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Thermal performance of modified polymeric heatsinks as an alternative for aluminum in heat rejection systems
Applied Thermal Engineering 159 (2019) 113823
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Research Paper
Thermal performance of modified polymeric heatsinks as an alternative for aluminum in heat rejection systems Hossein Rohania, Arash Badakhsha,b, Chan Woo Parka, a b
T
⁎
Division of Mechanical Design Engineering, Chonbuk National University, Jeonju, Jeonbuk 54896, Republic of Korea Multifunctional Carbon Materials Research Laboratory, Korea Institute of Carbon Convergence Technology KCTECH, Jeonju, Jeonbuk 54853, Republic of Korea
H I GH L IG H T S :
process for the two new polymeric composite material. • Manufacture effect of different filler ratios on the products thermal conductivity. • The the thermal performance of composite materials with the aluminum case. • Comparison • Heat diffusion observation inside the materials.
A R T I C LE I N FO
A B S T R A C T
Keywords: Polymer-matrix composite (PMC) Heat sink Infrared thermography Heat transfer
Many studies have been conducted to obtain a viable alternative for metallic compartments in heat exchangers. The present research has been conducted to distinctly reveal the performance characteristics of polymer based composite materials as heat transfer medium. Various prototypes of polymeric fins with different volume fractions for multiwalled-carbon nanotubes, copper as components, and high density polyethylene (HDPE) as a matrix, were fabricated and stacked around copper pipes as an air/water finned-pipe heat exchanger. The thermal performance of each case was tested and compared with those of aluminum fins under the same conditions in a miniature mechanically pumped cooling loop (MPCL). Results indicate that for the airside, the difference between total heat transfer rates of the aluminum heat sink increased more rapidly than that of polymeric cases at the beginning, and became constant when a certain Reynolds number is reached. Further, for the best case of new composite heat sinks, a 451% and 52% enhancement in thermal conductivity and heat rejection were achieved, respectively, compared with the pure HDPE. The thermographic method also was used for better visual comparison among the materials. Finally, numerical simulations using ANSYS Fluent, indicated consistency with the experimental results of heat distribution for each case.
1. Introduction The thermal management of sensitive units such as electronic control units (ECUs) require sensitive measures in system design [1] to achieve high reliability and safety. Weight, size, efficiency, and heat rejection mechanism are few significant features that must be considered. The chosen material is critical for this purpose. For long, polymers have been under the spotlight as the base material for many engineering applications [2] such as heat recovery, solar water heating, and cooling/refrigeration systems, [3–7]. However, metals are still the material of choice for heat exchangers hitherto; the feasibility of polymer-based heat rejection systems in an actual cooling application is deemed possible and viable [8,9]. Nonetheless, insufficient information
⁎
on polymeric stacked-fin heat sink deprive thermal engineers from a better and broader conclusion on the actual application of such significant units and their potential use in any other applications in which polymer characteristics may be required. The present study was conducted to reveal this actual applicability and to better understand the operational characteristics of such units. The closest study in terms of geometry was performed by Chen et al. [10]. They experimentally studied the application of thermally modified polypropylene (PP) in a finned-tube heat exchanger, and the results showed that a threshold value of material thermal conductivity existed that divided the performance curve into two parts. Below this value, improving the thermal conductivity can improve the heat exchanger performance considerably; meanwhile, over this value, the
Corresponding author. E-mail address: [email protected] (C.W. Park).
https://doi.org/10.1016/j.applthermaleng.2019.113823 Received 16 January 2019; Received in revised form 24 April 2019; Accepted 24 May 2019 Available online 25 May 2019 1359-4311/ © 2019 Published by Elsevier Ltd.
