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Applied Thermal Engineering 33-34 (2012) 246e252

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Applied Thermal Engineering

journal homepage: www.elsevier .com/locate/apthermeng

Development and experimental investigation of a novel spray cooling systemintegrated in refrigeration circuit

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

a Technical Institute of Physics and Chemistry, Chinese Academy of Sciences, P.O. Box 2711, Beijing 100190, PR ChinabGraduate University of Chinese Academy of Sciences, Beijing 100049, PR China

a r t i c l e i n f o

Article history:Received 6 April 2011Accepted 4 October 2011Available online 14 October 2011

Keywords:Spray coolingRefrigerationHeat transfer

* Corresponding author. Tel./fax: þ86 10 82543696E-mail address: chqtian@mail.ipc.ac.cn (T. Changq

1359-4311/$ e see front matter � 2011 Elsevier Ltd.doi:10.1016/j.applthermaleng.2011.10.005

a b s t r a c t

A novel spray cooling system integrated in the refrigeration circuit is proposed and its performance isinvestigated experimentally. In this system, the inverter compressor is used to replace the pump incommon spray systems, the nozzle plays the role of atomization and throttling, and the spray chamberhas function of the evaporator. The nozzle inlet pressure, the evaporation pressure and the degree ofsubcooling at nozzle inlet are all adjusted to testify the performance of the novel system in experiments.With 60 W/cm2 heat flux, the heat transfer coefficient observed is higher than 30 000 W/m2 K. Thecritical heat flux up to 110 W/cm2 is obtained, and heater surface temperature is only 31.5 �C under theheat load. Keeping the nozzle inlet pressure (Pin ¼ 390 kPa), the evaporation pressure (Pe ¼ 180 kPa) andthe heat flux (q ¼ 72W/cm2) constant, the experimental results show that the optimal subcooling degreeis 5.8 �C. The novel spray cooling system developed in this paper has simple structure and convenientregulation, and its performance can meet the heat removal requirements of most electronic devices inactual applications.

� 2011 Elsevier Ltd. All rights reserved.

1. Introduction

With the increase in requirements for heat removal in fields ofelectronics and aerospace, spray cooling as an effective coolingtechnology is gaining more attention and applications. By takingthe advantage of the working fluid’s high latent heat, spray coolingcan achieve high heat fluxes (up to 1200 W/cm2) [1] at low surfacesuperheat as well as high heat transfer coefficient with low massflow rate. In a high pressure die casting process, spray cooling willprovide a more thermally balanced die, improve quality and reducemanufacturing costs [2]. For high power electronic devices such aslaser-diode arrays pumped lasers, spray cooling is a better way tomeet heat removal requirements on the order of several kilowattsor more. NASA has also expressed the strong desire for this heatremoval technology in recent years [3].

As an appealing choice for many cooling systems, spray coolinghas been studied by many researchers. Although the theoreticalunderstanding about heat transfer mechanism of spay cooling isstill immature due to the complex interaction of liquid and vaporphases, liquid droplet impact and phase change, a lot of valuableconclusions and principles have been gained by experiments and

.ing).

All rights reserved.

simulations. Horacek et al. [4] found that forced convection playedan important role in heat transfer at low superheat conditions andthe disturbance of droplets impingement on the liquid film wouldenhance the heat transfer. They also determined that phase changewas the major reason for heat transfer enhancement at highsuperheat condition. Rini et al. [5] found that heat transfer wasrelated to the nucleate boiling, especially the phenomenon ofsecondary nucleation, which improved the heat transfer.

In regard to the heat removal capacity, the critical heat flux(CHF) was studied by researchers. Results indicated that spraycooling yields a heat flux approximately an order of magnitudehigher than pool boiling due to reduction in resistance to vaporremoval from the heater surface. For example, when using FC-72 asthe coolant, CHF of pool boiling and spray cooling are approxi-mately 20e30 W/cm2 and 100 W/cm2 respectively [6,7]. Mudawaret al. [8] studied the effect of the distance from the orifice to theheater surface and found that CHF can be maximized when thespray impact area just inscribed the square surface of the heater [8].The effects of surface structure on CHF were investigated withsurface enhancements consisting of cubic pin fins, pyramids, andstraight fins by Silk et al. [9]. In their study CHF up to 126 W/cm2

was attained with the straight finned surface.In addition to the above mentioned investigations, different

coolants were also studied by researchers. Lin et al. [10] usedmethanol as the coolant and achieved heat fluxes of 130 W/cm2 at

