Cooling of microprocessors with micro-evaporation: A novel two-phase cooling cycle

Jackson Braz Marcinichen ^a,*, John Richard Thome ^a, Bruno Michel ^b

^aLaboratory of Heat and Mass Transfer (LTCM), Faculty of Engineering (STI), E´cole Polytechnique Fe´de´rale de Lausanne (EPFL), Station 9, CH-1015 Lausanne, Switzerland

b IBM Research GmbH, Zurich Research Laboratory, Sa¨umerstrasse 4, CH-8803 Ru¨schlikon, Switzerland Dedicated to Professor Dr.-Ing. Dr.h.c.mult. Karl Stephan on the occasion of his 80th birthday.

a r t i c l e i n f o                                                   a b s t r a c t

Article history: Received 4 March 2010 Received in revised form 3 June 2010 Accepted 5 June 2010 Available online 12 June 2010 Keywords: Cooling Component Electronic Microprocessor-design Comparison Cooling circuit COP Heat recovery

Three micro-evaporator cooling cycles, one with a pump, one with a compressor and a hybrid of the two together, are proposed for cooling a computer blade server. The hybrid cycle is characterized by the interchangeability between the first two cycles, where the decision on the cycle to operate is based on the season (necessity or economical benefit for heat recovery) or the maintenance of cycle’s driver. The main characteristics of each cycle are presented as well as the details of the micro-evaporator cooler for the blade’s CPU. Analysis of the cycle overall efficiency and the potential for heat recovery shows that the best cycle to use depends mainly on the end application of the heat recovered. Four refrigerants were evaluated as the possible working fluids for cooling the microprocessors. HFC134a and HFC245fa were found to be the best choices for the desired application. ª 2010 Elsevier Ltd and IIR. All rights reserved.

Refroidissement des microprocesseurs a` l’aide de la microe´vaporation : un cycle de refroidissement diphasique et innovant

Mots cle´s : Refroidissement ; Composant ; E´lectronique ; Microprocesseur-conception ; Comparaison ; Circuit frigorifique ; COP ; Re´cupe´ration de chaleur

* Corresponding author. Tel.: þ41 21 693 5894; fax: þ41 21 693 5960. E-mail address: jackson.marcinichen@epfl.ch (J.B. Marcinichen).

0140-7007/$ e see front matter ª 2010 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2010.06.008

1.             Introduction

Cooling of data centers causes estimated annual electricity bills of 1.4 billion dollars in the United States and of 3.6 billion dollars world wide (Koomey, 2007). Currently, the most widely used cooling technology is refrigerated air cooling of the data centers’ numerous servers. According to recent articles published at ASHRAE Winter Annual Meeting at Dallas (January, 2007) typically 40% or more of the refrigerated air flow bypasses the server racks in data centers. Furthermore, servers that are turned off or on standby are cooled as if they were operating, wasting a significant amount of the energy for the unnecessary flow. This poor energetic performance in one of industries leading technological sectors is quite startling and motivates the search for a green thermal solution for future generations of higher performance servers that consume much less energy to operate and cool while they also provide the possibility to recover a large quantity of waste heat. This is the topic of research addressed here.

Current chip cooling technology consists of conducting the microprocessor’sJouleheatingawaythroughthesiliconchipdie itself,thenacross a thermal interface material (TIM) toa copper or aluminum heat spreader/finned cooling element and finally byconvectiontorefrigeratedairenteringat10e15C.Lookingat thisonamaterialbasis,themicroprocessorcircuitryhasamass of about 5 mg, the silicon die about 5 g and the metallic cooling element about 0.5 kg, representing about five orders of magnitudeintheratioofthematerialsinvolvedandthuspointsoutthe huge opportunity to improve this whole process.

Nomenclature Roman CHF COP GWP mr mw ODP Psuc Pdisc Q Sub Tdisc Tevap_inlet Tevap_outlet Tw_inlet Tw_outlet Wcompressor WCond_pump Wpump

critical heat flux [W cm2] coefficient of performance [-] global warming potential [-] refrigerant mass flow rate [kgh1] water mass flow rate [kg h1] ozone depleting potential [-] suction pressure [bar] discharge pressure [bar] cooling capacity or power generated by electronic components [W] subcooling [K] discharge temperature [C] inlet evaporating temperature [C] outlet evaporating temperature [C] inlet water temperature [C] outlet water temperature [C] compressor power [W] pumping power of water in the condenser [W] liquid pump power [W]

WSubcoole wv xoutlet Greek r Dhcomp DP DPME DhME hcycle hcycle_LP hcycle_VC Subscripts comp Cond disc evap ME suc w

r_pump pumping power of water in the subcondenser [W] volumic refrigerating effect [kJ m3] outlet vapor quality [%] specific mass [kg m3] compressor enthalpy difference [kJ kg1] pressureincreaseprovided by the liquid pump[Pa] micro-evaporator pressure drop [bar] specific cooling capacity [kJ kg1] cycle overall efficiency [-] liquid pump cycle overall efficiency [-] vapor compression cycle overall efficiency [-] compressor condenser discharge evaporating micro-evaporator suction water

Thermal designers of data centers and server manufacturers now agree about the long term need to improve the cooling process by implementing liquid or two-phase cooling directly in the server itself, eliminating the poor thermal performing air as a coolant all together (Greenberg et al., 2006; Hannemann and Chu, 2007; Samadiani et al., 2008). That said, there is a clear need for a detailed design and evaluation of these new cooling strategies in order to arrive at an improved solution. They should provide more efficient heat transfer from the chips, memories, etc. using water-cooled or boilingcooled elements, eliminating air as a means of heat transfer, while also reducing energy consumption for driving the cooling system by a significant amount. Some examples of design and evaluation of these new cooling strategies can be found in Scott (1974), Bash (2001), Peeples (2001), Heydari (2002), Schmidt and Notohardjono (2002), Maveety et al. (2002), Phelan and Swanson (2004) and Suman et al. (2004). Additionally, since data centers often dissipate on the order of

5e15 MW of heat, this makes heat recovery an important

Data center with 64 Blades (Refrigerated air cooling vs. Twophase on-chip cooling system)

Fig. 1 e Comparative power supply for a data center containing 64 blades for air cooled and two-phase on-chip cooling with and without heat recovery.

energetic and environmental issue to consider and will greatly reduce the CO₂ footprint of the data center.

