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International Journal of Thermal Sciences 86 (2014) 21e27

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International Journal of Thermal Sciences

journal homepage: www.elsevier .com/locate/ i j ts

Experimental investigation on heat transfer of spray coolingwith isobutane (R600a)

Hongbo Xu a, Chunqiang Si a, b, Shuangquan Shao a, *, Changqing Tian a

a Beijing Key Laboratory of Thermal Science and Technology, Technical Institute of Physics and Chemistry, Chinese Academy of Sciences, P.O. Box 2711,Beijing 100190, PR Chinab Internal Trade Engineering & Research Institute, Beijing 100069, PR China

a r t i c l e i n f o

Article history:Received 7 June 2013Received in revised form3 April 2014Accepted 23 June 2014Available online

Keywords:Heat transferSpray coolingRefrigerationIsobutane

* Corresponding author.E-mail address: shaoshq@mail.ipc.ac.cn (S. Shao).

http://dx.doi.org/10.1016/j.ijthermalsci.2014.06.0251290-0729/© 2014 Elsevier Masson SAS. All rights res

a b s t r a c t

In order to enhance the heat transfer performance of spray cooling system, an integrated system isdeveloped. The system is based on the refrigeration cycle, and uses isobutane (R600a) as the coolant. Theheat transfer performance, such as the heat surface temperature, the heat transfer coefficient and thesurface temperature distribution, are experimentally investigated in this paper. When the coolant massflow rate is nearly 6.9 kg/h, the surface temperature can be kept at 57.3 �C with the heat flux of 145 W/cm2. The heat transfer coefficient can be achieved up to 35,000 W/(m2 �C) when the nozzle inlet pressureand the chamber pressure are 4.9 bar and 2.3 bar, respectively. The surface temperature uniformity ismainly influenced by the mass flow rate, the heat flux and the nozzle inlet pressure together, and thestandard deviation of the surface temperature is less than 4.0 �C in this study. All the results indicate thatthis system is promising in the application for high heat flux removal.

© 2014 Elsevier Masson SAS. All rights reserved.

1. Introduction

Compared with the traditional cooling technologies such as aircooling and heat pipe cooling, the spray cooling has advantages ofhigh heat flux, small surface superheat and low coolant mass flowrate, therefore it has been applied in many processes such as metalquenching, microelectronics cooling, laser system cooling andhigh-power electronics cooling.

During last several decades, large amounts of experimental re-searches on spray cooling have been performed in order to un-derstand its mechanisms of heat removal and optimize the spraycooling technique [1,2]. Nucleate boiling and secondary nucleation,film evaporation and forced convection are the recognized heattransfer mechanisms of the spray cooling, and the complex heattransfer phenomena have resulted in a heat flux approximately anorder of magnitude higher than the pool boiling using the sameliquid [3e5]. For example, using water as the coolant the criticalheat flux (CHF) of the pool boiling and the spray cooling areapproximately 120 W/cm2 and 1000 W/cm2, respectively. In thecase of FC-72, the CHF of the spray cooling is approximately100 W/cm2 [6], while that of the pool boiling is only about20e30 W/cm2 [7,8].

erved.

Experimental investigations on the influence of hydrodynamicparameters, nozzle types and surface properties on the heattransfer efficiency were also performed by many researchers. Theresults showed that the volumetric flux played a dominant role inthe heat transfer compared to other hydrodynamic properties [9],and the mass flow rate had different effect on the film thickness forthe pressure atomizer and the air assisted nozzle [10,11]. Estes andMudawar [2] presented a CHF correlation depending on the volu-metric flux and the Sauter mean diameter which predicted thespray cooling performance of FC-72, FC-87 and water exactly. Hsiehand Yao [12] experimentally investigated the influence of the sur-face with micro-structures on silicon wavers to the spray coolingand found the heat flux on structured surface increase due to thecapillary force of the surface acting on the film. Sodtke and Stephan[13] studied the spray cooling on the micro-structured surfaceswith different spray distance and found that the spray cooling onthe micro-structured surface can get significantly better coolingperformances compared to the smooth surfaces at the same wallsuperheat. Silk et al. [14] carried out experiments on the surfaceswith cubic fins, pyramids and straight fins, respectively, and ob-tained similar conclusions with Sodtke and Stephan [13]. Utilizing asingle air assisted nozzle and spraying on the heat surface withroughness of 0.3 mm, Pais et al. [15] obtained the maximum heatflux of 1200 W/cm2.

