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High heat flux spray cooling with ammonia: Investigation of enhanced surfaces for CHF
International Journal of Heat and Mass Transfer 55 (2012) 3849–3856
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International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
High heat flux spray cooling with ammonia: Investigation of enhanced surfaces for CHF Huseyin Bostanci a,⇑, Daniel P. Rini a, John P. Kizito b, Virendra Singh c, Sudipta Seal c, Louis C. Chow c a
RINI Technologies, Inc., 582 S. Econ Circle, Oviedo, FL 32765, United States Department of Mechanical and Chemical Engineering, North Carolina A&T State University, Greensboro, NC 27411, United States c Department of Mechanical, Materials and Aerospace Engineering, University of Central Florida, Orlando, FL 32816, United States b
a r t i c l e
i n f o
Article history: Received 2 June 2011 Received in revised form 5 March 2012 Available online 20 April 2012 Keywords: Spray cooling Thermal management CHF Enhanced surfaces
a b s t r a c t A spray cooling study was conducted to investigate the effect of enhanced surfaces on Critical Heat Flux (CHF). Test surfaces involved micro-scale indentations and protrusions, macro (mm) scale pyramidal pin fins, and multi-scale structured surfaces, combining macro and micro-scale structures, along with a smooth surface that served as reference. Tests were conducted in a closed loop system using a vapor atomized spray nozzle with ammonia as the working fluid. Nominal flow rates were 1.6 ml/cm2 s of liquid and 13.8 ml/cm2 s of vapor, resulting in a pressure drop of 48 kPa. Results indicated that the multi-scale structured surface helped increase maximum heat flux limit by 18% over the reference smooth surface, to 910 W/cm2 at nominal flow rate. During the additional CHF testing at higher flow rates, most heaters experienced failures before reaching CHF at heat fluxes above 950 W/cm2. However, some enhanced surfaces can achieve CHF values of up to 1100 W/cm2 with 67% spray cooling efficiency based on liquid usage. The results also shed some light on the current understanding of the spray cooling heat transfer mechanisms. Enhanced surfaces are found to be capable of retaining more liquid compared to a smooth surface, and efficiently spread the liquid film via capillary force within the structures. This important advantage delays the occurrence of dry patches at high heat fluxes, and leads to higher CHF. The present work demonstrated ammonia spray cooling as a unique alternative for challenging thermal management tasks that call for high heat flux removal while maintaining a low device temperature with a compact and efficient cooling scheme. Ó 2012 Elsevier Ltd. All rights reserved.
1. Introduction Many critical applications today, in electronics, optics and aerospace fields, among others, demand advanced thermal management solutions for the acquisition of high heat loads they generate in order to operate reliably and efficiently. Current competing technologies for this challenging task include several single and two phase cooling options. When these cooling schemes are compared based on the high heat flux removal (100–1000 W/ cm2) and isothermal operation (within several °C across the cooled device) aspects, as well as system mass, volume and power consumption, spray cooling appears to be the best choice. Although a vast amount of research has been done on heat transfer enhancement in general, studies focusing on spray cooling enhancement are fairly limited. In early works, Pais et al. [1] and Sehmbey et al. [2] examined the effects of surface roughness and contact angle using water with air atomized nozzle, at flow rates up to 1.4 ml/cm2 s water and 400 ml/cm2 s air, and found enhance-
⇑ Corresponding author. E-mail address: [email protected] (H. Bostanci). 0017-9310/$ - see front matter Ó 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.ijheatmasstransfer.2012.03.040
ment in heat transfer coefficient with decreasing surface roughness and increasing contact angle. They obtained heat fluxes up to 1250 W/cm2 at 11 °C surface superheat on ultrasmooth (Ra = 0.3 lm) copper surface. Kim et al. [3] investigated spray cooling enhancement on microporous coated surfaces using water at flow rates up to 0.03 ml/ cm2 s. The porous layer was fabricated using a mixture of methyl–ethyl-ketone (MEK), epoxy, and aluminum powder, and its maximum thickness was 500 lm. They found that the CHF increased 50% relative to the uncoated surface. However, highest heat flux reached was 3.2 W/cm2 at 65 °C surface superheat due to very low flow rates. Stodke and Stephan [4] studied spray cooling on microstructured and micro-porous surfaces using water at 6 kPa system pressure and 1.4 ml/cm2 s flow rate. Pyramidal micro-grooves and micro-pyramids with 75 lm height increased the wetted area by a factor of 1.4. Their 100-lm thick porous layers were very similar to those used by Kim et al. [3] and created with the same ingredients of MEK, epoxy, and aluminum powder. A maximum heat flux of 97 W/cm2 was observed for the micro-pyramid surface compared to 30 W/cm2 for the flat surface, both at a surface superheat of 12 °C. This enhancement was much larger than the surface area
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Nomenclature A CHF CHFref cp d32 EFCHF h hfg I k N q00 Tl Tsat
area (cm2) Critical Heat Flux (W/cm2) Critical Heat Flux of reference surface (W/cm2) liquid specific heat (J/kg °C) Sauter-mean droplet size (m) CHF enhancement factor heat transfer coefficient (W/m2 °C) latent heat of vaporization (kJ/kg) current (A) heater wall thermal conductivity (W/m °C) mean droplet flux (#/m2 s) heat flux (W/cm2) liquid temperature (°C) saturation temperature (°C)
enhancement. However, when a micro-porous surface was used, a significant degradation in heat transfer occurred when compared to the plain surface resulting in 14 W/cm2 maximum heat flux at 12 °C superheat. The low heat flux was due to the poor thermal conductivity of the epoxy binder. Amon et al. [5] and Hsieh and Yao [6] performed spray cooling experiments with water on silicon substrates using 1 2 nozzle array at very low flow rates of up to 0.07 ml/cm2 s on square micro-studs with 160–480 lm heights. The surface texture was found to have little effect in the single phase and dry out regimes. The authors attributed the higher heat transfer observed for the micro-textured surfaces in the intermediate regimes to more effective spreading of the liquid by capillary forces. The maximum heat flux achieved was just over 50 W/cm2 at 55 °C surface superheat. Silk et al. [7] investigated the effect of surface geometry in spray cooling with PF-5060 using 2 2 nozzle array at 1.6 ml/cm2 s flow rate. They used embedded structures (dimples, pores and tunnels) and compound extended structures (straight fins, cubic pin fins and dimples) all in the order of 1 mm size. Of these macrostructured surfaces, straight fins and porous tunnels performed the best providing a CHF of 175 W/cm2 for gassy conditions at surface superheats of up to 36 °C and offered CHF enhancement of 62% over flat surface. Coursey et al. [8] performed spray cooling tests focusing on straight fin geometry with heights between 0.25 and 5 mm. They used PF-5060 at flow rates up to 1.0 ml/cm2 s and found that fin heights between 1 and 3 mm were optimum for heat fluxes up to 124 W/cm2 at 19 °C surface superheat. In the earlier work, Bostanci et al. [9] has evaluated two kinds of microstructured surfaces featuring indentations (micro-i) and protrusions (micro-p), in spray cooling tests with saturated anhydrous ammonia. They used vapor atomized spray nozzles with 1.6 ml/ cm2 s liquid and 13.8 ml/cm2 s vapor flow rates. At heat flux levels of up to 500 W/cm2, they observed 49% and 112% improvement in heat transfer coefficient with micro-i and micro-p surfaces, respectively, over a reference smooth surface. The current study also focused on high heat flux spray cooling with ammonia on enhanced surfaces. Compared to some other commonly used coolants, ammonia possesses important advantages such as low saturation temperature, and a relatively high latent heat of vaporization. Moreover, enhanced surfaces offer a potential to greatly improve heat transfer performance. The main objectives of the study were to investigate the effect of surface enhancement on spray cooling Critical Heat Flux (CHF) limit, and contribute to the current understanding of spray cooling heat transfer mechanisms. The experimental study used a set of optimized enhanced surfaces, as detailed by Bostanci [10], including micro-scale indentations and protrusions, macro (mm) scale pyramidal pin fins, and
Tsurf TCavg Ra V v V_ x
DTsat DTsub
e ql g
surface temperature (°C) average thermocouple reading (°C) average surface roughness (lm) voltage (V) mean droplet velocity (m/s) liquid flow rate of spray nozzle (ml/cm2 s) vertical distance between TC hole to spray surface in heater wall (m) surface superheat (°C) subcooling (°C) spray cooling effectiveness (J/ml) liquid density (kg/m3) spray cooling efficiency
multi-scale structured surfaces, combining macro and micro-scale structures. 2. Experimental setup and procedure Experiments were conducted in a closed loop spray cooling system. Fig. 1 is a schematic diagram of the system where the main components consist of a reservoir, 1 2 nozzle array, a subcooler, a condenser, and a pump. In the setup, the reservoir supplies ammonia liquid and vapor to the nozzle array. Liquid and vapor mix in the atomizer nozzles and the resulting sprays cool a 1 cm 2 cm heater where a mounted thick film resistor is the heat source. Exhaust from 1 2 nozzle array slightly subcools the incoming liquid supply in a small heat exchanger before flowing into the larger heat exchanger to condense. Finally, the two phase pump takes the liquid and vapor ammonia and transfers it back to the reservoir, providing the pressure difference that is needed to drive ammonia in the cycle and generate the spray. A separate air cooled R-22 cycle is employed to absorb the heat from the ammonia cycle and reject it to the ambient. System allows controlling flow rates and pressures across the nozzle array and is equipped with computer controlled data acquisition system for accurate data recording. 2.1. Working fluid In the present work, anhydrous ammonia (Refrigerant 717) was used as the working fluid. Ammonia has the second highest latent heat of vaporization after water among refrigerants (1368 and 2257 kJ/kg at atmospheric pressure, respectively). For applications that require low temperature operation, ammonia becomes advantageous compared to water by offering lower saturation temperature at a given pressure. Ammonia however, is not compatible with most of the commonly used engineering materials. The use of ammonia requires careful material selection and component design in cooling systems. 2.2. Spray nozzle RINI’s highly compact 1 2 vapor atomized nozzle array, shown in Fig. 2a, was used in all tests. In these nozzles a fine liquid stream is injected into a high velocity vapor stream. The shear force created by the vapor stream atomizes the liquid into fine droplets that are ejected from the nozzle orifice. The spray nozzles featured a 0.76 mm orifice size, and for the flow range of 1.5– 2.5 ml/cm2 s liquid and 12.0–20.0 ml/cm2 s vapor, they provided a full-cone spray pattern. Fig. 2a also shows the pressure and temperature measurement ports on the nozzle array that were
H. Bostanci et al. / International Journal of Heat and Mass Transfer 55 (2012) 3849–3856
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Fig. 1. Schematic diagram of the spray cooling setup.
