ; For 2300 Re 10 ; Nu Re Pr f l 2 3 ( Pr ) 2 / l 1 f Nu ( Re 1000) ( Pr ) 2 / l 1 Pr f l 2 f and for Re 2300 Were te friction factor is calculated by, Nu 4.36 (19) [ 1.58 ln ( Re ) 3.28] 2 f (20) Single pase or supereated vapor region: Te single pase eat transfer coefficient of te supereated vapor is obtained from Gnielinski correlation [3]: 6 For 2300 Re g 5 10 and 0.5 Prg 2000; and for Re 2300 Were te friction factor is given by, ( Re 1000) f 8 g Prg Nu (21) / f 8 ( Prg 1) Nu 4.36 (22) [ 0.790ln( Re ) 1.64] 2 f (23) g Te average eat transfer for te wole evaporator is calculated by: L LTP L 1 SP TPdz+ Ltot z 0 z 0 SP dz (24) Were L L + L tot TP SP Nu k TP l TP D (25) Nu SPk g SP D (26) 4. NUMERCAL ANALYSIS OF THE HEAT SPREADER Te computational fluid dynamics software package Fluent 6.0 and Gambit 2.0 were used to conduct a numerical analysis of te eat spreader. Te 3D double precision option was used wit te SIMPLE algoritm to solve te continuity equation. Te QUICK algoritm was used as te converging criteria for bot te momentum and energy equations. Te Gambit 2.0 software was used to generate te mes and Fluent was used to solve te problem numerically. Te surface area of te eat spreader is mm 2. Te size of te CPU eat source was mm 2. Te CPU was cated at te bottom-center of te eat spreader. To simplify te analysis, te contact resistance between te eat spreader and te CPU was neglected. In addition, any eat sses from te sides

7 6 and bottom of te eat spreader and CPU were neglected. Te eat from te top surface of te CPU dissipates to te cold plate evaporator via te eat spreader. Six steady state cases were run wit different eat spreader tickness of 2.5, 5.0, 7.5, 10.0, 12.5, and 15.0 mm. For eac tickness, CPU eat fluxes of 55, 62, and 69 W/cm 2 were analyzed Wcomp, e [W] Tcond 39.4 C Tcond 46.3 C Tcond 52.4 C Tevap [C] Figure 4: Compressor power input versus evaporating and condensing temperatures COP [ - ] Tcond 39.4 C Tcond 46.3 C Tcond 52.4 C Tevap [C] Figure 5: COP versus evaporating and condensing temperatures m refrig [g/s] Figure 3: Refrigeration system model Tcond 39.4 C Tcond 46.3 C Tcond 52.4 C Tevap [C] Figure 6: Refrigerant mass fw rate versus evaporating and condensing temperature. 5.1 System Simulation Model Validation 5. NUMERCAL RESULTS Te results of te system simulation model were compared to te experimental data as depicted in Table 1. Te model predicts te measured system performance reasonable well. However, tere is a significant error in te

8 7 prediction of te eat rejection rate of te condenser due to te use of te lump capacitance model in te condenser. In addition, te experimental bread board system experienced significant eat ss from te connecting piping. Table 1: Comparison of experimental results and modeling results Tcomp, i Pcomp, i Pcomp, o Qevap _r Tevap Error DTsup Error Tcond Error Test [ C] [kpa] [kpa] [W] [W/m 2 -K] [ C] [%] [ C] [%] [ C] [%] Input of Refrigeration Model Model Exp Model Exp Model Exp Model Tcomp,o Error mrefrig Error Qcond Error Wcomp, e Error COPsystem Error Test [ C] [%] [g/s] [%] [W] [%] [W] [%] [ - ] [%] Exp Model Exp Model Exp Model Exp Model Exp Model System Performance Te system performance was predicted using te system simulation model for evaporating temperature ranging from 10 to 20 C and condensing temperature ranging from 40 to 60 C. Figure 4 sows te compressor power input as a function of te evaporating and condensing temperatures. For a fixed condensing temperature, te compressor power input decreases as te evaporating temperature increases due to te decrease of te pressure ratio. For a fixed evaporating temperature, te compressor power input increases as te condensing temperature increases due to te increase of te pressure ratio. Te COP and refrigerant mass fw rates as a function of te evaporating and condensing temperatures are depicted in Figure 5 and Figure 6, respectively. For a fixed cooling capacity of 200 W, te COP increases wit increasing evaporating temperature or decreasing condensing temperature. For a fixed condensing temperature, te mass fw rate increases as te evaporating temperature increases. For a fixed evaporating temperature, te mass fw rate decreases as te condensing temperature increases. Te numerical values of te system simulation results are sown in Table 2. Table 2: System performance predictions by te system model Test T comp, i P comp, i P comp, o Q evap _r T evap DT sup T cond m refrig Q cond W comp, elec COP system [ C] [kpa] [kpa] [W] [W/m 2 -K] [ C] [ C] [ C] [g/s] [W] [W] [ - ]