Nomenclature

h heat transfer coefficient, W/(m2 K)P pressure, kPaq heat flux, W/cm2

T temperature, �Ci enthalpy, kJ/kgΔT temperature margin, �CΔx distance between two thermocouples, ml thermal conductivity, W/(mK)GWP global warming potentialODP ozone depression potential

Subscriptse evaporationin nozzle inletsub subcoolingsur surface

Receiver

Spray chamber

Pump

Heat exchanger

The open system

Pump

Nozzle

Receiver

Nozzle

Spray chamber

(Can be replaced

by vacuum pump)

Heat exchanger

Pump

Heat exchangerThe closed system

Pump

a

b

Fig. 1. Existing spray cooling system.

S. Chunqiang et al. / Applied Thermal Engineering 33-34 (2012) 246e252 247

78 �C surface temperature. PF-5060 was used in the experiments bySilk et al. [9]. A new expression with an interpolation method toconstruct the partially developed nucleate boiling curve wasdeveloped by Zhou et al. [11] in their experiments with R113.Bostanci et al. [12] sprayed ammonia on microstructured surfaces,and heat fluxes up to 500 W/cm2 (well below CHF limit) wereremoved from the heater surface. Hsieh et al. [13] used R134a asworking fluid and obtained cooling characteristics (i.e., boilingcurves) over a specific range of spray mass flux, Weber number,wall superheat and degree of subcooling.

Further studies on spray cooling have addressed inclination ofthe spray, nozzle type, droplets characteristics, volumetric flux andSauter Mean Diameter [9,14e16]. All the experiments presentedmainly focused on the impact factors of heat transfer and how toimprove heat transfer while ignoring the design of the systemconfiguration and how to adjust the parameters (i.e., mass flowrate, subcooling degree) to meet the heat removal requirements(i.e., heat flux, surface temperature) of devices. Existing experi-mental facilities reported can be divided into two types. One typethat the spray chamber connected with external environmentdirectly is open system and shown in Fig. 1 (a), and the other thatthe spray chamber isolated with external environment is closedsystem and shown in Fig. 1 (b).

In the open system, the coolant is directly sprayed on the heatersurface by the pump [1]. After heat transfer, the coolant vapor isreleased to the atmosphere, and the coolant liquid is recovered. Themain deficiencies of such a system are as follows: (1) The evapo-ration pressure cannot be adjusted and the surface temperaturemay be too high due to the high boiling point of coolant at theatmospheric pressure. If water was used as the coolant, the boilingwould not appear until the heater surface temperature was higherthan 100 �C. Most electronic devices cannot run well under thistemperature. (2) The coolant cannot be recovered completely andmuch coolant vapor would be released to the atmosphere, whichwastes the coolant and may destroy the environment. If R134a isused in the open system, it is neither economic nor environmentfriendly because its GWP is 1300. (3) Besides the spray coolingsystem, there are additional auxiliary pipes, heat exchangers andthe refrigeration system to guarantee the coolant temperature. Thisundoubtedly will increase the difficulty and complexity of thesystem.

In the closed system [9], the coolant can be used circularly.However, there are some disadvantages limiting its wide applica-tion: (1) In order to keep the evaporation pressure, it is essential to

set a vacuum pump or an additional evaporator on the spraychamber. The systemmust be sealed well, otherwise it cannot worknormally due to the break in of air. (2) It is necessary to set a pumpsupplying coolant and a refrigeration system adjusting the coolanttemperature. The complexity is not less than the open system. (3) Inthe system with the additional evaporator, the regulation of evap-oration pressure is fairly hard because the additional evaporator iscontrolled by the refrigeration system which also guarantees thetemperature of coolant.

In these two types of systems, there are a pump, a vacuumpump, a refrigeration system and several heat exchangers, whichnot only increase the cost but also make the system complex andinconvenient.

S. Chunqiang et al. / Applied Thermal Engineering 33-34 (2012) 246e252248

A new spray cooling system is proposed in this paper, whichcombines the refrigeration system and the spray cooling systemtogether. A series of experiments with different nozzle inlet pres-sure, evaporation pressure and degree of subcooling are carried outto verify the heat transfer performance of the novel system.

