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Heat transfer of nanofluids in the mini-rectangular fin heat sinks
International Communications in Heat and Mass Transfer 40 (2013) 25–31
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Heat transfer of nanofluids in the mini-rectangular fin heat sinks☆ Paisarn Naphon ⁎, Lursukd Nakharintr Thermo-Fluid and Heat Transfer Enhancement Lab. (TFHT), Department of Mechanical Engineering, Faculty of Engineering, Srinakharinwirot University, 63 Rangsit-Nakhornnayok Rd., Ongkharak, Nakhorn-Nayok, 26120, Thailand
a r t i c l e
i n f o
Available online 22 October 2012 Keywords: Nanofluids Minichannel heat sink Convective heat transfer
a b s t r a c t In the present study, the heat transfer characteristics of nanofluids cooling in the mini-rectangular fin heat sink are studied. The heat sinks with three different channel heights are fabricated from the aluminum by the wire electrical discharge machine with the length, width and base thickness of 110, 60, and 2 mm, respectively. The nanofluids are the mixture of de-ionized water and nanoscale TiO2 particles. The results obtained from the nanofluids cooling in mini-rectangular fin heat sink are compared with those from the de-ionized water cooling method. Effects of the inlet temperature of nanofluids, nanofluid Reynolds number, and heat flux on the heat transfer characteristics of mini-rectangular fin heat sink are considered. It is found that average heat transfer rates for nanofluids as coolant are higher than those for the de-ionized water as coolant. The results of this study are of technological importance for the efficient design of cooling systems of electronic devices to enhance cooling performance. © 2012 Elsevier Ltd. All rights reserved.
1. Introduction In order to ensure reliable operation, the personal computer or electronic devices must be operated in specific temperature ranges. The exceeding maximum allowable temperature is a serious problem for these devices. Therefore, in order to keep the operated temperature constant, the cooling system must dissipate the heat equal or higher than the generated heat from these devices. Due to the conventional air or liquid cooling systems limitation and the small physical size of the electronic devices, the development of the miniaturized technology, mini and micro-components has been increased especially in the electronic devices. The development of the miniaturized technology, mini and micro-components has been introduced as one of the heat transfer enhancement techniques. The heat transfer and pressure drop in the mini and micro-channel have been widely studied by researchers. Peng and Peterson [1] experimentally investigated on the single-phase forced convective heat transfer and the flow characteristics of water in microchannel. Adams et al. [2] investigated the single-phase turbulent flow of water in circular microchannels. A generalized correlation for the Nusselt number for single-phase turbulent flow in circular microchannels was proposed. Gillot et al. [3] applied the single-phase and two-phase micro heat sinks for cooling of power components. Qu and Mudawar [4] experimentally and numerically investigated the pressure drop and heat transfer characteristics of a single-phase micro-channel heat sink. The results obtained from the numerical prediction were
☆ Communicated by W.J. Minkowycz. ⁎ Corresponding author. E-mail address: [email protected] (P. Naphon). 0735-1933/$ – see front matter © 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.icheatmasstransfer.2012.10.012
reasonable agreement with the measured data. Hetsroni et al. [5] applied the heat sink for cooling the electronic devices. Agostini et al. [6] experimentally studied on the friction factor and heat transfer coefficient for a vertical liquid up-flow of R134a. Owhaib and Palm [7] studied the heat transfer characteristics of single-phase forced convection of R134a through single circular micro-channels. The results were compared with those from the proposed correlations in macroscale and microscale channels. Agostini et al. [8] investigated on the friction factor and heat transfer coefficient of R134a flowing through the rectangular mini-channels. Yue et al. [9] considered the pressure drop characteristics of single and two-phase flows through two T-type rectangular microchannel mixers. Abdelall et al. [10] experimentally investigated the single-phase and two-phase pressure drops caused by abrupt flow area expansion and contraction in small circular channels. Lee et al. [11] verified the thermal behavior obtained from the classical correlations with the experimental data in the rectangular microchannels. Zhang et al. [12] reported a single-phase liquid cooling in the microchannel heat sink for electronic packages. Shen et al. [13] investigated the single phase convective heat transfer of de-ionized water in a compact heat sink consisting of 26 rectangular microchannels. Timothy and Brackbill [14] studied on the roughness elements affect internal flows in different ways. The channel used in this study is rectangular, with varying separation between walls that have machined roughness elements. Steinke and Kandlikar [15] considered the validity of friction factor theory based upon conventional sized passages for microchannel flows is still an active area of research. Hrnjak and Tu [16] focused on investigating fully developed liquid and vapor flow through rectangular microchannels. R134a liquid and vapor were used as the testing fluids. Xie et al. [17] studied the development of the information technology industry, the heat flux in integrated circuit chips cooled by air flowing in the microchannels.
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Nomenclature As Dh h knf kw kp khs L • mnf Nu Qnf qin Re Rth Tb T ht Tin,nf Tout,nf ΔTLMTD um
total heat transfer surface area of the heat sink, m 2 hydraulic diameter, m average heat transfer coefficient, W/(m 2 K) thermal conductivity of the nanofluids, W/(m K) thermal conductivity of the base fluid, W/(m K) thermal conductivity of the nanoparticles, W/(m K) thermal conductivity of the heat sink, W/(m K) the base thickness of the heat sink, m mass flow rate of the nanofluids, kg/s Nusselt number heat transfer rate by the nanofluids, W input heat flux from the electric heater, W/m 2 nanofluid Reynolds number thermal resistance, K/W the base temperature of the heat sink, °C heat sink temperature, °C inlet temperatures of the nanofluids, °C outlet temperatures of the nanofluids, °C logarithm mean temperature difference, °C nanofluid velocity, m/s
thermal conductivity models of nanofluids on the thermal performance in a trapezoidal microchannel. Duangtongsuk and Wongwises [29– 31] studied on the heat transfer and pressure drop characteristics of TiO2-water nanofluids in a double-tube counter flow heat exchanger. Effect of thermophysical property models on the predicting of the convective heat transfer coefficient for low concentration nanofluids was considered. As mentioned above, the numerous papers presented the study on heat transfer and pressure drop of nanofluids in the mini-and microchannel with wide variety of nanoparticles. However, only few work reported on the application of nanofluids with suspending TiO2 nanoparticles for the cooling electronic components. In addition, the main reasons for choosing TiO2 as the particles are summarized as follows (Duangtongsuk and Wongwises [30]): (1) it is cheap because it is produced in commercially available products, (2) it is generally regarded as a safe material for human beings and animals, and (3) it has excellent chemical and physical stability even without an additional stabilizer. Therefore, the paper focuses on the experimental study on the cooling characteristics in the mini-rectangular fin heat sinks by nanofluids with suspending TiO2 particles in the base fluid. Effects of channel height, heat flux, and coolant flow rate on the cooling performance are considered. The results obtained from the nanofluid cooling system are compared with those from the conventional liquid cooling system. 2. Experimental apparatus and method