Fig. 1 shows the comparative energy consumption required by a data center with 64 blades (325 W per blade) when using traditional air cooling, two-phase on-chip cooling and twophase on-chip cooling when the energy is recovered, using a vapor compression cycle. For air cooling, it is assumed that the power required to cool the data center is the same as that required to run the information technology equipment (Koomey, 2007; Ishimine et al., 2009). This is plotted as a function of the compressor overall efficiency. It is seen that, if no heatwasrecovered,thecostofcoolingthedatacenterwouldbe approximately 59% that of traditional air cooling when operatingatacompressorwithanoverallefficiencyof60%,whichis typical of light commercial systems. However, if the heat was toberecoveredandconsidering60%ofrecoveryefficiency,this value drops down to about 24% that of traditional air cooling. These results show that the cost of cooling could be drastically decreased when using on-chip cooling, representing a huge potential for the next generation of data centers cooling systems.

Recent publications show the development of primarily four competing technologies for cooling chips (all with their own pros and cons): microchannel single-phase flow, porous media flow, jet impingement cooling and microchannel twophase flow (Agostini et al., 2007). Single-phase liquid cooling is now fairly well known and can be used to remove high heat fluxes. Leonard and Phillips (2005) showed that the use of such new technology for cooling of chips could produce savings in energy consumption of over 60%. Despite the potential of this technology,itsapplicationseemslimitedsofarduetotheneed of a high pumping power to keep the temperature gradient in the fluid from inlet to outlet within acceptable limits. Moreover,theworkingfluidmodeledinmoststudiesisusuallypure water, which presents a problem with its high freezing point, and hence the even higher pressure drop/pumping power of watereglycol mixtures, the real liquid working fluid, has to be dealt with for realistic evaluations. Furthermore, manufacturers are reluctant to use water-based fluids directly in their servers and mainframes due to reliability questions.

The use of a porous media with single-phase and twophase flow has as the main advantage of the large surface area for heat transfer. Nevertheless, a high pumping power remains as a limitation. Jet impingement is another promising cooling solution that can reach low thermal resistances without any thermal interface and yield nearly uniform surface temperatures with multiple jets requiring potentially high pumping power. However, possible problems related to surface erosion as a consequence of the continuous impingement of the high velocity jets needs to be further investigated.

Finally, two-phase flow in microchannels, i.e. evaporation of dielectric refrigerants, is a promising medium to long term solution, despite the higher complexity involved. This solution consumes a low pumping power (only 1/10 as much as water cooling according to Hannemann et al., 2004), has good temperature uniformity (Agostini et al., 2008), very high heat transfer coefficients (as high as 270 000 W m^[1] K¹ according to Madhour et al., in press), and provides high heat flux dissipation. Studies demonstrate that the thermal resistance decreases with increasing heat flux and with decreasing hydraulic diameter. Possible problems with flow instabilities have been resolved using micro-orifices at the channel inlets (Agostini et al., 2007) while the prediction methods of local heat transfer coefficients (Consolini and Thome, 2010), the critical heat flux (Mauro et al., 2010), and pressure loss (Cioncolini et al., 2009) in the two-phase region are still improving. On the other hand, published tests using water evaporating in microchannels to cool computer chips are not a viable solution (too low absolute pressure and vapor density at 60 C relative to the ensuing pressure drop and speed of sound) and hence this fluid is not considered here.

Marcinichen and Thome (2010) showed results of a simulationcodeforsingle-phaseliquidwaterandtwo-phaseHFC134a cooling cycle, both with a liquid pump as the driver. The liquid water cooling cycle presented a pumping power consumption 16.5 times that obtained for the two-phase HFC134a cooling cycle, considering a design of components and piping so that thetotalpressuredropinthecyclewasabout1bar.Theresults permitted them to conclude that the two-phase HFC134a coolingcyclecanoperateatamuchlowerenergyconsumption compared with a single-phase liquid water cooling cycle. This result can be considered a differential when compared with demonstration projects, such as that for the new supercomputer called AQUASAR (Ganapati, 2009), which considers the implementation of a liquid water cooling cycle on a rack cabinet with power consumption of around 10 kW.

In this context, the objective of the present study is to proposeandanalyzepotentialtwo-phasecoolingcyclesableto maintain the temperature of the chip below its upper operating limit and to recover energy from the cycle’s condenser forexternalapplications,suchasheatingabuilding,residence, hospital,preheatingofboilerfeedwater,etc.Themainfocusis to work with two-phase flow of dielectric refrigerants, using a liquid pump or a vapor compressor to drive the fluid, which can reduce the demand for cooling energy by an impressive amount compared to the large refrigeration chillers currently used to cool air in data centers. A specific analysis of the potential working fluids for this application and the results of critical heat flux obtained with mathematical models developed for micro-evaporators are presented. Also, qualitative and quantitative comparisons of the cooling cycles proposed are made here, using the cycle overall efficiency and the potential for heat recovery as the factors of merit.

had a mass flow rate, a pumping power and a condenser size that were 4.6, 10 and 2 times smaller than the water-cooled system. The coolant temperature rise was 10 C for the water but negligible for HFC134a. In their study, a demonstration radar cooling unit was also designed and built for a 6.4 kW heat load (sixteen 400 W cold plates with convoluted fins). For 25 C ambient air temperature, the working fluid saturation temperature was maintained at 32 C with a total volumetric liquid flow rate of 376 L h¹ and a cold plate outlet vapor quality of 30%, providing a safety factor for dry-out. The system was stable, easily controllable and provided essentially isothermal conditions for all the cold plates. They emphasized the significant benefits from efficiency, size and weight that were provided with the PLMC solution.