a
Vapor inlet Liquid inlet
Vapor-atomized nozzles Pressure / temperature measurement ports
Exhaust outlet
1 cm
b Thick film resistor TC#1
c
TC#2
d
Fig. 2. RINI 1 2 vapor atomized nozzle array (a), heater-spray side (b), heater-resistor side (c), and complete 1 2 nozzle array and heater assembly (d).
needed to determine the driving pressures across the nozzle, and saturation condition of the coolant. Nozzle-to-surface distance of 11 mm was found to be adequate for each spray to cover 1 cm2 area, and the distance was held constant. 2.3. Heaters and spray surfaces The heater design employed in this study is shown in Fig. 2b and c. Heater body and enhanced structures were made of copper material, and featured a nickel coated spray surface for corrosion protection. The current study is the continuation of a previous ef-
fort that focused on investigation of the highest heat transfer coefficient on enhanced surfaces using the same experimental setup, working fluid and nozzle array. Heater design however, was modified to accommodate the higher heat loads required in the present CHF tests. Total heated area on the current design was reduced to 1 cm 1 cm featuring a 1 cm2 thick film resistor that was soldered onto heater to simulate heat generating device. Heat flux was determined from the total power supplied into the thick film resistor per unit resistor base area ðq00 ¼ V I=AÞ. Heater temperature was monitored with two type-T thermocouples embedded halfway in the heater wall and spaced equally across the 1 cm2 area as
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Table 1 Heater descriptions. Heater ID
Surface condition Enhancement type, structure geometry, structure size
s1 s2 mi-c mp-c Mpf-0.50 Mpf-0.50mi-c
Smooth, plain, machine-finished (Ra 0.3 lm) Smooth, plain, machine-finished (Ra 0.3 lm) Microstructured, indentations, coarse (Ra 5 lm) Micro-structured, protrusions, coarse (Ra 20 lm) Macro structured, pyramidal fins, 0.50 mm Multi-scale structured, pyramidal fins/indentations, 0.50 mm/coarse (Ra 5 lm) Multi-scale structured, pyramidal fins/protrusions, 0.50 mm/coarse (Ra 20 lm)
Mpf-0.50mp-c
illustrated in Fig. 2c. Temperature spread resulted from the two thermocouples did not exceed ±1 °C at 200 W/cm2, and ±1.5 °C at 500 W/cm2 indicating uniform cooling across the surface. Spray surface temperature was calculated by extrapolating average of two thermocouple readings through the known distance to the surface and assuming steady 1-D conduction through the heater wall ðT surf ¼ TCav g ðq00 xÞ=kÞ. The description and surface characteristics of the test heaters investigated in this study are provided in Table 1. Two of the heaters, with machine finished-smooth surface, served as reference for all enhanced surfaces in obtaining quantitative performance comparison. Enhanced surfaces, that can be grouped as micro-, macroand multi-scale structured surfaces, reflected the optimized structure geometry and size based on a study by Bostanci [10]. Fig. 3 includes some schematics, scanning electron microscope (SEM) images, and solid models of micro- and macrostructured surfaces to characterize the surface structures. The microstructured spray surfaces featuring indentations and protrusions were fabricated using a particle blasting and a thermal spray coating
process, respectively. While the smooth surface had a Ra roughness level of 0.3–0.5 lm, the surfaces with indentations and protrusions resulted in Ra of 5 lm and 20 lm. Although not obvious from the SEM images, the surface mi has open cavities, while the surface mp has many randomly sized re-entrant cavities. The macrostructured surface featured 0.50 mm high pyramidal fins (with 0.59 mm base length and 0.59 mm pitch) that were fabricated on a precise CNC mill. 2.4. Test conditions and procedure All the tests were conducted at 550–570 kPa (65–68 psig) and 7–8 °C conditions using saturated ammonia as the working fluid. Cooling curves that were in the form of surface superheat (DTsat = Tsurf Tsat) vs. heat flux (q00 ) were generated for each test heater by increasing the heat flux gradually and recording the corresponding heater temperatures at certain sampling rates. A series of tests were performed in order to find the optimum flow rates that balances high heat transfer rate, and reasonably low coolant usage and pumping power, as detailed in [10]. In these tests, liquid flow rate was first varied between 1.4 and 1.8 ml/ cm2 s, at a constant vapor flow rate of 15.7 ml/cm2 s. Data illustrated that the increase in liquid flow rate up to 1.6 ml/cm2 s improves the heat transfer, but further increase results in practically the same cooling curve. In the next step, effect of vapor flow rate was examined for the constant 1.6 ml/cm2 s liquid flow rate. Data then suggested that varying vapor flow rate in the range of 11.8–15.7 ml/cm2 s have no considerable effect on the heat transfer. Based on these data, optimum flow rates were determined to be 1.6 ml/cm2 s for liquid, and 13.8 ml/cm2 s for vapor (corresponding to a pressure drop of 48 kPa (7 psi) across the nozzle), and these conditions were applied in the initial tests to investigate CHF performance of enhanced surfaces. However, higher flow rates
Fig. 3. Schematics and SEM images of microstructured surfaces (a), and solid model of macrostructured surface (b).
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q" (W/cm )
of up to 2.1 ml/cm2 s for liquid, and 17.7 ml/cm2 s for vapor (corresponding to a pressure drop of 83 kPa (12 psi) across the nozzle), were employed as well to determine the effect of flow rate on CHF limits. Additional tests were performed to determine whether the magnitude and duration of the heat flux steps in the generation of cooling curves were a factor. Comparison of data from these tests matched very well and indicated flexibility in the selection of the heat flux steps in tests. Heat flux was therefore gradually increased in steps of 10–100 W/cm2 up to CHF and corresponding heater temperatures are recorded every 1–3 s over 1–2 min long steps. Adjusting steps to be smaller and quicker for the high heat flux ranges enabled keeping the overall testing time at a reasonable level while approaching CHF in a smooth and controllable pace.
600 500 400 300 CHF
200
s1 Nominal (Low) Flow Rate (1.6 ml/cm2.s liquid, 13.8 ml/cm2.s vapor)
100
s2 Ref
0 0
10
20
2.5. Uncertainty analysis
40
50
60
o
Fig. 4. CHF performance of the reference surface based on two different heaters.
1000 900 800 700 2
q" (W/cm )
Uncertainties were estimated mainly for heat flux and temperature measurements that are critical in performance evaluation. Error involved in heat flux measurement (considering variations in voltage, current, and area) was ±2.2% at 750 W/cm2. Error in temperature measurements from the embedded thermocouples in the heater wall was calibrated to be ±0.2 °C. Spray surface temperature had an uncertainty of approximately ±1.0 °C involving uncertainty in temperature extrapolation from the TC hole to the heater surface at 750 W/cm2. Heat transfer coefficients included ±3.5, 4.1 and 4.4% uncertainty at 750 W/cm2 and nominal flow rates for the surfaces s, mp-c, and Mpf-0.50, respectively. Flow rate measurements had ±5% uncertainty dictated by the flow meter characteristics. Heat loss from the thick film resistor to surrounding heater body via conduction was estimated to be approximately 9% based on finite element analysis results. Heat loss to the ambient environment was negligibly small (<1 W) based on calculations considering natural convection and black body radiation from 150 °C heater surface to 20 °C stagnant air.
