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Applied Thermal Engineering 62 (2014) 613e621

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Applied Thermal Engineering

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

The effects of micro-structured surfaces on multi-nozzle spray cooling

Yan Hou a,b, Yujia Tao a, Xiulan Huai a,*a Institute of Engineering Thermophysics, University of Chinese Academy of Sciences, P.O. Box 2706, Beijing 100190, ChinabUniversity of Chinese Academy of Sciences, Beijing 100049, China

h i g h l i g h t s

� The optimal micro-structured surfaces were straight fins2 and 3 in zone I.� The heat transfer performance of cubic pin fins was the best one in zone II.� A dimensionless number was proposed to scale heat transfer enhancement.� Temperature uniformity was discussed.

a r t i c l e i n f o

Article history:Received 13 August 2013Accepted 16 October 2013Available online 25 October 2013

Keywords:Micro-structured surfaceMulti-nozzle spray coolingHeat transferTemperature uniformity

* Corresponding author. Tel./fax: þ86 10 8254 3108E-mail address: hxl@mail.etp.ac.cn (X. Huai).

1359-4311/$ e see front matter � 2013 Elsevier Ltd.http://dx.doi.org/10.1016/j.applthermaleng.2013.10.03

a b s t r a c t

Experiments were conducted to investigate heat transfer characteristics of spray cooling with eightnozzles for micro-structured surfaces included cubic pin fins and straight pin fins of different sizes.Liquid volume flow rate ranged from 2.46 � 10�2 m3/s/m2 to 3.91 � 10�2 m3/s/m2 and the correspondedinlet pressures changed from 0.28 MPa to 0.6 MPa by keeping the inlet water temperature between20.4 �C and 24.31 �C. And the input power of heat block varied from 180 W to 1080 W. The results showthat the heat transfer performances of straight fins2 and straight fins3 are the best in single phase zone,but the cubic pin fins is better in two phase zone. Notably, the critical point between single phase zoneand two phase zone shifts to left with the increasing of liquid volume flow rate. Moreover, with the liquidvolume flow rate increasing, the heat transfer coefficient increases as well, but straight fins1 and pol-ished surface are not sensitive to this change. For a deeper analysis of the heat transfer enhancement, adimensionless number (DM) is created to characterize heat transfer performance of different micro-structures in single phase heat transfer. We verified the dimensionless number using experimental re-sults in this study and previous literature. Furthermore, the micro-structured surfaces have negligibleeffects on temperature distribution except for cubic pin fins.

� 2013 Elsevier Ltd. All rights reserved.

1. Introduction

Among alternative solutions for cooling high-powered devices,spray cooling has the advantages of removing high heat flux andproviding uniform temperature distribution in a confined space.Traditionally, spray cooling was utilized to cool highly heated sur-faces for equipments and processes in metallurgy, chemical andnuclear industry [1,2]. Recently, it has received increasing atten-tions in the development of modern technologies, such as thecooling of electronic devices and high power solid-state lasers. Ithas been reported that spray cooling had been applied in thecooling of Cray X1 vector supercomputers [3].

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However, the maximum CHF was limited to 1000 W/cm2. Inorder to obtain higher heat flux and more uniform temperature,multi-nozzle spray cooling has been widely studied. The heattransfer performance of multiple nozzle sprays significantly de-pends on geometry and spacing [3,4]. Lin et al. [5,6] carried out anexperimental investigation with eight miniature nozzles at spraypressure drops greater than 1.72 bar. It showed that the CHF levelswith eight-nozzle sprays were 90 W/cm2 with pure FC-72 and490 W/cm2 with pure methanol respectively. Y. B. Tan et al. [7]developed a correlation based on their multi-nozzle spray coolingexperimental results to predict the dimensionless heat flux. Jia andQiu [8] made an experiment to investigate the advantage of sur-factant addition in spray cooling with five nozzle arrays and indi-cated that the surfactant addition could result in a relativelyconstant heat removal rate near the CHF regime. Panao et al. [9]presented a thermal assessment of a multi-jet strategy for spraycooling system. The assessment considered a diverse number of