Mongia et al. (2006) designed and built a small-scale refrigeration system applicable for a notebook computer. The system basically included a minicompressor, a microchannel condenser, a microchannel evaporator and a capillary tube as the throttling device and is considered to be the first refrigeration system developed that can fit within the tight confines of a notebook and operate with high refrigeration efficiencies. HC600a (isobutane) was the working fluid, chosen from an evaluation of 40 candidate refrigerants. According to them, HC600a presented the best efficiency at a low pressure ratio and was readily available, although flammable, but the system required only a very small fluid charge (a few milliliters). Two evaporators were used, the first one a microchannel evaporator to cool the high heat flux component (chip) and the second one a superheater (conventional finned evaporator) to cool lower heat flux components, such as memories, which also guaranteed that superheated vapor was delivered to the minicompressor inlet. Two thermal test vehicles were used to simulate the chip and the power components. For a baseline operating condition, when the evaporator and condenser temperatures and the heat load were 50 C, 90 C and 50 W, the coefficient of performance (COP) obtained was 2.25. The COP reached 3.70 when the evaporator and condenser temperatures increased and decreased by 10 C from the baseline conditions and the heat load was reduced to 44 W. The smallscale refrigeration system achieved 25e30% of the Carnot efficiency (ideal COP for a Carnot cycle), values comparable with those obtained in today’s household refrigerators.

Trutassanawinetal.(2006)designed,builtandevaluatedthe performance of a miniature-scale refrigeration system (MSRS) suitableforelectronicscoolingapplications.TheirMSRShadthe following components: a commercial small-scale compressor, a microchannel condenser, a manual needle valve as the expansion device, a cold plate microchannel evaporator, a heat spreader and two compressor cooling fans. A suction accumulatortoavoidliquidflowtothecompressor,anoilfiltertoreturn oil to the compressor and guarantee good lubrication, and heat sources to simulate the chips were also installed. HFC134a was the working fluid. System performance measurements were conducted at evaporator temperatures from 10 C to 20 C and condenser temperatures from 40 C to 60 C. The cooling capacityofthesystemvariedfrom121Wto268WwithaCOPof 1.9e3.2 at pressure ratios of 1.9e3.2. Their MSRS was able to dissipate CPU heat fluxes of approximately 40e75 W cm² and keep the junction temperature below 85 C for a chip size of

1.9 cm². It was concluded that a new compressor design for electronics cooling applications was needed to achieve better performanceofthesystem(themostsignificantlossesoccurred in the compressor, which was not designed for the operating conditions of electronics cooling). It was also recommended to study the development of an automatic expansion device and a suitable control strategy for the MSRS.

Trutassanawin et al. (2006) also mentioned some alternative cooling approaches such as heat pipes, liquid immersion, jet impingement and sprays, thermoelectrics and refrigeration. For refrigeration, the following possible advantages were cited: (i) one of the only methods which can work at a high ambient temperature, (ii) chip to fluid thermal resistances are considerably lower, resulting in lower junction temperatures, which could lead to higher heat fluxes being dissipated, and (iii) lower junction temperatures can also increase the microprocessor’s performance and increase the chip’s reliability. Possible “disadvantages” were characterized to be: (i) an increase in the complexity and cost, (ii) possible increase in the cooling system volume and (iii) uncertainties in the system reliability (moving parts in the compressor).

Thome et al. (2007) surveyed the advances in thermal modeling for flow boiling of low-pressure refrigerants in multimicrochannel evaporators for cooling of microprocessors. According to them, multi-microchannel evaporators hold promise to replace the actual air-cooling systems and can compete with water cooling to remove high heat fluxes, higher than 300 W cm², while maintaining the chip safely below its maximum working temperature, providing a nearly uniform chip base temperature (Agostini et al., 2008) and minimizing energyconsumption.Variablessuchascriticalheatfluxes,flow boilingheattransfercoefficientsandtwo-phasefrictionfactors were evaluated and characterized as important design parameters to the micro-evaporator for high heat flux applications.

Thome and Bruch (2008) simulated two-phase cooling elements for microprocessors with micro-evaporation. Heat fluxesof50Wcm² and150Wcm² inamicro-evaporator with channels 75 mm wide, 680 mm high and 6 mm long with 100 mm thick fins were simulated for flow boiling. The size of the chip was assumed to be 12 mm by 18 mm and the micro-evaporator was considered with the fluid inlet at the centerline of the chip and outlets at both sides, i.e. a split flow design to reduce the pressure drop but increase the critical heat flux. Results of pumping power, critical heat flux, and junction and fluid temperaturesweregeneratedforHFC134aataninletsaturation temperature of 55 C (chosen to allow for heat recovery). The followingconclusionswerereached:i)theinfluenceofmassflux on the fluid, chip and wall temperatures was small, ii) for the heatflux of 150 W cm², the chip temperature was 70C orless, i.e.wellbelowitsoperationallimitof85C,iii)fortheheatfluxof 150 W cm², the junction-to-fluid temperature difference was only 15 K, which is lower than that with liquid cooling systems, iv)thefluidtemperaturecouldstillberaisedby10Ktoajunction temperature of 80 C while rejecting heat at 65 C for reuse, and v) the critical heat flux increased with the mass flux and the lower limit was about 150 W cm² for 250 kg m² s¹. The channel width had a significant effect on the wall and junction temperatures, and there was a turning point at about 100 mm when considering 1000 kg m² s¹ of mass flux and 150 W cm² of base heat flux, at which these temperatures reached a minimum. For the same mass flux and base heat flux, the reduction of channel width also reduced the energy consumption to drive the flow (pumping power).

