1. Introduction
The microelectronics sector, which manufactures electronic circuits, has grown quickly and has also manufactured more powerful, more compact, and more advantageous components for portable and fixed applications. These include data centers, electronic boxes, and cars. However, overheating of these electronic components could affect their performance and reduce their life cycles, particularly when they are used in regions suffering hot weather conditions [
1]. To enhance thermal conduction in electronic devices operating in these conditions it is essential to apply cooling systems that can appropriately dissipate the high heat flux released. The selection and design of a specific cooling system depends mainly on the heat flux dissipated by the heating component, the climatic conditions of the site, and the energy parameter for the cooling capacity required from the system installed. Among these methods, forced convection air cooling has been the one most commonly employed to maintain the working temperature of electronic components at a safe level. This is because of its simplicity and inherent cost-effectiveness. However, this cooling system remains limited for implementation in high-performance microelectronic equipment, and as a consequence, there is a rising interest in designing other methods of high-performance liquid cooling systems [
2,
3].
Single-phase microchannel heat sinks, consisting of a stream of water passing through a micro heat exchanger, are an efficient means of dissipating heat fluxes. These fluxes, usually in the order of hundreds of W/cm
2, exceed the rates that air-cooled systems are able to reach [
4]. In terms of cooling efficiency, micro heat exchangers are particularly promising components for usage in small-scale and advanced applications. They are compact, light, have low energy requirements, and have a relatively low cost. Therefore, they provide vast windows of opportunity for integration into microelectronic components. The functional features of micro heat exchangers were originally explored by Tuckerman and Pease in the early 1980s [
4]. This concept led to the investigation of heat transfer and fluid flow in rectangular microchannels. Microscopic channels 50 µm wide and 300 µm deep were etched in silicon, and deionized water was pushed through as the coolant. This design permitted the dissipation of 790 W/cm
2 with a temperature increase of 71°C for a single chip. Gao et al. [
5] studied the influence of channel height on the thermo-hydraulic characteristics of microchannels and minichannels. Friction factors agreed with those estimated by the conventional laminar theory, regardless of channel height. However, the authors noticed a heat transfer enhancement when the size of the channels was decreased. Mala and Li [
6] performed an experimental study into how microchannel diameter variation affected the pressure drop. The authors reported significant deviations between the flow characteristics and the theoretical predictions for microchannels with a reduced diameter.
Several numerical investigations evaluated the impact of the microchannel geometrical design on heat transfer mechanisms and flow patterns [
7,
8,
9,
10,
11,
12,
13]. One of the most relevant results found that corrugated microchannels demonstrated a high potential for dissipating high heat fluxes. Other recent studies [
14,
15,
16,
17,
18,
19] are based on the tree design reported by Bejan [
20]. This design is useful for selecting the best geometric layout for heat sinks to allow for maximum heat transfer between the cooling fluid and the wall of the microchannels.
Investigations are currently being carried out into the improvement of the thermophysical properties of coolants, and have attracted the interest of various researchers over the last few decades [
21]. The most promising technique to improve heat transfer is to employ nanofluids with metal nanoparticles in the base fluid. This mixture is more effective than conventional working fluids, such as water, ethylene glycol, or oil [
22,
23,
24]. Several computational and experimental studies, available in the literature, are focused on the heat transfer and fluid flow of nanofluids on macro and micro-scales [
25,
26,
27,
28,
29,
30,
31,
32,
33,
34,
35,
36,
37,
38,
39,
40,
41,
42,
43,
44,
45,
46]. Kalteh et al. [
47] investigated numerically and experimentally alumina-water nanofluid (0-5%) flow inside a wide heat sink microchannel. The two-phase Eulerian method was utilized to model the nanofluid flow. In addition, homogeneous modeling was carried out to compare the experimental results with those of the two-phase simulation approach. The authors' numerical results demonstrated that the two-phase technique was more appropriate than the homogeneous model for modeling nanofluid flow. The maximum deviations with experimental results were 12.6% and 7.4% for homogeneous and two-phase methods, respectively. Mohammed et al. [
48] analysed the performance of microchannel heat sinks employing Al
2O