Nomenclature

A area of micro-structured surface, m2

A0 area of polished surface, m2

a the height of each fin, mmBo Bond numberb the groove width, mmc the fin width, mmd normal distance between the two rows of

thermocouples, mDM a dimensionless numberg gravitational constant, m/s2

h total heat transfer coefficient of heated surface, W/m2/K

k conductivity of the heat block, W/m/Kl normal distance between the heated surface and the

upper row of thermocouples, mNu the averaged Nusselt number in natural convectionPr the Prandtl numberq0 averaged heat flux of the heated surface, W/m2

Q 0loss the heat transferred to surrounding air from the part of

heat block between two rows of thermocouples, WQv inlet volume flow rate, m3/s

RaL the Rayleigh numberTlower averaged temperature of the lower row of

thermocouples, �CTupper averaged temperature of the upper row of

thermocouples, �CTw the averaged temperature of the heated surface, �CTl temperature of inlet water, �CTc averaged vapor temperature in spray chamber, �C

Greek symbolsg surface tension of water, N/mdk error in conductivity, W/m/Kdq0 the uncertainty of the averaged heat flux, W/m2

dT error in temperature measurement, �CdTw the uncertainty of the averaged temperature, �Cdx error in thermocouple location, mεT heat loss percentage between two rows of

thermocouplesr density, kg/m3

SubscriptsV vapor phasel water phase

Y. Hou et al. / Applied Thermal Engineering 62 (2014) 613e621614

nozzles. Their results supported that themulti-nozzle spray coolingwas a good choice for smaller processors with more adequatethermal management system. However multi-nozzle spray coolingcannot meet the increasing cooling requirements, so more effectivemethods are required.

Recently researchers have been paying more attentions to theenhanced surfaces that show great potential to further improve theperformances of spray cooling heat transfer. Various studies havebeen reported. Sehmbey et al. [10] discovered that increasing thesurface roughness had a positive effect on heat transfer through aspray cooling experiment with liquid nitrogen. Pais et al. [11]studied the surface roughness and its effect on the heat transfermechanism in spray cooling using an air atomizing nozzle. Theyfound that the nucleate boiling played a major role in the heattransfer when the surface roughness was greater than 1 mm.However, for films of the order of 0.1 mm, heat was conductedthrough the film and evaporated on the surface, yielding very highheat fluxes of the order of 1200 W/cm2 at very low superheat. Kimet al. [12] built the microporous structures on the heated surfacesand studied the effect of particle size on the heat transfer co-efficients experimentally using the air-atomized nozzle. The resultsshowed that the heat transfer coefficient increased by up to 400%relative to that of uncoated surface cooled by dry air, and thisenhancement was maintained at high heat fluxes. They attributedthis enhancement to the increased capillary forces betweenmicrostructures.

Bostanci et al. [13] conducted the experiments to investigatespray cooling onmicro-structured surface with ammonia using twovapor atomized spray nozzles, and a smooth surface was also testedfor comparison. Results suggested that the heat transfer coefficientsincreased by 112% and 49% for treated surfaces with protrusionsand indentations respectively, in comparison with smooth surface,when the heat flux over heated surface was 500 W/m2. Stephanet al. [14] studied the spray cooling heat transfer performance onmicro-structured surfaces consisted of micro pyramids withdifferent heights. They found that a significant enhancement in theheat transfer performance due to the surface structures could beobserved, especially at low coolant fluxes. The authors attributedthis to the increase of the three phase contact line, which leads to

more effective thin film evaporation. Bostanci et al. [15] had studiedspray cooling with ammonia on structured surfaces to determinethe CHF limits. The results showed that the maximum heat flux ofmulti-scale structured surface increased by 18% over smooth sur-face, up to 910 W/cm2 at nominal flow rate. And the multi-scalestructured surface with pyramidal fins and protrusions achievedthe highest CHF value of 1090 W/cm2, so did the surface withprotrusions. de Souza et al. [16] conducted an experiment to studythe spray cooling on copper-foam enhanced surface with R134a.The enhancement factor of copper-foam surface is as high as 1.39.In sum, treated surfaces can enhance heat transfer significantly. It isessential to understand the heat transfer mechanism to obtain theoptimized micro-structure, however, few studies were found.

