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Development and experimental investigation of a novel spray cooling system integrated in refrigeration circuit
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 system integrated in refrigeration circuit Si Chunqiang a, b, Shao Shuangquan a, Tian Changqing a, *, Xu Hongbo a a b
Technical Institute of Physics and Chemistry, Chinese Academy of Sciences, P.O. Box 2711, Beijing 100190, PR China Graduate University of Chinese Academy of Sciences, Beijing 100049, PR China
a r t i c l e i n f o
a b s t r a c t
Article history: Received 6 April 2011 Accepted 4 October 2011 Available online 14 October 2011
A novel spray cooling system integrated in the refrigeration circuit is proposed and its performance is investigated experimentally. In this system, the inverter compressor is used to replace the pump in common spray systems, the nozzle plays the role of atomization and throttling, and the spray chamber has function of the evaporator. The nozzle inlet pressure, the evaporation pressure and the degree of subcooling 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. The critical heat flux up to 110 W/cm2 is obtained, and heater surface temperature is only 31.5 C under the heat load. Keeping the nozzle inlet pressure (Pin ¼ 390 kPa), the evaporation pressure (Pe ¼ 180 kPa) and the heat flux (q ¼ 72 W/cm2) constant, the experimental results show that the optimal subcooling degree is 5.8 C. The novel spray cooling system developed in this paper has simple structure and convenient regulation, and its performance can meet the heat removal requirements of most electronic devices in actual applications. Ó 2011 Elsevier Ltd. All rights reserved.
Keywords: Spray cooling Refrigeration Heat transfer
1. Introduction With the increase in requirements for heat removal in fields of electronics and aerospace, spray cooling as an effective cooling technology is gaining more attention and applications. By taking the advantage of the working fluid’s high latent heat, spray cooling can achieve high heat fluxes (up to 1200 W/cm2) [1] at low surface superheat as well as high heat transfer coefficient with low mass flow rate. In a high pressure die casting process, spray cooling will provide a more thermally balanced die, improve quality and reduce manufacturing costs [2]. For high power electronic devices such as laser-diode arrays pumped lasers, spray cooling is a better way to meet heat removal requirements on the order of several kilowatts or more. NASA has also expressed the strong desire for this heat removal technology in recent years [3]. As an appealing choice for many cooling systems, spray cooling has been studied by many researchers. Although the theoretical understanding about heat transfer mechanism of spay cooling is still immature due to the complex interaction of liquid and vapor phases, liquid droplet impact and phase change, a lot of valuable conclusions and principles have been gained by experiments and
* Corresponding author. Tel./fax: þ86 10 82543696. E-mail address: [email protected] (T. Changqing). 1359-4311/$ e see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2011.10.005
simulations. Horacek et al. [4] found that forced convection played an important role in heat transfer at low superheat conditions and the disturbance of droplets impingement on the liquid film would enhance the heat transfer. They also determined that phase change was the major reason for heat transfer enhancement at high superheat condition. Rini et al. [5] found that heat transfer was related to the nucleate boiling, especially the phenomenon of secondary 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 spray cooling yields a heat flux approximately an order of magnitude higher than pool boiling due to reduction in resistance to vapor removal from the heater surface. For example, when using FC-72 as the coolant, CHF of pool boiling and spray cooling are approximately 20e30 W/cm2 and 100 W/cm2 respectively [6,7]. Mudawar et al. [8] studied the effect of the distance from the orifice to the heater surface and found that CHF can be maximized when the spray impact area just inscribed the square surface of the heater [8]. The effects of surface structure on CHF were investigated with surface enhancements consisting of cubic pin fins, pyramids, and straight 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] used methanol as the coolant and achieved heat fluxes of 130 W/cm2 at
S. Chunqiang et al. / Applied Thermal Engineering 33-34 (2012) 246e252
Nomenclature h P q T i ΔT Δx
l
GWP ODP
a
heat transfer coefficient, W/(m2 K) pressure, kPa heat flux, W/cm2 temperature, C enthalpy, kJ/kg temperature margin, C distance between two thermocouples, m thermal conductivity, W/(mK) global warming potential ozone depression potential
Subscripts e evaporation in nozzle inlet sub subcooling sur surface
78 C surface temperature. PF-5060 was used in the experiments by Silk et al. [9]. A new expression with an interpolation method to construct the partially developed nucleate boiling curve was developed 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) were removed from the heater surface. Hsieh et al. [13] used R134a as working fluid and obtained cooling characteristics (i.e., boiling curves) over a specific range of spray mass flux, Weber number, wall superheat and degree of subcooling. Further studies on spray cooling have addressed inclination of the spray, nozzle type, droplets characteristics, volumetric flux and Sauter Mean Diameter [9,14e16]. All the experiments presented mainly focused on the impact factors of heat transfer and how to improve heat transfer while ignoring the design of the system configuration and how to adjust the parameters (i.e., mass flow rate, subcooling degree) to meet the heat removal requirements (i.e., heat flux, surface temperature) of devices. Existing experimental facilities reported can be divided into two types. One type that the spray chamber connected with external environment directly is open system and shown in Fig. 1 (a), and the other that the spray chamber isolated with external environment is closed system and shown in Fig. 1 (b). In the open system, the coolant is directly sprayed on the heater surface by the pump [1]. After heat transfer, the coolant vapor is released to the atmosphere, and the coolant liquid is recovered. The main deficiencies of such a system are as follows: (1) The evaporation pressure cannot be adjusted and the surface temperature may be too high due to the high boiling point of coolant at the atmospheric pressure. If water was used as the coolant, the boiling would not appear until the heater surface temperature was higher than 100 C. Most electronic devices cannot run well under this temperature. (2) The coolant cannot be recovered completely and much coolant vapor would be released to the atmosphere, which wastes the coolant and may destroy the environment. If R134a is used in the open system, it is neither economic nor environment friendly because its GWP is 1300. (3) Besides the spray cooling system, there are additional auxiliary pipes, heat exchangers and the refrigeration system to guarantee the coolant temperature. This undoubtedly will increase the difficulty and complexity of the system. In the closed system [9], the coolant can be used circularly. However, there are some disadvantages limiting its wide application: (1) In order to keep the evaporation pressure, it is essential to
247
Nozzle Spray chamber
Receiver
Pump Pump
Heat exchanger The open system
b
Nozzle Heat exchanger (Can be replaced by vacuum pump)
Spray chamber
Receiver
Pump Pump
Heat exchanger The closed system Fig. 1. Existing spray cooling system.