30
ΔTsat ( C)
600 500
CHF
400
Ref, EF_CHF=1.00 mi-c, EF_CHF=1.01
300
mp-c, EF_CHF=1.18
200
Mpf-0.50, EF_CHF=1.10 Mpf-0.50mi-c, EF_CHF=1.06
Nominal (Low) Flow Rate (1.6 ml/cm2.s liquid, 13.8 ml/cm2.s vapor)
100
Mpf-0.50mp-c, EF_CHF=1.18
0 0
10
20
30
40
50
60
o
ΔTsat ( C)
3. Results and discussion Two heaters with the surface s were tested to obtain reference data. As shown in Fig. 4, the difference in performance between these two surfaces is negligible for the entire testing range. One of the heaters reached CHF at 760 W/cm2, while the other one attained a higher CHF at 780 W/cm2. A reference cooling curve was then established by averaging the two data sets. Study was continued by testing all the enhanced surfaces at fixed liquid/vapor flow rates to evaluate their CHF performance. Results are included in Fig. 5 along with the reference curve for comparison. Among two types of microstructured surfaces, the surface mi-c reached CHF at 780 W/cm2, while the surface mpc reached CHF at a much higher level at 910 W/cm2. These surfaces also exhibited different cooling curves. The surface mp-c entered two-phase regime at lower heat fluxes, and had higher heat transfer coefficients up to 500 W/cm2. At higher heat fluxes however, the surface mi-c performed better and reached CHF sooner. The surface mp-c transitioned to the last region of cooling curve much slower resulting in a higher CHF and surface superheat. The macrostructured surface Mpf-0.50, on the other hand reached CHF at 850 W/cm2, and indicated a superior performance throughout the testing range compared to the surface s. When it comes to multi-scale structured surfaces that combine micro- and macro-scale structures, the surface Mpf-0.50mi-c, performed better than the surface mi-c, and reached CHF at 820 W/ cm2. The other surface, Mpf-0.50mp-c, performed same as the surface mp-c, and provided a CHF of 910 W/cm2.
Fig. 5. CHF performance of the reference, and micro-, macro-, and multi-scale structured surfaces.
As far as the overall heat transfer performance is concerned, the surface mp-c was better than the surface mi-c up to 500 W/cm2, where two curves crossed over, and at higher heat fluxes the surface mi-c offered higher heat transfer coefficients. The contribution of 0.50 mm high pyramids to the performance of microstructures (indentations and protrusions) was not significant as evidenced by 1.5 °C lower superheat at 500 W/cm2. Fig. 5 also lists an enhancement factor, EFCHF, that is defined as
EF CHF ¼
CHF CHF ref
ð1Þ
to express each heater’s CHF enhancement over the reference surface. As can be noticed, while the surface mi-c offers a minimal 1% improvement, the surface mp-c provides 18% increase in CHF over the surface s. The surface Mpf-0.50 alone offers 10% enhancement. When the surface Mpf-0.50 is combined with the surfaces mi-c and mp-c, they result in 6% and 18% improvement, respectively. These data therefore suggest that CHF enhancement due to the multi-scale structures is not additive. Current results can provide some insights into spray cooling CHF enhancement mechanism. In general, rough, porous or textured surfaces would retain more liquid compared to a smooth surface for a given spray nozzle and flow rate. The textured surfaces also provide an efficient means to spread the liquid film via
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2o
h (W/m C)
300,000
200,000
Ref mi-c
100,000
mp-c Mpf-0.50 Nominal (Low) Flow Rate (1.6 ml/cm2.s liquid, 13.8 ml/cm2.s vapor)
Mpf-0.50mi-c Mpf-0.50mp-c
0 0
100
200
300
400
500
600
700
800
900
1000
2
q" (W/cm ) Fig. 6. Heat transfer coefficients as a function of heat flux for reference, and micro-, macro-, and multi-scale structured surfaces.
Table 2 Flow rates used in CHF tests.
a
Condition
Liquid flow rate (ml/cm2 s)
Vapor flow rate (ml/cm2 s)
Liquid-to-vapor loading ratioa
Nominal (low) Medium High
1.6
13.8
14.7
1.8 2.1
15.7 17.7
15.0 15.2
Based on mass flow rate.
2
q" (W/cm )
0
10
CHF
Heater Limit mi-c, low mi-c, high mp-c, low mp-c, high
10
20
30
40
50
60
ΔTsat (oC)
2
2
1100 1000 900 800 700 600 500 CHF 400 300 1s, low 1s, medium 200 100 1s, high 0 20 30 40 50 60 0 1100 ΔTsat (oC) 1000 900 800 700 600 500 400 CHF Heater Limit 300 200 Mpf-0.50, low Mpf-0.50, high 100 0 20 30 40 50 60 0
q" (W/cm )
1100 1000 900 800 700 600 500 400 300 200 100 0
2
q" (W/cm )
1100 1000 900 800 700 600 500 400 300 200 100 0
q" (W/cm )
capillary force within the micro-scale structures. These observations have also been reported by several experimental studies [1,3,6]. The capillary wicking delays the occurrence of dry patches at high heat fluxes and leads to higher CHF. In the current study, based on both qualitative visual observations, and quantitative surface roughness analysis results as detailed in [10], the surface mp-c has the highest roughness, overall structure height, and actual surface area with plenty of re-entrant
cavities. Therefore the surface mp-c is expected to hold more liquid, and the multitude of cavities can spread the liquid very efficiently, thus keeping the surface wet longer and achieving higher CHF values. The surface mi-c on the other hand, exhibits less roughness and open cavities, that can still hold more liquid than a smooth surface, but cannot resist liquid film break up as efficiently as the surface mp-c at high fluxes leading to only a slight CHF improvement over the smooth surface. The surface Mpf-0.50 naturally forms grooves between adjacent pyramids that can help manage the liquid distribution. However, as experimental data implied, its enhancement level is between the two aforementioned microstructured surfaces. The fluid retention and capillary wicking mechanism also explains why the surface mp-c performs better than the surface mi-c up to a certain heat flux, but the trend reverses afterwards during CHF tests. Although surface bubble nucleation is very effective at low to medium heat fluxes (<500 W/cm2), evaporative effects gradually become more important as the controlling factor at higher heat fluxes. Hence, the surface mi-c with thinner liquid film starts to offer higher heat transfer coefficients beyond 500 W/cm2. Eventually the thin film breaks up resulting in more pronounced dry patches, and the heater approaches CHF sooner. The surface mp-c, with a thicker liquid film, results in a lower evaporation rate and consequently lower heat transfer coefficients, but also extends the transition to CHF. Fig. 6 includes heat transfer coefficients as a function of heat flux. The curves peak at 500–700 W/cm2 heat flux, and then present a declining slope as they approach to CHF. The highest heat transfer coefficient with the surface s was 220,000 W/m2°C. Among others, surfaces mi-c, Mpf-0.50mi-c, and Mpf-0.50mp-c all reached heat transfer coefficient of approximately 300,000 W/m2°C. Once the CHF limits of the enhanced surfaces were determined using the nominal flow rates of 1.6 ml/cm2 s liquid, and 13.8 ml/ cm2 s vapor, the study was continued to investigate the effect of flow rate on CHF. For spray cooling, it has been established that increasing the liquid flow rate would help increase CHF for a given nozzle. However, the increase in CHF is not proportional to the flow rate, and the higher flow rate is only effective up to a certain level, beyond which CHF remains relatively the same regardless of
0
10
ΔTsat (oC)
CHF
Heater Limit
Mpf-0.50mi-c, low Mpf-0.50mi-c, high Mpf-0.50mp-c, low Mpf-0.50mp-c, high
10
20
30
40
50
60
ΔTsat (oC)
Fig. 7. CHF performance of the reference, and micro-, macro-, and multi-scale structured surfaces at various flow rate conditions.
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Heater limit
1100 1050
2
CHF (W/cm )
1000 950 900 Ref
850
mi-c 800
mp-c Mpf-0.50
750
Mpf-0.50mi-c
700
Mpf-0.50mp-c
650 1.5
1.6
1.7
1.8
1.9
2.0
2.1
2.2
2
Liquid Flow Rate (ml/cm .s) Fig. 8. Effect of liquid flow rate on CHF for the reference, and micro-, macro-, and multi-scale structured surfaces.