From a system viewpoint, Thome and Bruch (2008) showed an approximate comparison of performances of liquid water cooling versus two-phase cooling. For the same pumping power consumption to drive the fluids, two-phase cooling allowed the chip to operate about 13 K cooler than water cooling or it could operate at the same junction temperature but consume less pumping power using a lower refrigerant flow rate. The two-phase cooling system appeared to be more energy-efficient than classical air-cooling or direct liquid cooling systems while also exhausting the heat at higher reusable temperatures. Regarding the choice between a pump and a compressor as the driver for a micro-evaporation heat sink system, they emphasized that the choice depends on the economic value of the re-used energy. The system with a compressor is ideal for energy reuse because of the higher heat rejection temperature; however the additional energy consumed by the compressor compared to the pump has to be justified by the reuse application.

Mauro et al. (2010) evaluated the performance of a multimicrochannel copper heat sink with respect to critical heat flux (CHF ) and two-phase pressure drop. A heat sink with 29 parallel channels (199 mm wide and 756 mm deep) was tested experimentally with a split flow system with one central inlet at the middle of the channels and two outlets at either end. Three working fluids were tested (HFC134a, HFC236fa and HFC245fa) and also the parametric effects of mass velocity, saturation temperature and inlet temperature. The analysis of their results showed that a significantly higher CHF was obtainable with the split flow system compared to the single inlet-single outlet system (Park and Thome, 2010), providing also a much lower pressure drop. For the same mass velocity, the increase in CHF exceeded 80% for all working fluids evaluateddueto theshorterheatedlengthof asplitsystemdesign. For the same total refrigerant mass flow rate, an increase of 24% for HFC134a and 43% for HFC236fa were obtained (no comparabledatawereavailableforHFC245fa).Theyconcluded that the split flow system had the benefit of much larger CHF values with reduced pressuredrops and further developments in the design of split flow system could yield an interesting energetic solution for cooling of computer chips. Fig. 2 shows the details of the two configurations of multi-microchannel copper heat sink regarding the inlet and outlet flow system.

Fig. 2 e Schematic of micro-evaporator: a) one inlet/one outlet and b) one inlet/two outlets.

Itisworthnotingthatthefocusoftheabovestudieswasthe development of multi-microchannels evaporators able to

remove “in loco” the heat load generated by the microprocessors and also the development of two-phase cooling systems able to: i) control the operating conditions in the micro-evaporator, ii) maintain the microprocessor temperature at acceptable levels, iii) recover the heat for a secondary process and iv) operate at a much lower pump energy consumption compared with a single-phase liquid water cooling system.

3.              Present work

Three micro-evaporator cooling cycles are proposed here:

1.   one with a liquid pump as the driver of the working fluid,

2.   one with a vapor compressor as the driver of the workingfluid, and

3.   one hybrid cycle that is a combination of the first twocycles.

The main characteristics of each cycle are presented below with a focus on their advantages and the functions of the components. Additionally some simulations are presented showing the following: (i) performance of the vapor compression cooling cycle for four refrigerants as the possible working fluids for cooling the microprocessors, (ii) operational limits for one specific geometry of a micro-evaporator (critical heat flux, outlet vapor quality, pressure drop, etc) to demonstrate its suitability for this type of application, and (iii) potential for heat recovery and the cycle overall efficiency for the first two cycles proposed.

3.1.            Two-phase micro-evaporator cooling cycle

Figs. 3e5 depict the cycles with a liquid pump, a vapor compressor and hybrid of these two, respectively. The goal is to control the chip temperature to a pre-established level by controlling the inlet conditions of the micro-evaporator (pressure, subcooling and mass flow rate). It is imperative to keep the micro-evaporator (ME) outlet vapor quality below that of the critical vapor quality, which is associated with the critical heat flux. Due to this limitation, additional latent heat is available, which can be used by other heat generating components. The critical heat flux and outlet vapor quality are

Fig. 3 e Schematic of the liquid pump cooling cycle.

Fig. 4 e Schematic of the vapor compression cooling cycle.

predicted using methods developed by Revellin and Thome (2008), which are a function of micro-evaporator inlet conditions and microchannel dimensions.

Another parameter that must be controlled is the condensing pressure (condensing temperature). The aim is, during the winter, to recover the energy dissipated by the refrigerant in the condenser to heat buildings, residences, district heating, etc. In order to accomplish this, the idea is to use a variable speed compressor and an electric expansion valve, as will be discussed below.

Fig. 3 depicts the two-phase cooling cycle where the flow rate is controlled by a liquid pump. The P-h diagram (Fig. 6), which was drawn for low pressure refrigerant HFC245fa, shows the thermodynamic conditions for specific points along the cooling cycle, considering 9.9 K and 60 C for the subcooling and evaporating temperature at the ME inlet, respectively. The pressure drops in the micro-evaporator and microchannel cooling plate for the memory chips (MP_M) were simulated to be on the order of 0.5 bar and 0.0 bar (it is negligible), respectively, based on preliminary calculations. These values are representative and were defined only for cycle interpretation. The components considered and their main functions are presented below:

a)    Variable speed liquid pump: controls the mass flow ratecirculating in the system.

b)   Stepper motor valve: controls the liquid flow rate tocontrol the outlet vapor quality in each micro-evaporator (0%e100%).

c)    Micro-evaporator (ME): transfers the heat generated by themicroprocessor to the refrigerant.

d)   Microchannel cold plate for memories (MP_M): additional component used to cool the memories using the remaining latent heat, which is available due to the limitations enforced on the micro-evaporator.

e)    Pressure control valve (PCV): controls the condensingpressure.

Fig. 5 e Schematic of the hybrid cooling cycle highlighting the possibility of interchangeability between liquid pump and vapor compression cooling cycle.

f)    Condenser: counter-flow tube-in-tube exchanger or a micro-condenser.

g)   Liquid accumulator: guarantees that there is only satu-rated liquid at the subcooler inlet, independent of changes in thermal load.

h)   Temperature control valve (TCV): controls the subcoolingat the inlet of liquid pump.