3/water nanofluids at concentrations of 1–5%. They employed the finite volume approach, based on a hybrid discretization methodology and the SIMPLE algorithm to solve the velocity fields. The results revealed that the use of nanofluids could enhance heat transfer in heat sink microchannels, and that it was dependent on the volumetric fraction of nanoparticles dispersed in the base fluid. The authors also observed that the thermal resistance was lower for heat sink microchannels with nanofluids at a 5% nanoparticle volume fraction. TiO
2 nanoparticles have good heat transfer characteristics, are highly stable, easily available, ecologically safe, and have a low cost. Experimental research exploring the influence of high nanoparticle concentrations (= 2%) on convective heat transfer in microchannels is very scarce. Manay and Sahin [
49] studied how microchannel height and five different TiO
2 nanoparticle volume fractions of nanofluids (0.25%, 0.5%, 1.0%, 1.5% and 2.0%) in pure water impacted heat transfer and flow characteristics. The study indicated that a reduction in microchannel height significantly decreased heat duties and increased pressure drop. By increasing the nanosized TiO
2 particles concentration in the base fluid, heat transfer rates increased but there was no excessive increase in pressure drop when compared to using pure water. Furthermore, various studies are available in the literature on the use of high concentrations of TiO
2/water nanofluids in micro-heat exchangers. Martínez et al. [
50] investigated numerically the effect of microchannel height on the thermal performance of a heat sink subjected to a continuous heat flow of 50 W/cm
2 at the bottom surface. The dimensions of the microchannel were 283 μm width, 50 mm length, and three different heights (800 µm, 600 µm, and 400 µm). A laminar three-dimensional flow study was carried out using water and TiO
2/water nanofluids at weight concentrations of 1 wt% and 3 wt% and with Reynolds numbers ranging from 200 to 1200. The authors concluded that both the use of nanoparticles and the reduction of the microchannel height improved heat transfer at low Reynolds numbers of 200. The maximum increase obtained was 19.7% with a nanoparticle concentration of 3%.
To determine the heat transfer rate and friction factor, Nitiapiruk et al. [
51] investigated the effects of TiO
2/water nanofluids, at volume fractions of 0.5%, 1%, and 2%, in a microchannel heat sink with 40 flowing channels. The length, width, and height of each channel were 40 mm, 500 μm and 800 μm, respectively. The authors reported that the use of nanoparticles at a volume fraction of 2% with minimum amounts of heat flux and Reynolds number was more suitable than other volume fractions. Heydari et al. [
52] investigated the effect of a rib design on heat transfer characteristics and laminar flow for the nanofluid TiO
2/water in a three-dimensional rectangular microchannel at volume fractions of 0%, 2%, and 4%. Results indicated that the ribs significantly affected the pattern of fluid flow; however, it also varied depending on the Reynolds number of the flow.
The present study involved the use of titanium dioxide (TiO2) nanoparticles in a micro heat exchanger to investigate how this could enhance heat transfer for cooling of electronic components. There is scarce information in the open literature on the impact of using both micro heat exchangers and nanofluids for cooling of electronic components at high coolant inlet temperatures. Experiments and numerical simulations were conducted to explore how different parameters, such as the inlet temperature of the hot nanofluids, Reynolds number, and concentration of nanoparticles, affect heat transfer enhancement in the cooling of an electronic heating component. The results of this study could be very useful for the design of efficient cooling systems for electronic devices that operate at high temperatures. The use of nanofluids in micro heat exchangers can improve heat transfer performance and reduce energy consumption, which is essential in many industrial electronics applications.
4. Numerical Approach
This section presents the physical model used in the simulation analysis of the monophasic flow in the micro heat exchanger [
65]. The governing equations, boundary conditions, and meshing required for the numerical simulation are provided.
Figure 7 (A) depicts the configuration of the micro heat exchanger evaluated in the present study. The calculation model considered an elementary volume, shown in
Figure 7 (B). Computational Fluid Dynamics (CFD) ANSYS/FLUENT 14.0, employing the finite volume approach [
66] was used to simulate the conjugate heat transport within the microchannels.
4.1. Assumptions and Boundary Conditions
Cooling performance of the micro heat exchanger was investigated employing a three-dimensional fluid-solid model. The following assumptions were made in the model: (i) incompressible, laminar, and steady state flow; (ii) constant properties of solids and fluids; (iii) neglected effects of gravity in the momentum equation and viscous dissipation in the energy equation; (iv) adiabatic external surface boundaries, except on the bottom of the micro heat exchanger.