Chien et al. [17] investigated multi-nozzle jets cooling with FC-72 on cubic pin fins and straight pin fins. They indicated that theheat transfer performance increased with the increasing of liquidvolume flow rate or surface area enhancement ratio. Their datashows that the heat transfer performance of two-phase jets isdependent on Re, Bo and surface enhancement ratio. Hsieh et al.[18] investigated evaporative heat transfer characteristics of a wa-ter droplet spray on the plain and square micro-studs silicon sur-faces at very low spray mass fluxes up to 4.41 mL/cm2. Theyindicated that the Bond number of the microstructures was animportant factor to explain the heat transfer enhancement ofevaporative spray cooling on micro-structured silicon surfaces.Moita et al. [19] studied the impact of droplets onto micro-structured surfaces and scaled the effects of surface topographyon secondary atomization. They indicated that wetting propertieswere responsible for different characteristics of the thermal-induced atomization. The results also show good correlation be-tween the mean sizes of the secondary droplets generated bythermal-induce atomization and the ratio of the mean height of thepeaks and the pitch between them.

As above mentioned, the surface technologies improve heattransfer performance greatly. However the enhancement mecha-nisms of treated surface are not clear yet. Besides, previous studieson micro-structured surface enhancing multiple nozzle spraycooling and influencing temperature uniformity are limited. Themain objective of the current work is to investigate the effects of

Fig. 1. Experimental setup.

Y. Hou et al. / Applied Thermal Engineering 62 (2014) 613e621 615

micro-structured surface on heat transfer performance and tem-perature uniformity. A dimensionless number, DM, was proposedto scale the heat transfer enhancement on cubic-finned surfacesand straight-finned surfaces. We verified the dimensionless num-ber using experimental results in this study and previous literature.Surface modification techniques were used to obtain microscalecubic pin fins and the straight fins of different sizes on the heatersurfaces. A smooth surface was also tested to have baseline data forcomparison. Tests were conducted in an open loop system withwater as the working fluid, using pressure atomized spray nozzles.The heat fluxes and the heat transfer coefficients under differentliquid volume flow rates with micro-structured and smooth sur-faces were obtained.

2. Experimental system

2.1. Test facility

The experimental setup consists of four parts: liquid delivery,spray, heater assembly and data acquisition system. The open-loopfluid delivery system is comprised of two liquid reservoirs, a rotaryvane pump, a filter and the flow channels with control valves asshown in Fig. 1. The de-ionized water pumped from the reservoirpasses through the filter, the three-way valve, and then reaches thespray chamber. In the spray chamber, water is split into dropletswhen it passes through a multi-nozzle assembly, then the dropletsimpact onto the heated surface. A heat transfer process iscompleted when the liquid outflows from the spray chamber andreturns to the other reservoir. In this open-loop experimental sys-tem, the liquid volume flow rate is regulated by a three-way valve,and the heat power supplied to the surface is regulated by the ACvoltage, respectively.

2.2. Spray system

The structure of the multi-nozzle assembly is shown in Fig. 2.There are eight miniature nozzles in the multi-nozzle plate. Eachnozzle has a swirl insert, a swirl chamber and a discharge orifice.The swirl insert is mounted onto the multi-nozzle plate. The waterflows into the swirl chamber through three orifices with thediameter of 2 � 10�4 m, generates a swirl flow pattern, and thenflows out from the discharge orifice with diameter of 3 � 10�4 mand breaks up into fine droplets. Besides, the distance between thetwo discharge orifices is 8 � 10�3 m.