set a vacuum pump or an additional evaporator on the spray chamber. The system must be sealed well, otherwise it cannot work normally due to the break in of air. (2) It is necessary to set a pump supplying coolant and a refrigeration system adjusting the coolant temperature. The complexity is not less than the open system. (3) In the system with the additional evaporator, the regulation of evaporation pressure is fairly hard because the additional evaporator is controlled by the refrigeration system which also guarantees the temperature of coolant. In these two types of systems, there are a pump, a vacuum pump, a refrigeration system and several heat exchangers, which not only increase the cost but also make the system complex and inconvenient.
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A new spray cooling system is proposed in this paper, which combines the refrigeration system and the spray cooling system together. A series of experiments with different nozzle inlet pressure, evaporation pressure and degree of subcooling are carried out to verify the heat transfer performance of the novel system. 2. Configuration of the new system As mentioned above, the disadvantages of existing spray cooling systems have limited their widespread applications. Therefore, a new spray 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 some accessories. Compared to common refrigeration systems, the only difference is that the throttling device and the evaporator are replaced by the pressure nozzle and the spray chamber, respectively. The role of the nozzle in the system is throttling and atomizing. The complete cycle of the system is described as follows. Firstly, the inverter compressor exhausts high pressure and high temperature coolant vapor into the condenser. Then, the coolant vapor is condensed into high pressure liquid in the condenser. In order to test the performance, a water-cooled condenser is applied in the new system. The coolant from the condenser flows into the receiver and will be used to supply the working fluid for the nozzle. The coolant from the receiver enters into the nozzle through the expansion valve B and is throttled and atomized by the nozzle. Then the droplets of coolant spray onto the heater surface. After absorbing heat from the heater surface, the vapor coolant with excess liquid flows back to gasliquid separator, then flows into the compressor in the form of superheat vapor for next cycle. The characteristics of the new system can be elaborated from the following aspects: 2.1. Selection of coolant Different refrigerant has its own characteristics including GWP, ODP, polarity, conductivity, and corrosion, which will directly
determine whether it can be used on the high power devices well. For example, ammonia has amazing high latent heat, but it is not a proper working fluid due to its toxicity and corrosion. It also should be mentioned that different kinds of coolant can be used in the novel system according to the actual requirements. If R22 is used as working fluid, a pressure of higher than 1000 kPa can be supplied to the nozzle by the compressor. If R290 is used as coolant, surface temperature would be much low due to its low boiling point. Isobutane as green refrigerant is used as working fluid in the following performance experiments in this study. With isobutene, pressure up to 500 kPa can be attained, which satisfies the requirements of the pressure nozzle, and low evaporation temperature can achieve lower surface temperature according to high power devices. It should be mentioned that the novel system can be used under high temperature conditions with appropriate coolant (i.e., R30 or R123 can be used in the novel system, whose evaporation temperature is less than 40 C and the condensation temperature is more than 80 C). 2.2. Regulation and controlling A good cooling system must have proper regulation and controlling to adapt to the fluctuation of heat load, environmental change and requirements of surface temperature. The novel system can adjust the nozzle inlet pressure, evaporation pressure, mass flow rate and degree of subcooling. The nozzle inlet pressure can be controlled by the frequency of compressor and the cooling water temperature in the condenser. The nozzle inlet pressure can be also adjusted by regulating the fan speed if the condenser is an air cooled one. The evaporation pressure varies simultaneously according with the frequency changes of compressor. Increasing or decreasing the frequency of compressor will achieve the variation of mass flow rate correspondingly. The mass flow rate also can be regulated by adjusting valve A (see Fig. 2) opening in this system. The coolant degree of subcooling can be controlled by adjusting the opening of expansion valve A. 3. Experimental setup and procedures
Fig. 2. Schematic of novel spray cooling system.
As shown in Fig. 2, several measurement points are set on the system to analyze the system performance. In this study, an inverter rotary compressor is used, whose cooling capacity is 1500 W at the rating condition. A plate heat exchanger is applied as the water-cooled condenser in the test system. The condensation pressure and the nozzle inlet pressure can be controlled by adjusting the mass flow rate and temperature of the cooling water. The expansion valve A and the subcooler are used to adjust the degree of subcooling. In this paper, the degree of subcooling is the difference between the coolant saturation temperature corresponding to the nozzle inlet pressure and the coolant temperature of liquid into nozzle. The coolant state at the nozzle inlet is also regulated by expansion valve B in front of the nozzle. The valve A controls the mass flow rate of the nozzle by changing its opening. A full-cone pressure nozzle with spray cone angle 60 is installed in the system, and its Sauter Mean Diameter of droplets is approximately 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 well insulated by packaging insulation material and electrically heated by five cartridge heaters, which can generate 1100 W of heat load. According to Wang’s simulation results, the heat conduction of the heater unit can be seen as one-dimensional case [17]. Therefore, the heat conduction in the other directions will not be considered in the following calculation. Three T-type thermocouples with a diameter of 0.5 mm are embedded beneath the heat surface to
S. Chunqiang et al. / Applied Thermal Engineering 33-34 (2012) 246e252
249
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 variables to reflect the performance of the system. The surface temperature reflects whether the system meets the temperature requirements and heat transfer coefficient represent the heat transfer capacity. CHF indicates maximum heat removal capacity of the system and make sure whether the system meets the heat removal requirements of high power devices. 4.1. Nozzle inlet pressure
Fig. 3. Heater unit (partial sectional view).