Numerical simulations were run using water at temperatures (Tin) between 25ºC and 65ºC at the collector entrance and water Reynolds numbers ranging from 400 to 2000. The heating power applied to the bottom of the micro heat exchanger was 100 W. A second-order upwind approach was employed to solve the energy and momentum equations, and the SIMPLE algorithm was applied to manage the coupling of pressure force fields and velocity. Walls between solid and liquid regions were designated interfaces, and the inlet and outlet walls of the microchannels were taken as inner walls. The remaining walls were considered adiabatic walls. The iterative procedure was regarded as successful when the residuals of the continuity and momentum equations were below 10-4 and those of the energy equations were lower than 10-7.
4.2. Governing Equations
The governing equations employed in the model are the standard continuity equations for mass conservation, the Navier-Stokes equation for momentum conservation, and the energy equation to predict the conjugate heat transfer. The assumptions mentioned above were established in order to construct the following governing differential equations for fluid flow and heat transfer:
- -
Momentum conservation equation:
- -
Energy conservation equation:
4.3. Effect of Grid Refinement
The mesh density was analyzed before carrying out the final simulations to determine how it influenced the numerical solution for the entire micro heat exchanger. Accordingly, different numerical trials were performed for several numbers of mesh elements between 6x10
5 and 2x10
6, to ensure independency of the mesh size. The heating element temperature and the total pressure were used to assess how the mesh number affected the accuracy of the results.
Figure 8 illustrates the mesh sensitivity analysis performed at a Reynolds number of 400, an input heat power of 100 W and an inlet cooling nanofluid temperature of 25ºC. It can be observed that the results are not affected when the number of mesh elements is above 14x10
5.
Figure 1.
Test module: (a) Micro heat exchanger, (b) Microchannels.
Figure 1.
Test module: (a) Micro heat exchanger, (b) Microchannels.
Figure 2.
Schematic diagram of the micro channel heat exchanger assembly.
Figure 2.
Schematic diagram of the micro channel heat exchanger assembly.
Figure 3.
Experimental rig.
Figure 3.
Experimental rig.
Figure 4.
Embedded device in parallelepiped box.
Figure 4.
Embedded device in parallelepiped box.
Figure 5.
Codification of the thermocouples installed in the experimental setup.
Figure 5.
Codification of the thermocouples installed in the experimental setup.
Figure 6.
Photos of the TiO2/water nanofluids prepared at different concentrations.
Figure 6.
Photos of the TiO2/water nanofluids prepared at different concentrations.
Figure 7.
Micro heat exchanger: (a) Layout, (b) Calculation volume.
Figure 7.
Micro heat exchanger: (a) Layout, (b) Calculation volume.
Figure 8.
Influence of the grid size on the temperature variation along the microchannels centerline.
Figure 8.
Influence of the grid size on the temperature variation along the microchannels centerline.
Figure 9.
Comparison of the average Nusselt number with the experimental data of Lee et al. [
67], and of the values calculated by the Peng and Peterson correlation, using pure water [
68].
Figure 9.
Comparison of the average Nusselt number with the experimental data of Lee et al. [
67], and of the values calculated by the Peng and Peterson correlation, using pure water [
68].
Figure 10.
Comparison of the friction factor with the experimental data of Harms et al. [
70], and the values calculated by the Shah and London correlation [
69].
Figure 10.
Comparison of the friction factor with the experimental data of Harms et al. [
70], and the values calculated by the Shah and London correlation [
69].
Figure 11.
Comparison of numerical results with data reported by Kawano et al. [
71], Qu and Mudawar [
72], and Al-Neama et al. [
73].
Figure 11.
Comparison of numerical results with data reported by Kawano et al. [
71], Qu and Mudawar [
72], and Al-Neama et al. [
73].
Figure 12.
Average Nusselt number versus nanofluid concentration and inlet temperature at Reynolds numbers: (a) Re=400, (b) Re=800, (c) Re=1200, (d) Re=1600.
Figure 12.
Average Nusselt number versus nanofluid concentration and inlet temperature at Reynolds numbers: (a) Re=400, (b) Re=800, (c) Re=1200, (d) Re=1600.