In addition to the above investigations, it was also studied ondifferent coolants. Hsieh et al. [16] used R134a as work fluid and

Nomenclature

A area of heat surface, m2

d diameter, mh heat transfer coefficient, W/(m2 �C)_m mass flow rate, kg/h_N droplets generation frequencyP pressure, Paq heat flux, W/cm2

v velocity, m/sT temperature, �CDT temperature margin measured on side F, �CDx the distance margin measured on side F, ml thermal conductivity, W/(m �C)b spray cone angle, �

r coolant density, kg/m3

s surface tension, N/m

SubscriptsB backF frontL leftR rightc chamberin nozzle inletout nozzle outletsat saturationsd standard deviationsur surface

H. Xu et al. / International Journal of Thermal Sciences 86 (2014) 21e2722

obtained the cooling characteristic (boiling curves) over a range ofspray mass fluxes, Weber number, wall superheat and subcoolingdegree. When the heat flux is 5W/cm2, the heat transfer coefficientup to 5596 W/(m2 �C) was achieved with R134a in the experimentof Yan et al. [17]. Zhou and Ma [18] presented an experiment withR113 and developed a new expression with an interpolationmethod to construct the partially developed nucleate boiling curve.Bostanci et al. [19] sprayed ammonia on the micro-structuredsurfaces, and the heat flux was up to 500 W/cm2 (well belowcritical heat flux limit). By measuring the CHF for the same massflow rate at different nozzle-to-surface distances, Mudawar andEstes [20] got the principle of CHF: the spray impact area justinscribing the square surface of the heater make the CHF bemaximized when the spray is configured. Wang et al. [21] studiedthe effects of spray inclination angle through experiments, andgeneralized correlations were developed for local Nusselt numberas a function of the spray Reynolds number and the non-dimensional temperature in the non-boiling regime of the verti-cal and inclined sprays.

Most investigations were very essential, and some importantprinciples and conclusions have been obtained by researchers,however, most reported experimental systems of spray coolingwith water were driven by a pump, and heat transfer occurred inthe open space. Therefore, the environmental pressure and thecoolant boiling temperature were difficult to adjust. Usually thetemperature of the heat surface was very high. For example, theheat surface temperaturemust be up to 100 �C in order tomake thewater occur boiling heat transfer in the open space. However, mostdevices need to be kept at lower surface temperature. In the case ofmost electronic chips, surface temperature below 75 �C is better fortheir normal operation. In order to make the water boiling pointlower, which is the major condition to get lower heat surfacetemperature, it is necessary to achieve the higher vacuum envi-ronment. For the coolant such as ammonia and R134a, a close loopcan be constructed, and lower surface temperature can be attained.However some other disadvantages came out. At the standard at-mospheric pressure, the latent heat of Ammonia is 1366 kJ/kg, andits boiling temperature is �33 �C. However, it is unacceptable by itstoxicity and flammability in most applications. Its corrosive char-acter also limited its application. R134a has the latent heat of only214 kJ/kg and will be limited due to its greenhouse effects. Manyspray cooling system driven by pump consisted of the spray cycleand the refrigeration cycle to cool the coolant, so the system wereextremely complex.

Therefore, the combination of the spray system and the refrig-eration system is a better way to solve the above problems. Thethrottling device and the evaporator in the refrigeration system

were replaced by the nozzle and the spray chamber, respectively,which achieved the spray cooling in the refrigeration cycle. It alsoshould be emphasized that the isobutane (R600a) as anenvironment-friendly refrigerant has the latent heat of 366 kJ/kg atthe atmospheric pressure. Although R600a is flammable, thewidely application in refrigerators has proven that the security canbe guaranteed.

In this study, a close-loop spray cooling system is constructed,and the pressure nozzle is driven by a compressor. R600a containedlubricating oil is used as the coolant to cool a copper block heatedby five cartridge heaters. The temperature and the uniformity of thesurface and the heat transfer coefficient are investigated in thispaper, and the influences of the heat flux, the chamber pressure andthe nozzle inlet pressure are analyzed.