9 8 5.3 Heat Spreader Design Analysis Te temperature profiles of te eat spreader for a eat dissipation of 55 W/cm 2 and an average R-134a eat transfer coefficient of W/m 2 -K for a variety of eat spreader tickness are depicted in Figure 7. Figure 8 presents te temperature profile at te vertical-center of te eat spreader for a tickness of 7.5 mm and a eat dissipation of 55 W/cm 2 at two cations, i.e., at te top surface of te CPU and at te top surface of te eat spreader. Te temperature profile of te eat spreader wit a tickness of 7.5 mm for various eat dissipations is sown in Figure 9. It can be seen from Figure 7 (c) and Figure 9 (a) to (b) tat te maximum temperature increases as te eat dissipation increases from 55 to 62, and ten to 69 W/cm 2 for te same eat spreader tickness of 7.5 mm, due to te additional eat flux. Te minimum and maximum temperatures of te eat spreader for a eat dissipation of 55 W/cm 2 as a function of different eat spreader ticknesses and average refrigerant eat transfer coefficients are presented in Table 3. Te eat spreader tickness of 7.5 mm is te optimum tickness, since tere is a more uniform temperature across te tickness of te eat spreader. (a) Heat spreader tickness of 2.5 mm. (b) Heat spreader tickness of 5.0 mm. (c) Heat spreader tickness of 7.5 mm. (d) Heat spreader tickness of 10.0 mm. Figure 7: Temperature profile of te eat spreader for eat dissipation of 55 W/cm 2 and average eat transfer coefficient of R-134a of W/m 2 -K. (a) Top surface of CPU. (b) Top surface of eat spreader. Figure 8: Temperature profile at te centerline of te eat spreader for a tickness of 7.5 mm and a eat dissipation of 55 W/cm 2.

10 9 (a) Heat dissipation of 62 W/cm 2. (b) Heat dissipation of 69 W/cm 2. Figure 9: Temperature profile of eat spreader wit tickness of 7.5 mm for various eat dissipations. Table 3: Temperature range at top surface of eat spreader for various conditions. Tickness [mm] Average R-134a eat transfer coefficient [W/m 2 -K] Minimum temperature [K] Maximum temperature [K] Tickness [mm] Average R-134a eat transfer coefficient [W/m 2 -K] Minimum temperature [K] Maximum temperature [K] CONCLUSIONS A numerical simulation of a miniature-scale vapor compression refrigeration system was deveped to predict its system performance for electronics cooling. In addition, te Fluent software was empyed to simulate te temperature profile in te eat spreader connecting te eat source CPU to te cold plate evaporator. Te results of te system simulation model sow reasonable agreement wit nine experimental test results taken wit a bread board system. Te model will be used in future studies to improve te design of te experimental setup. In addition, te analysis of te eat spreader indicated tat an optimum tickness of te eat spreader can be found for te given operating conditions. NOMENCLATURE Bo Boiling number ( - ) Co Convection number ( - ) D Diameter (m) F fl Fluid-surface parameter ( - ) Fr Froude number ( - ) f Q Heat ss coefficient of compressor ( - ) G Mass flux (kg/s-m 2 ) Entalpy (J/kg-K)

11 10 L Lengt (m) ṁ Refrigerant mass fw rate (kg/s) N Compressor speed (RPM) Nu Nusselt number ( - ) Pr Prandtl number ( - ) Q Heat ss or eat transfer rate (W) q Heat flux (W/m 2 ) Re Reynolds number ( - ) Power consumption or power input (W) x Quality ( - ) Greek symbol η Efficiency (%) ρ Density (kg/m 3 ) Subscript CBD comp cond Convective boiling dominant Compressor Condenser dis Discarge e Electricity evap Evaporator exp Expansion f Fluid or liquid or friction factor g Gas or vapor i, in Inlet isen Isentropic Liquid only mec Mecanical motor Motor NBD Nucleate boiling dominant o Overall refrig Refrigerant SP Single pase suct Suction TP Two-pase vol Volumetric REFERENCES [1] Pelan, P.E., 2001, Current and Future Miniature Refrigeration Cooling Tecnogies for Hig Power Microelectronics, Semiconductor Termal Measurement and Management Symposium: Seventeent Annual IEEE, San Jose, CA USA, Marc 20-22, 2001, pp [2] Kandlikar, S.G., 1999, Fw Boiling Heat Transfer Coefficient in Microcannels-Correlation and Trends, ttp:// [3] Incropera, F.P. and Dewitt, D.P., 2002, Fundamentals of Heat and Mass Transfer, 5 t edition, Jon Wiley &Sons, Inc. [4] Braun, E.J., 2001, Lecture notes of ME 518: Analysis of Energy Utilization Systems, Evaporator and Condenser analysis. [5] Mudawar, I., 2002, Lecture notes of ME 506: Two-pase Fw and Heat Transfer, Homogeneous two-pase fw model. ACKNOWLEDGEMENT Te financial support of te Cooling Tecnogies Researc Center (CTRC) at Purdue University is greatly appreciated as well as all discussions and suggestions from Prof. Sures V. Garimella, Prof. Jayati Y. Murty, and te members of te CTRC. Valuable contributions of te Ray W. Herrick Laboratories sop tecnicians in te design and construction of te experimental bread board system are also acknowledged.