This cycle is characterized in having low initial costs, a low vapor quality at the ME outlet, a high overall efficiency, low maintenance costs and a low condensing temperature. This is a good operating option when the energy dissipated in the condenser is not recovered, typically during the summer season. However, the heat can still be recovered if there is an appropriate demand for low quality heat (low exergy).

Fig. 6 e HFC245fa P-h diagram showing the thermodynamic conditions for specific points of the liquid pump cooling cycle.

Fig. 4 shows a two-phase cooling cycle where a vapor compressor is the driver of the working fluid. The P-h diagram (Fig. 7), which was also drawn for low pressure refrigerant HFC245fa, shows the thermodynamic conditions for specific points along the cooling cycle, considering 0.69 K and 60 C for the subcooling and evaporating temperature at the ME inlet, respectively. The pressure drops in the ME and MP_M were considered to be the same as for the liquid pump cycle above. The components considered and their main functions are:

a)    Variable speed compressor: controls the ME inlet pressureand consequently the level of inlet subcooling.

b)    Pressure control valve (PCV): controls the condensingpressure.

c)    Condenser:       counter-flow             tube-in-tube              exchanger or

a micro-condenser.

Fig. 7 e HFC245fa P-h diagram showing the thermodynamic conditions for specific points of the vapor compression cooling cycle.

Fig. 8 e Effect of superheating at the inlet of the VSC on the isentropic COP.

d)    Liquid accumulator: guarantees that there is only satu-rated liquid at the internal heat exchanger (iHEx1) inlet.

e)    Internal heat exchanger liquid line/suction line (iHEx1):increases the performance of the cooling system. Fig. 8 shows the ratio of the isentropic COP with superheating at the inlet of the VSC relative to the saturation COP_sat (as defined by Gosney, 1982). Condensing and evaporating temperatures of 60 C and 90 C were considered, respectively. It is worth noting that for the four potential working fluids analyzed, the ratio increases with superheating, although some fluids, such as ammonia, shows decreasing performance (Gosney, 1982).

f)     Electric expansion valve (EEV): controls the low-pressurereceiver level.

g)    Lowpressurereceiver (LPR):this componentcan be seen asa second internal heat exchanger liquid line/suction line, which increases the EEV inlet subcooling and allows an overfeed to the ME since the ME outlet returns to this receiver. The same analysis considered for the iHEx1 can be considered here, i.e. the LPR increases the performance of the cooling system as it, together with the iHEx1, generates the superheating and the subcooling at the inlet of the VSC and EEV, respectively.

h)    Stepper motor valve: controls the liquid flow rate tocontrol the outlet vapor quality in each micro-evaporator (0%e100%).

i)     Micro-evaporator (ME): transfers the heat away from the microprocessor.

j)     Microchannel cold plate for memories (MPM): cools thememories.

This cycle is characterized by a high condensing temperature (high heat recovery potential), a high range of controllabilityof theME inletsubcooling (characteristic ofsystemswith VSC and EEV), a medium overall efficiency when compared with the liquid pumping cooling cycle (uncertain, evaluate potential for heat recovery in the condenser). This is a good operating option when the energy dissipated in the condenser is recovered for other use, typically during the winter season when considering a district heating application (high exergy).

Fig. 5 considers a hybrid two-phase cooling cycle, i.e. this multi-purpose cooling cycle makes it possible to interchange the cycles driven by the liquid pump and the vapor compressor. The change of cycle would be accomplished through the shut off valves 1e7 (SOV). The decision on the cooling cycle to operate would depend on demand for the heat recovered, or allow for cycle maintenance (repair of the compressor or pump). The microprocessors cannot operate without cooling and thus the interchangeability of the cycles represents a safety mechanism in case of failure of the pump or compressor. The “cons” of the hybrid cycle would be mainly the higher initial cost but certainly the advantages (system online reliability, controllability, cycle interchangeability and flexibility in heat recovery) may prove to justify the higher initial cost. Furthermore, this hybrid cycle represents a plugand-play option where any one of the three cycles can be installed based on the particular application, minimizing engineering costs.

Fig. 9 e Typical blade with two microprocessors and a heat generation capacity higher than 300 W.

It is worth mentioning that the applicability of these cooling cycles is not restricted to only one microprocessor but can be applied to blade servers and clusters, which may have up to 64 blades per rack cabinet. Each blade, such as that shown in Fig. 9, can have two (or more) microprocessors with a heat generation capacity higher than 150 W. If the auxiliary electronics (memories, etc.) on the blade are included, the total heat generation per blade can be higher than 300 W. Thus, the microchannel cold plate (MP_M) described in the cooling cycles has the function to cool the auxiliary electronics that can represent about 60% of the total heat load on the blade, but have a larger surface area compared to the CPU and thus a lower heat flux.

Finally, when considering an entire rack, a very sizable heat loadis generated,which representsa good opportunity to recover the heat rejected. In this case, reuse of the heat removed from the blades for a secondary application will greatly reduce the CO₂ “footprint” of the system. For example, if we consider a data center with 50 vertical racks, where each rack has 64 blades and each blade dissipates 300 W, the total potential amount of heat to be recovered will be 0.96 MW. Such a heat recovery system requires a secondary heat transfer fluid to pass through all the condensers (either water or a refrigerant) and then transport the heat to its destination. 3.2. Working fluids

Thepresenceofoil inthe coolingcycleswouldadverselyaffect the performance of the heat exchangers and also possibly lead toproblemsofcloggingofsmallcomponentsandgenerationof contaminants (Marcinichen, 2006). So, for this reason, these cyclesshoulduse drivercomponents that do not requireoil for lubrication purposes (that is, an oil free liquid pump and/or an oil free vapor compressor should be used).