2.3. Heater assembly

A copper block heater consisted of six rod-type heaters isemployed as the heat source in the experiment as shown in Fig. 3.The maximum power of each rod-type heater is 250 W. Each hasindependent wires and plug, so it is convenient to set series orparallel electric circuit. The upper surface of the heater with an areaof 3.2� 10�2 m� 1.6� 10�2 m is used as the test surface. Sixteen K-type thermocouples with probe diameter of 0.5 mm are embeddedinto the holes drilled along the two planes in the block heater asshown in Fig. 3. The distance between the two thermocouplelocation planes is 5 mm. The distance between the test surface andthe upper plane of the thermocouple locations is 31 mm. Theheated surface is sealed in the bottom plate of the spray chamber.The heater block is placed in a stainless steel shell with fiberfrax tominimize the heat transfer to the ambience.

2.4. Enhanced surfaces

Four enhanced surface geometries are shown in Fig. 4. Themicro-structure in Fig. 4(a) is consisted of 512 cubic pin fins, and themicro-structures in Fig. 4(bed) are consisted of straight fins ofdifferent sizes. Fig. 5 shows the cross sectional views of the micro-structures. The corresponding dimensions are presented in Table 1.

2.5. Data acquisition system

The liquid volume flow rate, the liquid inlet pressure, the aver-aged vapor temperature in spray chamber and two rows of tem-peratures in block heater could be measured directly in theexperiments. And temperatures were collected by 34970A dataacquisition unit. However, the test surface temperature was ac-quired by indirect measurement due to the thin liquid film on thesurface and the continuous impact of droplets. As the heat lossbetween the two rows of thermocouples is small, the heat con-duction between lower row of thermocouples and heated surfacecan be simplified into one-dimensional condition. The averagedheat flux of heated surface can be calculated by expression (1) onthe basis of Fourier law of heat conduction:

q0 ¼ k�Tlower � Tupper

��d (1)

The averaged temperature of heated surface can be calculatedby:

Fig. 2. Multi-nozzle assembly (dimensions in �10�3 m).

Y. Hou et al. / Applied Thermal Engineering 62 (2014) 613e621616

Tw ¼ Tupper � l=d�Tlower � Tupper

�(2)

3. Test conditions and procedure

All tests were conducted using the de-ionized water. Liquidvolume flow rate was increased from 2.46 � 10�2 m3/s/m2 to3.91 �10�2 m3/s/m2, and the corresponded inlet pressure changedfrom 0.28MPa to 0.6MPa. The liquid inlet temperature ranged from

Fig. 3. The structure of heater (dimensions in �10�3 m).

Fig. 4. The micro-structured surfa

20.4 �C to 24.31 �C, and the input power of the block heater wasgradually increased from 180 W to 1080 W.

Since the main goal was to investigate spray cooling enhance-ment on micro-structured surfaces. Cooling curves (heat flux vs.surface superheat) for all surfaces were generated to compare theirperformances. The operating steps are as follows:

(1) Before each test, we adjust the liquid volume flow rate andthe liquid inlet temperature to the desirable values.

(2) The AC voltage is turned on when the spray becomes stablein order to avoid an over-high temperature at the test sur-face. The voltage is set at a low starting level.

ces (dimensions in �10�3 m).

Fig. 5. Geometry cross sectional view.

Table 1Parameters of test surfaces.

Surface C (mm) b (mm) a (mm)

Cubic pin fins 500 500 600Straight fins1 200 400 400Straight fins2 200 300 400Straight fins3 200 200 200

Y. Hou et al. / Applied Thermal Engineering 62 (2014) 613e621 617

(3) When the steady state heat transfer is achieved, the liquidinlet pressure, averaged vapor temperature in the spraychamber and the temperatures of the sixteen thermocouplesare measured respectively.

(4) Increase the AC voltage to each preset level and repeat step(3).