Greek symbols ρnf density of the nanofluids ρp density of the nanoparticles ρw density of the base fluid ϕ volume fraction of the nanoparticles ρcp nf heat capacity of the nanofluids ρc heat capacity of the base fluid p w ρcp p heat capacity of the nanoparticles μnf viscosity of the nanofluids μw viscosity of the base fluid
The most frequently used coolants in the heat transfer devices study are air, water, and fluoro-chemicals. However, the heat transfer capability is limited by the working fluid transport properties. One of the methods for the heat transfer enhancement is the application of additives to the working fluids to change the fluid transport properties and flow features. Therefore, in order to further enhance thermal performance of heat transfer devices, the use of nanofluids is proposed. Koo and Kleinstreuer [18] simulated the steady laminar liquid nanofluid flow in microchannels. Chein and Huang [19] analyzed the silicon microchannel heat sink performance using nanofluids as coolants. The nanofluid was a mixture of pure water and nanoscale Cu particles with various volume fractions. Jang and Choi [20] numerically investigated the cooling performance of a microchannel heat sink with nanoparticle-fluid suspensions. Kang et al. [21] employed the nano-fluid as the working medium for a grooved circular heat pipe. Chein and Chuang [22] studied the microchannel heat sink performance using nanofluids as coolants. Experiments were performed to verify the theoretical predictions. Lee and Mudawar [23] explored the micro-channel cooling benefits of water-based nanofluids containing small concentrations of Al2O3. Nguyen et al. [24] experimentally investigated the behavior and heat transfer enhancement of a particular nanofluids flowing inside a closed system for cooling the electronic components. Chen [25] analyzed the forced convection heat transfer in microchannel heat sinks for electronic system cooling. Tsai and Chein [26] studied the microchannel heat sink performance using copper-water and carbon nanotube-water nanofluids as coolants. Kleinstreuer et al. [27] reviewed the microfluidics, a bio-MEMS application in terms of nanofluid flow in microchannels. Li and Kleinstreuer [28] considered effective
2.1. Test loop A schematic diagram of the experimental apparatus is shown in Fig. 1. The test loop consists of a set of ultrasonic system, cooling nanofluids loop and data acquisition system. The close-loop of nanofluids consists of a 10−3 m3 storage tank, pump, and the flow rate measurement system. After the temperatures of the nanofluids are cooled to achieve the desired level, the nanofluids are pumped out of the storage tank, passed through the mini-rectangular fin heat sink, and returned to the storage tank. The flow rates are varied during a set of experiments with the help of a dimmerstat connected to the pump. The flow rates of the cooling water are controlled by adjusting the valve and measured by collecting the fluid with the precise cylinder for a period of time during 15 min and the fluid mass is measured by an electronic weight scale. The base fluids are the de-ionized water. The nanofluids with suspending TiO2 nanoparticles in base fluids were used as working fluids. The average spherical particle size of TiO2 nanoparticles is about 21 nm. The nanofluids are prepared by ultrasonic method with the constant nanoparticle concentration of 0.4% by volume without using surfactant. The pressure drops across the test section are measured by the differential pressure transducer with an accuracy of 0.02% of full scale. There are two pressure taps on the cover plate upstream and downstream of the test section. The type-T copper-constantan thermocouples with an accuracy of 0.1% of full scale are employed to measure nanofluid temperatures. The inlet and outlet temperatures of nanofluids are measured by thermocouples with 1 mm diameter probes extending into the cover plate in which the nanofluids flow, respectively. All the thermocouple probes are precalibrated by dry-box temperature calibrator with 0.01 °C precision. 2.2. Test section The schematic diagram of the test section is shown in Fig. 2. The test sections are fabricated from the copper by a wire electrical discharge machine (WEDM) with the widths∗ length of 60 ∗ 110 mm. Three heat sinks with fin height of 1.00, 1.50, and 2.00 mm are tested with the details as shown in Table 1. An AC power supply is used as the source of power for the plate type heaters. Rear face of the test section is insulated with a 2 mm thick heat-resistant Mica standard sheet, 10 mm thick Acrylic standard sheet, and then 5 mm thick Aeroflex standard sheet,
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27
Fig. 1. Schematic diagram of experimental apparatus.
respectively. While front face of the test section is covered with 10 mm thick Acrylic standard sheet, and the 5 mm thick Aeroflex standard sheet, respectively.
2.3. Experimental method Nanofluids were used as coolant. The nanofluids were pumped into the mini-rectangular fin heat sink passed through the mini-rectangular fin heat sink, and returned to the storage tank. The inlet temperature of nanofluids before entering the cooling section was kept nearly constant with the refrigerant system. Experiments were conducted with various cooling nanofluid flow rates, channel height of fin, inlet temperature of the nanofluids, and heat flux. The supplied heat into the test section was adjusted to achieve the desired level by using the dimmerstat. The supplied power was calculated using the measured voltage and current supplied to the heaters. The supplied voltage and current to the heaters were measured by the digital clamp meter. Data collection was carried out using a data acquisition system (DataTaker, DT800).
The energy balance between the supplied heat by the heater plate and the absorbed heat by the nanofluids is ±10%. The average value of heat transfer rate is obtained by the supplied heat by heater plate and the absorbed heat by the nanofluids. The uncertainty and accuracy of the measurement are given in Table 2. The uncertainties of measurements data and the relevant parameters obtained from the data reduction process are calculated. The maximum uncertainties of the relevant parameters in the data calculation are based on Coleman and Steel method [32]. The maximum uncertainties of relevant parameters are ±5% for Reynolds number, ±10% for heat transfer rate and ±7.5% for the Nusselt number. 3. Data reduction The density of the nanofluids is calculated from Pak and Cho [33] using the following equation: ρnf ¼ ϕρP þ ð1−ϕÞρw
ð1Þ
Fluid inlet
Fluid outlet
Cover plate Gasket Bottom Electric heater Minichannel heat sink
Gasket
Fig. 2. Schematic diagram of the test section of the minichannel heat sinks.