2. Experimental setup and procedures

2.1. Experimental setup

Fig. 1 shows the schematic of the experimental system, whichuses R600a as the working fluid. The spray cooling system as aclosed loop is similar with the common refrigeration cycle. Thesuperheated coolant vapor from the compressor (point 1) iscondensed by water-cooled condenser and enters the receiver(point 2). Through the filter (point 3), the liquid is divided into twoparts (points 4 and 4a). One part coolant from point 4a is throttledby the expansion valve and flows into the subcooler (point 5a),which will be used to cool the liquid flowing into the nozzle andkeep it saturated. After the subcooler, the coolant is inhaled by thecompressor in the form of vapor. The other part coolant from point4 enters the subcooler (point 5) through the flow meter. Then thecoolant sprays onto the heat surface via of the nozzle. The coolantflows into to the accumulator (point 6) in the form of vapor andliquid. The liquid evaporates in the accumulator, and the entirevapor is inhaled into the compressor (point 7) to continue nextcycle.

The system mainly consists of a fluid supply portion, a spraychamber and the data acquisition system.

For the fluid supply portion, an inverter compressor is used asthe hydrodynamic force and the source of the nozzle inlet pressure.The inverter compressor can also control the chamber pressure bychanging its frequency. The condenser can adjust the nozzle inletpressure by changing the temperature of cooling water. Theexpansion valve and the subcooler guarantee that the liquid issaturated before entering the nozzle. Valve 1 plays the role ofthrottling, and it also plays the role of controlling the nozzle inletpressure. In order to observe the coolant flow pattern before nozzle

Fig. 1. Systemic schematics of spray cooling.

Fig. 2. Heat unit (partial sectional view).

H. Xu et al. / International Journal of Thermal Sciences 86 (2014) 21e27 23

inlet, a hyaline tube was installed at the upstream of the nozzleinlet.

The spray chamber is the most important part of the systemwhich contains a heat block simulating the heat device and a full-cone pressure nozzle. The spray cone angle of the nozzle is 60�,the outlet diameter of the nozzle is 0.83 mm, and it is installed10.4 mm right above the heat surface. The heat surface is a circularand its diameter is 1.2 cm. A thermocouple is installed in the spraychamber to measure its temperature.

According to Ghodbane and Holman's [22] research, the massmean diameter of droplets after the nozzle is calculated as follows:

d50 ¼ 9:5doutDP0:37 sin b

2

(1)

where, dout is the nozzle outlet diameter, DP is the pressure dropbetween the inlet and outlet of the nozzle, and b is spray coneangle.

So we can get the range of mass mean diameter of droplets isapproximately 70e90 mm.

According to the energy conservation equation (Eq. (2)) andcontinuity equation (Eq. (3)), as follows:

rinvinAin

v2in2

þ Pinrin

!¼ _N

proutd3506

�v2out2

þ Poutrout

�þ _Nspd250 (2)

where, r is the coolant density, v is the flow speed of coolant, A isthe cross-sectional area of the nozzle, P is the coolant pressure, s isthe surface tension of droplets, d50 is the quality average diameterof droplets, _N is the droplets generation frequency.

rinvinAin ¼ _Nproutd350

6(3)

In our system, the spray chamber and spray cooling process arevery similar to Ghodbane and Holman's [22] research, so we can

use Eqs. (2) and (3), and get the droplets velocity calculationequation, as follows:

vout ¼ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiv2in þ 2ðPin � PoutÞ

r� 12srd50

s(4)

Therefore, we can get the range of droplets velocity is16e31 m/s.

The nozzle plays a role of atomization and throttling effect.According to the rule of CHF, the distance from the orifice to heatsurface is determined at 10 mm [20]. The copper heater is wellinsulated by package insulation material and is electrically heatedby five cartridge heaters, which can generate 1100 W of heat load.According toWang et al.'s [23] simulation and experimental results,the heat conduction of this kind of heater unit can be seen as one-dimensional case. Six T-type thermocouples are embedded beneaththe heat surface to measure the surface temperature, and eachthermocouple bead size is 0.5 mm (see Fig. 2). The thermocouples

H. Xu et al. / International Journal of Thermal Sciences 86 (2014) 21e2724

are placed on the four sides of copper heater named F (front), B(back), L (left) and R (right). The three thermocouples on side Fspaces 3 mm apart from each other along the axis of the copperheater, and the distance from the surface to the first thermocoupleis also 3 mm. The other three thermocouples are all apart from theheat surface 3 mm on each side. On side F, the depth of threethermocouples into the heater are 6 mm, and the depth are 5 mm,4 mm and 3 mm on side L, B and R, respectively.