(5) Change the liquid volume flow rate and the test surface,respectively, and repeat the above process.

(6) When the tests were completed, turn off the AC voltagebefore the spray system.

Fig. 7. The heat loss percentages between two rows of thermocouples under differentheat fluxes and the mean heat loss percentage between two rows of thermocouples.

Table 2Important parameters of test surfaces.

Polishedsurface

Cubic pinfins

Straightfins1

Straightfins2

Straightfins3

4. Uncertainty and heat loss analysis

4.1. Uncertainty analysis

The uncertainties for the physical quantities directly measuredare as follow: �2 L/h for inlet volume flow rate; �0.02 MPa forliquid inlet pressure; �0.1 V for input voltage; �1 �C for tempera-ture measurement; �2 W/m/K for thermal conductivity; and�0.5 mm for the location of thermocouples. The uncertainties ofthe averaged temperature and the heat flux were calculated byEquations (3) And (4) respectively.

dTw¼�ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi��

vTwvTupper

�2

þ�

vTwvTlower

�2�d2Tþ

��vTwvl

�2

þ�vTwvd

�2�d2x

s

(3)

dq0 ¼�ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi�vq0

vd

�2

d2xþ�vq0

vk

�2

d2kþ��

vq0

vTlower

�2

þ�

vq0

vTupper

�2�d2T

s

(4)The results of calculations showed that the maximum uncer-

tainty of the averaged temperature and the heat flux were �5.4 �Cand �1.33 W/cm2.

Heat transferarea (�10�4 m2)

5.120 11.264 12.032 13.312 10.24

Bo e 0.1834 0.1467 0.1100 0.0734DM e 17.79 24.56 38.11 38.53Space between

fins (�10�6 m3)0 280.4 138.24 122.88 51.2

4.2. Heat loss analysis

In the experiments, about 82% of the input power could bedissipated through the multi-nozzle spray cooling system under

Fig. 6. The numerical result of temperature fi

different liquid volume flow rates. A three dimensional numericalsimulation was performed to determine the heat loss between tworows of thermocouples in the block heater, which is critical forcalculating the averaged temperature of heated surface.

Fig. 6(a) shows the geometric model of the calculation domainand boundary conditions. The black field corresponds to the partof heat block between two rows of thermocouples and the grayfield corresponds to the surrounding air. The yellow face is definedas the symmetry boundary. The upper and lower sides of the blackfield are defined as specific temperature boundary and the othersides of the black field were defined as natural convectionboundary. The natural convection heat transfer coefficient wascalculated by Equation (5) based on the correlation for verticalplate. Fig. 6(b) shows one of the temperature contours of calcu-lation domain.

eld between two rows of thermocouples.

Qv =2.46×10-2 m3/s/m2 Qv =3.05×10-2 m3/s/m2

Qv =3.48×10-2 m3/s/m2 Qv =3.91×10-2 m3/s/m2

Fig. 8. Heat transfer characteristics under various inlet volume flow rates.

Y. Hou et al. / Applied Thermal Engineering 62 (2014) 613e621618

Nu ¼(0:825þ 0:387Ra1=6Lh

1þ ð0:492=PrÞ9=16i8=27

)2

(5)

The heat loss percentage between two rows of thermocoupleswas calculated through Equation (6)

εT ¼ Q 0loss=q

0A0 (6)Fig. 7 shows the heat losses percentage between two rows of

thermocouples computed by Fluent for polished surface under theinlet volume flow rate of 2.46 � 10�2 m3/s/m2. Results show thatthe maximum and mean heat loss percentages between two rowsof thermocouples are 1.4% and 1.3% respectively.

5. Results and discussion

The present experiment focuses on studying the effects ofmicro-structured surfaces on multi-nozzle spray cooling. We pro-posed a dimensionless number to scale the heat transferenhancement of micro-structured surfaces and investigated theeffects of liquid volume flow rates on heat transfer coefficient.Furthermore, the temperature distributions of various micro-structured surfaces were presented.