2. Configuration of the new system

As mentioned above, the disadvantages of existing spray coolingsystemshave limited theirwidespread applications. Therefore, a newspray cooling system integrated in a refrigeration circuit is proposed,as shown in Fig. 2. It mainly includes an inverter compressor,a condenser, a spray chamber with nozzle, a receiver and someaccessories. Compared to common refrigeration systems, the onlydifference is that the throttling device and the evaporator arereplaced by the pressure nozzle and the spray chamber, respectively.The role of the nozzle in the system is throttling and atomizing. Thecomplete cycle of the system is described as follows. Firstly, theinverter compressor exhausts high pressure and high temperaturecoolant vapor into the condenser. Then, the coolant vapor iscondensed into high pressure liquid in the condenser. In order to testthe performance, a water-cooled condenser is applied in the newsystem. The coolant from the condenser flows into the receiver andwill be used to supply the working fluid for the nozzle. The coolantfrom the receiver enters into the nozzle through the expansion valveB and is throttled and atomized by the nozzle. Then the droplets ofcoolant spray onto the heater surface. After absorbing heat from theheater surface, the vapor coolantwith excess liquidflowsback to gas-liquid separator, then flows into the compressor in the form ofsuperheat vapor for next cycle.

The characteristics of the new system can be elaborated fromthe following aspects:

2.1. Selection of coolant

Different refrigerant has its own characteristics including GWP,ODP, polarity, conductivity, and corrosion, which will directly

Fig. 2. Schematic of novel spray cooling system.

determinewhether it can beusedon the highpowerdeviceswell. Forexample, ammoniahas amazinghigh latent heat, but it is not aproperworking fluid due to its toxicity and corrosion. It also should bementioned that different kinds of coolant can be used in the novelsystem according to the actual requirements. If R22 is used asworking fluid, a pressure of higher than 1000 kPa can be supplied tothe nozzle by the compressor. If R290 is used as coolant, surfacetemperature would be much low due to its low boiling point.Isobutane as green refrigerant is used as working fluid in thefollowing performance experiments in this study. With isobutene,pressure up to 500 kPa can be attained, which satisfies the require-ments of the pressure nozzle, and low evaporation temperature canachieve lower surface temperature according to high power devices.It should bementioned that the novel system can be used under hightemperature conditions with appropriate coolant (i.e., R30 or R123can be used in the novel system, whose evaporation temperature isless than 40 �C and the condensation temperature is more than80 �C).

2.2. Regulation and controlling

A good cooling system must have proper regulation andcontrolling to adapt to the fluctuation of heat load, environmentalchange and requirements of surface temperature. The novel systemcan adjust the nozzle inlet pressure, evaporation pressure, massflow rate and degree of subcooling. The nozzle inlet pressure can becontrolled by the frequency of compressor and the cooling watertemperature in the condenser. The nozzle inlet pressure can be alsoadjusted by regulating the fan speed if the condenser is an aircooled one. The evaporation pressure varies simultaneouslyaccording with the frequency changes of compressor. Increasing ordecreasing the frequency of compressor will achieve the variationof mass flow rate correspondingly. The mass flow rate also can beregulated by adjusting valve A (see Fig. 2) opening in this system.The coolant degree of subcooling can be controlled by adjusting theopening of expansion valve A.

3. Experimental setup and procedures

As shown in Fig. 2, several measurement points are set on thesystem to analyze the system performance. In this study, aninverter rotary compressor is used, whose cooling capacity is1500Wat the rating condition. A plate heat exchanger is applied asthe water-cooled condenser in the test system. The condensationpressure and the nozzle inlet pressure can be controlled byadjusting the mass flow rate and temperature of the cooling water.The expansion valve A and the subcooler are used to adjust thedegree of subcooling. In this paper, the degree of subcooling is thedifference between the coolant saturation temperature corre-sponding to the nozzle inlet pressure and the coolant temperatureof liquid into nozzle. The coolant state at the nozzle inlet is alsoregulated by expansion valve B in front of the nozzle. The valve Acontrols the mass flow rate of the nozzle by changing its opening. Afull-cone pressure nozzle with spray cone angle 60� is installed inthe system, and its Sauter Mean Diameter of droplets is approxi-mately 60e90 mm.