The data acquisition system includes a computer, a data logger,several pressure sensors, a flow meter and a few thermocouples.The measuring oil point at Valve 2 is used to determine oil contentby gravimetric method in this study. All the measuring points areillustrated in Fig. 1. And all the information of experimental devicesis shown in Table 1.

2.2. Experimental procedure

The experiments are conducted according to the followingprocedure:

1) Prior to each set of experiments, the temperature of coolingwater used in the condenser should be kept at the requiredvalue firstly.

2) The refrigeration system is turned on and operates for approx-imately 5 min to assure the values of pressure sensors reach in asteady state.

3) Adjust AC voltage regulator and input prospective heat load onthe heat unit.

4) When the temperatures' deviations are less than 0.5 �C and keptstable for more than 5 min, all the values of temperatures, massflow rate and pressures are recorded. The oil ratio in the coolantis also measured at the same time. In this study, the oil ratiodeviated between 2.4% and 3.2%, which can be seen as constantin the following experiments.

5) According to the experimental requirement, the nozzle inletpressure can be adjusted by controlling valve 1 and the coolingwater temperature, and the chamber pressure also can beadjusted by changing the frequency of the compressor.

2.3. Data processing and uncertainty analysis

The heat flux q on the heat surface is calculated by Fourier heatconduction rule as the following equation:

q ¼ lDTDx

(5)

By Eq. (1) and the measured values of the thermocouples shownin Fig. 2, four surface temperatures of four sides (TF, TL, TB, TR) can becalculated, and the average of these temperatures is defined as thesurface temperature Tsur.

The heat transfer coefficient h is calculated based on thefollowing definition:

Table 1The detailed information of devices.

Components Type Range Accuracy

Compressor Rotary and inverter 30e120 Hz e

Power meter PF9830 0e2500 W ±1.5%Flow meter Coriolis 0e40 kg/h ±0.25%Pressure senor e 0e10 bar ±0.25%Thermocouple T �40 to 350 �C ±0.1 �CBalance e 0e2000 g ±0.01 g

h ¼ qTsur � Tsat

(6)

To study the factors influencing the temperature distribution onthe heater surface, a parameter Tsd named temperature standarddeviation was defined as:

Tsd ¼ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi14

hðTF� TsurÞ2 þðTL � TsurÞ2 þðTB � TsurÞ2 þðTR � TsurÞ2

ir(7)

The surface temperature was extrapolated by the measuringvalue of thermocouples and Fourier heat conduction law. The un-certainty of thermocouples is about ±0.1 �C. The pressure sensorcan obtain all the pressures including the chamber pressure and thenozzle inlet pressure. All the pressure sensors' uncertainties are±0.25%. The Coriolis flow meter has a measurement uncertainty of±0.25% for total mass flow meter. According to the traditionalmethod of error analysis [24,25], the uncertainty of the heat fluxand the heat transfer coefficient are calculated by Eqs. (8) and (9) asfollows:

4q ¼ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiX2i¼1

����vln qvTi

����2

ε2T þ

����vln qvx

����2

ε2x

vuut (8)

4h ¼����4q �

����� TsurTsur � Tsat

� 4T

����þ���� TsatTsur � Tsat

� 4Ts

��������� (9)

where, εx is the absolute error of distance, εT is the absolute error oftemperature, 4T is the uncertainty of surface temperature, 4Ts is theuncertainty of saturation temperature.

Then, the calculation results of 4q and 4h are 5.5% and 6.4%,respectively.