of the actual heat removed to the total heat capacity of the liquid used, including the required heat to bring the liquid from subcooled to saturation condition (sensible heat), and then to complete vaporization (latent heat). This liquid usage efficiency g, can be expressed as
q00
g¼ _ V ql ðcp DT sub þ hfg Þ
ð2Þ
where V_ is volumetric flow rate, ql is liquid density, cp is liquid specific heat, DTsub = Tsat Tl is subcooling, and hfg is latent heat of vaporization. Since saturated spray conditions were maintained throughout this study, total heat capacity of the used liquid is equivalent to its latent heat of vaporization. Spray cooling efficiencies of all surfaces at their respective CHF values (or heater limits) are summarized in Fig. 9 as a function of liquid flow rate. At low flow rate condition, efficiency can be seen to range from 63.4%, for the surface s, to 74.9% for the surfaces mp-c and Mpf-0.50mpc. Complete data set from the reference surface was again used to determine the overall trend, which indicates that when flow rate is increased from low to medium, efficiency slightly goes up (1%), but with further flow rate increase efficiency starts to decrease since CHF remains nearly the same. The highest efficiency reached during this study was therefore approximately 75% for both the surfaces mp-c and Mpf-0.50mp-c. Besides efficiency, another performance parameter, spray cooling effectiveness e, with the resulting units of J/ml is defined as
q00 A
e¼ _ V
ð3Þ
693
80
η and ε at heater limit
70
60
50
40 1.5
616
539
462
Ref mi-c mp-c Mpf-0.50 Mpf-0.50mi-c Mpf-0.50mp-c 1.6
ε at CHF (J/ml)
90
η at CHF (%)
flow rate [11,12]. This can be attributed to the counterbalance of various spray cooling heat transfer mechanisms driven by the advantage of higher droplet velocity and the disadvantage of higher film thickness. Other studies [13,14] use the mean droplet velocity (v), the mean spray droplet flux (N), and the Sauter-mean droplet diameter (d32) as three independent spray parameters which affect the CHF. These studies, utilizing extensive experimental data, found that CHF varies with V1/4 and N1/6, and is relatively independent of d32. Although these spray parameters are not specifically measured in the current work, the parameters are proportional to flow rate, where an increase in flow rate would increase V, N and d32 simultaneously. The use of higher flow rates for improved CHF performance also brings system level implications, such as higher pumping power, affects cooling efficiency, and thus requires further optimization. When CHF is the main design consideration, however, higher flow rates would be preferred. The present work therefore investigated the CHF limits of the enhanced surfaces at the flow rates listed in Table 2. Fig. 7 shows all the data from CHF tests at low, medium, and high flow rates. The figure includes four separate plots for easier comparison between reference, and micro-, macro-, and multiscale structured surfaces. The surface s was tested at all three flow rate conditions, and the CHF values at medium and high flow rates were the same at 930 W/cm2. Data thus suggested that increasing flow rate beyond the medium level has no considerable effect on CHF. Once this trend was established, the other surfaces were tested only at high flow rates in addition to nominal/low flow rates. All heaters with enhanced surfaces, however, consistently failed during high flow rate CHF tests at heat fluxes starting at approximately 960 W/cm2. These data points are marked as ‘‘heater limit’’ in the plots, to distinguish them from the CHF condition. Considering the elevated temperatures at these heat flux levels, cracking of the thick film resistors was most likely due to the stresses induced by thermal expansion mismatch. As a result, true CHF value for these conditions could not be experimentally obtained. The surfaces Mpf-0.50mp-c and mp-c attained the highest heat flux, but not CHF, of 1090 W/cm2 before the heater damage occurred. As far as the overall heat transfer performance, besides CHF, is concerned, higher flow rates only aided the surface s where higher heat transfer coefficients were achieved. For other surfaces, high flow rate generally resulted in slightly higher superheats at heat fluxes lower than 700 W/cm2. In an effort to further confirm the effect of flow rate on CHF, some additional tests were conducted with the surface s at medium and high flow rates. Data indicated that CHF varies between 760 and 780 W/cm2 for low, 890–930 W/cm2 for medium, and 920–930 W/cm2 for the high flow rates, and suggest that CHF is not necessarily a very repeatable quantity, and can occur over a narrow heat flux range. Based on this set of data in Fig. 8, a curve was fitted to represent the effect of flow rate on the CHF for the outlined conditions. The trend once again implies that increasing flow rate beyond a certain level (the medium flow rate in this case) has a minimal effect on CHF, and is consistent with observations made by earlier studies. Fig. 8 also incorporates CHF data, or highest recorded heat fluxes at heater limit otherwise, from all other tests. Using the curve fit for the surface s as a guideline, it was offset to estimate the effect of flow rate on CHF for other surfaces. The amount of offset was determined based on the experimental CHF values at low flow rate. During a close examination of the data, this approach actually seems reasonable since heat fluxes at heater limit, for the surfaces Mpf-0.50 and Mpf-0.50mp-c at high flow rate, match or exceed the estimated CHF values. As mentioned before, when attempting to achieve higher CHF with higher flow rates, another performance aspect to consider is the spray cooling efficiency. In general it can be defined as the ratio
385
1.7
1.8
1.9
2.0
2.1
308 2.2
Liquid Flow Rate (ml/cm2.s) Fig. 9. Effect of liquid flow rate on spray cooling efficiency and effectiveness for the reference, and micro-, macro-, and multi-scale structured surfaces.
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and can be useful in comparison of alternative coolants and spray nozzle designs. Effectiveness values from all CHF tests are also included in Fig. 9. 4. Conclusions The present work focused on high heat flux spray cooling with ammonia. A set of optimized enhanced surfaces, including microscale indentations and protrusions, macro (mm) scale pyramidal pin fins, and multi-scale structured surfaces, combining macro and micro-scale structures, was investigated to determine the CHF limits and gain a better understanding on the underlying mechanism. Based on the results, the following conclusions are drawn: The surfaces Mpf-0.50mp-c and mp-c had the highest CHF value of 910 W/cm2 with the nominal flow rates of 1.6 ml/cm2 s liquid and 13.8 ml/cm2 s vapor, corresponding to 18% increase over the reference smooth surface. When the effect of higher liquid flow rate was investigated, most of the heaters experienced resistor failures at heat fluxes >950 W/cm2 before they reach CHF. However, the effect of flow rate was still captured, and it was estimated that the surface Mpf-0.50mp-c can attain a CHF value of approximately 1100 W/cm2. Enhanced surfaces are capable of retaining more liquid compared to a smooth surface, and efficiently spread the liquid film via capillary force within the structures. Fluid retention delays the occurrence of dry patches at high heat fluxes, and leads to higher CHF. Overall, the present study demonstrated that the spray cooling with ammonia on enhanced surfaces offers a very high level of heat flux removing capability, which previously considered to be achievable by using water only. Having ammonia as a highly capable coolant alternative would enable unique thermal management applications that demand operation of high power devices at low temperatures, such as laser systems.
Acknowledgments We acknowledge US Air Force Research Laboratory (AFRL) Propulsion Directorate and Universal Technology Corporation for their financial support. References [1] M.R. Pais, L.C. Chow, E.T. Mahefkey, Surface roughness and its effects on the heat transfer mechanism in spray cooling, J. Heat Transfer 114 (1992) 211– 219. [2] M.S. Sehmbey, M.R. Pais, L.C. Chow, Effect of surface material properties and surface characteristics in evaporative spray cooling, J. Thermophys. Heat Transfer 6 (1992) 505–511. [3] J.H. Kim, S.M. You, U.S. Choi, Evaporative spray cooling of plain and microporous coated surfaces, Int. J. Heat Mass Transfer 47 (2004) 3307–3315. [4] C. Stodke, P. Stephan, Spray cooling heat transfer on microstructured surfaces, in: Proceedings of the 6th World Conference on Experimental Heat Transfer, Fluid Mechanics, and Thermodynamics, Matsushima, Miyagi, Japan, 2005. [5] C. Amon, S.C. Yao, C.F. Wu, C.C. Hsieh, Microelectromechanical system-based evaporative thermal management of high heat flux electronics, J. Heat Transfer 127 (2005) 66–75. [6] C.C. Hsieh, S.C. Yao, Evaporative heat transfer characteristics of a water spray on micro-structured silicon surfaces, Int. J. Heat Mass Transfer 49 (2006) 962– 974. [7] E. Silk, J. Kim, K.T. Kiger, Enhanced surface spray cooling with embedded and compound extended surface structures, in: Proceedings of ITHERM 2006, San Diego, CA, 2006. [8] J. Coursey, J. Kim, K.T. Kiger, Spray cooling of high aspect ratio open microchannels, in: Proceedings of ITHERM 2006, San Diego, CA, 2006. [9] H. Bostanci, D.P. Rini, J.P. Kizito, L.C. Chow, Spray cooling with ammonia on microstructured surfaces: performance enhancement and hysteresis effect, J. Heat Transfer 131 (2009). 071401-1-9. [10] H. Bostanci, High Heat Flux Spray Cooling with Ammonia on Enhanced Surfaces, Ph.D. Dissertation, University of Central Florida, Orlando, Florida, 2010. [11] L.C. Chow, M.S. Sehmbey, M.R. Pais, High heat flux spray cooling, in: C.-L. Tien (Ed.), Annual Review of Heat Transfer, vol. 8, Hemisphere Pub. Corp., 1997, pp. 291–318. [12] J. Yang, M.R. Pais, L.C. Chow, Critical heat flux limits in secondary gas atomized liquid spray cooling, Exp. Heat Transfer 6 (1993) 55–67. [13] R.-H. Chen, L.C. Chow, J.E. Navedo, Effects of spray characteristics on critical in subcooled water spray cooling, Int. J. Heat Mass Transfer 45 (2002) 4033– 4043. [14] R.-H. Chen, L.C. Chow, J.E. Navedo, Optimal spray characteristics in water spray cooling, Int. J. Heat Mass Transfer 47 (2004) 5095–5099.