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The heat transfer coefficient of the nanofluids is calculated from the following equation:
Table 1 Dimensions of the mini-channel heat sinks. Heat sinks
Fins size (mm)
Fin height (mm)
Heat sink size (mm)
Base thickness (mm)
Number of fins
Heat transfer surface area (m2)
Type 1 Type 2 Type 3
0.4 ∗ 1.8 0.4 ∗ 1.8 0.4 ∗ 1.8
1.0 1.5 2.0
60 ∗ 110 60 ∗ 110 60 ∗ 110
2.0 2.0 2.0
2500 2500 2500
0.0158 0.0513 0.0268
nf
¼ ϕ ρcp þ ð1−ϕÞ ρcp p
ð2Þ
w
is the heat capacity of the nanofluids, ρcp is the heat where ρcp nf w capacity of the base fluid and ρcp is the heat capacity of the p
nanoparticles. Drew and Passman [35] suggested the well-known equation for calculating the viscosity, which is applicable to spherical particles in volume fractions of less than 5.0 vol.% and is defined as follows: μ nf ¼ ð1 þ 2:5ϕÞμ w
ð3Þ
where μnf is the viscosity of the nanofluids and μw is the viscosity of the base fluid. The thermal conductivity of the nanofluids is calculated from Yu and Choi [36] using the following equation:
knf
3 2 kp þ 2kw −2ϕ kw −kp 5kw ¼4 kp þ 2kw þ ϕ kw −kp
ð4Þ
ρum Dh μ nf
ð5Þ
where Re is the nanofluid Reynolds number, um is the nanofluid velocity, and Dh is the hydraulic diameter. The heat transfer rate into the nanofluids is calculated from: Q nf ¼
• mnf
cpnf ðT in −T out Þnf
ð6Þ
where Qnf is the heat transfer rate by the nanofluids, m•nf is the mass flow rate of the nanofluids and Tin,nf and Tout,nf are the inlet and outlet temperatures of the nanofluids. The average heat transfer rate is calculated from: Q ave ¼
ð8Þ
ΔT LMTD
T b −T in;nf − T b −T out;nf : ¼ ðT −T Þ ln T b−T in;nf ð b out; nf Þ
ð9Þ
The average base temperature of heat sink can be calculated from: q L T b ¼ T ht − in khs
ð10Þ
where Tb is the base temperature of the heat sink, T ht is the heat sink temperature, qin is the input heat flux from the electric heater, L is the base thickness of the heat sink and khs is the thermal conductivity of the heat sink. The Nusselt number is calculated from the following equation: Nu ¼
hD h knf
ð11Þ
where Nu is the Nusselt number and Dh is the hydraulic diameter of the channel. 4. Results and discussion
where knf is the thermal conductivity of the nanofluids, kw is the thermal conductivity of the base fluid and kp is the thermal conductivity of the nanoparticles. The nanofluid Reynolds number is defined as: Re ¼
Q ave As ðΔT LMTD Þ
is the average heat transfer coefficient, A is the total heat transfer where h s surface area of the heat sink and ΔTLMTD is the logarithm mean temperature difference as follows:
where ρnf is the density of the nanofluids, ρp is the density of the nanoparticles, ρw is the density of the base fluid and ϕ is the volume fraction of the nanoparticles. The specific heat is calculated from Xuan and Roetzel [34] as follows: ρcp
h ¼
Q elec þ Q nf : 2
ð7Þ
Table 2 Accuracy and uncertainty of measurements. Instruments
Accuracy
Uncertainty
Voltage supplied by power source, voltage Current supplied by power source, ampere Thermocouple type T, data logger, °C Differential pressure transducer
0.2% 0.2% 0.1% 0.02%
±0.5 ±0.5 ±0.1 ±0.02
Initially experiments are performed with the de-ionized water which forms the basis for comparison of the results with nanofluids. Constant particle volume fraction of 0.4% is used in this experiment. The Reynolds numbers based on the hydraulic diameter of the channel are tested in the range between 80 and 200. Fig. 3 shows the variation of the outlet nanofluid temperature with nanofluid Reynolds number for different channel heights. It is found that when the inlet nanofluid temperature and heat flux are kept constant, higher heat transfer rate as nanofluid mass flow rate increase. However, the increase of heat transfer rate is less than that of the nanofluid mass flow rate. Therefore, the outlet nanofluid temperature tends to decrease as nanofluid mass flow rate 50 Nanofluids, Tin = 22 oC
Average outlet temperature (oC)
28
45 40 35 30 25 Channel height = 1.0 mm Channel height = 1.5 mm Channel height = 2.0 mm
20 15 10 60
80
100
120
140
160
180
200
220
Reynolds Number Fig. 3. Variation of average outlet temperature of the nanofluids for different channel height.
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350
60 qin = 31.70 kW/m qin = 33.11 kW/m2 qin = 34.51 kW/m2
Channel height = 2.0 mm Nanofluids, Tin = 24oC
55
Water, channel height = 1.0 mm Water, channel height = 1.5 mm Water, channel height = 2.0 mm Nanofluids, channel height = 1.0 mm Nanofluids, channel height = 1.5 mm Nanofluids, channel height = 2.0 mm
2
Average heat transfer rate (W)
Average surface temperature (oC)
29
50
45
40
35
300 250
Tin = 20 oC
200 150 100 50 0
30 70
80
90
100
110
120
130
140
4.5
150
5.0
5.5
6.0
6.5
7.0
7.5
8.0
Mass flow rate (g/s)
Reynolds number Fig. 4. Variation of average surface temperature of the heat sink for different heat flux.
Fig. 6. Comparison of average heat transfer rate between the base fluid and the nanofluids.
increases. Higher surface area and surface roughness result in increase heat transfer rate, therefore, the outlet nanofluid temperatures of channel with of w=2 mm are higher than those of w=1.5, 1.0 mm, respectively. Fig. 4 shows the variation of the average plate temperature with nanofluid Reynolds number for different heat flux. It can be seen that the heat transfer rate depends on the cooling capacity rate of nanofluids. Therefore, the average plate temperature decreases as nanofluid Reynolds number increases. In addition, for a given nanofluids flow rate, average plate temperatures at higher heat flux are higher than those from lower ones. Fig. 5 shows the variation of the heat transfer rate with nanofluid mass flow rate. As expected, the heat transfer rate is directly proportional to the nanofluid mass flow rate. In addition, it can be noted that the fin height strongly affects on the heat transfer rate. Fig. 6 shows the comparison of the average heat transfer rate obtained from the nanofluids as the coolant and from the de-ionized water as coolant. As expected, the average heat transfer rate increases with increasing nanofluid mass flow rate. Effect of channel height on the enhancement of average heat transfer is shown in Fig. 5. Due to higher heat transfer area and higher surface roughness, the average heat transfer rate of the heat sink with w = 2.0 mm is higher than those with w = 1.5, 1.0 mm. For the heat sink with w = 1.0 mm and mass flow rate of
nanofluids =4.5 g/s, the average heat transfer rate for the nanofluids as coolant is significantly 11.0% higher than that for the de-ionized water as coolant. Similar trend in the average heat transfer rate is obtained for the different nanofluid mass flow rates as shown in Fig. 6. In addition, for the heat sink with w=1.5 and 2.0 mm, appreciable increase more than 27.0, and 42.3% in average heat transfer rate for the nanofluids compared to the de-ionized water are obtained, respectively. Fig. 7 shows the variation of average Nusselt number with nanofluid Reynolds number for the different channel height. It can be seen that with the increasing nanofluid Reynolds number, the average Nusselt number tends to increase. The thermal conductivity of nanofluids depends on the sizes and species of the nanoparticles, the concentration of nanoparticles in the based fluid and the combination of the nanoparticles and the based fluid properties. However, at present, there are various correlations that are proposed to predict the thermal conductivity and other properties of the nanofluids. Due to particle migration inside the heat sink, non-uniformity of the nanoparticle concentration has significant effect on the thermal conductivity and viscosity of the fluid. Therefore, in the present study, the average heat transfer characteristics are presented. It can be seen that the average Nusselt numbers of nanofluids are greater than of their based fluid (de-ionized water).