3. Results and discussions

3.1. Effects of the mass flow rate and the inlet pressure

In the spray cooling system, themass flow rate of the coolant is avery important parameter which has directly effects on the processof heat transfer. Maintaining the chamber pressure of 2.0 bar (withthe evaporation temperature of 7.7 �C) and the heat flux of 145 W/cm2, the mass flow rate changes from 4.4 kg/h to 6.9 kg/h, and itseffects on the heat transfer coefficient and the surface temperaturestandard deviation are shown in Fig. 3. In accordancewith themassflow rate, the nozzle inlet pressure varies from 2.9 bar to 4.9 barFig. 3(a) indicates that themass flow rate has a positive influence onheat transfer coefficient, and the similar phenomenonwas found inother investigations [16,17]. When the mass flow rate is 6.9 kg/h,the heat transfer coefficient is up to 25,373W/(m2 �C) at the surfacetemperature of 57.3 �C. This value is an order of magnitude higherthan the result of 3556 W/(m2 �C) with mass flow rate of 46.8 kg/h[17], but in considering of the per surface area flow rate in our studyis 6.1 kg/(h cm2), that is more than an order of magnitude higherthan the per surface area flow rate in Ref. [17], which is about0.35 kg/(h cm2), so we think our experimental results are reason-able. As mentioned before, in this study, the main heat transfermechanisms are nucleate boiling, secondary nucleation and forcedconvection. The increase of mass flow rate results in the increases ofboth the droplet amount and the droplet velocity impinging on theheat surface at the same time. When the droplets impinge on aliquid surface, they entrain vapor as bubbles into the liquid, whichsever as nuclei for additional bubbles to grow [3]. The amount ofsecondary nucleation sites is proportional to the number of

Fig. 4. Effects of nozzle inlet pressure on surface temperature.

Fig. 3. Effects of mass flow rate. (a) heat transfer coefficient; (b) surface temperaturestandard deviation.

H. Xu et al. / International Journal of Thermal Sciences 86 (2014) 21e27 25

droplets, which means that the increase of the mass flow rateenhance the heat transfer. The droplets velocity increased due tothe increase of mass flow firmly improve the droplets momentum,which benefits the forced convection. It is also highlighted that thehigher latent heat of R600a plays an important role in the heattransfer coefficient.

Fig. 3(b) indicates that the surface temperature differences isfirstly increased and then tends to be constant as the mass flowincreasing. This result is opposite to other research [17]. In theirstudy, there is only about 60% of the entire heat surface impingedby the coolant droplets, so the increase of the mass flow rate canreduce the uneven distribution of the surface temperature natu-rally. However, the area of heat surface is only about 1.1 cm2 in thisstudy, which is all covered by the sprayed coolant. It can beexplained that the enhancement of the droplets velocity due to theincrease of the mass flow rate may result in splashing of thedroplets impinged on the surface. The severe agitation leads touneven distribution of surface temperature, especially on the smallheat surface. When the mass flow rate is high enough, the inter-action of agitation among lots of droplets and the droplets to heatsurface can be equilibrium state, and the role of the forced con-vection increases due to the enough liquid supply. All these tworeasons will keep the standard deviation of the surface temperatureat constant.

Fig. 4 shows that the surface temperature declines obviouslyfrom 84.0 �C to 60.0 �C with the improvement of the nozzle inlet

pressure, when the chamber pressure is kept at 2.0 bar and the heatflux is maintained at 120 W/cm2, while the nozzle inlet pressurevaries from 2.8 bar to 4.9 bar. The increase of the nozzle inletpressure will increase the mass flow rate and enhance the atomi-zation, both of which are good to the heat transfer. This can beexplained that the increase of mass flow ratewhich is caused by theincrease of nozzle inlet pressure will result in the increase of thedroplet amount and the droplet velocity, and that will enhance theheat transfer according to the research of Pais et al. [15]. The nozzleinlet pressure not only increases the velocity of droplets, but alsoenhances the coolant atomization. The increase of the droplet ve-locity will enhance the forced convection, and the advanced at-omization benefits the evaporation of the liquid film. Thecombination of two effects enhances the heat transfer of the spraycooling significantly.

3.2. Effects of the heat flux

Keeping the chamber pressure at 2.3 bar (with the evaporationtemperature of 11.7 �C) and nozzle inlet pressure at 4.9 bar (withthemass flow rate of 5.3 kg/h), the effects of heat flux on the surfacetemperature, temperature distribution and heat transfer coefficientare shown in Fig. 5.