measure the surface temperature (see Fig. 3). The three thermocouples spaces 3 mm apart from each other along the center axis of the copper heater, and the distance from the heater surface to the first thermocouple is also 3 mm. A computer, a data logger, two pressure sensors, a flow meter and a few thermocouples compose the data acquisition system together. The heat flux on the heat surface is calculated by Fourier Heat Conduction Law as the following equation: /
q ¼ l$grad T
(1)
Nozzle inlet pressure affects atomization (including droplet size and distribution) and the mass flow rate, and all the effects will further impact the heat removal process and be reflected by heater surface temperature and heat transfer coefficient. With different heat flux, the effects of the nozzle inlet pressure on the heater surface temperature and heat transfer coefficient are analyzed. The three curves in Fig. 4 (a) all indicate that the heater surface temperature reduces with increasing the nozzle inlet pressure. The trend is especially obvious when the heat flux is much higher. When the heat flux is 60 W/cm2 and the evaporation pressure is 205 kPa (Te ¼ 7.8 C), the heater surface temperature reduces from 32.4 C to 29.5 C with the nozzle inlet pressure increasing from 330 kPa to 477 kPa (from point A to point B). The heater surface temperature decreases 2.9 C, and the nozzle inlet pressure increases approximately 150 kPa. Based on the two curves (q ¼ 50 W/cm2 & 60 W/cm2), it can be determined that changing the nozzle inlet pressure is an effective way to get the same heater
a
50
Pe =205 kPa Tin =21.5 oC
For one-dimensional case, it can be written as
45
(3)
The heater surface temperature and the heat flux are all computed according to Fourier Heat Conduction Law by the measured temperature gradient to the surface from the thermocouples in the heater. The accuracy of the thermocouple is 0.3 C in this experiment. The range of pressure sensor is 0w10 bar, and its accuracy is 0.25% for the full scale. The Coriolis flow meter with the range of 0e40 kg/h has a measurement accuracy of 0.25% for total mass flow rate. According to the traditional method of error analysis, the uncertainty of the heat fluxes, the surface temperature and the 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, evaporation pressure and degree of subcooling are all regulated respectively, and their effects on heater surface temperature and heat
T sur (oC)
In the experiments, DT represents temperature difference between measured temperatures inside heater unit and heat surface, and Dx is distance for different measured locations (in the case of presented heater unit Dx ¼ 3 mm). By equation (2) and the measured values of the thermocouples shown in Fig. 3, the surface temperature can be calculated. The heat transfer coefficient is calculated based on the following definition:
q h ¼ Tsur Tsat
q =50 W/cm2 q =60 W/cm2 q =85 W/cm2
(2)
40 35
A
B
30
C
25 300
b
350
400
450
500
600
700
Pin (kPa)
35000
q =50 W/cm2 Pe=205 kPa q =60 W/cm2 =21.5 oC q =85 W/cm2 T in
31000
h (W/m 2 K)
DT Dx
q ¼ l
27000 23000 19000 15000 300
400
500
P in (kPa) Fig. 4. Effect of nozzle inlet pressure on heat transfer.
250
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surface temperature under different heat fluxes. The 29.5 C heater surface temperature can be obtained with the heat fluxes from 50 to 60 W/cm2 by changing the nozzle inlet pressure from 410 kPa to 477 kPa (from point C to point B). Fig. 4 (b) indicates that the increasing of the nozzle inlet pressure plays a positive role on the improvement of heat transfer coefficient. When the heat flux is high, the effect is more obvious due to the enhancement of atomization, which makes the working fluid evaporate sufficiently. It can be seen that the curve of heat transfer coefficient is nearly linear when the heat flux is 50 W/cm2. The convection is mainly way of heat transfer when the heat flux is low. With the increasing heat flux, the role of boiling is also more important besides convection. The curve of heat transfer coefficient turns into curved when the heat flux is higher. Improving the heat transfer coefficient will decrease the area of heater surface at constant heat flux, which will benefit to the actual application. In this system, the nozzle inlet pressure can be changed easily and the largest heat transfer coefficient observed is higher than 30 000 W/m2 K, which is much higher than that achieved by Zhou et al. [11] in their experiments with R113. 4.2. Evaporation pressure and temperature In the new system, evaporation temperature changes with the evaporation pressure increasing or decreasing. Changing the evaporation temperature will directly affect the surface superheat and the surface temperature. The heat transfer coefficient also changes along with the evaporation temperature. The heater surface temperature changing trend according with the evaporation pressure is obtained with different heat flux (Fig. 5 (a)) while keeping the nozzle inlet pressure at 390 kPa and the coolant degree of subcooling at 7.2 C. Results indicate that the heater surface temperature increases with the increase of the evaporation pressure. When the heat flux is 65 W/cm2, changing the evaporation pressure from 211 kPa (evaporation temperature Te ¼ 8.7 C) to 161 kPa (Te ¼ 0.8 C), results in a variation of the heater surface temperature from 34.6 C to 26.5 C (from point A to point B). The temperature on the heater surface decreases nearly 8 C due to the pressure reduction with 50 kPa. This means that changing the evaporation pressure can provide the same heater surface temperature with different heat fluxes. When the heat flux is 65, 55 and 45 W/cm2, a heater surface temperature of 30 C can be achieved by changing the evaporation pressure from 178 kPa to 206 kPa (Point C to Point E). The main reason for this is that the evaporation temperature reduces according to the reduction of the evaporation pressure. Compared with the increasing of nozzle inlet pressure, the reduction in evaporation temperature is more effective in reducing the heater surface temperature. The relation between the heat transfer coefficient and the evaporation pressure is studied as well, which is shown in Fig. 5 (b). The plot indicates that the heat transfer coefficient declines with an increase of evaporation pressure. There are two reasons for this phenomenon. One is that the increasing of the evaporation pressure reduces the latent heat of coolant, and the other is that increase of evaporation pressure delays the appearance of boiling due to the increasing of evaporation temperature accordingly. When the mass flow rate is constant under high heat flux, the superheat will increase due to the decreasing of heat transfer coefficient made by the decrease of latent heat. Comparing three curves, the effect of the evaporation pressure on the heat transfer coefficient is more obvious, especially when the heat flux is higher. It can be seen that the curve is nearly linear when the heat flux is 90 W/cm2. When the heat flux is high, the boiling effect plays an important role in heat transfer. The increasing of evaporation pressure will directly affect the boiling. Therefore, heat transfer coefficient decreases quickly and the trend is nearly linear. When
Fig. 5. Effect of evaporation pressure on heat transfer.
