Effects of particle volume fraction on spray heat transfer performance of Al2O3–water nanofluid

International Journal of Heat and Mass Transfer 55 (2012) 1014–1021

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Effects of particle volume fraction on spray heat transfer performance of Al2O3–water nanofluid Tong-Bou Chang ⇑, Siou-Ci Syu, Yen-Kai Yang Department of Mechanical Engineering, Southern Taiwan University, Tainan, Taiwan

a r t i c l e

i n f o

Article history: Received 18 March 2011 Received in revised form 26 July 2011 Accepted 26 July 2011 Available online 1 November 2011 Keywords: Spray cooling Nanofluid Heat transfer enhancement Nano-sorption layer

a b s t r a c t An experimental investigation is conducted into the effects of the particle volume fraction on the spray heat transfer performance of a nanofluid comprising de-ionized water and Al2O3 particles with a diameter of 35 nm. The tests are performed with a flat, horizontal heated surface using a nozzle with an orifice diameter of 0.7 mm and a nozzle-to-heated surface distance of 17 mm. The spray mass flux is varied in the range of 26.433–176.751 kg/m2 s, while the particle volume fraction is specified as 0%, 0.001%, 0.025%, or 0.05%. It is found that the optimal heat transfer performance is obtained using a particle volume fraction of 0.001%. The surface compositions of the sprayed samples are observed using scanning electron microscopy. The results show that the surfaces sprayed with a nanofluid containing 0.025 Vol% or 0.05 Vol% of nanoparticles contain a small amount of Al. However, those cooled using a nanofluid with a particle volume fraction of 0% or 0.001% show no traces of Al. Ó 2011 Elsevier Ltd. All rights reserved.

1. Introduction Advances in nano-fabrication technology in recent years have led to many innovations in the photonics, biotechnology, electronics, medicine, chemistry, energy, materials and mechanics fields. For example, it has been shown that nano-structures increase the convection heat transfer coefficient of the liquid phase and inhibit pool boiling phase-change heat transfer. Spray cooling, in which the heated surface is cooled via the evaporation of an impinging jet of droplets, is a highly effective means of transferring large amounts of energy at low temperatures. Accordingly, the present study examines the heat transfer performance of an enhanced cooling method in which the heated surface is spray cooled using a working fluid containing nano-scale metallic particles. The addition of metal or metal oxide nanoparticles to traditional coolants (e.g., water, glycol and refrigerants) in order to improve their thermal conductivity and enhance the heat transfer performance was first demonstrated by Choi and Eastman [1] in 1995. Wang and Xu [2] showed that the addition of nanoparticles to liquid coolants increases the thermal conductivity, improves the overall cooling efficiency of the system, reduces the operation costs, and increases the maximum heat flux. Das et al. [3] analyzed the pool boiling heat transfer performance of an Al2O3–water nanofluid with nanoparticle additions ranging from 1 to 4 Vol%. ⇑ Corresponding author. Address: Department of Mechanical Engineering, Southern Taiwan University, No. 1, Nan-Tai Street, YongKang Dist., Tainan 710, Taiwan. Tel.: +886 6 2533131x3533; fax: +886 6 2425092. E-mail address: [email protected] (T.-B. Chang). 0017-9310/$ - see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.ijheatmasstransfer.2011.10.009

The experimental results showed that some of the nanoparticles were deposited on the heated surface during boiling, thereby reducing the boiling heat transfer performance. Moreover, the extent to which the heat transfer performance was reduced was found to increase with an increasing nanoparticle addition. You et al. [4] showed that there was no obvious change in the thermal conductivity of an Al2O3–water nanofluid when the level of particle addition was specified as 0.001 Vol%. Bang and Chang [5] studied the boiling heat transfer performance of an Al2O3–water nanofluid on a horizontal heated surface given an average nanoparticle diameter of 47 nm and particle volume fractions of 0.5%, 1%, 2% and 4%, respectively. The results showed that the use of a nanofluid reduced the pool boiling heat transfer performance relative to that achieved using pure deionized (DI) water. As in [3], it was suggested that the lower heat transfer performance was the result of the formation of a thin nano-film on the heated surface during boiling. Kim et al. [6] investigated the pool boiling heat transfer performance of three different types of nanofluid (Al2O3–water, ZrO2–water and SiO2–water) for particle volume fractions of 0.001%, 0.01% and 0.1%. The experimental results showed that for all three nanofluids, a thin nano-porous layer was formed on the heated surface, which enhanced the wettability of the surface, reduced the static contact angle and increased the CHF (critical heat flux). Kwark et al. [7] conducted nucleate pool boiling experiments using a low-concentration Al2O3 nanofluid under atmospheric pressure conditions. It was shown that the deposition of a thin nanoparticle film on the heated surface not only increased the CHF, but also resulted in transient characteristics in the nucleate boiling heat transfer performance. Recently, Abu-Nada and Oztop