350 300
Nanofluids, Tin = 22 oC Tin = 20oC
2.5
Average Nusselt number
Average heat transfer rate (W)
3.0 Channel height = 1.0 mm Channel height = 1.5 mm Channel height = 2.0 mm
250 200 150 100
Water, channel height = 1.0 mm Water, channel height = 1.5 mm Water, channel height = 2.0 mm Nanofluids, channel height = 1.0 mm Nanofluids, channel height = 1.5 mm Nanofluids, channel height = 2.0 mm
2.0
1.5
1.0
0.5
50 0
0.0 4.5
5.0
5.5
6.0
6.5
7.0
7.5
8.0
Mass flow rate (g/s) Fig. 5. Variation of average heat transfer rate for different channel height.
40
60
80
100
120
140
160
180
200
220
Reynolds number Fig. 7. Comparison of average Nusselt number between the base fluid and the nanofluids.
30
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0.30
4 Tin = 20 oC
Water, channel height = 1.0 mm Water, channel height = 1.5 mm Water, channel height = 2.0 mm Nanofluids, channel height = 1.0 mm Nanofluids, channel height = 1.5 mm Nanofluids, channel height = 2.0 mm
Water, channel height = 1.0 mm Tin = 20 oC Water, channel height = 1.5 mm Water, channel height = 2.0 mm Nanofluids, channel height = 1.0 mm Nanofluids, channel height = 1.5 mm Nanofluids, channel height = 2.0 mm
3
Pressure drop (kPa)
Thermal resistance (oC/W)
0.25
0.20
0.15
0.10
0.05
2
1
0
0.00 60
80
100
120
140
160
180
200
Reynolds number
-1 60
80
100
120
140
160
180
200
Reynolds number
Fig. 8. Comparison of thermal resistance between the base fluid and the nanofluids.
Fig. 10. Comparison pressure drop between the base fluid and the nanofluids.
In this study, the overall heat sink performance can be shown in the thermal resistance form. The heat sink thermal resistance is defined as Rth ¼
1 h As
¼
1
ð12Þ
Q ave ðΔT LMTD Þ
where Rth is the thermal resistance. Using this definition, thermal resistance of heat sink is function of heat transfer coefficient. It is found that the heat sink thermal resistance decreases with increasing heat transfer rate (increasing Reynolds number). Based on the results shown in Fig. 8, it is seen that the heat sink with nanofluids as coolant give the thermal resistance lower than as de-ionized water for the same heat sink. Figs. 9–10 show the variation of pressure drop with nanofluid Reynolds number for different heat flux and for different heat sinks, respectively. The test section as shown in Fig. 2, the pressure drop across the test section is produced from: inlet port to inlet plenum due to flow sudden expansion, inlet plenum to channel inlets due to flow contraction, channel inlet to exit due to wall friction, channel exits to outlet plenum due to flow sudden expansion, and outlet plenum to outlet port due to flow contraction. It can be seen that the heat flux has an insignificant effect on the pressure drop of the nanofluids. This is due to the increase in the heat flux having a slight effect on the viscosity of the nanofluids,
which leads to a tiny change in the measured pressure drop of the nanofluids. As compare to the de-ionized water, the pressure drop in the nanofluids is very close to that in the de-ionized water for the given conditions.
5. Conclusions In the present study, the heat transfer characteristics of nanofluids cooling in the mini-rectangular fin heat sink are studied. The TiO2 nanofluids are used as the working fluid with de-ionized water as the base fluid. The results obtained from the nanofluids cooling in minirectangular fin heat sink are compared with the other from the deionized water cooling method. Effects of inlet temperature of coolant, coolant Reynolds number, and heat flux on the heat transfer characteristics of mini-rectangular fin heat sink are considered. It is found that the nanofluids have significant effect on the enhancement of heat transfer of the heat sinks. In addition, the pressure drops of the nanofluids are approximately the same as those of de-ionized water in the given conditions. The results of this study are of technological importance for the efficient design of cooling systems of electronic devices to enhance cooling performance.
4.0 3.5
Pressure drop (kPa)
3.0
qin = 14.80 kW/m
2
qin = 15.76 kW/m
2
qin = 16.75 kW/m
2
Acknowledgments
Channel high = 1.0 mm o
Nanofluids, Tin = 20 C
The authors would like to express their appreciation to the Excellent Center for Sustainable Engineering (ECSE) of the Srinakharinwirot University (SWU) for providing financial support for this study.
2.5
References
2.0 1.5 1.0 0.5 0.0 100
120
140
160
180
Reynolds number Fig. 9. Variation of pressure drop for different heat flux.