the heat flux is low, the effect of decreasing evaporation pressure is not obvious on heat transfer because the convection is mainly way of heat transfer as mentioned above. Thus, the trend of heat transfer coefficient is different with different heat flux. 4.3. Degree of subcooling In the new system, the degree of subcooling at the inlet of nozzle is adjusted by changing the opening of expansion valve A. The effects of subcooling on heat transfer are investigated by regulating degree of subcooling gradually. As shown in Fig. 6 (a), when keeping the nozzle inlet pressure (Pin ¼ 390 kPa), the evaporation pressure (Pe ¼ 180 kPa) and the heat flux (q ¼ 72 W/cm2) constant, the heater surface temperature first decreases and then increases when the subcooling degree of the liquid coolant (from point A to point C) decreases. This variation means that there is an optimal subcooling degree of 5.8 C in the experiment, which makes the boiling heat transfer stronger. This can be explained from the diagram of pressure-enthalpy shown in Fig. 6 (b). The throttling process of the nozzle can be described by the lines AD, BE and CF respectively (each having different subcooling degrees). It is known that the throttling effect of the nozzle is identified when the nozzle inlet pressure and the evaporation pressure are kept as constant. The difference in the state of the coolant at points D, E, F is the vapor 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 can be concluded that the appropriate vapor quality (point E) benefits for the heat transfer. For a spray, a certain vapor quality (vaporeliquid mass ratio) can strengthen the atomization. The vapor has
S. Chunqiang et al. / Applied Thermal Engineering 33-34 (2012) 246e252
a
35
Pin=390 kPa
Tsur (oC)
34
A
Pe=180 kPa q=72 W/cm2
33
C
32
B
31 30 0
3
6
9
12
15
Tsub (oC)
b
can meet the requirement of heat removal for high power devices. Keeping the evaporation pressure (Pe ¼ 177 kPa) and the nozzle inlet pressure (Pin ¼ 285 kPa) constant, the heat flux is increased gradually until the rapid increase of surface temperature appears. The results are shown in Fig. 7. The surface temperature increases gradually according with the enhancement of heat flux. When the heat flux is up to approximately 110 W/cm2, the surface temperature increases acutely. This value is CHF, which reflects that the heat transfer departs from nucleate boiling. It represents the maximum heat transfer capacity of the system under these conditions. It should be highlighted that the heater surface temperature is only 31.5 C when the heat flux is up to 110 W/cm2. Most electronic devices can work normally when the temperature is below 55 C, and their heat removal is less than 100 W/cm2. This means the novel system can meet the heat removal demand of most devices. 5. Conclusions
P (kPa) B A
251
Pin=390kPa
C
D EF
Pe=180kPa i (kJ/kg)
Fig. 6. Effect of degree of subcooling on heat transfer.
the role of assisting the atomization, which is similar to air assist atomization [18]. At point E, the quality of vapor is fit to the liquid to complete atomization, and the droplets absorb heat from the heater surface sufficiently. It is not far to see that the optimal degree of subcooling is important to the heat transfer. However, the optimal subcooling may be different for different conditions and nozzles. Hence, it is necessary to regulate the subcooling for the actual applications. It also should be noticed that the optimal mechanism and vaporeliquid mass ratio need further study for this new system. In the new system, the degree of subcooling can be adjusted by changing opening of expansion valve A. The easily way to change subcooling makes the new system have advantages for application compared to the traditional spray cooling system.
In order to satisfy the heat removal requirements for applications and improve heat transfer capacity, a novel spray cooling system was constructed. The new system incorporated a refrigeration circuit, the evaporator was replaced by the spray chamber consisting of a nozzle. The nozzle inlet pressure and the evaporation pressure can be adjusted according to desired level of heat removal by changing the frequency of compressor. The optimal degree of subcooling can be obtained as well according to the opening of expansion valve A. Based on the new system, a series of experiments were carried out to investigate the effects of pressure, degree of subcooling on the system performance. The results showed that the heat transfer coefficient up to 30 000 W/cm2 was obtained, and the heater surface temperature can be kept below 30 C when the heat flux is 50 W/cm2. The optimal degree of subcooling is 5.8 C in this experiment. The stable operation, the convenient adjustment, the simple structure and the better performance of heat dissipation reflect that the novel system has promising applications in high power devices. The way integrated the novels with electronic devices will be studied in the future in order to make the system more flexible to respond effectively to rapidly changing heat flux scenarios [19,20]. In addition, further research is needed for the influence of lubricant on the heat transfer in the novel spray cooling system. Acknowledgement
4.4. CHF of the new system The critical heat flux (CHF) reflects the maximum heat transfer capacity, and the value of CHF directly decides whether the system
The authors appreciate the financial supports from the National Natural Science Foundation of China (No. 51106170) for the work reported in this paper. References
Fig. 7. Measured curve of heat flux.
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[8] I. Mudawar, K.A. Estes, Optimization and predicting CHF in spray cooling of a square surface, J. Heat Transf.-Trans. ASME 118 (1996) 672e680. [9] E.A. Silk, J. Kim, K. Kiger, Spray cooling of enhanced surfaces: impact of structured surface geometry and spray axis inclination, Int. J. Heat Mass Transf. 49 (2006) 4910e4920. [10] L. Lin, R. Ponnappa, Heat transfer characteristics of spray cooling in a closed loop, Int. J. Heat Mass Transf. 46 (2003) 3737e3746. [11] D.W. Zhou, C.F. Ma, Local jet impingement boiling heat transfer with R113, Heat Mass Transf. 40 (2004) 539e549. [12] H. Bostanci, D.P. Rini, J.P. Kizito, et al., Spray cooling with ammonia on microstructured surfaces: performance enhancement and hysteresis effect, J. Heat Transf.-Trans. ASME 131 (2009) 1401e1409. [13] S.S. Hsieh, T.C. Fan, H.H. Tsai, Spray cooling characteristics of water and R134a. Part I: nucleate boiling, Int. J. Heat Mass Transf. 47 (2004) 5703e5712. [14] J. Yang, L.C. Chow, M.R. Pais, et al., Liquid film thickness and topography determination using Fresnel diffraction and holography, Exp. Heat Transf. 5 (1992) 239e252.