Table 1 shows a comparison of four refrigerants in relation to their environmental parameters (BNCR35, 2008), where GWP is the global warming potential (ratio of the warming caused by the substance to the warming caused by a similar mass of carbon dioxide, GWP ¼ 1 for CO₂) and ODP is the ozone depleting potential (ratio of the impact on ozone of a chemical compared to the impactof a similarmassof CFC11, ODP¼ 1 for CFC11). It is worth noting that the refrigerants considered have an ODP of zero, but still have rather high values of GWP.

The four working fluids (HFC236fa, HFC245fa, HFC134a and HC600a-isobutane) were evaluated with regard to COP and the volumic refrigerating effect for the vapor compression cooling cycle proposed (Fig. 4). The cycle considers two microchannel cooling components, ME and MP_M, the first to cool the microprocessor (outlet vapor quality set to 30%) and the second to cool the auxiliary electronics (memories, DC/DC, etc) on the blade microprocessor (outlet vapor quality set to 90%, which is the estimated value that considers the blade manufacturer’s information that the auxiliary components

Table1 eEnvironmental parameters GWPandODP for the four potential working fluids.

Refrigerant                             GWP (100 year)                         ODP

HFC236fa 6300 0 HFC245fa 950 0 HFC134a                                                               1300                                                           0 HC600a                                                                         3                                                           0

Table 2 e Boundary conditions for the working fluids analysis on the vapor compression cooling cycle.

1)  Condenser > condensing temperature ¼ 90 C, outlet vapor quality ¼ 0% 2)  Micro-evaporator on chip (ME) > inlet saturation temperature ¼ 60 C, >outlet vapor quality ¼ 30% 3)  Microchannel cold plate on memories (MPM) > outlet vapor quality ¼ 90% 4)  Effectiveness of iHEx1 ¼ 90% 5)  Input data > fluids: HFC245fa, HFC236fa, HFC134a and HC600a > total pressure drop in the two evaporators (ME and MPM) ¼ 0.5 bar 6)  Outlet data > discharge temperature (isentropic compression) > enthalpy difference in the two evaporators and in the compressor > volumic refrigerating effect (qualitative idea of compressor size) >COP

can represent up to 65% of the total heat load). It was also considered that iHEx1 has an effectiveness of 90% and the two microchannel cooling components have a total pressure drop of 0.5 bar.

The volumic refrigerating effect (w_v) is determined by calculating the ratio between the sum of the ME and MP_M enthalpy differences and the specific volume in the compressor suction (Gosney, 1982). This parameter indicates comparatively the size of compressor for the different working fluids, i.e. a higher volumetric refrigerating effect means that a smaller swept volume rate is required for a particular cooling capacity.

Table 3 shows the results considering the conditions in Table 2. For this cycle, the COP was determined by dividing the sum of the ME and MP_M enthalpy differences (Dh_ME) by the compressor enthalpy difference (Dh_comp). It can be observed that HFC245fa has the lowest suction and discharge pressures

(P_suc and P_disc), which is advantageous for the compressor and cooling system (allows a less robust construction that enables material cost savings). However, it also has a lower volumic refrigerating effect, meaning that a larger compressor will be necessary. The best working fluid, when looking at the volumic refrigerating effect, is HFC134a since its value is more than 2 times higher than that of HFC245fa, but requires operation at a higher P_suc and P_disc. It is worth noting that HC600a (isobutane) has the highest specific cooling capacity (Dh_ME), as shown in Fig. 10, implying lower mass flow rates for

the same cooling capacity.

Relatively small differences in COP are observed in Table 3 for the four fluids, showing no significant effect on the choice of the working fluid. The same can be said about the compressor discharge temperature (T_disc). The high values of COP observed are justified by the fact that the thermodynamic analysis does not consider the irreversibilities inherent in the cycle. However, due to the high evaporating temperatures considered here (for attaining a high performance green

HC600a

h,kJ/kg

Fig. 10 e HC600a P-h diagram highlighting the large specific cooling capacity.

computing solution), the COP values are higher than those found in household refrigerators and light commercial systems (actual COP about 2 or 3). Finally, it can be observed that HC600a and HFC134a present the lowest pressure ratios, which is an advantage because they represent compressors with high volumetric efficiencies.

Fig. 11 shows the effects of iHEx1 effectiveness and condensing temperature on the cycle COP. The same conditions described in Table 2 were considered and HFC134a was used as working fluid. It can be observed that the COP increases when the iHEX1 effectiveness increases. However, the condensing temperature has a greater effect on the COP, decreasing with an increase of condensing temperature. It is worth mentioning that there might be an optimal condensing temperature to obtain the maximum economical value of recovered heat for the penalty paid in compressor power consumption.

ME must be able to maintain the microprocessor’s operating temperature from 70 C to 75 C (83 C is the nominal maximum operating temperature).

Basedontheaforementionedinformation,Table4showsthe results obtained by the methods developed to evaluate the performance of the ME’s. The three-zone model (Thome et al., 2004) was used for two-phase heat transfer since it was shown to predict many fluids and geometries with good accuracy (Dupont et al., 2004), the numerically based model of Revellin and Thome (2008) was used for critical heat flux calculations and the homogeneous model was used for two-phase pressure dropssinceitwasfoundtopredictmicrochannelpressuredrops withrelativelygoodaccuracy(Ribatskietal.,2006).Themethods werealsousedtoestimatethemassflowrateintheME.Theheat load was considered to vary between 90% and 100%, i.e. from 146.25W to162.50 W. The ME inlet subcooling, Sub, wasfixed at

5 K and two inlet evaporating temperatures, T_{evap_inlet}, were considered, 50 C and 60 C. The dimensions of the ME were 170mmoffinwidthandchannelwidthand1700mmoffinheight, with a heated “footprint” of 18.5 mm by 13.5 mm. The working fluid selected for the present analysis was HFC134a.