200
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[21] S.W. Kang, W.C. Wei, S.H. Tsai, S.Y. Yang, Experimental investigation of silver nano-fluid on heat pipe thermal performance, Applied Thermal Engineering 26 (2006) 2377–2382. [22] R. Chein, J. Chuang, Experimental microchannel heat sink performance studies using nanofluids, International Journal of Thermal Sciences 46 (2007) 57–66. [23] J. Lee, I. Mudawar, Assessment of the effectiveness of nanofluids for single-phase and two-phase heat transfer in micro-channels, International Journal of Heat and Mass Transfer 50 (2007) 452–463. [24] C.T. Nguyen, G. Roy, C. Gauthier, N. Galanis, Heat transfer enhancement using Al2O3-water nanofluids for an electronic liquid cooling system, Applied Thermal Engineering 27 (2007) 1501–1506. [25] C.H. Chen, Forced convection heat transfer in microchannel heat sinks, International Journal of Heat and Mass Transfer 50 (2007) 2182–2189. [26] T.H. Tsai, R. Chein, Performance analysis of nanofluid-cooled microchannel heat sinks, International Journal of Heat and Fluid Flow 28 (2007) 1013–1026. [27] C. Kleinstreuer, J. Li, J. Koo, Microfluidics of nano-drug delivery, International Journal of Heat and Mass Transfer 51 (2008) 5590–5597. [28] J. Li, C. Kleinstreuer, Thermal performance of nanofluid flow in microchannels, International Journal of Heat and Fluid Flow 29 (2008) 1221–1232. [29] W. Duangtongsuk, S. Wongwises, Effect of thermophysical properties models on the predicting of the convective heat transfer coefficient for low concentration nanofluid, International Communications in Heat and Mass Transfer 35 (2008) 1320–1326. [30] W. Duangtongsuk, S. Wongwises, Heat transfer enhancement and pressure drop characteristics of TiO2-water nanofluid in a double-tube counter flow heat exchanger, International Journal of Heat and Mass Transfer 52 (2009) 2059–2067. [31] W. Duangtongsuk, S. Wongwises, An experimental study on the heat transfer performance and pressure drop of TiO2-water nanofluids flowing under a turbulent flow regime, International Journal of Heat and Mass Transfer 53 (2010) 334–344. [32] H.W. Coleman, W.G. Steele, Experimental and Uncertainty Analysis for Engineers, John Wiley & Sons, New York, 1989. [33] B.C. Pak, Y.I. Cho, Hydrodynamic and heat transfer study of dispersed fluids with submicron metallic oxide particles, Experimental Heat Transfer 11 (1998) 151–170. [34] Y. Xuan, Roetzel, Conceptions of heat transfer correlation of nanofluids, International Journal of Heat and Mass Transfer 43 (2000) 3701–3707. [35] D.A. Drew, S.L. Passman, Theory of Multi Component Fluids, Springer, Berlin, 1999. [36] W. Yu, S.U.S. Choi, The role of interfacial layers in the enhanced thermal conductivity of nanofluids: a renovated Maxvell model, Journal of Nanoparticle Research 5 (2003) 167–171.
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International Communications in Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ichmt
Heat transfer of nanofluids in the mini-rectangular fin heat sinks☆ Paisarn Naphon ⁎, Lursukd Nakharintr Thermo-Fluid and Heat Transfer Enhancement Lab. (TFHT), Department of Mechanical Engineering, Faculty of Engineering, Srinakharinwirot University, 63 Rangsit-Nakhornnayok Rd., Ongkharak, Nakhorn-Nayok, 26120, Thailand
a r t i c l e
i n f o
Available online 22 October 2012 Keywords: Nanofluids Minichannel heat sink Convective heat transfer
a b s t r a c t In the present study, the heat transfer characteristics of nanofluids cooling in the mini-rectangular fin heat sink are studied. The heat sinks with three different channel heights are fabricated from the aluminum by the wire electrical discharge machine with the length, width and base thickness of 110, 60, and 2 mm, respectively. The nanofluids are the mixture of de-ionized water and nanoscale TiO2 particles. The results obtained from the nanofluids cooling in mini-rectangular fin heat sink are compared with those from the de-ionized water cooling method. Effects of the inlet temperature of nanofluids, nanofluid Reynolds number, and heat flux on the heat transfer characteristics of mini-rectangular fin heat sink are considered. It is found that average heat transfer rates for nanofluids as coolant are higher than those for the de-ionized water as coolant. The results of this study are of technological importance for the efficient design of cooling systems of electronic devices to enhance cooling performance. © 2012 Elsevier Ltd. All rights reserved.
1. Introduction In order to ensure reliable operation, the personal computer or electronic devices must be operated in specific temperature ranges. The exceeding maximum allowable temperature is a serious problem for these devices. Therefore, in order to keep the operated temperature constant, the cooling system must dissipate the heat equal or higher than the generated heat from these devices. Due to the conventional air or liquid cooling systems limitation and the small physical size of the electronic devices, the development of the miniaturized technology, mini and micro-components has been increased especially in the electronic devices. The development of the miniaturized technology, mini and micro-components has been introduced as one of the heat transfer enhancement techniques. The heat transfer and pressure drop in the mini and micro-channel have been widely studied by researchers. Peng and Peterson [1] experimentally investigated on the single-phase forced convective heat transfer and the flow characteristics of water in microchannel. Adams et al. [2] investigated the single-phase turbulent flow of water in circular microchannels. A generalized correlation for the Nusselt number for single-phase turbulent flow in circular microchannels was proposed. Gillot et al. [3] applied the single-phase and two-phase micro heat sinks for cooling of power components. Qu and Mudawar [4] experimentally and numerically investigated the pressure drop and heat transfer characteristics of a single-phase micro-channel heat sink. The results obtained from the numerical prediction were
☆ Communicated by W.J. Minkowycz. ⁎ Corresponding author. E-mail address: [email protected] (P. Naphon). 0735-1933/$ – see front matter © 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.icheatmasstransfer.2012.10.012
reasonable agreement with the measured data. Hetsroni et al. [5] applied the heat sink for cooling the electronic devices. Agostini et al. [6] experimentally studied on the friction factor and heat transfer coefficient for a vertical liquid up-flow of R134a. Owhaib and Palm [7] studied the heat transfer characteristics of single-phase forced convection of R134a through single circular micro-channels. The results were compared with those from the proposed correlations in macroscale and microscale channels. Agostini et al. [8] investigated on the friction factor and heat transfer coefficient of R134a flowing through the rectangular mini-channels. Yue et al. [9] considered the pressure drop characteristics of single and two-phase flows through two T-type rectangular microchannel mixers. Abdelall et al. [10] experimentally investigated the single-phase and two-phase pressure drops caused by abrupt flow area expansion and contraction in small circular channels. Lee et al. [11] verified the thermal behavior obtained from the classical correlations with the experimental data in the rectangular microchannels. Zhang et al. [12] reported a single-phase liquid cooling in the microchannel heat sink for electronic packages. Shen et al. [13] investigated the single phase convective heat transfer of de-ionized water in a compact heat sink consisting of 26 rectangular microchannels. Timothy and Brackbill [14] studied on the roughness elements affect internal flows in different ways. The channel used in this study is rectangular, with varying separation between walls that have machined roughness elements. Steinke and Kandlikar [15] considered the validity of friction factor theory based upon conventional sized passages for microchannel flows is still an active area of research. Hrnjak and Tu [16] focused on investigating fully developed liquid and vapor flow through rectangular microchannels. R134a liquid and vapor were used as the testing fluids. Xie et al. [17] studied the development of the information technology industry, the heat flux in integrated circuit chips cooled by air flowing in the microchannels.