[15] M.R.O. Panao, A.L.N. Moreira, Heat transfer correlation for intermittent spray impingement: a dynamic approach, Int. J. Therm. Sci. 48 (2009) 1853e1862. [16] K.A. Estes, I. Mudawar, Correlation of Sauter mean diameter and critical heat flux for spray cooling of small surfaces, Int. J. Heat Mass Transf. 38 (1995) 2985e2996. [17] Y.Q. Wang, M.H. Liu, D. Liu, et al., Experiment study on non-boiling heat transfer performance in spray cooling for high-power laser, Chin. J. Lasers 36 (2009) 1973e1979 (In Chinese). [18] X. Jiang, G.A. Siamas, K. Jagus, et al., Physical modeling and advanced simulations of gas-liquid two-phase jet flows in atomization and sprays, Prog. Energy Combust. Sci. 36 (2010) 131e167. [19] S.G. Singh, S.P. Duttagupta, A. Agrawal, In-situ impact analysis of very high heat flux transients on non-linear p-n diode characteristics and mitigation using on-chip single-phase and two-phase microfluidics, J. Microelectromech S. 18 (2009) 1208e1219. [20] S.G. Singh, A. Agrawal, S.P. Duttagupta. Reliable MOSFET operation using twophase microfluidics in presence of high heat flux transients, J. Micromech. Microeng. in press.
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Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Development and experimental investigation of a novel spray cooling system integrated in refrigeration circuit Si Chunqiang a, b, Shao Shuangquan a, Tian Changqing a, *, Xu Hongbo a a b
Technical Institute of Physics and Chemistry, Chinese Academy of Sciences, P.O. Box 2711, Beijing 100190, PR China Graduate University of Chinese Academy of Sciences, Beijing 100049, PR China
a r t i c l e i n f o
a b s t r a c t
Article history: Received 6 April 2011 Accepted 4 October 2011 Available online 14 October 2011
A novel spray cooling system integrated in the refrigeration circuit is proposed and its performance is investigated experimentally. In this system, the inverter compressor is used to replace the pump in common spray systems, the nozzle plays the role of atomization and throttling, and the spray chamber has function of the evaporator. The nozzle inlet pressure, the evaporation pressure and the degree of subcooling 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. The critical heat flux up to 110 W/cm2 is obtained, and heater surface temperature is only 31.5 C under the heat load. Keeping the nozzle inlet pressure (Pin ¼ 390 kPa), the evaporation pressure (Pe ¼ 180 kPa) and the heat flux (q ¼ 72 W/cm2) constant, the experimental results show that the optimal subcooling degree is 5.8 C. The novel spray cooling system developed in this paper has simple structure and convenient regulation, and its performance can meet the heat removal requirements of most electronic devices in actual applications. Ó 2011 Elsevier Ltd. All rights reserved.
Keywords: Spray cooling Refrigeration Heat transfer
1. Introduction With the increase in requirements for heat removal in fields of electronics and aerospace, spray cooling as an effective cooling technology is gaining more attention and applications. By taking the advantage of the working fluid’s high latent heat, spray cooling can achieve high heat fluxes (up to 1200 W/cm2) [1] at low surface superheat as well as high heat transfer coefficient with low mass flow rate. In a high pressure die casting process, spray cooling will provide a more thermally balanced die, improve quality and reduce manufacturing costs [2]. For high power electronic devices such as laser-diode arrays pumped lasers, spray cooling is a better way to meet heat removal requirements on the order of several kilowatts or more. NASA has also expressed the strong desire for this heat removal technology in recent years [3]. As an appealing choice for many cooling systems, spray cooling has been studied by many researchers. Although the theoretical understanding about heat transfer mechanism of spay cooling is still immature due to the complex interaction of liquid and vapor phases, liquid droplet impact and phase change, a lot of valuable conclusions and principles have been gained by experiments and
* Corresponding author. Tel./fax: þ86 10 82543696. E-mail address: [email protected] (T. Changqing). 1359-4311/$ e see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2011.10.005
simulations. Horacek et al. [4] found that forced convection played an important role in heat transfer at low superheat conditions and the disturbance of droplets impingement on the liquid film would enhance the heat transfer. They also determined that phase change was the major reason for heat transfer enhancement at high superheat condition. Rini et al. [5] found that heat transfer was related to the nucleate boiling, especially the phenomenon of secondary 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 spray cooling yields a heat flux approximately an order of magnitude higher than pool boiling due to reduction in resistance to vapor removal from the heater surface. For example, when using FC-72 as the coolant, CHF of pool boiling and spray cooling are approximately 20e30 W/cm2 and 100 W/cm2 respectively [6,7]. Mudawar et al. [8] studied the effect of the distance from the orifice to the heater surface and found that CHF can be maximized when the spray impact area just inscribed the square surface of the heater [8]. The effects of surface structure on CHF were investigated with surface enhancements consisting of cubic pin fins, pyramids, and straight 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] used methanol as the coolant and achieved heat fluxes of 130 W/cm2 at
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Nomenclature h P q T i ΔT Δx
l
GWP ODP
a
heat transfer coefficient, W/(m2 K) pressure, kPa heat flux, W/cm2 temperature, C enthalpy, kJ/kg temperature margin, C distance between two thermocouples, m thermal conductivity, W/(mK) global warming potential ozone depression potential
Subscripts e evaporation in nozzle inlet sub subcooling sur surface
78 C surface temperature. PF-5060 was used in the experiments by Silk et al. [9]. A new expression with an interpolation method to construct the partially developed nucleate boiling curve was developed 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) were removed from the heater surface. Hsieh et al. [13] used R134a as working fluid and obtained cooling characteristics (i.e., boiling curves) over a specific range of spray mass flux, Weber number, wall superheat and degree of subcooling. Further studies on spray cooling have addressed inclination of the spray, nozzle type, droplets characteristics, volumetric flux and Sauter Mean Diameter [9,14e16]. All the experiments presented mainly focused on the impact factors of heat transfer and how to improve heat transfer while ignoring the design of the system configuration and how to adjust the parameters (i.e., mass flow rate, subcooling degree) to meet the heat removal requirements (i.e., heat flux, surface temperature) of devices. Existing experimental facilities reported can be divided into two types. One type that the spray chamber connected with external environment directly is open system and shown in Fig. 1 (a), and the other that the spray chamber isolated with external environment is closed system and shown in Fig. 1 (b). In the open system, the coolant is directly sprayed on the heater surface by the pump [1]. After heat transfer, the coolant vapor is released to the atmosphere, and the coolant liquid is recovered. The main deficiencies of such a system are as follows: (1) The evaporation pressure cannot be adjusted and the surface temperature may be too high due to the high boiling point of coolant at the atmospheric pressure. If water was used as the coolant, the boiling would not appear until the heater surface temperature was higher than 100 C. Most electronic devices cannot run well under this temperature. (2) The coolant cannot be recovered completely and much coolant vapor would be released to the atmosphere, which wastes the coolant and may destroy the environment. If R134a is used in the open system, it is neither economic nor environment friendly because its GWP is 1300. (3) Besides the spray cooling system, there are additional auxiliary pipes, heat exchangers and the refrigeration system to guarantee the coolant temperature. This undoubtedly will increase the difficulty and complexity of the system. In the closed system [9], the coolant can be used circularly. However, there are some disadvantages limiting its wide application: (1) In order to keep the evaporation pressure, it is essential to
247
Nozzle Spray chamber
Receiver
Pump Pump
Heat exchanger The open system
b
Nozzle Heat exchanger (Can be replaced by vacuum pump)
Spray chamber
Receiver
Pump Pump
Heat exchanger The closed system Fig. 1. Existing spray cooling system.