The results show that the mass flow rate, m_r, to guarantee the cooling capacity must be from 10.82 kg h¹ to 11.90 kg h¹ when the outlet vapor quality, x_outlet, is 30% and the inlet evaporating temperature is 60 C. For this case the lowest critical heat flux, CHF, was 141.2 W cm², a value well above the actual value of 65 W cm² (safety factor of 2.2). However, when the outlet vapor quality was set to 50%, the smallest CHF was then only 83.1 W cm², a value judgedto be too nearto the actual value for the standard blade (65 W cm²) since the accuracy in predicting CHF is about 20%. Thus, it is best to consider an outlet vapor quality of 30%. While not done here, it is also possible to use the one-dimensional numerical method of Revellin and Thome (2008) to analyze the CPU die’s power dissipation map to verify the local safety factors in CHF with respect to the local hot spots.

3.4. Analysis of the cycle overall efficiency and potential for energy recovery

Table 3 e Results of simulations on the vapor compression cooling cycle/potential working fluids.
	COP         Tdisc (C)           Pdisc (bar)          Psuc (bar)          Pdisc/Psuc                 DhME (kJ kg1)            wv (kJ m[2])							Dhcomp (kJ kg1)
HFC236fa	8.0	110.9	15.65	7.14	2.19	104.0	4333	13.0
HFC245fa	8.3	110.9	10.04	4.14	2.43	150.0	3010	18.1
HFC134a	7.0	119.2	32.47	16.33	1.99	112.7	7736	16.1
HC600a	8.4	111.1	16.14	8.10	1.99	250.5	4566	30.0

The overall efficiencies (h_cycle) of the proposed cycles were evaluated considering the potential for energy recovery. This is determined by the ratio of the recovered energy in the condenser and subcooler to the energy consumed to drive the working fluid. Some additional terms were also considered to take into account the pumping power of the secondary fluid in the condenser and subcooler. Thus, considering the possible heat recovery in the heat exchangers, h_cycle will be influenced by the type of heat recovery application, since different types

Effects of iHEx1 and condenser on the cycle COP

Fig. 11 e Effects of iHEx1 effectiveness and condensing temperature on the cycle COP.

of condensers, subcoolers and condensing temperatures could be chosen to maximize h_cycle for the particular situation.

For an ideal case, the power dissipated by the microprocessor and memories, Q_MEþMPM, and the power consumed by the compressor, W_compressor, or the liquid pump, W_pump, are fully recovered in the condenser and subcooler. This also holds for the power consumed by the pumps associated with the secondary fluid in the condenser, W_{Cond_pump}, and subcooler, W_{Subcooler_pump}. The cycle overall efficiency for the liquid pump and vapor compression cycle can, therefore, be written as:

a)   Liquid pump cyclehcycle LP ¼ QMEþMPM þ Wþpump þ WCond pump þ WSubcooler pump

                          Wpump                WCond pump þ WSubcooler pump

               ¼ þ              þ          QMEþMPM                                                                                                                                  (1)

1

                        Wpump                WCond pump þ WSubcooler pump

b)  Vapor compression cycle

QMEþMP_M þ WCompressor þ WCond pump

Presently, we are not concerned with the performance of the secondary system heat exchanger, which will be a function of its unknown (for now) mass flow rate and fluid properties. As noted in Eqs. (1) and (2), the h_cycle will depend mainly on the pumping power of the secondary fluid, which in itself is a function of the type of application of the secondary system (heat exchanger size, type and properties of the secondary fluid). It is worth noting that the difference in cycle overall efficiency for the two cycles is in the denominator. In general, the compressor power is higher than the liquid pump power, due to the work needed to obtain a differential pressure associated with the compressor. This could lead to the conclusion that the h_cycle for the liquid pump cycle is always higher than for the vapor compression cycle. However, the pumping power of the secondary fluid through the condenser

is higher for the liquid pump cycle than for the other cycle because of the lower condensing temperature, with the possibility of the opposite to be true. Furthermore, the h_cycle will depend on the efficiency of each component and on the end use of the energy recovered in the condenser and subcooler.

The results of a simplified analysis evaluating the potential of heat recovery for the cycles with the liquid pump and the vapor compressor are depicted in Table 5. To develop this analysis, the results in the second line of Table 4 were taken into consideration as well as the following assumptions:

a)   water was considered as the secondary fluid,

b)   the condenser and subcooler water pumping powers werenot considered,

c)   the HFC134a liquid pumping power was determinedthrough Eq. (3) for 100% liquid pump overall efficiency. The liquid pump inlet subcooling was considered 10 K and the inlet pressure was considered that at the ME outlet,

d)   the compressor suction pressure was considered to be thesame pressure as at the ME outlet and with 10 K of

superheating,

e)   the vapor compression was considered isentropic and100% compressor overall efficiency, vapor compression cycle,

Table 4 e Operational limits for one micro-evaporator. HFC134a as working fluid.

Boundary conditions in the micro-evaporator (working fluid: HFC134a) Tevap_inlet (C)                Sub (K)            xoutlet (%)              Qw                      mr (Kg h1)             DPME (bar)              Tevap_outlet (C)                CHF (W cm2)

50                                                     5                                 30                         162.50                         10.82                               0.0092                                    50.0                                          145.7 60                                                     5                                 30                         162.50                         11.90                               0.0096                                    59.9                                          148.9 50                                                     5                                 30                         146.25                         10.28                               0.0082                                    50.0                                          141.2 60                                                     5                                 30                         146.25                         10.82                               0.0082                                    59.9                                          141.2 50                                                     5                                 50                         162.50                            7.03                               0.005                                       50.0                                             91.9 60                                                     5                                 50                         162.50                            7.57                               0.005                                       59.9                                             91.2 50                                                     5                                 50                         146.25                            6.28                               0.0045                                    50.0                                             83.3 60                                                     5                                 50                         146.25                            6.76                               0.0043                                    59.9                                             83.1

Table 5 e Comparative analysis for the liquid pump and vapor compression cooling cycle regarding heat recovery.