The heater is made of copper, and its structure is shown in Fig. 3.The heater surface area is p � 0.62 cm2. The copper heater is wellinsulated by packaging insulation material and electrically heatedby five cartridge heaters, which can generate 1100 W of heat load.According to Wang’s simulation results, the heat conduction of theheater unit can be seen as one-dimensional case [17]. Therefore, theheat conduction in the other directions will not be considered inthe following calculation. Three T-type thermocouples witha diameter of 0.5 mm are embedded beneath the heat surface to

25

30

35

40

45

50

300 350 400 450 500 P in (kPa)

T su

r ( o C

)

P e =205 kPa T in =21.5 o C

A B

C

q =50 W/cm q =60 W/cm q =85 W/cm

2 2

2

h (W

/m 2 K

)

15000

19000

23000

27000

31000

35000

3 0 0 40 0 50 0 6 0 0 7 0 0

q =85 W/cm 2 q =60 W/cm 2 q =50 W/cm 2

P e =205 kPa T in =21.5 o

P in (kPa)

C

a

b

Fig. 4. Effect of nozzle inlet pressure on heat transfer.

Fig. 3. Heater unit (partial sectional view).

S. Chunqiang et al. / Applied Thermal Engineering 33-34 (2012) 246e252 249

measure the surface temperature (see Fig. 3). The three thermo-couples spaces 3 mm apart from each other along the center axis ofthe copper heater, and the distance from the heater surface to thefirst thermocouple is also 3 mm. A computer, a data logger, twopressure sensors, a flow meter and a few thermocouples composethe data acquisition system together.

The heat flux on the heat surface is calculated by Fourier HeatConduction Law as the following equation:

q/ ¼ �l$grad T (1)

For one-dimensional case, it can be written as

q ¼ lDTDx

(2)

In the experiments, DT represents temperature differencebetween measured temperatures inside heater unit and heatsurface, and Dx is distance for different measured locations (in thecase of presented heater unit Dx ¼ 3 mm). By equation (2) and themeasured values of the thermocouples shown in Fig. 3, the surfacetemperature can be calculated.

The heat transfer coefficient is calculated based on the followingdefinition:

h ¼ qTsur � Tsat

(3)

The heater surface temperature and the heat flux are allcomputed according to Fourier Heat Conduction Law by themeasured temperature gradient to the surface from the thermo-couples in the heater. The accuracy of the thermocouple is �0.3 �Cin this experiment. The range of pressure sensor is 0w10 bar, and itsaccuracy is 0.25% for the full scale. The Coriolis flowmeter with therange of 0e40 kg/h has a measurement accuracy of 0.25% for totalmass flow rate. According to the traditional method of error anal-ysis, the uncertainty of the heat fluxes, the surface temperature andthe heat transfer coefficient are 3.3%, 4.5% and 2.7%, respectively.

4. Experimental results and discussion

In the following experiments, the nozzle inlet pressure, evapo-ration pressure and degree of subcooling are all regulated respec-tively, and their effects on heater surface temperature and heat

transfer coefficient are analyzed with different experiments. At last,CHF is further studied by experiments. The surface temperature,heat transfer coefficient and CHF are mainly three physical vari-ables to reflect the performance of the system. The surfacetemperature reflects whether the system meets the temperaturerequirements and heat transfer coefficient represent the heattransfer capacity. CHF indicates maximum heat removal capacity ofthe system and make sure whether the system meets the heatremoval requirements of high power devices.

4.1. Nozzle inlet pressure

Nozzle inlet pressure affects atomization (including droplet sizeand distribution) and the mass flow rate, and all the effects willfurther impact the heat removal process and be reflected by heatersurface temperature and heat transfer coefficient. With differentheat flux, the effects of the nozzle inlet pressure on the heatersurface temperature and heat transfer coefficient are analyzed. Thethree curves in Fig. 4 (a) all indicate that the heater surfacetemperature reduces with increasing the nozzle inlet pressure. Thetrend is especially obvious when the heat flux is much higher.When the heat flux is 60 W/cm2 and the evaporation pressure is205 kPa (Te ¼ 7.8 �C), the heater surface temperature reduces from32.4 �C to 29.5 �C with the nozzle inlet pressure increasing from330 kPa to 477 kPa (from point A to point B). The heater surfacetemperature decreases 2.9 �C, and the nozzle inlet pressureincreases approximately 150 kPa. Based on the two curves(q ¼ 50 W/cm2 & 60 W/cm2), it can be determined that changingthe nozzle inlet pressure is an effective way to get the same heater

Fig. 5. Effect of evaporation pressure on heat transfer.