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Nomenclature h kbase kNFs Kparticles m q00

heat transfer coefficient (W/m2 K) thermal conductivity of base fluid thermal conductivity of nanofluid thermal conductivity of particles spray mass flux (kg/m2 s) wall heat flux (W/m2)

[8] performed a numerical investigation into the heat transfer performance of Al2O3–water nanofluid and showed that the addition of Al2O3 nanoparticles yielded an effective increase in the mean Nusselt number compared to that obtained when using pure DI water alone. Hodgson and Sutherland [9] investigated the heat transfer performance of direct spray cooling given a heated surface with a temperature higher than the Leidenfrost temperature. Grissom and Wierum [10] found that in the liquid spray cooling of a heated flat surface, the impinging droplets form a liquid film, whose thickness directly determines the boiling heat transfer performance. Choi and Yao [11] and Pais et al. [12] showed that a high spray rate results in the breakup of the liquid layer on the heated surface, and is therefore beneficial in improving the heat transfer performance. Chow et al. [13] performed an experimental investigation into the heat transfer performance of high power density evaporative spray cooling. The results were used to identify the optimal heat transfer conditions and to design an improved spray cooling system. It was shown that the enhanced cooling system yielded an effective improvement in both the CHF and the heat transfer efficiency. Lin and Linhard [14] conducted experimental measurements of the heat flux on a thin metal plate heated from one side with a plasma arc and cooled from the other by an impinging water jet. It was shown that no CHF existed and that the heat flux was limited only by the power of the heating source. Hsieh and Yao [15] studied the evaporative heat transfer characteristics of a water spray impinging on a flat micro-structured silicon surface. The experimental results showed that the capillary force induced by the micro-structured surface enhanced the ability of the liquid film to coat the heated surface and increased the evaporation efficiency as a result. Chang et al. [16] investigated the heat transfer performance of a shell-and-tube evaporator system incorporating interior spray nozzles, and showed that the impinging droplets postponed the dry-out phenomenon on the lower surfaces of the tubes and therefore resulted in a better heat transfer performance than pool boiling. In general, the results presented in the studies above indicate that nanofluid coolants reduce the boiling heat transfer performance as a result of the formation of a thin nanoparticle layer on the heated surface, whereas spray cooling improves the heat transfer performance due to an enhanced latent evaporation cooling effect. Thus, the present study examines the heat transfer performance of a cooling system in which a low-concentration Al2O3–water nanofluid is sprayed onto the heated surface. It is shown that the impinging droplets suppress the formation of a thin nanoparticle layer on the surface and improve the heat transfer performance as a result.

T Tsat Tw Vol%

temperature (°C) saturation temperature (°C) wall temperature (°C) particle volume fraction (%)

module and a spray nozzle, a cooling water flow loop, a sub-cooler, a pre-heating system, and a data acquisition system. The nanofluid flow loop had the form of a closed-type circulation system and was designed in such a way as to ensure that the liquid entered the test section at the desired flow rate and temperature. During the experiments, the sprayed mass flux was varied in the range of 0.0208–0.0555 kg/s by regulating the differential pressure

Fig. 1. Schematic illustration of experimental system.

Tw T1 O ring

x1

T2

x2

T3

x3

T4

x4

2. Experimental apparatus Fig. 1 presents a schematic illustration of the experimental spray cooling system used in this study. As shown, the system comprised a nanofluid flow loop, a test section containing a heating

heater Fig. 2. Schematic illustration of heating module.