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Nomenclature As Dh h knf kw kp khs L • mnf Nu Qnf qin Re Rth Tb T ht Tin,nf Tout,nf ΔTLMTD um
total heat transfer surface area of the heat sink, m 2 hydraulic diameter, m average heat transfer coefficient, W/(m 2 K) thermal conductivity of the nanofluids, W/(m K) thermal conductivity of the base fluid, W/(m K) thermal conductivity of the nanoparticles, W/(m K) thermal conductivity of the heat sink, W/(m K) the base thickness of the heat sink, m mass flow rate of the nanofluids, kg/s Nusselt number heat transfer rate by the nanofluids, W input heat flux from the electric heater, W/m 2 nanofluid Reynolds number thermal resistance, K/W the base temperature of the heat sink, °C heat sink temperature, °C inlet temperatures of the nanofluids, °C outlet temperatures of the nanofluids, °C logarithm mean temperature difference, °C nanofluid velocity, m/s
thermal conductivity models of nanofluids on the thermal performance in a trapezoidal microchannel. Duangtongsuk and Wongwises [29– 31] studied on the heat transfer and pressure drop characteristics of TiO2-water nanofluids in a double-tube counter flow heat exchanger. Effect of thermophysical property models on the predicting of the convective heat transfer coefficient for low concentration nanofluids was considered. As mentioned above, the numerous papers presented the study on heat transfer and pressure drop of nanofluids in the mini-and microchannel with wide variety of nanoparticles. However, only few work reported on the application of nanofluids with suspending TiO2 nanoparticles for the cooling electronic components. In addition, the main reasons for choosing TiO2 as the particles are summarized as follows (Duangtongsuk and Wongwises [30]): (1) it is cheap because it is produced in commercially available products, (2) it is generally regarded as a safe material for human beings and animals, and (3) it has excellent chemical and physical stability even without an additional stabilizer. Therefore, the paper focuses on the experimental study on the cooling characteristics in the mini-rectangular fin heat sinks by nanofluids with suspending TiO2 particles in the base fluid. Effects of channel height, heat flux, and coolant flow rate on the cooling performance are considered. The results obtained from the nanofluid cooling system are compared with those from the conventional liquid cooling system. 2. Experimental apparatus and method
Greek symbols ρnf density of the nanofluids ρp density of the nanoparticles ρw density of the base fluid ϕ volume fraction of the nanoparticles ρcp nf heat capacity of the nanofluids ρc heat capacity of the base fluid p w ρcp p heat capacity of the nanoparticles μnf viscosity of the nanofluids μw viscosity of the base fluid
The most frequently used coolants in the heat transfer devices study are air, water, and fluoro-chemicals. However, the heat transfer capability is limited by the working fluid transport properties. One of the methods for the heat transfer enhancement is the application of additives to the working fluids to change the fluid transport properties and flow features. Therefore, in order to further enhance thermal performance of heat transfer devices, the use of nanofluids is proposed. Koo and Kleinstreuer [18] simulated the steady laminar liquid nanofluid flow in microchannels. Chein and Huang [19] analyzed the silicon microchannel heat sink performance using nanofluids as coolants. The nanofluid was a mixture of pure water and nanoscale Cu particles with various volume fractions. Jang and Choi [20] numerically investigated the cooling performance of a microchannel heat sink with nanoparticle-fluid suspensions. Kang et al. [21] employed the nano-fluid as the working medium for a grooved circular heat pipe. Chein and Chuang [22] studied the microchannel heat sink performance using nanofluids as coolants. Experiments were performed to verify the theoretical predictions. Lee and Mudawar [23] explored the micro-channel cooling benefits of water-based nanofluids containing small concentrations of Al2O3. Nguyen et al. [24] experimentally investigated the behavior and heat transfer enhancement of a particular nanofluids flowing inside a closed system for cooling the electronic components. Chen [25] analyzed the forced convection heat transfer in microchannel heat sinks for electronic system cooling. Tsai and Chein [26] studied the microchannel heat sink performance using copper-water and carbon nanotube-water nanofluids as coolants. Kleinstreuer et al. [27] reviewed the microfluidics, a bio-MEMS application in terms of nanofluid flow in microchannels. Li and Kleinstreuer [28] considered effective
2.1. Test loop A schematic diagram of the experimental apparatus is shown in Fig. 1. The test loop consists of a set of ultrasonic system, cooling nanofluids loop and data acquisition system. The close-loop of nanofluids consists of a 10−3 m3 storage tank, pump, and the flow rate measurement system. After the temperatures of the nanofluids are cooled to achieve the desired level, the nanofluids are pumped out of the storage tank, passed through the mini-rectangular fin heat sink, and returned to the storage tank. The flow rates are varied during a set of experiments with the help of a dimmerstat connected to the pump. The flow rates of the cooling water are controlled by adjusting the valve and measured by collecting the fluid with the precise cylinder for a period of time during 15 min and the fluid mass is measured by an electronic weight scale. The base fluids are the de-ionized water. The nanofluids with suspending TiO2 nanoparticles in base fluids were used as working fluids. The average spherical particle size of TiO2 nanoparticles is about 21 nm. The nanofluids are prepared by ultrasonic method with the constant nanoparticle concentration of 0.4% by volume without using surfactant. The pressure drops across the test section are measured by the differential pressure transducer with an accuracy of 0.02% of full scale. There are two pressure taps on the cover plate upstream and downstream of the test section. The type-T copper-constantan thermocouples with an accuracy of 0.1% of full scale are employed to measure nanofluid temperatures. The inlet and outlet temperatures of nanofluids are measured by thermocouples with 1 mm diameter probes extending into the cover plate in which the nanofluids flow, respectively. All the thermocouple probes are precalibrated by dry-box temperature calibrator with 0.01 °C precision. 2.2. Test section The schematic diagram of the test section is shown in Fig. 2. The test sections are fabricated from the copper by a wire electrical discharge machine (WEDM) with the widths∗ length of 60 ∗ 110 mm. Three heat sinks with fin height of 1.00, 1.50, and 2.00 mm are tested with the details as shown in Table 1. An AC power supply is used as the source of power for the plate type heaters. Rear face of the test section is insulated with a 2 mm thick heat-resistant Mica standard sheet, 10 mm thick Acrylic standard sheet, and then 5 mm thick Aeroflex standard sheet,
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27
Fig. 1. Schematic diagram of experimental apparatus.
respectively. While front face of the test section is covered with 10 mm thick Acrylic standard sheet, and the 5 mm thick Aeroflex standard sheet, respectively.
2.3. Experimental method Nanofluids were used as coolant. The nanofluids were pumped into the mini-rectangular fin heat sink passed through the mini-rectangular fin heat sink, and returned to the storage tank. The inlet temperature of nanofluids before entering the cooling section was kept nearly constant with the refrigerant system. Experiments were conducted with various cooling nanofluid flow rates, channel height of fin, inlet temperature of the nanofluids, and heat flux. The supplied heat into the test section was adjusted to achieve the desired level by using the dimmerstat. The supplied power was calculated using the measured voltage and current supplied to the heaters. The supplied voltage and current to the heaters were measured by the digital clamp meter. Data collection was carried out using a data acquisition system (DataTaker, DT800).