5.1. The effects of micro-structured surfaces on the heat transferenhancement

Many researchers have tried to reveal the mechanisms of heattransfer enhancement on micro-structured surface. Hsieh believedthat the capillary force is the primary factor responsible for the heat

transfer enhancement of evaporative spray cooling. The capillaryforce is characterized by a dimensionless number, Bond number(Bo), which is calculated by the Equation (7).

Bo ¼ b=ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffig=ðrl � rvÞg

p(7)

Apparently, narrower grooves correspond to smaller Bondnumber. Chien et al. [17] indicated that effective heat transfer areais another important factor influencing the heat transferenhancement of micro-structured surfaces. The heat exchangeareas and Bond numbers for all surfaces in this study are listed inTable 2. The space between micro-structures indicates to the vol-ume between fins on test surface. Based on these data, a dimen-sionless number, DM, is proposed. The 1.5 exponent in Equation (8)is obtained through the Rayleigh method using partial experi-mental results. Generally, bigger DM corresponds to better heattransfer performance.

DM ¼ 1=BoðA=A0Þ1:5 (8)

To prove that this dimensionless number is correct and practical,we calculated DM numbers for every test surfaces as shown inTable 2 and compared them with the experimental results.

Fig. 8 shows the spray cooling curves, q0 vs. Tw � Tl, for diversesurfaces under different inlet volume flow rates. Obviously, the heattransfer for polished surface is worst under any conditions. This isattributed to its smallest heat exchange area as shown in Table 2.For each inlet volume flow rate, the spray cooling curves can bedivided into two zones. In zone I, the heat flux increases linearlywith the increase of surface superheat, the heat transfer is mainlyruled by single phase heat transfer. In zone II, the slope of the spray

Input power of 360 W Input power of 540 W

Input power of 720 W Input power of 900 W

Fig. 9. The heat transfer coefficients vs. liquid volume flow rates under diverse input powers.

Fig. 10. The temperature measurement mapping points on heated surface and thelocations of nozzles.

Y. Hou et al. / Applied Thermal Engineering 62 (2014) 613e621 619

cooling curve increases for cubic pin fins, the nucleate boiling heattransfer is the primary part of heat transfer for cubic pin fins. Alsowhen the inlet volume flow rate is 2.46 � 10�2 m3/s/m2, nucleateboiling will appear in the thin liquid film when critical surfacesuperheat is 86.7 �C. Critical surface superheat decreases with theincreasing of inlet volume flow rate. When the inlet volume flowrate is increased to 3.91 � 10�2 m3/s/m2, the critical surface su-perheat is reduced to 51 �C.

As shown in Fig. 8, when the liquid volume flow rate is in therange from 2.46 � 10�2 to 3.48 � 10�2 m3/s/m2, in zone I, thestraight fins2 and straight fins3 having biggest DM numbers arebest, followed by straight fins1, cubic pin fins and polished surface.Although the heat exchange area of straight fins1 is the second-biggest, its heat transfer performance is not. This is because thestructure of straight fins1 has a bigger Bond number, smallercapillary force, which is not helpful in spreading liquid over theheated surface, resulting in a thicker liquid film, and deterioratesthe heat transfer enhancement. DM number of straight fins1 issmallest among all treated surfaces, which agrees well with presentexperimental results. In zone II, the heat transfer ability of cubic pinfins become the best among all of surfaces, because the space be-tweenmicro-structures of cubic pin fins is the biggest and the crosschannel of cubic pin fins enhances boiling heat transfer. Moreover,when the liquid volume flow rate is higher than 3.48 � 10�2 m3/s/m2, the heat transfer performance of cubic pin fins is improved inzone I. In other words, with the increasing of liquid volume flowrate, the heat transfer performance of cubic pin fins is improvedsignificantly. As stated above, compared to straight fins surfaces,

the heat flux of cubic pin fins is more sensitive to the surface su-perheat temperature and the liquid volume flow rate. Chien et al.[17] indicated that c-822 (straight-finned surface) and p-822 (pin-finned surface) have similar performance at low heat flux (up to400 kW/m2) and pin-finned surface yields a better cooling perfor-mance at high heat fluxes, which agrees with the present study.Under given experimental conditions, straight fins2 and straightfins3 have similar performance. Because the DM number of straightfins2 is almost equal to that of straight fins3.