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across the spray nozzle using a bypass valve installed on the outlet side of the fluid pump. The test section comprised a cylindrical stainless steel vessel with a height of 50 cm, an internal diameter of 25 cm, and a thickness of 1 cm. The evaporated vapor was condensed by a coil condenser installed in the upper part of the test section. The resulting condensate dripped off the coil, mixed with the non-evaporated liquid, and then flowed out of the test section via an outlet pipe fitted to its lower surface. For observation purposes, the test section was fitted with two view-glass ports. The temperature of the refrigerant vapor was measured using an RTD meter, while its stability was monitored using an electrical pressure sensor. The spray tests were performed using a circular, horizontal flat plate as the test surface and a full-cone circular hydraulic nozzle with an orifice diameter of 0.7 mm, an outlet diameter of 2.3 mm and a cone angle of 15°. The nozzle was fixed at a vertical distance of 17 mm above the heated test surface. Fig. 2 presents a schematic illustration of the heating module; comprising a cylindrical heated block, an exposed circular flat plate, and three heaters. The heated block was fabricated of copper and measured 20 mm in diameter and 52 mm in height. The block was heated electrically by three 300 W resistor-type heaters installed within the base. The test plate was mounted on the upper surface of the heated block and had a diameter of 20 mm. To evaluate the effects of surface roughness on the heat transfer performance, the test plates were polished to a finish of 0.2 lm or 1.40 lm prior to installation on the block. Each spray test was performed using a new test plate. Having examined the surface composition of the sprayed surface, the plate was sealed in a glass bottle and retained for further examination if required. The spray tests were conducted using a nanofluid comprising DI water and Al2O3 particles with a diameter of 35 nm. To investigate the effect of the particle volume fraction on the heat transfer efficiency, the tests were conducted using nanofluids with 0 Vol%, 0.001 Vol%, 0.025 Vol%, and 0.05 Vol% nanoparticle addition, respectively. The data of interest in the experimental system, e.g., the mass flow rate and temperature of the inlet refrigerant, the heater temperature, the surface heat flux, and so forth, were interfaced to a PC, converted into digitized readable values and displayed on a monitor. Having waited for approximately 30 min for the system to reach steady-state conditions (as indicated by a variation of less than 0.1 °C/min in the saturation temperature), each data point of interest was obtained by averaging a minimum of 20 data-acquisition scans. 3. Experimental procedure and data reduction 3.1. Water–Al2O3 nanofluid preparation The nanofluids used in the spray tests were prepared by mixing commercial Al2O3 nano powder in DI water. According to Wen et al. [17], nanoparticles distributed in a base fluid tend to aggregate under the effects of Brownian motion and the Van der Waals force and then sink to the bottom of the container. Therefore, to ensure a good suspension stability, the present nanofluids were prepared using an ultrasonic bath and a magnetic stirrer. As discussed above, nanofluids were prepared with 0 Vol%, 0.001 Vol%, 0.025 Vol%, and 0.05 Vol% Al2O3 nanoparticle addition, respectively, where the particle volume fraction, /, was defined as wal

/¼

where wal2 o3 : mass of Al2O3 (kg)

wwater: mass of deionized water (kg)

qwater: density of deionized water (kg/m3) 3.2. Thermal conductivity measurement The thermal conductivity of the Al2O3–water nanofluids was measured using a Decagon Pro KD2 analyzer equipped with a hot-wire measuring probe. Before the measurement tests, the KD2 analyzer was calibrated by inserting the measuring probe into a calibration liquid bottle. 3.3. Experimental data reduction In the spray tests, the spray mass flux of the nanofluid was varied in the range of 26.433–176.751 kg/m2 s, while the heat flux was varied from 105 W/m2 to more than 106 W/m2. In accordance with the empirical correlation developed originally by Longwell [18] and subsequently simplified by Bonacina and Comini [19], the droplet diameter was computed as

dp ¼

9:5dn ðDpÞ

0:37

sinðb=2Þ

ð2Þ

;

where dn is the nozzle orifice diameter (mm), Dp is the pressure difference across the nozzle (Pa), and b is the nozzle spray angle (°). Table 1 summarizes the pressure difference, droplet diameter and spray velocity for each of the spray mass flux conditions considered in the present study. The mean heat transfer coefficient, h, was determined in accordance with Newton’s cooling law as

h¼

q00 ; T w  T sat

ð3Þ

where q00 is the heat flux on the test surface (W/m2), Tw is the temperature of the test surface (°C), and Tsat is the saturation temperature (°C). The temperature of the test surface was calculated from the temperature readings obtained from two thermocouples (TCs) installed within the interior of the heated block under the assumption of a one-dimensional heat conduction process, i.e.,

Tw ¼ T2 

x2 ðT 2  T 1 Þ; x2  x1

ð4Þ

where Tw: temperature of the test surface (°C). x1: distance from test surface to first temperature measurement position (mm). x2: distance from test surface to second temperature measurement position (mm). T1: temperature at first measurement position (°C). T2: temperature at second measurement position (°C). To ensure the accuracy of the measured data, all of the measurement devices within the experimental setup were calibrated prior to use. Following calibration, the accuracies of the temperature sensors (RTD and TC), power meter and flow meter were found to be ±0.1 °C, ±0.1% and ±0.001 kg/s, respectively. Based upon these

Table 1 Summary of spray parameters.