The energy balance between the supplied heat by the heater plate and the absorbed heat by the nanofluids is ±10%. The average value of heat transfer rate is obtained by the supplied heat by heater plate and the absorbed heat by the nanofluids. The uncertainty and accuracy of the measurement are given in Table 2. The uncertainties of measurements data and the relevant parameters obtained from the data reduction process are calculated. The maximum uncertainties of the relevant parameters in the data calculation are based on Coleman and Steel method [32]. The maximum uncertainties of relevant parameters are ±5% for Reynolds number, ±10% for heat transfer rate and ±7.5% for the Nusselt number. 3. Data reduction The density of the nanofluids is calculated from Pak and Cho [33] using the following equation: ρnf ¼ ϕρP þ ð1−ϕÞρw
ð1Þ
Fluid inlet
Fluid outlet
Cover plate Gasket Bottom Electric heater Minichannel heat sink
Gasket
Fig. 2. Schematic diagram of the test section of the minichannel heat sinks.
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The heat transfer coefficient of the nanofluids is calculated from the following equation:
Table 1 Dimensions of the mini-channel heat sinks. Heat sinks
Fins size (mm)
Fin height (mm)
Heat sink size (mm)
Base thickness (mm)
Number of fins
Heat transfer surface area (m2)
Type 1 Type 2 Type 3
0.4 ∗ 1.8 0.4 ∗ 1.8 0.4 ∗ 1.8
1.0 1.5 2.0
60 ∗ 110 60 ∗ 110 60 ∗ 110
2.0 2.0 2.0
2500 2500 2500
0.0158 0.0513 0.0268
nf
¼ ϕ ρcp þ ð1−ϕÞ ρcp p
ð2Þ
w
is the heat capacity of the nanofluids, ρcp is the heat where ρcp nf w capacity of the base fluid and ρcp is the heat capacity of the p
nanoparticles. Drew and Passman [35] suggested the well-known equation for calculating the viscosity, which is applicable to spherical particles in volume fractions of less than 5.0 vol.% and is defined as follows: μ nf ¼ ð1 þ 2:5ϕÞμ w
ð3Þ
where μnf is the viscosity of the nanofluids and μw is the viscosity of the base fluid. The thermal conductivity of the nanofluids is calculated from Yu and Choi [36] using the following equation:
knf
3 2 kp þ 2kw −2ϕ kw −kp 5kw ¼4 kp þ 2kw þ ϕ kw −kp
ð4Þ
ρum Dh μ nf
ð5Þ
where Re is the nanofluid Reynolds number, um is the nanofluid velocity, and Dh is the hydraulic diameter. The heat transfer rate into the nanofluids is calculated from: Q nf ¼
• mnf
cpnf ðT in −T out Þnf
ð6Þ
where Qnf is the heat transfer rate by the nanofluids, m•nf is the mass flow rate of the nanofluids and Tin,nf and Tout,nf are the inlet and outlet temperatures of the nanofluids. The average heat transfer rate is calculated from: Q ave ¼
ð8Þ
ΔT LMTD
T b −T in;nf − T b −T out;nf : ¼ ðT −T Þ ln T b−T in;nf ð b out; nf Þ
ð9Þ
The average base temperature of heat sink can be calculated from: q L T b ¼ T ht − in khs
ð10Þ
where Tb is the base temperature of the heat sink, T ht is the heat sink temperature, qin is the input heat flux from the electric heater, L is the base thickness of the heat sink and khs is the thermal conductivity of the heat sink. The Nusselt number is calculated from the following equation: Nu ¼
hD h knf
ð11Þ
where Nu is the Nusselt number and Dh is the hydraulic diameter of the channel. 4. Results and discussion
where knf is the thermal conductivity of the nanofluids, kw is the thermal conductivity of the base fluid and kp is the thermal conductivity of the nanoparticles. The nanofluid Reynolds number is defined as: Re ¼
Q ave As ðΔT LMTD Þ
is the average heat transfer coefficient, A is the total heat transfer where h s surface area of the heat sink and ΔTLMTD is the logarithm mean temperature difference as follows:
where ρnf is the density of the nanofluids, ρp is the density of the nanoparticles, ρw is the density of the base fluid and ϕ is the volume fraction of the nanoparticles. The specific heat is calculated from Xuan and Roetzel [34] as follows: ρcp
h ¼
Q elec þ Q nf : 2
ð7Þ
Table 2 Accuracy and uncertainty of measurements. Instruments
Accuracy
Uncertainty
Voltage supplied by power source, voltage Current supplied by power source, ampere Thermocouple type T, data logger, °C Differential pressure transducer
0.2% 0.2% 0.1% 0.02%
±0.5 ±0.5 ±0.1 ±0.02
Initially experiments are performed with the de-ionized water which forms the basis for comparison of the results with nanofluids. Constant particle volume fraction of 0.4% is used in this experiment. The Reynolds numbers based on the hydraulic diameter of the channel are tested in the range between 80 and 200. Fig. 3 shows the variation of the outlet nanofluid temperature with nanofluid Reynolds number for different channel heights. It is found that when the inlet nanofluid temperature and heat flux are kept constant, higher heat transfer rate as nanofluid mass flow rate increase. However, the increase of heat transfer rate is less than that of the nanofluid mass flow rate. Therefore, the outlet nanofluid temperature tends to decrease as nanofluid mass flow rate 50 Nanofluids, Tin = 22 oC
Average outlet temperature (oC)
28
45 40 35 30 25 Channel height = 1.0 mm Channel height = 1.5 mm Channel height = 2.0 mm
20 15 10 60
80
100
120
140
160
180
200
220
Reynolds Number Fig. 3. Variation of average outlet temperature of the nanofluids for different channel height.
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350
60 qin = 31.70 kW/m qin = 33.11 kW/m2 qin = 34.51 kW/m2
Channel height = 2.0 mm Nanofluids, Tin = 24oC
55
Water, channel height = 1.0 mm Water, channel height = 1.5 mm Water, channel height = 2.0 mm Nanofluids, channel height = 1.0 mm Nanofluids, channel height = 1.5 mm Nanofluids, channel height = 2.0 mm
2
Average heat transfer rate (W)
Average surface temperature (oC)
29
50
45
40
35
300 250
Tin = 20 oC
200 150 100 50 0
30 70
80
90
100
110
120
130
140
4.5
150
5.0
5.5
6.0
6.5
7.0
7.5
8.0
Mass flow rate (g/s)
Reynolds number Fig. 4. Variation of average surface temperature of the heat sink for different heat flux.