Figure 13.
Temperature profile of the heating component versus the Reynolds number and the nanoparticle concentration at nanofluid inlet temperatures: (a) Tin = 35°C, (b) Tin = 45°C, (c) Tin = 55°C, (d) Tin = 65°C.
Figure 13.
Temperature profile of the heating component versus the Reynolds number and the nanoparticle concentration at nanofluid inlet temperatures: (a) Tin = 35°C, (b) Tin = 45°C, (c) Tin = 55°C, (d) Tin = 65°C.
Figure 14.
Temperature contours (°C) at Re=1200, inlet temperature of 45ºC, and nanoparticle concentration: (a) pure water, (b) Ø=1%, and (c) Ø=5%.
Figure 14.
Temperature contours (°C) at Re=1200, inlet temperature of 45ºC, and nanoparticle concentration: (a) pure water, (b) Ø=1%, and (c) Ø=5%.
Figure 15.
Pressure drop versus the Reynolds number for both TiO2 nanofluids and pure water.
Figure 15.
Pressure drop versus the Reynolds number for both TiO2 nanofluids and pure water.
Figure 16.
Pumping power versus the Reynolds number for both TiO2 nanofluids and pure water.
Figure 16.
Pumping power versus the Reynolds number for both TiO2 nanofluids and pure water.
Figure 17.
Performance evaluation criterion (PEC) versus the Reynolds number and nanofluid inlet temperature at two TiO2 concentrations: (a) Ø= 1%, (b) Ø= 5%.
Figure 17.
Performance evaluation criterion (PEC) versus the Reynolds number and nanofluid inlet temperature at two TiO2 concentrations: (a) Ø= 1%, (b) Ø= 5%.
Table 1.
Geometric parameters of the micro heat exchanger shown in
Figure 2.
Table 1.
Geometric parameters of the micro heat exchanger shown in
Figure 2.
Geometric parameter |
Dimension (mm) / Number (-) |
Heat sink width (W) |
16 |
Heat sink height (H) |
1.63 |
Heat sink length (Lmc) |
40 |
Microchannel width (Wmc) |
0.7 |
Microchannel height (Hmc) |
1 |
Half thickness of the solid (es) |
0.35 |
Thickness of fins (e) |
0.25 |
Collector tube length (Lc) |
40 |
Hydraulic diameter (Dh) |
0.8 |
Collector tube diameter (Dc) |
5 |
Number of channels (N) |
17 |
Table 2.
Thermophysical properties of TiO
2 nanoparticles [
53,
54,
55].
Table 2.
Thermophysical properties of TiO
2 nanoparticles [
53,
54,
55].
Properties |
TiO2 nanoparticles |
Mean diameter, dp |
20 nm |
Thermal conductivity, k |
8.4 W/m.K |
Specific heat, Cp |
710 J/kg.K |
Density, |
4157 kg/m3
|
Table 3.
Uncertainties of the various sensors used in the experiments, as well as the parameters calculated.
Table 3.
Uncertainties of the various sensors used in the experiments, as well as the parameters calculated.
Sensor |
Uncertainty |
K-type thermocouple |
± 0.1°C |
Pressure sensors |
±2.5 % FS |
Peristaltic pump |
±1% |
Heater power supply voltage and current |
0.01% and 0.1% |
L (mm) |
2.5 |
W (mm) |
1.25 |
|
Parameter |
Uncertainty (%) |
Re |
1.54 |
∆P (Pa) |
0.5 |
h (W/m2.°C) |
2 |
Nu |
3 |
Table 4.
Heat transfer enhancement in the micro heat exchanger due to the addition of nanoparticles to pure water.
Table 4.
Heat transfer enhancement in the micro heat exchanger due to the addition of nanoparticles to pure water.
Φ (%) |
1 |
5 |
|
Tin (°C) |
35 |
45 |
55 |
65 |
35 |
45 |
55 |
65 |
Re |
|
400 |
26% |
24% |
23% |
21% |
37% |
35% |
32% |
30% |
800 |
35% |
32% |
30% |
28% |
52% |
49% |
47% |
45% |
1200 |
44% |
42% |
40% |
39% |
60% |
57% |
55% |
53% |
2000 |
57% |
56% |
54% |
49% |
70% |
65% |
63% |
59% |