Moreover, we use experimental results in literature to verify thepracticability of DM number. Hsieh et al. [18] investigated the effectof micro-structured surface on spray cooling using water as work-ing fluid. There are three type of surfaces: 120G � 160S,

Qv =2.46×10-2 m3/s/m2 Qv =3.05×10-2 m3/s/m2

Qv =3.48×10-2 m3/s/m2 Qv =3.91×10-2 m3/s/m2

Fig. 11. The curves of temperature distributions for diverse micro-structured surfaces.

Y. Hou et al. / Applied Thermal Engineering 62 (2014) 613e621620

120G� 480S, 360G� 480S, the heat transfer performances of thesesurfaces decrease in turn, while DM numbers of these surfaces are102, 54, 11 respectively. Zhen Zhang et al. [20] studied on the heattransfer enhancement of cubic pin fins with different sizes. Theyindicated that the heat transfer performance of surface S3 was best,followed by S2 and S1, while the DM numbers of the three surfaceswere 64, 38, 12.5 respectively, and decreases in turn. Chien et al.[17] investigated the effects of pin-finned surfaces and straight-finned surfaces on multiple jet-cooling using FC-72 as workingfluid. It indicated that the heat transfer performances of P-822 andC-822 are better than P-4 and C-4. The DM numbers of P-822, C-822, P-4, C-4 are 40, 47, 5, 5.8 respectively, which can explain thisresults well. As above mentioned, the DM numbers of micro-structured surfaces agree well with experimental results, so thisdimensionless number is credible to characterize the heat transferenhancement of straight fins and cubic pin fins.

5.2. The effect of the liquid volume flow rate on the heat transfercoefficient

For deeper understanding of themulti-nozzle spray cooling heattransfer for micro-structured surfaces, the total heat transfer co-efficient of heated surface was calculated through Equation (9). Thetotal heat transfer is determined by adding the contributions of allheat transfer mechanisms, such as convection, evaporation fromliquid film surface, nucleate boiling and so on.

h ¼ q0=ðTw � TcÞ (9)

Fig. 9 shows the heat transfer coefficients vs. liquid volume flowrates under various input powers. For straight fins2 and straightfins3, the heat transfer coefficients increase with the increasing of

the liquid volume flow rates generally. Heat transfer coefficient ofthe straight fins3 is 2.35 times as large as that of the polishedsurfaces when the liquid volume flow rate is 2.46 � 10�2 m3/s/m2

under the input power of 720W. And the straight fins3 is 2.54 timesthe heat transfer coefficient of the polished surface when the liquidvolume flow rate is 3.91 � 10�2 m3/s/m2 with the input power of720 W.

However, the heat transfer coefficients of straight fins1 andpolished surface are not sensitive to the liquid volumeflow rates. Onone hand, increasing liquid volume flow rate leads to higher inletpressure, smaller droplet Sauter Mean Diameter and faster dropletvelocity [21,22], which is favorable for spray cooling heat transfer.On the other hand, increasing liquid volume flow ratewill result in athicker liquid film on the surface which has negative effect on heattransfer. However, compared to straight fins1, the Bond numbers ofstraight fins2 and straight fins3 are smaller. That is to say, there isstronger capillary force and better liquid spread ability for straightfins2 and straight fins3. This will diminish the increase of liquid filmthickness caused by the increase of liquid volume flow rate and leadto a better heat transfer performance for straight fins2 and straightfins3. The heat transfer coefficient for cubic pin fins increasesdramatically with the increasing of liquid volume flow rate. Thisresult consists with the prior conclusion that the heat transfer per-formance of cubic pin fins is improved in zone I when the liquidvolume flow rate is higher than 3.48 � 10�2 m3/s/m2.