set a vacuum pump or an additional evaporator on the spray chamber. The system must be sealed well, otherwise it cannot work normally due to the break in of air. (2) It is necessary to set a pump supplying coolant and a refrigeration system adjusting the coolant temperature. The complexity is not less than the open system. (3) In the system with the additional evaporator, the regulation of evaporation pressure is fairly hard because the additional evaporator is controlled by the refrigeration system which also guarantees the temperature of coolant. In these two types of systems, there are a pump, a vacuum pump, a refrigeration system and several heat exchangers, which not only increase the cost but also make the system complex and inconvenient.
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A new spray cooling system is proposed in this paper, which combines the refrigeration system and the spray cooling system together. A series of experiments with different nozzle inlet pressure, evaporation pressure and degree of subcooling are carried out to verify the heat transfer performance of the novel system. 2. Configuration of the new system As mentioned above, the disadvantages of existing spray cooling systems have limited their widespread applications. Therefore, a new spray 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 some accessories. Compared to common refrigeration systems, the only difference is that the throttling device and the evaporator are replaced by the pressure nozzle and the spray chamber, respectively. The role of the nozzle in the system is throttling and atomizing. The complete cycle of the system is described as follows. Firstly, the inverter compressor exhausts high pressure and high temperature coolant vapor into the condenser. Then, the coolant vapor is condensed into high pressure liquid in the condenser. In order to test the performance, a water-cooled condenser is applied in the new system. The coolant from the condenser flows into the receiver and will be used to supply the working fluid for the nozzle. The coolant from the receiver enters into the nozzle through the expansion valve B and is throttled and atomized by the nozzle. Then the droplets of coolant spray onto the heater surface. After absorbing heat from the heater surface, the vapor coolant with excess liquid flows back to gasliquid separator, then flows into the compressor in the form of superheat vapor for next cycle. The characteristics of the new system can be elaborated from the following aspects: 2.1. Selection of coolant Different refrigerant has its own characteristics including GWP, ODP, polarity, conductivity, and corrosion, which will directly
determine whether it can be used on the high power devices well. For example, ammonia has amazing high latent heat, but it is not a proper working fluid due to its toxicity and corrosion. It also should be mentioned that different kinds of coolant can be used in the novel system according to the actual requirements. If R22 is used as working fluid, a pressure of higher than 1000 kPa can be supplied to the nozzle by the compressor. If R290 is used as coolant, surface temperature would be much low due to its low boiling point. Isobutane as green refrigerant is used as working fluid in the following performance experiments in this study. With isobutene, pressure up to 500 kPa can be attained, which satisfies the requirements of the pressure nozzle, and low evaporation temperature can achieve lower surface temperature according to high power devices. It should be mentioned that the novel system can be used under high temperature conditions with appropriate coolant (i.e., R30 or R123 can be used in the novel system, whose evaporation temperature is less than 40 C and the condensation temperature is more than 80 C). 2.2. Regulation and controlling A good cooling system must have proper regulation and controlling to adapt to the fluctuation of heat load, environmental change and requirements of surface temperature. The novel system can adjust the nozzle inlet pressure, evaporation pressure, mass flow rate and degree of subcooling. The nozzle inlet pressure can be controlled by the frequency of compressor and the cooling water temperature in the condenser. The nozzle inlet pressure can be also adjusted by regulating the fan speed if the condenser is an air cooled one. The evaporation pressure varies simultaneously according with the frequency changes of compressor. Increasing or decreasing the frequency of compressor will achieve the variation of mass flow rate correspondingly. The mass flow rate also can be regulated by adjusting valve A (see Fig. 2) opening in this system. The coolant degree of subcooling can be controlled by adjusting the opening of expansion valve A. 3. Experimental setup and procedures
Fig. 2. Schematic of novel spray cooling system.