2 o3

qal2 o3 Volume of Al2 O3 ¼ Volume of Al2 O3 þ Volume of DI water wal2 o3 þ wwater qal2 o3

qal2 o3 : density of Al2O3 (kg/m3)

qwater

ð1Þ

Spray mass flux (kg/m2 s)

Dp (Pa)

dp (mm)

Vp (m/s)

26.43 66.24 128.03 176.75

1.54  106 1.84  106 2.0  106 2.10  106

0.613 0.574 0.557 0.547

2.0 5.01 9.68 13.36

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1.6

1.08

Ra=0.20um m=128.03 kg/m 2s 0.001 Vol%

Present data Bhattacharya[21]

DI Water 0.05 Vol%

1.06 1.4

h(W/cm2.k)

Thermal conductivity ratio (knf/kbf)

0.025 Vol%

1.04

1.2

1.02

1

1

0.02

0.04

0.06

0.08

0.1

20

40

60

80

100

120

q"(W/cm2)

Vol% Fig. 3. Variation of thermal conductivity ratio with particle volume fraction. (Note that the theoretical results obtained using Bhattacharya’s equation [21] are also presented for comparison purposes.)

Fig. 5. Variation of heat transfer coefficient with surface heat flux as function of particle volume fraction for test surface roughness of 0.2 lm.

the Brownian motion-based equation proposed by Bhattacharya et al. [21], i.e., calibration results, and applying the propagation-of-error method presented in [20], the uncertainty in the experimental heat transfer coefficient measurements was determined to be less than 4.3%.

4. Results and discussion 4.1. Verification of thermal conductivity measurements Fig. 3 compares the measured values of the thermal conductivity of the present nanofluids with the theoretical values predicted using

K NFs ¼ ð1  /ÞK base þ K particles /;

where Kparticles is the thermal conductivity of the Al2O3 nanoparticles, Kbase is the thermal conductivity of the base fluid (DI water), and / is the volume fraction of the Al2O3 nanoparticles.As expected, Fig. 3 shows that the thermal conductivity of the nanofluid increases with an increasing nanoparticle addition. A good general agreement is observed between the experimental and theoretical results. However, it is noted that the experimental values of the thermal conductivity are approximately 1.5% higher than the theoretical values. This discrepancy arises since the theoretical equation 1.6

250

Ra=1.40 um m=128.03 kg/m 2s 0.001 Vol%

DI-water m=26.43 kg/m2 s Tsat= 66.3 oC Lin et al[14] Present data

200

ð4Þ

0.025 Vol% DI Water 0.05 Vol%

h(W/cm2.k)

q"(W/cm2)

1.4 150

100

1.2

50

1 0

20 0

10

20

Tw-Tsat Fig. 4. Spray heat transfer performance achieved using DI water.

30

40

60

80

100

120

q"(W/cm2) Fig. 6. Variation of heat transfer coefficient with surface heat flux as function of particle volume fraction for test surface roughness of 1.4 lm.

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Fig. 7. SEM images and energy spectrum charts of nanoparticle deposits/coatings on heated surface with roughness of 0.20 lm given particle volume fractions of (a) 0%, (b) 0.001%, (c) 0.025%, and (d) 0.05%.

neglects the effects of the particle properties (e.g., the material, nature and size) and the aggregation of the individual particles. 4.2. Verification of heat flux measurements Fig. 4 compares the experimental results for the heat flux on the test surface with the equivalent results presented by Lin and

Ponnappan [22]. Note that the working fluid is pure DI water in both cases. Note also that the roughness of the test surface is 0.20 lm. It can be seen that both spray cooling curves exhibit a linear dependency on the surface temperature. However, for a given value of Tw  Tsat, the heat flux reported by Lin et al. is around 10–15% higher than that observed in the present study. This difference arises because the roughness of the heated surface and the

T.-B. Chang et al. / International Journal of Heat and Mass Transfer 55 (2012) 1014–1021

nozzle-to-heated surface distance are not explicitly stated in [22] and may therefore differ from those in the present study. 4.3. Effect of nanoparticle volume fraction on spray cooling heat transfer Fig. 5 shows the variation of the heat transfer coefficient with the surface heat flux given a heated surface roughness of 0.20 lm

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and nanoparticle volume fractions of 0% (DI water), 0.001%, 0.025% and 0.05%, respectively. It can be seen that the nanofluid with 0.001 Vol% particle addition yields the best heat transfer performance, followed by the nanofluids with 0.025 Vol%, 0 Vol%, and 0.05 Vol% particle addition, respectively. On average, the heat transfer coefficients obtained using nanofluids with particle volume fractions of 0.001% and 0.025%, respectively, are around

Fig. 8. SEM images and energy spectrum charts of nanoparticle deposits/coatings on heated surface with roughness of 1.4 lm given particle volume fractions of (a) 0%, (b) 0.001%, (c) 0.025%, and (d) 0.05%.