5.3. The effect of liquid volume flow rate on temperaturedistribution of micro-structured surface

Beside how to raise the heat flux on heated surface, how toimprove the temperature uniformity on heated surface is another

Table 3The maximum differences between local superheated temperatures under variousliquid volume flow rates.

Liquid volumeflow rate(�10�2 m3/s/m2)

The maximum difference between local superheatedtemperatures on heated surface (�C)

Polishedsurface

Cubic pinfins

Straightfins1

Straightfins2

Straightfins3

2.46 20.8 16.4 16.5 22.7 23.53.05 20 76.5 16.7 21.9 233.48 20.7 16.8 16.9 23.4 23.633.91 20.1 16.7 17.5 23.9 24.3

Y. Hou et al. / Applied Thermal Engineering 62 (2014) 613e621 621

difficult problem. Compared to single nozzle spray cooling, multi-nozzle spray cooling can further improve the temperature unifor-mity on heated surface. Micro-structured surface can enhance heattransfer, but rare researches focus on its effect on temperaturedistribution. In view of this, the effects of micro-structured surfaceson temperature distributions under various conditions are pre-sented in this article.

Fig. 10 shows the locations of nozzles and the temperaturemeasurement mapping points on heated surface. Fig. 11 presentsthe temperature distributions of various micro-structured surfaceswhen the liquid volume flow rate ranges from 2.46 � 10�2 m3/s/m2

to 3.91 � 10�2 m3/s/m2, and the input power is kept at 180 W. Asseen in Figs. 10 and 11, it is obviously that local superheated tem-peratures of T2, T4, T5 and T7 are higher than that of the neighboringlocations. The reason is that the liquid film is relatively thicker andthe heat transfer performance is worse in the multi-nozzle spraycooling stagnation zone [23].

As seen in Table 3, for polished surface and straight fins surfaces,the increase of liquid volume flow rate leads to slight change oftemperature distribution, the structure size has negligible effect ontemperature distribution. However, for cubic pin fins, the temper-ature uniformity decreases firstly, and then increases with theincreasing of the liquid volume flow rate. The turning point of thisprocess is right corresponded to the critical point where the slopeof heat transfer curve increases dramatically.

6. Conclusions

Multi-nozzle spray cooling experiments for micro-structuredsurfaces have been carried out to investigate the effects of liquidvolume flow rate and surface structure on heat transfer perfor-mances and temperature distributions. Based on the results, thefollowing conclusions are drawn:

(1) Compared with polished surface, the micro-structured sur-face enhances the cooling performance. The heat transferperformances of the straight fins2 and 3 are the best in zone Iwhen liquid volume flow rate is smaller than 3.48� 10�2 m3/s/m2. This is attributed to checks and balances between heatexchange area and capillary force between grooves, which ischaracterized by DM. Cubic pin fins is the best one in zone II.

(2) The straight fins3 is 2.54 times the heat transfer coefficient ofthe polished surface when the liquid volume flow rate is3.91 � 10�2 m3/s/m2 with input power of 720 W. Heattransfer coefficient rises with the increase of liquid volumeflow rate for straight fins2 and straight fins3. But straightfins1 and polished surface are not sensitive to the change ofliquid volume flow rate. The curve of cubic pin fins risesrapidly when the liquid volume flow rate is 3.48� 10�2 m3/s/m2 as shown in Fig. 9.

(3) The temperatures of the multi-nozzle spray cooling stagna-tion zones are higher than that of the neighboring locations.The structure size and the liquid volume flow rate havenegligible effects on temperature distribution except forcubic pin fins.

Acknowledgements

This research has been supported by the National NaturalScience Foundation of China under Grant No. 51276181 and Na-tional Basic Research Program of China under Grant No.2011CB710705.

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