S. Chunqiang et al. / Applied Thermal Engineering 33-34 (2012) 246e252250

surface temperature under different heat fluxes. The 29.5 �C heatersurface temperature can be obtained with the heat fluxes from 50to 60 W/cm2 by changing the nozzle inlet pressure from 410 kPa to477 kPa (from point C to point B). Fig. 4 (b) indicates that theincreasing of the nozzle inlet pressure plays a positive role on theimprovement of heat transfer coefficient. When the heat flux ishigh, the effect is more obvious due to the enhancement of atom-ization, whichmakes theworking fluid evaporate sufficiently. It canbe seen that the curve of heat transfer coefficient is nearly linearwhen the heat flux is 50 W/cm2. The convection is mainly way ofheat transfer when the heat flux is low. With the increasing heatflux, the role of boiling is also more important besides convection.The curve of heat transfer coefficient turns into curved when theheat flux is higher. Improving the heat transfer coefficient willdecrease the area of heater surface at constant heat flux, which willbenefit to the actual application. In this system, the nozzle inletpressure can be changed easily and the largest heat transfer coef-ficient observed is higher than 30 000 W/m2 K, which is muchhigher than that achieved by Zhou et al. [11] in their experimentswith R113.

4.2. Evaporation pressure and temperature

In the new system, evaporation temperature changes with theevaporation pressure increasing or decreasing. Changing the evap-oration temperature will directly affect the surface superheat andthe surface temperature. The heat transfer coefficient also changesalong with the evaporation temperature. The heater surfacetemperature changing trend according with the evaporation pres-sure is obtained with different heat flux (Fig. 5 (a)) while keepingthe nozzle inlet pressure at 390 kPa and the coolant degree ofsubcooling at 7.2 �C. Results indicate that the heater surfacetemperature increases with the increase of the evaporation pres-sure. When the heat flux is 65 W/cm2, changing the evaporationpressure from 211 kPa (evaporation temperature Te ¼ 8.7 �C) to161 kPa (Te ¼ 0.8 �C), results in a variation of the heater surfacetemperature from 34.6 �C to 26.5 �C (from point A to point B). Thetemperature on the heater surface decreases nearly 8 �C due to thepressure reduction with 50 kPa. This means that changing theevaporation pressure can provide the same heater surface temper-ature with different heat fluxes. When the heat flux is 65, 55 and45W/cm2, a heater surface temperature of 30 �C can be achieved bychanging the evaporation pressure from 178 kPa to 206 kPa (Point Cto Point E). The main reason for this is that the evaporationtemperature reduces according to the reduction of the evaporationpressure. Comparedwith the increasing of nozzle inlet pressure, thereduction in evaporation temperature is more effective in reducingthe heater surface temperature. The relation between the heattransfer coefficient and the evaporation pressure is studied as well,which is shown in Fig. 5 (b). The plot indicates that the heat transfercoefficient declines with an increase of evaporation pressure. Thereare two reasons for this phenomenon. One is that the increasing ofthe evaporation pressure reduces the latent heat of coolant, and theother is that increase of evaporation pressure delays the appearanceof boiling due to the increasing of evaporation temperatureaccordingly. When the mass flow rate is constant under high heatflux, the superheat will increase due to the decreasing of heattransfer coefficient made by the decrease of latent heat. Comparingthree curves, the effect of the evaporation pressure on the heattransfer coefficient is more obvious, especially when the heat flux ishigher. It can be seen that the curve is nearly linear when the heatflux is 90W/cm2. When the heat flux is high, the boiling effect playsan important role in heat transfer. The increasing of evaporationpressure will directly affect the boiling. Therefore, heat transfercoefficient decreases quickly and the trend is nearly linear. When

the heat flux is low, the effect of decreasing evaporation pressure isnot obvious on heat transfer because the convection is mainly wayof heat transfer as mentioned above. Thus, the trend of heat transfercoefficient is different with different heat flux.