Fig. 5(a) illustrates that the surface temperature increasesalmost linearly with the increase of the heat flux. When the heatflux is up to 132W/cm2, the surface temperature can bemaintainedbelow 52.0 �C. Lin and Ponnappa [26] used methanol as the coolantin a closed loop, and the surface temperature is nearly 78 �C whenthe heat flux is 130W/cm2. Although the latent heat of R600a is lessthan methanol, and its boiling point is much lower than methanolat the same pressure, it can obtain lower surface temperature atsame heat flux. For the surface temperature standard deviation,Fig. 5(b) illustrates that the increases of heat fluxwill make the heattransfer more severely and the droplets move more fiercely. Thedisorder degree of droplet movement increased, which enhancedthe surface temperature difference accordingly. Fig. 5(c) indicatesthat the heat transfer coefficient has a little improvement as theincrease of the heat flux. According to Eq. (2), the linear increase ofthe heat flux will inevitably cause linear increases of the heattransfer coefficient and the temperature difference. Fig. 5(a) showsthat the slope of the curve increases with the heat flux improve-ment. It indicates that the surface temperature increases more andmore quickly. Therefore the heat transfer coefficient should

Fig. 5. Effects of heat flux. (a) Surface temperature; (b) surface temperature standarddeviation; (c) heat transfer coefficient.

Fig. 6. Effects of chamber pressure. (a) Surface temperature; (b) heat transfer coeffi-cient; (c) surface temperature standard deviation.

H. Xu et al. / International Journal of Thermal Sciences 86 (2014) 21e2726

improve slowly. It should be highlighted that the heat transfercoefficient is up to 35,000W/(m2 �C), which is more than two timesof the water force convection coefficient and far higher than thevalue (5596 W/(m2 �C)) of R134a in Yan's [17] report. Based on theresult, it can be concluded that this spray cooling system can satisfymost heat removal requirement, especially for high-power elec-tronic products.

3.3. Effects of the chamber pressure

Compared with the open loop, this system has an obviousadvantage that the evaporation temperature can be changed bycontrolling the chamber pressure. The character can meet theneeds of lower surface temperature. The nozzle inlet pressure iskept at 3.4 bar, and the heat flux is 110 W/cm2. By regulating thefrequency of the inverter compressor, the chamber pressure ischanged from 1.8 bar to 2.2 bar, which also causes the evaporationtemperature changing from 3.8 �C to 10.3 �C. Fig. 6 reflects the

H. Xu et al. / International Journal of Thermal Sciences 86 (2014) 21e27 27

effect of the chamber pressure on the heat surface temperature, thetemperature deviation and the heat transfer coefficient. It can beseen that the surface temperature increased from 55.2 �C to 59.0 �Cby increasing the chamber pressure from 1.8 bar to 2.2 bar fromFig. 6(a). The increase of the chamber pressure leads to thedecrease of the difference between the saturation temperature andthe surface temperature. However, the increase of the surfacetemperature (from 55.2 �C to 59.0 �C) is less than the increase ofthe saturation temperature (from 3.8 �C to 10.3 �C). These phe-nomena indicate that the increase of chamber pressure is good atthe heat transfer, which can be found in Fig. 6(b). With the increaseof the chamber pressure, the heat transfer coefficient is enhanced(from 19,000 W/(m2 �C) to 23,000 W/(m2 �C)) nearly 4000 W/(m2 �C). The reason is that the increase of the chamber pressurewill decrease the pressure drop of the nozzle and the mass flowrate, which will weaken the forced convection and strengthen thephase change. The heat transfer coefficient improves inevitably,and the capacity of heat removal increases synchronously. To meetthe condition of heat removal, appropriate increasing the evapo-ration pressure can enhance the heat transfer coefficient. Thegreater heat flux will show these phenomena more obvious.Fig. 6(c) illustrates that the increase of chamber pressure has littleeffect on the temperature standard deviation, which is similar withother report [17].

4. Conclusions

A spray cooling systemwith refrigeration cycle was constructedwith the coolant of R600a to improve the heat transfer perfor-mance. The system not only simplifies the spray cooling system butalso makes the chamber pressure be adjusted easily, which can getlower surface temperature. By cooling the simulated heat surface, itis investigated the effects on the heat transfer performance by themass flow rate, the heat flux, the chamber pressure, and the nozzleinlet pressure. The heat transfer coefficient of up to 35,000 W/(m2 �C) is obtained, and a better uniform surface temperaturedistribution is achieved to keep the surface temperature standarddeviation less than 4.0 �C. Therefore, the spray cooling is a verypromising technology for high-power devices cooling, and thecombination of the refrigeration system and the spray system,especially R600a used as coolant will enhance its application.

Acknowledgement

The authors appreciate the financial supports from both theNational Natural Science Foundation of China (no. 51106170) andthe National “Twelfth Five-Year” Plan for Science & TechnologySupport of China (2012BAA13B00) for the work reported in thispaper.

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