Contents lists available at SciVerse ScienceDirect
International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
High heat flux spray cooling with ammonia: Investigation of enhanced surfaces for CHF Huseyin Bostanci a,⇑, Daniel P. Rini a, John P. Kizito b, Virendra Singh c, Sudipta Seal c, Louis C. Chow c a
RINI Technologies, Inc., 582 S. Econ Circle, Oviedo, FL 32765, United States Department of Mechanical and Chemical Engineering, North Carolina A&T State University, Greensboro, NC 27411, United States c Department of Mechanical, Materials and Aerospace Engineering, University of Central Florida, Orlando, FL 32816, United States b
a r t i c l e
i n f o
Article history: Received 2 June 2011 Received in revised form 5 March 2012 Available online 20 April 2012 Keywords: Spray cooling Thermal management CHF Enhanced surfaces
a b s t r a c t A spray cooling study was conducted to investigate the effect of enhanced surfaces on Critical Heat Flux (CHF). Test surfaces involved micro-scale indentations and protrusions, macro (mm) scale pyramidal pin fins, and multi-scale structured surfaces, combining macro and micro-scale structures, along with a smooth surface that served as reference. Tests were conducted in a closed loop system using a vapor atomized spray nozzle with ammonia as the working fluid. Nominal flow rates were 1.6 ml/cm2 s of liquid and 13.8 ml/cm2 s of vapor, resulting in a pressure drop of 48 kPa. Results indicated that the multi-scale structured surface helped increase maximum heat flux limit by 18% over the reference smooth surface, to 910 W/cm2 at nominal flow rate. During the additional CHF testing at higher flow rates, most heaters experienced failures before reaching CHF at heat fluxes above 950 W/cm2. However, some enhanced surfaces can achieve CHF values of up to 1100 W/cm2 with 67% spray cooling efficiency based on liquid usage. The results also shed some light on the current understanding of the spray cooling heat transfer mechanisms. Enhanced surfaces are found to be capable of retaining more liquid compared to a smooth surface, and efficiently spread the liquid film via capillary force within the structures. This important advantage delays the occurrence of dry patches at high heat fluxes, and leads to higher CHF. The present work demonstrated ammonia spray cooling as a unique alternative for challenging thermal management tasks that call for high heat flux removal while maintaining a low device temperature with a compact and efficient cooling scheme. Ó 2012 Elsevier Ltd. All rights reserved.
1. Introduction Many critical applications today, in electronics, optics and aerospace fields, among others, demand advanced thermal management solutions for the acquisition of high heat loads they generate in order to operate reliably and efficiently. Current competing technologies for this challenging task include several single and two phase cooling options. When these cooling schemes are compared based on the high heat flux removal (100–1000 W/ cm2) and isothermal operation (within several °C across the cooled device) aspects, as well as system mass, volume and power consumption, spray cooling appears to be the best choice. Although a vast amount of research has been done on heat transfer enhancement in general, studies focusing on spray cooling enhancement are fairly limited. In early works, Pais et al. [1] and Sehmbey et al. [2] examined the effects of surface roughness and contact angle using water with air atomized nozzle, at flow rates up to 1.4 ml/cm2 s water and 400 ml/cm2 s air, and found enhance-
⇑ Corresponding author. E-mail address: [email protected] (H. Bostanci). 0017-9310/$ - see front matter Ó 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.ijheatmasstransfer.2012.03.040
ment in heat transfer coefficient with decreasing surface roughness and increasing contact angle. They obtained heat fluxes up to 1250 W/cm2 at 11 °C surface superheat on ultrasmooth (Ra = 0.3 lm) copper surface. Kim et al. [3] investigated spray cooling enhancement on microporous coated surfaces using water at flow rates up to 0.03 ml/ cm2 s. The porous layer was fabricated using a mixture of methyl–ethyl-ketone (MEK), epoxy, and aluminum powder, and its maximum thickness was 500 lm. They found that the CHF increased 50% relative to the uncoated surface. However, highest heat flux reached was 3.2 W/cm2 at 65 °C surface superheat due to very low flow rates. Stodke and Stephan [4] studied spray cooling on microstructured and micro-porous surfaces using water at 6 kPa system pressure and 1.4 ml/cm2 s flow rate. Pyramidal micro-grooves and micro-pyramids with 75 lm height increased the wetted area by a factor of 1.4. Their 100-lm thick porous layers were very similar to those used by Kim et al. [3] and created with the same ingredients of MEK, epoxy, and aluminum powder. A maximum heat flux of 97 W/cm2 was observed for the micro-pyramid surface compared to 30 W/cm2 for the flat surface, both at a surface superheat of 12 °C. This enhancement was much larger than the surface area
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Nomenclature A CHF CHFref cp d32 EFCHF h hfg I k N q00 Tl Tsat
area (cm2) Critical Heat Flux (W/cm2) Critical Heat Flux of reference surface (W/cm2) liquid specific heat (J/kg °C) Sauter-mean droplet size (m) CHF enhancement factor heat transfer coefficient (W/m2 °C) latent heat of vaporization (kJ/kg) current (A) heater wall thermal conductivity (W/m °C) mean droplet flux (#/m2 s) heat flux (W/cm2) liquid temperature (°C) saturation temperature (°C)
enhancement. However, when a micro-porous surface was used, a significant degradation in heat transfer occurred when compared to the plain surface resulting in 14 W/cm2 maximum heat flux at 12 °C superheat. The low heat flux was due to the poor thermal conductivity of the epoxy binder. Amon et al. [5] and Hsieh and Yao [6] performed spray cooling experiments with water on silicon substrates using 1 2 nozzle array at very low flow rates of up to 0.07 ml/cm2 s on square micro-studs with 160–480 lm heights. The surface texture was found to have little effect in the single phase and dry out regimes. The authors attributed the higher heat transfer observed for the micro-textured surfaces in the intermediate regimes to more effective spreading of the liquid by capillary forces. The maximum heat flux achieved was just over 50 W/cm2 at 55 °C surface superheat. Silk et al. [7] investigated the effect of surface geometry in spray cooling with PF-5060 using 2 2 nozzle array at 1.6 ml/cm2 s flow rate. They used embedded structures (dimples, pores and tunnels) and compound extended structures (straight fins, cubic pin fins and dimples) all in the order of 1 mm size. Of these macrostructured surfaces, straight fins and porous tunnels performed the best providing a CHF of 175 W/cm2 for gassy conditions at surface superheats of up to 36 °C and offered CHF enhancement of 62% over flat surface. Coursey et al. [8] performed spray cooling tests focusing on straight fin geometry with heights between 0.25 and 5 mm. They used PF-5060 at flow rates up to 1.0 ml/cm2 s and found that fin heights between 1 and 3 mm were optimum for heat fluxes up to 124 W/cm2 at 19 °C surface superheat. In the earlier work, Bostanci et al. [9] has evaluated two kinds of microstructured surfaces featuring indentations (micro-i) and protrusions (micro-p), in spray cooling tests with saturated anhydrous ammonia. They used vapor atomized spray nozzles with 1.6 ml/ cm2 s liquid and 13.8 ml/cm2 s vapor flow rates. At heat flux levels of up to 500 W/cm2, they observed 49% and 112% improvement in heat transfer coefficient with micro-i and micro-p surfaces, respectively, over a reference smooth surface. The current study also focused on high heat flux spray cooling with ammonia on enhanced surfaces. Compared to some other commonly used coolants, ammonia possesses important advantages such as low saturation temperature, and a relatively high latent heat of vaporization. Moreover, enhanced surfaces offer a potential to greatly improve heat transfer performance. The main objectives of the study were to investigate the effect of surface enhancement on spray cooling Critical Heat Flux (CHF) limit, and contribute to the current understanding of spray cooling heat transfer mechanisms. The experimental study used a set of optimized enhanced surfaces, as detailed by Bostanci [10], including micro-scale indentations and protrusions, macro (mm) scale pyramidal pin fins, and
Tsurf TCavg Ra V v V_ x
DTsat DTsub
e ql g
surface temperature (°C) average thermocouple reading (°C) average surface roughness (lm) voltage (V) mean droplet velocity (m/s) liquid flow rate of spray nozzle (ml/cm2 s) vertical distance between TC hole to spray surface in heater wall (m) surface superheat (°C) subcooling (°C) spray cooling effectiveness (J/ml) liquid density (kg/m3) spray cooling efficiency
multi-scale structured surfaces, combining macro and micro-scale structures. 2. Experimental setup and procedure Experiments were conducted in a closed loop spray cooling system. Fig. 1 is a schematic diagram of the system where the main components consist of a reservoir, 1 2 nozzle array, a subcooler, a condenser, and a pump. In the setup, the reservoir supplies ammonia liquid and vapor to the nozzle array. Liquid and vapor mix in the atomizer nozzles and the resulting sprays cool a 1 cm 2 cm heater where a mounted thick film resistor is the heat source. Exhaust from 1 2 nozzle array slightly subcools the incoming liquid supply in a small heat exchanger before flowing into the larger heat exchanger to condense. Finally, the two phase pump takes the liquid and vapor ammonia and transfers it back to the reservoir, providing the pressure difference that is needed to drive ammonia in the cycle and generate the spray. A separate air cooled R-22 cycle is employed to absorb the heat from the ammonia cycle and reject it to the ambient. System allows controlling flow rates and pressures across the nozzle array and is equipped with computer controlled data acquisition system for accurate data recording. 2.1. Working fluid In the present work, anhydrous ammonia (Refrigerant 717) was used as the working fluid. Ammonia has the second highest latent heat of vaporization after water among refrigerants (1368 and 2257 kJ/kg at atmospheric pressure, respectively). For applications that require low temperature operation, ammonia becomes advantageous compared to water by offering lower saturation temperature at a given pressure. Ammonia however, is not compatible with most of the commonly used engineering materials. The use of ammonia requires careful material selection and component design in cooling systems. 2.2. Spray nozzle RINI’s highly compact 1 2 vapor atomized nozzle array, shown in Fig. 2a, was used in all tests. In these nozzles a fine liquid stream is injected into a high velocity vapor stream. The shear force created by the vapor stream atomizes the liquid into fine droplets that are ejected from the nozzle orifice. The spray nozzles featured a 0.76 mm orifice size, and for the flow range of 1.5– 2.5 ml/cm2 s liquid and 12.0–20.0 ml/cm2 s vapor, they provided a full-cone spray pattern. Fig. 2a also shows the pressure and temperature measurement ports on the nozzle array that were
H. Bostanci et al. / International Journal of Heat and Mass Transfer 55 (2012) 3849–3856
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Fig. 1. Schematic diagram of the spray cooling setup.