T.-B. Chang et al. / International Journal of Heat and Mass Transfer 55 (2012) 1014–1021

11.5% and 5.8% higher than that obtained when using DI water as the working fluid. By contrast, the heat transfer coefficient obtained using a nanofluid with 0.05 Vol% particle addition is around 2.6% lower than that obtained when using pure DI water. In other words, a very low level of Al2O3 addition (i.e., 0.001 Vol%) enhances the spray heat transfer performance, while a high level of Al2O3 addition (i.e., 0.05 Vol%) degrades the heat transfer performance. Fig. 6 shows the variation of the heat transfer coefficient with the surface heat flux as a function of the nanoparticle volume fraction for a test surface with a roughness of 1.40 lm. As in the previous case, the nanofluid with a particle volume fraction of 0.001% results in the best heat transfer performance, followed by the nanofluids with 0.025 Vol%, 0 Vol% and 0.05 Vol% nanoparticle addition, respectively. Compared to the heat transfer coefficient obtained when using pure DI water as the working fluid, the heat transfer coefficients obtained when using nanofluids with particle volume fractions of 0.025% or 0.05% are around 14.3% and 8.7% higher, respectively, while that obtained using a nanofluid with 0.05 Vol% particle addition is around 0.3% lower. It is noted that these values are all higher than the equivalent values obtained for the heated surface with a roughness of 0.2 lm (i.e., 11.5%, 5.8% and 2.6%, respectively). In other words, the results suggest that the spray cooling heat transfer performance of low-concentration nanofluids improves as the surface roughness of the heated surface increases. To clarify the enhanced spray cooling heat transfer performance of the low-concentration nanofluids, the surface compositions and surface profiles of the various spray cooled specimens were examined using scanning electron microscopy (PHILIPS Co., XL-40FEG), field emission scanning electron microscopy and energy-dispersive X-ray spectrometry. Fig. 7(a–d) presents the SEM images and surface composition analysis results for test surfaces with a surface roughness of 0.20 lm spray cooled using nanofluids with 0 Vol%, 0.001 Vol%, 0.025 Vol%, and 0.05 Vol% nanoparticle addition, respectively. Fig. 7(a) and (b) shows that the test surfaces spray cooled using DI water or a nanofluid with a particle volume fraction of 0.001% contain no traces of Al. Thus, it is inferred that when spray cooling is performed using a low-concentration nanofluid, most of the nanoparticles rebound directly from the heated surface, while those few particles which remain are washed away by subsequently-arriving droplets. Fig. 7(c) and (d) shows that the test surfaces cooled using nanofluids with 0.025 Vol% and 0.05 Vol% particle addition, respectively, contain 2.67% and 8.14% Al. In other words, given a higher volume fraction of nanoparticles, the impinging jet spray does not remove all of the nanoparticles from the heated surface. The deposited nanoparticles form an irregularshaped nano-sorption layer [23], which obstructs convective heat transfer between the heated surface and the impinging fluid and reduces the number of effective nucleation points on the surface. As a result, the spray cooling efficiency is significantly reduced. As the volume fraction of nanoparticles in the nanofluid increases, the thickness of the nano-sorption layer also increases. Consequently, the spray cooling heat transfer performance deteriorates with an increasing level of nanoparticle addition. Fig. 8(a–d) presents the SEM images and spectrum analysis results for the spray-cooled test surfaces with a surface roughness of 1.40 lm. The surfaces spray cooled with DI water or a nanofluid with a volume fraction of 0.001% contain no Al elements (see Fig. 8(a) and (b)). However, the surfaces sprayed with nanofluids containing 0.025 Vol% and 0.05 Vol% particle addition, respectively, have an Al content of 3.84% and 5.83%. The results presented above show that the nanofluid containing 0.001 Vol% Al2O3 nanoparticles yields the best spray cooling heat transfer performance. Accordingly, Fig. 9 shows the variation of the heat transfer coefficient with the surface heat flux for a nanofluid with a particle volume fraction of 0.001% and spray mass flux