4.3. Degree of subcooling

In the new system, the degree of subcooling at the inlet of nozzleis adjusted by changing the opening of expansion valve A. Theeffects of subcooling on heat transfer are investigated by regulatingdegree of subcooling gradually. As shown in Fig. 6 (a), whenkeeping the nozzle inlet pressure (Pin ¼ 390 kPa), the evaporationpressure (Pe ¼ 180 kPa) and the heat flux (q ¼ 72 W/cm2) constant,the heater surface temperature first decreases and then increaseswhen the subcooling degree of the liquid coolant (from point A topoint C) decreases. This variation means that there is an optimalsubcooling degree of 5.8 �C in the experiment, which makes theboiling heat transfer stronger. This can be explained from thediagram of pressure-enthalpy shown in Fig. 6 (b). The throttlingprocess of the nozzle can be described by the lines AD, BE and CFrespectively (each having different subcooling degrees). It is knownthat the throttling effect of the nozzle is identified when the nozzleinlet pressure and the evaporation pressure are kept as constant.The difference in the state of the coolant at points D, E, F is thevapor quality. The coolant vapor quality at point D is the minimum,and that of point F is the maximum. Combined with Fig. 6 (a), it canbe concluded that the appropriate vapor quality (point E) benefitsfor the heat transfer. For a spray, a certain vapor quality (vapor-eliquid mass ratio) can strengthen the atomization. The vapor has

30

31

32

33

34

35

0 3 6 9 12 15

P in =390 kPa

P e =180 kPa

q =72 W/cm 2

T sub ( C)

T su

r ( o C

)

B

C

o

A

P in =390kPa

P =180kPa

A B

D E F i (kJ/kg)

e

C

P (kPa)

a

b

Fig. 6. Effect of degree of subcooling on heat transfer.

S. Chunqiang et al. / Applied Thermal Engineering 33-34 (2012) 246e252 251

the role of assisting the atomization, which is similar to air assistatomization [18]. At point E, the quality of vapor is fit to the liquid tocomplete atomization, and the droplets absorb heat from theheater surface sufficiently. It is not far to see that the optimal degreeof subcooling is important to the heat transfer. However, theoptimal subcooling may be different for different conditions andnozzles. Hence, it is necessary to regulate the subcooling for theactual applications. It also should be noticed that the optimalmechanism and vaporeliquid mass ratio need further study for thisnew system. In the new system, the degree of subcooling can beadjusted by changing opening of expansion valve A. The easily wayto change subcooling makes the new system have advantages forapplication compared to the traditional spray cooling system.

4.4. CHF of the new system

The critical heat flux (CHF) reflects the maximum heat transfercapacity, and the value of CHF directly decides whether the system

Fig. 7. Measured curve of heat flux.

can meet the requirement of heat removal for high power devices.Keeping the evaporation pressure (Pe ¼ 177 kPa) and the nozzleinlet pressure (Pin ¼ 285 kPa) constant, the heat flux is increasedgradually until the rapid increase of surface temperature appears.The results are shown in Fig. 7. The surface temperature increasesgradually according with the enhancement of heat flux. When theheat flux is up to approximately 110 W/cm2, the surface tempera-ture increases acutely. This value is CHF, which reflects that the heattransfer departs from nucleate boiling. It represents the maximumheat transfer capacity of the system under these conditions. Itshould be highlighted that the heater surface temperature is only31.5 �C when the heat flux is up to 110 W/cm2. Most electronicdevices can work normally when the temperature is below 55 �C,and their heat removal is less than 100 W/cm2. This means thenovel system can meet the heat removal demand of most devices.

5. Conclusions

In order to satisfy the heat removal requirements for applica-tions and improve heat transfer capacity, a novel spray coolingsystem was constructed. The new system incorporated a refrigera-tion circuit, the evaporator was replaced by the spray chamberconsisting of a nozzle. The nozzle inlet pressure and the evapora-tion pressure can be adjusted according to desired level of heatremoval by changing the frequency of compressor. The optimaldegree of subcooling can be obtained as well according to theopening of expansion valve A. Based on the new system, a series ofexperiments were carried out to investigate the effects of pressure,degree of subcooling on the system performance. The resultsshowed that the heat transfer coefficient up to 30 000 W/cm2 wasobtained, and the heater surface temperature can be kept below30 �C when the heat flux is 50 W/cm2. The optimal degree ofsubcooling is 5.8 �C in this experiment. The stable operation, theconvenient adjustment, the simple structure and the betterperformance of heat dissipation reflect that the novel system haspromising applications in high power devices.

The way integrated the novels with electronic devices will bestudied in the future in order to make the system more flexible torespond effectively to rapidly changing heat flux scenarios [19,20].In addition, further research is needed for the influence of lubricanton the heat transfer in the novel spray cooling system.

Acknowledgement

The authors appreciate the financial supports from the NationalNatural Science Foundation of China (No. 51106170) for the workreported in this paper.

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