a
Vapor inlet Liquid inlet
Vapor-atomized nozzles Pressure / temperature measurement ports
Exhaust outlet
1 cm
b Thick film resistor TC#1
c
TC#2
d
Fig. 2. RINI 1 2 vapor atomized nozzle array (a), heater-spray side (b), heater-resistor side (c), and complete 1 2 nozzle array and heater assembly (d).
needed to determine the driving pressures across the nozzle, and saturation condition of the coolant. Nozzle-to-surface distance of 11 mm was found to be adequate for each spray to cover 1 cm2 area, and the distance was held constant. 2.3. Heaters and spray surfaces The heater design employed in this study is shown in Fig. 2b and c. Heater body and enhanced structures were made of copper material, and featured a nickel coated spray surface for corrosion protection. The current study is the continuation of a previous ef-
fort that focused on investigation of the highest heat transfer coefficient on enhanced surfaces using the same experimental setup, working fluid and nozzle array. Heater design however, was modified to accommodate the higher heat loads required in the present CHF tests. Total heated area on the current design was reduced to 1 cm 1 cm featuring a 1 cm2 thick film resistor that was soldered onto heater to simulate heat generating device. Heat flux was determined from the total power supplied into the thick film resistor per unit resistor base area ðq00 ¼ V I=AÞ. Heater temperature was monitored with two type-T thermocouples embedded halfway in the heater wall and spaced equally across the 1 cm2 area as
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Table 1 Heater descriptions. Heater ID
Surface condition Enhancement type, structure geometry, structure size
s1 s2 mi-c mp-c Mpf-0.50 Mpf-0.50mi-c
Smooth, plain, machine-finished (Ra 0.3 lm) Smooth, plain, machine-finished (Ra 0.3 lm) Microstructured, indentations, coarse (Ra 5 lm) Micro-structured, protrusions, coarse (Ra 20 lm) Macro structured, pyramidal fins, 0.50 mm Multi-scale structured, pyramidal fins/indentations, 0.50 mm/coarse (Ra 5 lm) Multi-scale structured, pyramidal fins/protrusions, 0.50 mm/coarse (Ra 20 lm)
Mpf-0.50mp-c
illustrated in Fig. 2c. Temperature spread resulted from the two thermocouples did not exceed ±1 °C at 200 W/cm2, and ±1.5 °C at 500 W/cm2 indicating uniform cooling across the surface. Spray surface temperature was calculated by extrapolating average of two thermocouple readings through the known distance to the surface and assuming steady 1-D conduction through the heater wall ðT surf ¼ TCav g ðq00 xÞ=kÞ. The description and surface characteristics of the test heaters investigated in this study are provided in Table 1. Two of the heaters, with machine finished-smooth surface, served as reference for all enhanced surfaces in obtaining quantitative performance comparison. Enhanced surfaces, that can be grouped as micro-, macroand multi-scale structured surfaces, reflected the optimized structure geometry and size based on a study by Bostanci [10]. Fig. 3 includes some schematics, scanning electron microscope (SEM) images, and solid models of micro- and macrostructured surfaces to characterize the surface structures. The microstructured spray surfaces featuring indentations and protrusions were fabricated using a particle blasting and a thermal spray coating
process, respectively. While the smooth surface had a Ra roughness level of 0.3–0.5 lm, the surfaces with indentations and protrusions resulted in Ra of 5 lm and 20 lm. Although not obvious from the SEM images, the surface mi has open cavities, while the surface mp has many randomly sized re-entrant cavities. The macrostructured surface featured 0.50 mm high pyramidal fins (with 0.59 mm base length and 0.59 mm pitch) that were fabricated on a precise CNC mill. 2.4. Test conditions and procedure All the tests were conducted at 550–570 kPa (65–68 psig) and 7–8 °C conditions using saturated ammonia as the working fluid. Cooling curves that were in the form of surface superheat (DTsat = Tsurf Tsat) vs. heat flux (q00 ) were generated for each test heater by increasing the heat flux gradually and recording the corresponding heater temperatures at certain sampling rates. A series of tests were performed in order to find the optimum flow rates that balances high heat transfer rate, and reasonably low coolant usage and pumping power, as detailed in [10]. In these tests, liquid flow rate was first varied between 1.4 and 1.8 ml/ cm2 s, at a constant vapor flow rate of 15.7 ml/cm2 s. Data illustrated that the increase in liquid flow rate up to 1.6 ml/cm2 s improves the heat transfer, but further increase results in practically the same cooling curve. In the next step, effect of vapor flow rate was examined for the constant 1.6 ml/cm2 s liquid flow rate. Data then suggested that varying vapor flow rate in the range of 11.8–15.7 ml/cm2 s have no considerable effect on the heat transfer. Based on these data, optimum flow rates were determined to be 1.6 ml/cm2 s for liquid, and 13.8 ml/cm2 s for vapor (corresponding to a pressure drop of 48 kPa (7 psi) across the nozzle), and these conditions were applied in the initial tests to investigate CHF performance of enhanced surfaces. However, higher flow rates
Fig. 3. Schematics and SEM images of microstructured surfaces (a), and solid model of macrostructured surface (b).
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q" (W/cm )
of up to 2.1 ml/cm2 s for liquid, and 17.7 ml/cm2 s for vapor (corresponding to a pressure drop of 83 kPa (12 psi) across the nozzle), were employed as well to determine the effect of flow rate on CHF limits. Additional tests were performed to determine whether the magnitude and duration of the heat flux steps in the generation of cooling curves were a factor. Comparison of data from these tests matched very well and indicated flexibility in the selection of the heat flux steps in tests. Heat flux was therefore gradually increased in steps of 10–100 W/cm2 up to CHF and corresponding heater temperatures are recorded every 1–3 s over 1–2 min long steps. Adjusting steps to be smaller and quicker for the high heat flux ranges enabled keeping the overall testing time at a reasonable level while approaching CHF in a smooth and controllable pace.
600 500 400 300 CHF
200
s1 Nominal (Low) Flow Rate (1.6 ml/cm2.s liquid, 13.8 ml/cm2.s vapor)
100
s2 Ref
0 0
10
20
2.5. Uncertainty analysis
40
50
60
o
Fig. 4. CHF performance of the reference surface based on two different heaters.
1000 900 800 700 2
q" (W/cm )
Uncertainties were estimated mainly for heat flux and temperature measurements that are critical in performance evaluation. Error involved in heat flux measurement (considering variations in voltage, current, and area) was ±2.2% at 750 W/cm2. Error in temperature measurements from the embedded thermocouples in the heater wall was calibrated to be ±0.2 °C. Spray surface temperature had an uncertainty of approximately ±1.0 °C involving uncertainty in temperature extrapolation from the TC hole to the heater surface at 750 W/cm2. Heat transfer coefficients included ±3.5, 4.1 and 4.4% uncertainty at 750 W/cm2 and nominal flow rates for the surfaces s, mp-c, and Mpf-0.50, respectively. Flow rate measurements had ±5% uncertainty dictated by the flow meter characteristics. Heat loss from the thick film resistor to surrounding heater body via conduction was estimated to be approximately 9% based on finite element analysis results. Heat loss to the ambient environment was negligibly small (<1 W) based on calculations considering natural convection and black body radiation from 150 °C heater surface to 20 °C stagnant air.