Fig. 6. Comparison of average heat transfer rate between the base fluid and the nanofluids.
increases. Higher surface area and surface roughness result in increase heat transfer rate, therefore, the outlet nanofluid temperatures of channel with of w=2 mm are higher than those of w=1.5, 1.0 mm, respectively. Fig. 4 shows the variation of the average plate temperature with nanofluid Reynolds number for different heat flux. It can be seen that the heat transfer rate depends on the cooling capacity rate of nanofluids. Therefore, the average plate temperature decreases as nanofluid Reynolds number increases. In addition, for a given nanofluids flow rate, average plate temperatures at higher heat flux are higher than those from lower ones. Fig. 5 shows the variation of the heat transfer rate with nanofluid mass flow rate. As expected, the heat transfer rate is directly proportional to the nanofluid mass flow rate. In addition, it can be noted that the fin height strongly affects on the heat transfer rate. Fig. 6 shows the comparison of the average heat transfer rate obtained from the nanofluids as the coolant and from the de-ionized water as coolant. As expected, the average heat transfer rate increases with increasing nanofluid mass flow rate. Effect of channel height on the enhancement of average heat transfer is shown in Fig. 5. Due to higher heat transfer area and higher surface roughness, the average heat transfer rate of the heat sink with w = 2.0 mm is higher than those with w = 1.5, 1.0 mm. For the heat sink with w = 1.0 mm and mass flow rate of
nanofluids =4.5 g/s, the average heat transfer rate for the nanofluids as coolant is significantly 11.0% higher than that for the de-ionized water as coolant. Similar trend in the average heat transfer rate is obtained for the different nanofluid mass flow rates as shown in Fig. 6. In addition, for the heat sink with w=1.5 and 2.0 mm, appreciable increase more than 27.0, and 42.3% in average heat transfer rate for the nanofluids compared to the de-ionized water are obtained, respectively. Fig. 7 shows the variation of average Nusselt number with nanofluid Reynolds number for the different channel height. It can be seen that with the increasing nanofluid Reynolds number, the average Nusselt number tends to increase. The thermal conductivity of nanofluids depends on the sizes and species of the nanoparticles, the concentration of nanoparticles in the based fluid and the combination of the nanoparticles and the based fluid properties. However, at present, there are various correlations that are proposed to predict the thermal conductivity and other properties of the nanofluids. Due to particle migration inside the heat sink, non-uniformity of the nanoparticle concentration has significant effect on the thermal conductivity and viscosity of the fluid. Therefore, in the present study, the average heat transfer characteristics are presented. It can be seen that the average Nusselt numbers of nanofluids are greater than of their based fluid (de-ionized water).
350 300
Nanofluids, Tin = 22 oC Tin = 20oC
2.5
Average Nusselt number
Average heat transfer rate (W)
3.0 Channel height = 1.0 mm Channel height = 1.5 mm Channel height = 2.0 mm
250 200 150 100
Water, channel height = 1.0 mm Water, channel height = 1.5 mm Water, channel height = 2.0 mm Nanofluids, channel height = 1.0 mm Nanofluids, channel height = 1.5 mm Nanofluids, channel height = 2.0 mm
2.0
1.5
1.0
0.5
50 0
0.0 4.5
5.0
5.5
6.0
6.5
7.0
7.5
8.0
Mass flow rate (g/s) Fig. 5. Variation of average heat transfer rate for different channel height.
40
60
80
100
120
140
160
180
200
220
Reynolds number Fig. 7. Comparison of average Nusselt number between the base fluid and the nanofluids.
30
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0.30
4 Tin = 20 oC
Water, channel height = 1.0 mm Water, channel height = 1.5 mm Water, channel height = 2.0 mm Nanofluids, channel height = 1.0 mm Nanofluids, channel height = 1.5 mm Nanofluids, channel height = 2.0 mm
Water, channel height = 1.0 mm Tin = 20 oC Water, channel height = 1.5 mm Water, channel height = 2.0 mm Nanofluids, channel height = 1.0 mm Nanofluids, channel height = 1.5 mm Nanofluids, channel height = 2.0 mm
3
Pressure drop (kPa)
Thermal resistance (oC/W)
0.25
0.20
0.15
0.10
0.05
2
1
0
0.00 60
80
100
120
140
160
180
200
Reynolds number
-1 60
80
100
120
140
160
180
200
Reynolds number
Fig. 8. Comparison of thermal resistance between the base fluid and the nanofluids.
Fig. 10. Comparison pressure drop between the base fluid and the nanofluids.
In this study, the overall heat sink performance can be shown in the thermal resistance form. The heat sink thermal resistance is defined as Rth ¼
1 h As
¼
1
ð12Þ
Q ave ðΔT LMTD Þ
where Rth is the thermal resistance. Using this definition, thermal resistance of heat sink is function of heat transfer coefficient. It is found that the heat sink thermal resistance decreases with increasing heat transfer rate (increasing Reynolds number). Based on the results shown in Fig. 8, it is seen that the heat sink with nanofluids as coolant give the thermal resistance lower than as de-ionized water for the same heat sink. Figs. 9–10 show the variation of pressure drop with nanofluid Reynolds number for different heat flux and for different heat sinks, respectively. The test section as shown in Fig. 2, the pressure drop across the test section is produced from: inlet port to inlet plenum due to flow sudden expansion, inlet plenum to channel inlets due to flow contraction, channel inlet to exit due to wall friction, channel exits to outlet plenum due to flow sudden expansion, and outlet plenum to outlet port due to flow contraction. It can be seen that the heat flux has an insignificant effect on the pressure drop of the nanofluids. This is due to the increase in the heat flux having a slight effect on the viscosity of the nanofluids,
which leads to a tiny change in the measured pressure drop of the nanofluids. As compare to the de-ionized water, the pressure drop in the nanofluids is very close to that in the de-ionized water for the given conditions.
5. Conclusions In the present study, the heat transfer characteristics of nanofluids cooling in the mini-rectangular fin heat sink are studied. The TiO2 nanofluids are used as the working fluid with de-ionized water as the base fluid. The results obtained from the nanofluids cooling in minirectangular fin heat sink are compared with the other from the deionized water cooling method. Effects of inlet temperature of coolant, coolant Reynolds number, and heat flux on the heat transfer characteristics of mini-rectangular fin heat sink are considered. It is found that the nanofluids have significant effect on the enhancement of heat transfer of the heat sinks. In addition, the pressure drops of the nanofluids are approximately the same as those of de-ionized water in the given conditions. The results of this study are of technological importance for the efficient design of cooling systems of electronic devices to enhance cooling performance.
4.0 3.5
Pressure drop (kPa)
3.0
qin = 14.80 kW/m
2
qin = 15.76 kW/m
2
qin = 16.75 kW/m
2
Acknowledgments
Channel high = 1.0 mm o
Nanofluids, Tin = 20 C
The authors would like to express their appreciation to the Excellent Center for Sustainable Engineering (ECSE) of the Srinakharinwirot University (SWU) for providing financial support for this study.
2.5
References
2.0 1.5 1.0 0.5 0.0 100
120
140
160
180
Reynolds number Fig. 9. Variation of pressure drop for different heat flux.
200
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