As shown in Fig. 2, several measurement points are set on the system to analyze the system performance. In this study, an inverter rotary compressor is used, whose cooling capacity is 1500 W at the rating condition. A plate heat exchanger is applied as the water-cooled condenser in the test system. The condensation pressure and the nozzle inlet pressure can be controlled by adjusting the mass flow rate and temperature of the cooling water. The expansion valve A and the subcooler are used to adjust the degree of subcooling. In this paper, the degree of subcooling is the difference between the coolant saturation temperature corresponding to the nozzle inlet pressure and the coolant temperature of liquid into nozzle. The coolant state at the nozzle inlet is also regulated by expansion valve B in front of the nozzle. The valve A controls the mass flow rate of the nozzle by changing its opening. A full-cone pressure nozzle with spray cone angle 60 is installed in the system, and its Sauter Mean Diameter of droplets is approximately 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 well insulated by packaging insulation material and electrically heated by five cartridge heaters, which can generate 1100 W of heat load. According to Wang’s simulation results, the heat conduction of the heater unit can be seen as one-dimensional case [17]. Therefore, the heat conduction in the other directions will not be considered in the following calculation. Three T-type thermocouples with a diameter of 0.5 mm are embedded beneath the heat surface to
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249
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 variables to reflect the performance of the system. The surface temperature reflects whether the system meets the temperature requirements and heat transfer coefficient represent the heat transfer capacity. CHF indicates maximum heat removal capacity of the system and make sure whether the system meets the heat removal requirements of high power devices. 4.1. Nozzle inlet pressure
Fig. 3. Heater unit (partial sectional view).
measure the surface temperature (see Fig. 3). The three thermocouples spaces 3 mm apart from each other along the center axis of the copper heater, and the distance from the heater surface to the first thermocouple is also 3 mm. A computer, a data logger, two pressure sensors, a flow meter and a few thermocouples compose the data acquisition system together. The heat flux on the heat surface is calculated by Fourier Heat Conduction Law as the following equation: /
q ¼ l$grad T
(1)
Nozzle inlet pressure affects atomization (including droplet size and distribution) and the mass flow rate, and all the effects will further impact the heat removal process and be reflected by heater surface temperature and heat transfer coefficient. With different heat flux, the effects of the nozzle inlet pressure on the heater surface temperature and heat transfer coefficient are analyzed. The three curves in Fig. 4 (a) all indicate that the heater surface temperature reduces with increasing the nozzle inlet pressure. The trend is especially obvious when the heat flux is much higher. When the heat flux is 60 W/cm2 and the evaporation pressure is 205 kPa (Te ¼ 7.8 C), the heater surface temperature reduces from 32.4 C to 29.5 C with the nozzle inlet pressure increasing from 330 kPa to 477 kPa (from point A to point B). The heater surface temperature decreases 2.9 C, and the nozzle inlet pressure increases approximately 150 kPa. Based on the two curves (q ¼ 50 W/cm2 & 60 W/cm2), it can be determined that changing the nozzle inlet pressure is an effective way to get the same heater
a
50
Pe =205 kPa Tin =21.5 oC
For one-dimensional case, it can be written as
45
(3)
The heater surface temperature and the heat flux are all computed according to Fourier Heat Conduction Law by the measured temperature gradient to the surface from the thermocouples in the heater. The accuracy of the thermocouple is 0.3 C in this experiment. The range of pressure sensor is 0w10 bar, and its accuracy is 0.25% for the full scale. The Coriolis flow meter with the range of 0e40 kg/h has a measurement accuracy of 0.25% for total mass flow rate. According to the traditional method of error analysis, the uncertainty of the heat fluxes, the surface temperature and the 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, evaporation pressure and degree of subcooling are all regulated respectively, and their effects on heater surface temperature and heat
T sur (oC)
In the experiments, DT represents temperature difference between measured temperatures inside heater unit and heat surface, and Dx is distance for different measured locations (in the case of presented heater unit Dx ¼ 3 mm). By equation (2) and the measured values of the thermocouples shown in Fig. 3, the surface temperature can be calculated. The heat transfer coefficient is calculated based on the following definition:
q h ¼ Tsur Tsat
q =50 W/cm2 q =60 W/cm2 q =85 W/cm2
(2)
40 35
A
B
30
C
25 300
b
350
400
450
500
600
700
Pin (kPa)
35000
q =50 W/cm2 Pe=205 kPa q =60 W/cm2 =21.5 oC q =85 W/cm2 T in
31000
h (W/m 2 K)
DT Dx
q ¼ l
27000 23000 19000 15000 300
400
500
P in (kPa) Fig. 4. Effect of nozzle inlet pressure on heat transfer.
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surface temperature under different heat fluxes. The 29.5 C heater surface temperature can be obtained with the heat fluxes from 50 to 60 W/cm2 by changing the nozzle inlet pressure from 410 kPa to 477 kPa (from point C to point B). Fig. 4 (b) indicates that the increasing of the nozzle inlet pressure plays a positive role on the improvement of heat transfer coefficient. When the heat flux is high, the effect is more obvious due to the enhancement of atomization, which makes the working fluid evaporate sufficiently. It can be seen that the curve of heat transfer coefficient is nearly linear when the heat flux is 50 W/cm2. The convection is mainly way of heat transfer when the heat flux is low. With the increasing heat flux, the role of boiling is also more important besides convection. The curve of heat transfer coefficient turns into curved when the heat flux is higher. Improving the heat transfer coefficient will decrease the area of heater surface at constant heat flux, which will benefit to the actual application. In this system, the nozzle inlet pressure can be changed easily and the largest heat transfer coefficient observed is higher than 30 000 W/m2 K, which is much higher than that achieved by Zhou et al. [11] in their experiments with R113. 4.2. Evaporation pressure and temperature In the new system, evaporation temperature changes with the evaporation pressure increasing or decreasing. Changing the evaporation temperature will directly affect the surface superheat and the surface temperature. The heat transfer coefficient also changes along with the evaporation temperature. The heater surface temperature changing trend according with the evaporation pressure is obtained with different heat flux (Fig. 5 (a)) while keeping the nozzle inlet pressure at 390 kPa and the coolant degree of subcooling at 7.2 C. Results indicate that the heater surface temperature increases with the increase of the evaporation pressure. When the heat flux is 65 W/cm2, changing the evaporation pressure from 211 kPa (evaporation temperature Te ¼ 8.7 C) to 161 kPa (Te ¼ 0.8 C), results in a variation of the heater surface temperature from 34.6 C to 26.5 C (from point A to point B). The temperature on the heater surface decreases nearly 8 C due to the pressure reduction with 50 kPa. This means that changing the evaporation pressure can provide the same heater surface temperature with different heat fluxes. When the heat flux is 65, 55 and 45 W/cm2, a heater surface temperature of 30 C can be achieved by changing the evaporation pressure from 178 kPa to 206 kPa (Point C to Point E). The main reason for this is that the evaporation temperature reduces according to the reduction of the evaporation pressure. Compared with the increasing of nozzle inlet pressure, the reduction in evaporation temperature is more effective in reducing the heater surface temperature. The relation between the heat transfer coefficient and the evaporation pressure is studied as well, which is shown in Fig. 5 (b). The plot indicates that the heat transfer coefficient declines with an increase of evaporation pressure. There are two reasons for this phenomenon. One is that the increasing of the evaporation pressure reduces the latent heat of coolant, and the other is that increase of evaporation pressure delays the appearance of boiling due to the increasing of evaporation temperature accordingly. When the mass flow rate is constant under high heat flux, the superheat will increase due to the decreasing of heat transfer coefficient made by the decrease of latent heat. Comparing three curves, the effect of the evaporation pressure on the heat transfer coefficient is more obvious, especially when the heat flux is higher. It can be seen that the curve is nearly linear when the heat flux is 90 W/cm2. When the heat flux is high, the boiling effect plays an important role in heat transfer. The increasing of evaporation pressure will directly affect the boiling. Therefore, heat transfer coefficient decreases quickly and the trend is nearly linear. When
Fig. 5. Effect of evaporation pressure on heat transfer.