30
ΔTsat ( C)
600 500
CHF
400
Ref, EF_CHF=1.00 mi-c, EF_CHF=1.01
300
mp-c, EF_CHF=1.18
200
Mpf-0.50, EF_CHF=1.10 Mpf-0.50mi-c, EF_CHF=1.06
Nominal (Low) Flow Rate (1.6 ml/cm2.s liquid, 13.8 ml/cm2.s vapor)
100
Mpf-0.50mp-c, EF_CHF=1.18
0 0
10
20
30
40
50
60
o
ΔTsat ( C)
3. Results and discussion Two heaters with the surface s were tested to obtain reference data. As shown in Fig. 4, the difference in performance between these two surfaces is negligible for the entire testing range. One of the heaters reached CHF at 760 W/cm2, while the other one attained a higher CHF at 780 W/cm2. A reference cooling curve was then established by averaging the two data sets. Study was continued by testing all the enhanced surfaces at fixed liquid/vapor flow rates to evaluate their CHF performance. Results are included in Fig. 5 along with the reference curve for comparison. Among two types of microstructured surfaces, the surface mi-c reached CHF at 780 W/cm2, while the surface mpc reached CHF at a much higher level at 910 W/cm2. These surfaces also exhibited different cooling curves. The surface mp-c entered two-phase regime at lower heat fluxes, and had higher heat transfer coefficients up to 500 W/cm2. At higher heat fluxes however, the surface mi-c performed better and reached CHF sooner. The surface mp-c transitioned to the last region of cooling curve much slower resulting in a higher CHF and surface superheat. The macrostructured surface Mpf-0.50, on the other hand reached CHF at 850 W/cm2, and indicated a superior performance throughout the testing range compared to the surface s. When it comes to multi-scale structured surfaces that combine micro- and macro-scale structures, the surface Mpf-0.50mi-c, performed better than the surface mi-c, and reached CHF at 820 W/ cm2. The other surface, Mpf-0.50mp-c, performed same as the surface mp-c, and provided a CHF of 910 W/cm2.
Fig. 5. CHF performance of the reference, and micro-, macro-, and multi-scale structured surfaces.
As far as the overall heat transfer performance is concerned, the surface mp-c was better than the surface mi-c up to 500 W/cm2, where two curves crossed over, and at higher heat fluxes the surface mi-c offered higher heat transfer coefficients. The contribution of 0.50 mm high pyramids to the performance of microstructures (indentations and protrusions) was not significant as evidenced by 1.5 °C lower superheat at 500 W/cm2. Fig. 5 also lists an enhancement factor, EFCHF, that is defined as
EF CHF ¼
CHF CHF ref
ð1Þ
to express each heater’s CHF enhancement over the reference surface. As can be noticed, while the surface mi-c offers a minimal 1% improvement, the surface mp-c provides 18% increase in CHF over the surface s. The surface Mpf-0.50 alone offers 10% enhancement. When the surface Mpf-0.50 is combined with the surfaces mi-c and mp-c, they result in 6% and 18% improvement, respectively. These data therefore suggest that CHF enhancement due to the multi-scale structures is not additive. Current results can provide some insights into spray cooling CHF enhancement mechanism. In general, rough, porous or textured surfaces would retain more liquid compared to a smooth surface for a given spray nozzle and flow rate. The textured surfaces also provide an efficient means to spread the liquid film via
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H. Bostanci et al. / International Journal of Heat and Mass Transfer 55 (2012) 3849–3856 400,000
2o
h (W/m C)
300,000
200,000
Ref mi-c
100,000
mp-c Mpf-0.50 Nominal (Low) Flow Rate (1.6 ml/cm2.s liquid, 13.8 ml/cm2.s vapor)
Mpf-0.50mi-c Mpf-0.50mp-c
0 0
100
200
300
400
500
600
700
800
900
1000
2
q" (W/cm ) Fig. 6. Heat transfer coefficients as a function of heat flux for reference, and micro-, macro-, and multi-scale structured surfaces.
Table 2 Flow rates used in CHF tests.
a
Condition
Liquid flow rate (ml/cm2 s)
Vapor flow rate (ml/cm2 s)
Liquid-to-vapor loading ratioa
Nominal (low) Medium High
1.6
13.8
14.7
1.8 2.1
15.7 17.7
15.0 15.2
Based on mass flow rate.
2
q" (W/cm )
0
10
CHF
Heater Limit mi-c, low mi-c, high mp-c, low mp-c, high
10
20
30
40
50
60
ΔTsat (oC)
2
2
1100 1000 900 800 700 600 500 CHF 400 300 1s, low 1s, medium 200 100 1s, high 0 20 30 40 50 60 0 1100 ΔTsat (oC) 1000 900 800 700 600 500 400 CHF Heater Limit 300 200 Mpf-0.50, low Mpf-0.50, high 100 0 20 30 40 50 60 0
q" (W/cm )
1100 1000 900 800 700 600 500 400 300 200 100 0
2
q" (W/cm )
1100 1000 900 800 700 600 500 400 300 200 100 0
q" (W/cm )
capillary force within the micro-scale structures. These observations have also been reported by several experimental studies [1,3,6]. The capillary wicking delays the occurrence of dry patches at high heat fluxes and leads to higher CHF. In the current study, based on both qualitative visual observations, and quantitative surface roughness analysis results as detailed in [10], the surface mp-c has the highest roughness, overall structure height, and actual surface area with plenty of re-entrant
cavities. Therefore the surface mp-c is expected to hold more liquid, and the multitude of cavities can spread the liquid very efficiently, thus keeping the surface wet longer and achieving higher CHF values. The surface mi-c on the other hand, exhibits less roughness and open cavities, that can still hold more liquid than a smooth surface, but cannot resist liquid film break up as efficiently as the surface mp-c at high fluxes leading to only a slight CHF improvement over the smooth surface. The surface Mpf-0.50 naturally forms grooves between adjacent pyramids that can help manage the liquid distribution. However, as experimental data implied, its enhancement level is between the two aforementioned microstructured surfaces. The fluid retention and capillary wicking mechanism also explains why the surface mp-c performs better than the surface mi-c up to a certain heat flux, but the trend reverses afterwards during CHF tests. Although surface bubble nucleation is very effective at low to medium heat fluxes (<500 W/cm2), evaporative effects gradually become more important as the controlling factor at higher heat fluxes. Hence, the surface mi-c with thinner liquid film starts to offer higher heat transfer coefficients beyond 500 W/cm2. Eventually the thin film breaks up resulting in more pronounced dry patches, and the heater approaches CHF sooner. The surface mp-c, with a thicker liquid film, results in a lower evaporation rate and consequently lower heat transfer coefficients, but also extends the transition to CHF. Fig. 6 includes heat transfer coefficients as a function of heat flux. The curves peak at 500–700 W/cm2 heat flux, and then present a declining slope as they approach to CHF. The highest heat transfer coefficient with the surface s was 220,000 W/m2°C. Among others, surfaces mi-c, Mpf-0.50mi-c, and Mpf-0.50mp-c all reached heat transfer coefficient of approximately 300,000 W/m2°C. Once the CHF limits of the enhanced surfaces were determined using the nominal flow rates of 1.6 ml/cm2 s liquid, and 13.8 ml/ cm2 s vapor, the study was continued to investigate the effect of flow rate on CHF. For spray cooling, it has been established that increasing the liquid flow rate would help increase CHF for a given nozzle. However, the increase in CHF is not proportional to the flow rate, and the higher flow rate is only effective up to a certain level, beyond which CHF remains relatively the same regardless of
0
10
ΔTsat (oC)
CHF
Heater Limit
Mpf-0.50mi-c, low Mpf-0.50mi-c, high Mpf-0.50mp-c, low Mpf-0.50mp-c, high
10
20
30
40
50
60
ΔTsat (oC)
Fig. 7. CHF performance of the reference, and micro-, macro-, and multi-scale structured surfaces at various flow rate conditions.
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Heater limit
1100 1050
2
CHF (W/cm )
1000 950 900 Ref
850
mi-c 800
mp-c Mpf-0.50
750
Mpf-0.50mi-c
700
Mpf-0.50mp-c
650 1.5
1.6
1.7
1.8
1.9
2.0
2.1
2.2
2
Liquid Flow Rate (ml/cm .s) Fig. 8. Effect of liquid flow rate on CHF for the reference, and micro-, macro-, and multi-scale structured surfaces.