Cycle                               Energy recovery (W)             Tw_inlet (C)                            Condenser                                           Subcooler                                                                                                               Tw_outlet (C)             mw (Kg h1)             Tw_outlet (C)               mw (Kg h1)

Liquid pump                                                    162.5                                              30                                    49.9                                    4.85                                    49.9                                    2.18 Vapor compressor                                          206.7                                              30                                    80.0                                    3.56                                       e                                          e

h) the condenser and subcooler outlet water temperature was assumed to be 10 K less than the condensing

temperature,

The refrigerant pumping power is thus calculated as:

m

W_pump ¼DP           (3) r

where m is the mass flow rate, r is the specific mass and DP is the pressure increase provided by the pump.

In Table 5, it can be observed that there is an increase of 27.2% of heat recovery for the vapor compression cycle (this additional heat is associated with that imparted by the compressor) while there is an increase of 98% of total water mass flow rate, m_w, for the liquid pump cycle that is associated with a lower water temperature difference in the condenser and subcooler. The final result shows that the liquid pump cycle will require a larger pump to circulate water in the condenser and subcooler, i.e. a higher pumping power and possibly a larger heat exchanger (condenser þ subcooler). Finally, it is important to remember that the results presented above are only for a simple example case and that a more detailed analysis considering all components and their thermal efficiencies needs to be done to better understand the behaviorand performance of eachcycle and its particular heat recovery application.

4.             Conclusion

1.   Three two-phase cooling cycles for cooling of data centerservers have been proposed for more energy-efficient cooling of blade server microprocessors and their memories. The cycles use two-phase boiling in microchannels for removing the heat from the microprocessors and memories and the heat can be dissipated either to the ambient or, better, it will be recovered for heating of buildings, preheating of boiler feed water, etc. This second solution has the potential to be a key step in the realization of a new generation of green, high performance data centers.

2.   To integrate operating flexibility and higher system operating reliability into one cooling cycle, include the possibility to recover heat or not, and to facilitate maintenance while still operating the server’s cooling system, a hybrid cooling cycle was proposed with interchangeability between the liquid pump and vapor compression driven cooling cycles. As the cooling of the servers should have a very high online availability, the interchangeability will

also guarantee uninterrupted operation in case of forced maintenance of the compressor or the pump.

3.   The vapor compression cooling cycle proposed was considered to determine the best working fluid for cooling applications of microprocessors and memories. The analysis took into consideration the following variables: suction and discharge pressures, volumic refrigerating effect, pressure ratio and COP. Of the four refrigerants considered, HFC134a and HFC245fa appear to be the best choices.

4.   Methods taken from the literature to evaluate the thermalperformance of the ME’s were used here to estimate the CHF of the ME and compare to the total heat flux of a specific blade. For an evaporating temperature, subcooling and outlet vapor quality of 60 C, 5 K and 30%, respectively, the predicted CHF was about 2.2 times the actual maximum heat flux of the blade server using fins that were 1700 mm high, 170 mm thick and channels 170 mm wide. This safety factor was considered sufficient since the accuracy in predicting CHF is about 20%. For an outlet vapor quality of 50% the factor decrease to only 1.3 times, a value judged to be too low to guarantee problem free operation.

5.   The micro-evaporator cooling cycles proposed were analyzedinrelationtothecycleoverallefficiency(h_cycle)and the potential for energy recovery, after the aforementioned constraint of critical heat flux was taken into account. The qualitative comparison showed that the best cycle, i.e. that with the highest h_cycle, will depend mainly on the end application of the energy recovered in the condenser and subcooler, which will influence the design of the cooling cycle and the thermodynamic conditions. A quantitative comparison showed that the vapor compression cycle is capable of recovering more energy for a lower water mass flow rate. Also, it was shown that a higher water temperatureisachievedwiththevaporcompressioncycleduetothe higher condensing temperature.

Acknowledgements

The Commission for Technology and Innovation (CTI) contract number 6862.2 DCS-NM entitled “Micro-Evaporation Cooling System for High Performance Micro-Processors: Development of Prototype Units and Performance Testing” directed by the LTCM laboratory sponsored this work along with the project’s industrial partners: IBM Zu¨rich Research Laboratory (Switzerland) and Embraco (Brazil). J.B. Marcinichen wishes to thank CAPES (“Coordenac¸a˜o de Aperfeic¸oamento de Pessoal de N´ıvel Superior”) for a one year fellowship to work at the LTCM laboratory.

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[1] 2.                                                                                                Literature review

Hannemann et al. (2004) have proposed a pumped liquid multiphase cooling system (PLMC) to cool microprocessors and microcontrollers of high-end devices such as computers, telecommunications switches, high-energy laser arrays and high-power radars. According to them, their system could handle applications with 100 W heat loads (single computer chip) as well as applications with short time periods of kW heat loads (radar). Their PLMC consisted basically of a liquid pump, a high performance cold plate (evaporator) and a condenser with a low acoustic noise air mover to dissipate the heat in the ambient air. A comparison between a singlephase liquid loop (water) and the system proposed with HFC134a wasmadefor a 200 Wheat load. The HFC134asystem

[2] .3.        Microprocessor heat load and micro-evaporator model

To evaluate the performance of the cooling cycles proposed, the first step was to define a standard blade model where we need to control the microprocessor and memory temperatures. A photograph of the standard blade is shown in Fig. 9. According to the manufacturer, when we consider only one microprocessor and the auxiliary electronics associated with it, the maximum heat flux will be about 65 W cm² and the area for heat transfer is about 2.5 cm² (this is the worst case, i.e. the entire heat load assumed to be concentrated on the small area of the chip and its ME). Thus the maximum heat transfer rate will be 162.5 W per ME. The cooling capacity per