2 0.001 Vol% Ra=0.20um m=176.75 kg/m2s m=128.03 kg/m2s m=66.24 kg/m2s

1.8

1.6

h(W/cm2.k)

1020

1.4

1.2

1

0.8 20

40

60

80

100

120

q"(W/cm2) Fig. 9. Variation of heat transfer coefficient with surface heat flux as function of spray mass flux for surface roughness of 0.2 lm.

in the range of 66.24176.75 kg/m2s. As expected, the results show that the cooling performance improves as the spray mass flux increases. Specifically, the heat transfer coefficient given a mass flux of 176.75 kg/m2s is around 10.9% higher than that obtained with a mass flux of 128.03 kg/m2s, which in turn is approximately 11.2% higher than that obtained for a mass flux of 66.24 kg/ m2s. The enhanced heat transfer performance at a higher spray mass flux can be attributed to two main factors. First, as the spray mass flux increases, the kinetic energy of the droplets also increases, and thus the forced convection effect between the nanofluid and the heated surface is enhanced. Second, a higher spray mass flux increases the droplet flux of the nozzle orifice. Consequently, the number of nucleation points on the heated surface increases, and therefore the heat transfer efficiency is improved. 5. Conclusions It is well-known that the heat transfer performance of nanofluids is significantly better than that of traditional working fluids such as water or glycol. However, previous studies have shown that in the pool boiling heat transfer mode, the nanoparticles form an irregular-shaped nano-sorption layer on the heated surface which obstructs the heat transfer between the surface and the fluid. Accordingly, the present study has examined the heat transfer performance of an enhanced cooling system in which Al2O3– water nanofluid is sprayed onto the heated surface, thereby suppressing the formation of the nanoparticle layer. The experimental results and observations support the following conclusions: 1. In spray cooling with high-volume-fraction nanofluids (i.e., 0.025 Vol% and 0.05 Vol%), the nanoparticles are easily deposited on the heated surface, thereby reducing the number of nucleation points and hindering the convection heat transfer mechanism between the surface and the nanofluid. As a result, high-volume-fraction nanofluids are unsuitable for spray cooling applications. 2. Low-volume-fraction nanofluids (i.e., 0.001 Vol%) yield a significant improvement in the spray cooling efficiency since most of the nanoparticles rebound from the heated surface directly or

T.-B. Chang et al. / International Journal of Heat and Mass Transfer 55 (2012) 1014–1021

are washed away by subsequently-arriving droplets. In other words, a nano-sorption film is not formed, and thus the thermal resistance at the fluid–surface interface is reduced.

Acknowledgment This study was supported by the National Science Council of Taiwan under Grant Nos. NSC 98-2221-E-218-044 and NSC 992221-E-218-013. References [1] S. Choi, J.A. Eastman, Enhancing thermal conductivity of fluids with nanoparticles, ASME International Mechanical Engineering Congress and Exposition, San Francisco, CA, 1995, pp. 99–105. [2] X.W. Wang, X.F. Xu, Thermal conductivity of nanoparticle–fluid mixture, Journal of Thermophysics and Heat Transfer 13 (4) (1999) 474–480. [3] S.K. Das, N. Putra, W. Roetzel, Pool boiling characteristics of nanofluids, International Journal of Heat and Mass Transfer 46 (2003) 851–862. [4] S.M. You, J.H. Kim, K.H. Kim, Effect of nanoparticles on critical heat flux of water in pool boiling heat transfer, Applied Physics Letters 83 (2003) 3374– 3376. [5] I.C. Bang, S.H. Chang, Boiling heat transfer performance and phenomena of Al2O3–water nanofluids from a plain surface in a pool, International Journal of Heat and Mass Transfer 48 (12) (2005) 2407–2419. [6] S.J. Kim, I.C. Bang, J. Buongiorno, L.W. Hu, Surface wettability change during pool boiling of nanofluids and its effect on critical heat flux, International Journal of Heat and Mass Transfer 12 (2007) 4105–4116. [7] S.M. Kwark, R. Kumar, G. Moreno, J. Yoo, S.M. You, Pool boiling characteristics of low concentration nanofluids, International Journal of Heat and Mass Transfer 53 (5–6) (2010) 972–981. [8] E. Abu-Nada, H.F. Oztop, Numerical analysis of Al2O3/water nanofluids natural convection in a wavy walled cavity, Numerical Heat Transfer, Part A: Applications 59 (5) (2011) 403–419.

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