the heat flux is low, the effect of decreasing evaporation pressure is not obvious on heat transfer because the convection is mainly way of heat transfer as mentioned above. Thus, the trend of heat transfer coefficient is different with different heat flux. 4.3. Degree of subcooling In the new system, the degree of subcooling at the inlet of nozzle is adjusted by changing the opening of expansion valve A. The effects of subcooling on heat transfer are investigated by regulating degree of subcooling gradually. As shown in Fig. 6 (a), when keeping the nozzle inlet pressure (Pin ¼ 390 kPa), the evaporation pressure (Pe ¼ 180 kPa) and the heat flux (q ¼ 72 W/cm2) constant, the heater surface temperature first decreases and then increases when the subcooling degree of the liquid coolant (from point A to point C) decreases. This variation means that there is an optimal subcooling degree of 5.8 C in the experiment, which makes the boiling heat transfer stronger. This can be explained from the diagram of pressure-enthalpy shown in Fig. 6 (b). The throttling process of the nozzle can be described by the lines AD, BE and CF respectively (each having different subcooling degrees). It is known that the throttling effect of the nozzle is identified when the nozzle inlet pressure and the evaporation pressure are kept as constant. The difference in the state of the coolant at points D, E, F is the vapor 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 can be concluded that the appropriate vapor quality (point E) benefits for the heat transfer. For a spray, a certain vapor quality (vaporeliquid mass ratio) can strengthen the atomization. The vapor has
S. Chunqiang et al. / Applied Thermal Engineering 33-34 (2012) 246e252
a
35
Pin=390 kPa
Tsur (oC)
34
A
Pe=180 kPa q=72 W/cm2
33
C
32
B
31 30 0
3
6
9
12
15
Tsub (oC)
b
can meet the requirement of heat removal for high power devices. Keeping the evaporation pressure (Pe ¼ 177 kPa) and the nozzle inlet pressure (Pin ¼ 285 kPa) constant, the heat flux is increased gradually until the rapid increase of surface temperature appears. The results are shown in Fig. 7. The surface temperature increases gradually according with the enhancement of heat flux. When the heat flux is up to approximately 110 W/cm2, the surface temperature increases acutely. This value is CHF, which reflects that the heat transfer departs from nucleate boiling. It represents the maximum heat transfer capacity of the system under these conditions. It should be highlighted that the heater surface temperature is only 31.5 C when the heat flux is up to 110 W/cm2. Most electronic devices can work normally when the temperature is below 55 C, and their heat removal is less than 100 W/cm2. This means the novel system can meet the heat removal demand of most devices. 5. Conclusions
P (kPa) B A
251
Pin=390kPa
C
D EF
Pe=180kPa i (kJ/kg)
Fig. 6. Effect of degree of subcooling on heat transfer.
the role of assisting the atomization, which is similar to air assist atomization [18]. At point E, the quality of vapor is fit to the liquid to complete atomization, and the droplets absorb heat from the heater surface sufficiently. It is not far to see that the optimal degree of subcooling is important to the heat transfer. However, the optimal subcooling may be different for different conditions and nozzles. Hence, it is necessary to regulate the subcooling for the actual applications. It also should be noticed that the optimal mechanism and vaporeliquid mass ratio need further study for this new system. In the new system, the degree of subcooling can be adjusted by changing opening of expansion valve A. The easily way to change subcooling makes the new system have advantages for application compared to the traditional spray cooling system.
In order to satisfy the heat removal requirements for applications and improve heat transfer capacity, a novel spray cooling system was constructed. The new system incorporated a refrigeration circuit, the evaporator was replaced by the spray chamber consisting of a nozzle. The nozzle inlet pressure and the evaporation pressure can be adjusted according to desired level of heat removal by changing the frequency of compressor. The optimal degree of subcooling can be obtained as well according to the opening of expansion valve A. Based on the new system, a series of experiments were carried out to investigate the effects of pressure, degree of subcooling on the system performance. The results showed that the heat transfer coefficient up to 30 000 W/cm2 was obtained, and the heater surface temperature can be kept below 30 C when the heat flux is 50 W/cm2. The optimal degree of subcooling is 5.8 C in this experiment. The stable operation, the convenient adjustment, the simple structure and the better performance of heat dissipation reflect that the novel system has promising applications in high power devices. The way integrated the novels with electronic devices will be studied in the future in order to make the system more flexible to respond effectively to rapidly changing heat flux scenarios [19,20]. In addition, further research is needed for the influence of lubricant on the heat transfer in the novel spray cooling system. Acknowledgement
4.4. CHF of the new system The critical heat flux (CHF) reflects the maximum heat transfer capacity, and the value of CHF directly decides whether the system
The authors appreciate the financial supports from the National Natural Science Foundation of China (No. 51106170) for the work reported in this paper. References
Fig. 7. Measured curve of heat flux.
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