of the actual heat removed to the total heat capacity of the liquid used, including the required heat to bring the liquid from subcooled to saturation condition (sensible heat), and then to complete vaporization (latent heat). This liquid usage efficiency g, can be expressed as
q00
g¼ _ V ql ðcp DT sub þ hfg Þ
ð2Þ
where V_ is volumetric flow rate, ql is liquid density, cp is liquid specific heat, DTsub = Tsat Tl is subcooling, and hfg is latent heat of vaporization. Since saturated spray conditions were maintained throughout this study, total heat capacity of the used liquid is equivalent to its latent heat of vaporization. Spray cooling efficiencies of all surfaces at their respective CHF values (or heater limits) are summarized in Fig. 9 as a function of liquid flow rate. At low flow rate condition, efficiency can be seen to range from 63.4%, for the surface s, to 74.9% for the surfaces mp-c and Mpf-0.50mpc. Complete data set from the reference surface was again used to determine the overall trend, which indicates that when flow rate is increased from low to medium, efficiency slightly goes up (1%), but with further flow rate increase efficiency starts to decrease since CHF remains nearly the same. The highest efficiency reached during this study was therefore approximately 75% for both the surfaces mp-c and Mpf-0.50mp-c. Besides efficiency, another performance parameter, spray cooling effectiveness e, with the resulting units of J/ml is defined as
q00 A
e¼ _ V
ð3Þ
693
80
η and ε at heater limit
70
60
50
40 1.5
616
539
462
Ref mi-c mp-c Mpf-0.50 Mpf-0.50mi-c Mpf-0.50mp-c 1.6
ε at CHF (J/ml)
90
η at CHF (%)
flow rate [11,12]. This can be attributed to the counterbalance of various spray cooling heat transfer mechanisms driven by the advantage of higher droplet velocity and the disadvantage of higher film thickness. Other studies [13,14] use the mean droplet velocity (v), the mean spray droplet flux (N), and the Sauter-mean droplet diameter (d32) as three independent spray parameters which affect the CHF. These studies, utilizing extensive experimental data, found that CHF varies with V1/4 and N1/6, and is relatively independent of d32. Although these spray parameters are not specifically measured in the current work, the parameters are proportional to flow rate, where an increase in flow rate would increase V, N and d32 simultaneously. The use of higher flow rates for improved CHF performance also brings system level implications, such as higher pumping power, affects cooling efficiency, and thus requires further optimization. When CHF is the main design consideration, however, higher flow rates would be preferred. The present work therefore investigated the CHF limits of the enhanced surfaces at the flow rates listed in Table 2. Fig. 7 shows all the data from CHF tests at low, medium, and high flow rates. The figure includes four separate plots for easier comparison between reference, and micro-, macro-, and multiscale structured surfaces. The surface s was tested at all three flow rate conditions, and the CHF values at medium and high flow rates were the same at 930 W/cm2. Data thus suggested that increasing flow rate beyond the medium level has no considerable effect on CHF. Once this trend was established, the other surfaces were tested only at high flow rates in addition to nominal/low flow rates. All heaters with enhanced surfaces, however, consistently failed during high flow rate CHF tests at heat fluxes starting at approximately 960 W/cm2. These data points are marked as ‘‘heater limit’’ in the plots, to distinguish them from the CHF condition. Considering the elevated temperatures at these heat flux levels, cracking of the thick film resistors was most likely due to the stresses induced by thermal expansion mismatch. As a result, true CHF value for these conditions could not be experimentally obtained. The surfaces Mpf-0.50mp-c and mp-c attained the highest heat flux, but not CHF, of 1090 W/cm2 before the heater damage occurred. As far as the overall heat transfer performance, besides CHF, is concerned, higher flow rates only aided the surface s where higher heat transfer coefficients were achieved. For other surfaces, high flow rate generally resulted in slightly higher superheats at heat fluxes lower than 700 W/cm2. In an effort to further confirm the effect of flow rate on CHF, some additional tests were conducted with the surface s at medium and high flow rates. Data indicated that CHF varies between 760 and 780 W/cm2 for low, 890–930 W/cm2 for medium, and 920–930 W/cm2 for the high flow rates, and suggest that CHF is not necessarily a very repeatable quantity, and can occur over a narrow heat flux range. Based on this set of data in Fig. 8, a curve was fitted to represent the effect of flow rate on the CHF for the outlined conditions. The trend once again implies that increasing flow rate beyond a certain level (the medium flow rate in this case) has a minimal effect on CHF, and is consistent with observations made by earlier studies. Fig. 8 also incorporates CHF data, or highest recorded heat fluxes at heater limit otherwise, from all other tests. Using the curve fit for the surface s as a guideline, it was offset to estimate the effect of flow rate on CHF for other surfaces. The amount of offset was determined based on the experimental CHF values at low flow rate. During a close examination of the data, this approach actually seems reasonable since heat fluxes at heater limit, for the surfaces Mpf-0.50 and Mpf-0.50mp-c at high flow rate, match or exceed the estimated CHF values. As mentioned before, when attempting to achieve higher CHF with higher flow rates, another performance aspect to consider is the spray cooling efficiency. In general it can be defined as the ratio
385
1.7
1.8
1.9
2.0
2.1
308 2.2
Liquid Flow Rate (ml/cm2.s) Fig. 9. Effect of liquid flow rate on spray cooling efficiency and effectiveness for the reference, and micro-, macro-, and multi-scale structured surfaces.
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and can be useful in comparison of alternative coolants and spray nozzle designs. Effectiveness values from all CHF tests are also included in Fig. 9. 4. Conclusions The present work focused on high heat flux spray cooling with ammonia. A set of optimized enhanced surfaces, including microscale indentations and protrusions, macro (mm) scale pyramidal pin fins, and multi-scale structured surfaces, combining macro and micro-scale structures, was investigated to determine the CHF limits and gain a better understanding on the underlying mechanism. Based on the results, the following conclusions are drawn: The surfaces Mpf-0.50mp-c and mp-c had the highest CHF value of 910 W/cm2 with the nominal flow rates of 1.6 ml/cm2 s liquid and 13.8 ml/cm2 s vapor, corresponding to 18% increase over the reference smooth surface. When the effect of higher liquid flow rate was investigated, most of the heaters experienced resistor failures at heat fluxes >950 W/cm2 before they reach CHF. However, the effect of flow rate was still captured, and it was estimated that the surface Mpf-0.50mp-c can attain a CHF value of approximately 1100 W/cm2. Enhanced surfaces are capable of retaining more liquid compared to a smooth surface, and efficiently spread the liquid film via capillary force within the structures. Fluid retention delays the occurrence of dry patches at high heat fluxes, and leads to higher CHF. Overall, the present study demonstrated that the spray cooling with ammonia on enhanced surfaces offers a very high level of heat flux removing capability, which previously considered to be achievable by using water only. Having ammonia as a highly capable coolant alternative would enable unique thermal management applications that demand operation of high power devices at low temperatures, such as laser systems.
Acknowledgments We acknowledge US Air Force Research Laboratory (AFRL) Propulsion Directorate and Universal Technology Corporation for their financial support. References [1] M.R. Pais, L.C. Chow, E.T. Mahefkey, Surface roughness and its effects on the heat transfer mechanism in spray cooling, J. Heat Transfer 114 (1992) 211– 219. [2] M.S. Sehmbey, M.R. Pais, L.C. Chow, Effect of surface material properties and surface characteristics in evaporative spray cooling, J. Thermophys. Heat Transfer 6 (1992) 505–511. [3] J.H. Kim, S.M. You, U.S. Choi, Evaporative spray cooling of plain and microporous coated surfaces, Int. J. Heat Mass Transfer 47 (2004) 3307–3315. [4] C. Stodke, P. Stephan, Spray cooling heat transfer on microstructured surfaces, in: Proceedings of the 6th World Conference on Experimental Heat Transfer, Fluid Mechanics, and Thermodynamics, Matsushima, Miyagi, Japan, 2005. [5] C. Amon, S.C. Yao, C.F. Wu, C.C. Hsieh, Microelectromechanical system-based evaporative thermal management of high heat flux electronics, J. Heat Transfer 127 (2005) 66–75. [6] C.C. Hsieh, S.C. Yao, Evaporative heat transfer characteristics of a water spray on micro-structured silicon surfaces, Int. J. Heat Mass Transfer 49 (2006) 962– 974. [7] E. Silk, J. Kim, K.T. Kiger, Enhanced surface spray cooling with embedded and compound extended surface structures, in: Proceedings of ITHERM 2006, San Diego, CA, 2006. [8] J. Coursey, J. Kim, K.T. Kiger, Spray cooling of high aspect ratio open microchannels, in: Proceedings of ITHERM 2006, San Diego, CA, 2006. [9] H. Bostanci, D.P. Rini, J.P. Kizito, L.C. Chow, Spray cooling with ammonia on microstructured surfaces: performance enhancement and hysteresis effect, J. Heat Transfer 131 (2009). 071401-1-9. [10] H. Bostanci, High Heat Flux Spray Cooling with Ammonia on Enhanced Surfaces, Ph.D. Dissertation, University of Central Florida, Orlando, Florida, 2010. [11] L.C. Chow, M.S. Sehmbey, M.R. Pais, High heat flux spray cooling, in: C.-L. Tien (Ed.), Annual Review of Heat Transfer, vol. 8, Hemisphere Pub. Corp., 1997, pp. 291–318. [12] J. Yang, M.R. Pais, L.C. Chow, Critical heat flux limits in secondary gas atomized liquid spray cooling, Exp. Heat Transfer 6 (1993) 55–67. [13] R.-H. Chen, L.C. Chow, J.E. Navedo, Effects of spray characteristics on critical in subcooled water spray cooling, Int. J. Heat Mass Transfer 45 (2002) 4033– 4043. [14] R.-H. Chen, L.C. Chow, J.E. Navedo, Optimal spray characteristics in water spray cooling, Int. J. Heat Mass Transfer 47 (2004) 5095–5099.