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Nomenclature A Atot Cp d dae F f K ṁ Nu Pr Q̇ Re SR T tl tq
tR U umax
area (m2) total surface area per meter (m) specific heat at constant pressure (j/kg.K) diameter (m) hydraulic diameter for single flow channel (m) LMTD Correction factor fanning friction factor thermal conductivity (W/mK) mass flow rate (kg/s) Nusselt number Prandtl number heat transfer rate (W) Reynolds number fin thickness (m) temperature (°C) longitudinal pitch (m) transvers pitch (m)
fin pitch (m) overall heat transfer coefficient (W/m2.K) maximum air speed (m/s)
Subscribes a b c h i Lm,cf o
air side bulk cold hot inside logarithmic mean temperature for cross-counter flow outside
Greek letters µ ρ
dynamic viscosity (m2/s) fluid density (kg/m3)
fillers were chosen and the best result (in case of highest thermal conductivity) was selected for the heat sink material [15,16]. Finally, in the MWCNT/HDPE case, the filler (MWCNT) volume fraction was chosen to be 6% and for composite reinforced with hybrid filler (Cu +MWCNT/HDPE), this number was 10% (5 vol% for Cu and 5 vol% for MWCNT). All composite materials and pristine HDPE (LUTENE-HPE0235, LG Chemical Co., Ltd., Korea) were molded into a plate-shaped fin, and dimensionally corresponding aluminum fins were purchased from ThermoLab Co., Ltd., Korea. The melt blending method [14] was employed to prepare the granular pristine polymers for the shaping process. Polymer granules were placed in a heated chamber (180 °C) and after complete melting and minimization of air voids, the polymer was stirred for 30 min by two rotating shafts. Then, fillers were added gently to the melted HDPE and allowed to mix for 60 min. The pasty material removed from the mixer was formed into the desired shape by a hot press machine. Due to the small thickness of the fin, a male-and-female paired mold with high dimensional accuracy was specifically designed for this purpose. The composite material preparation, molding process, developed pair molds view, and manufactured fin batch are shown in Fig. 1. Narrow channels connecting all six casting sections and allowed a uniform and continues flow and distribution of the molten material. The ultimate application of the miniature heat rejection system is in small-scale thermally sensitive units. To achieve an efficiently large heat transfer area, the polymeric heat sink was schemed as an aircooled heat exchanger formed by six parallel multi-pass copper tubes (t = 0.65 mm, dout = 6.35 mm) dressed with 0.43 mm-thick plateshaped fins. Fig. 2a illustrates the geometrical specifications of a single fin. To ensure a firm attachment of the fins to the tube walls, and an equal fin spacing, the fins have 2 mm protruded edges around the holes. Moreover, a high thermal conductivity paste (Omegatherm 201, Omega Engineering, Inc., USA) was applied on the contact areas in all the cases to ensure an equal contact condition. The final fabricated heat sink consists of 40 fins with 2.43 mm for the fin pitch, as shown in Fig. 2b.
thermal performance enhancement is little. Moreover, thermally conductive polyphenylene sulphide (PPS), which is among the typical polymer materials for heat exchange applications [11], has been used in some doubly-finned plate liquid-liquid heat exchangers [12] and pin fin heat sinks [13], and was also examined under the same conditions. Consequently, the PPS polymer heat exchangers were found to provide comparable thermal properties to those achieved with aluminum heat sinks. Several commercially available polymeric materials were tested to investigate the performance of polymers in harsh chemical environment [14]. The results indicated their potential for use in gas-fired condensing heat exchangers with no evidence of degradation after 10,000 cycles. Likewise, in the author’s previous study, the thermal and mechanical properties of polymer-based composites were studied for the mixture of high density polyethylene (HDPE) with multiwalled-carbon nanotubes (MWCNTs) and copper powder under different volume fractions and ball milling time [15–17]. The result indicated enhancement in thermal conductivity for all cases especially for the Cu +MWCNT/HDPE hybrid, which was almost thrice higher than that of pure HDPE. In this study, an air/water finned-tube heat exchanger was selected to compare the thermal performances of several novel polymer-matrix composites (PMCs): MWCNT/HDPE (MH) and Cu+MWCNT/HDPE (CMH) with pure HDPE, and aluminum heat sink. The comparison results of all the polymeric cases with typically used aluminum heat sink clarify the operational possibility of the fabricated composites in heat rejection systems. Lastly, the experimental results were verified versus the numerical simulation of the setup.
2. Experimental procedure 2.1. Experimental apparatus 2.1.1. Design and fabrication of the polymeric heat sink (PHS) In this research, high density polyethylene (HDPE) was chosen for the matrix material as it possesses the highest thermal conductivity (0.57 W/mK) among the widely used polymers. Likewise, 60-min ball milled multi-walled-carbon nanotubes “MWCNTs” (Carbon Nano-material Technology) and copper “Cu” powder (Chang Sung Corporation, ACU-325) were used as fillers owing to their excellent thermal conductivity. Because filler volume fraction and particle size are the most important and effective factors in composite thermal conductivity and crucial in composite design [18–25], different volumetric ratios for
2.1.2. Mechanically pumped cooling loop (MPCL) With the rapid progress of electronics in numerous electromechanical systems and the consequent ongoing demand for size and weight reduction, a compatible thermal management solution is inevitably a challenge. A mechanically pumped cooling loop (MPCL) is a key theme in developing a reliable alternative for conventional aluminum and/or copper cooling units. 2
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Fig. 1. (a) Composites material preparation, molding steps and (b) Developed pair molds illustrations.
thickness were used to create a square tunnel of 15 cm internal length. Two U-shaped-finned-pipe-heaters were installed at the air-inlet hatch and were connected to a digital-controller (OKE-2002, im Tech) to set the inlet-air temperature constant. To achieve a uniform airstream, the air passed through a honeycomb straightener before approaching the fins. After the heat sink, the air-duct cross-section was changed to a circular shape with an inner diameter of 15.4 cm (polyethylene pipe), using an adaptor, such that a high accuracy thermal mass flow meter (Fastflo 620G, Sierra Instruments, Inc. Netherland) could be used. Since the mass flow meter measurement could be influenced by the air stream profile, adequate distances from the adaptor and suction were adhered to avoid a non-uniform stream. The air flow was controlled by a digital inverter (Hyundai N700E Inventor, Ltd., Korea) connected to a suction pump (HSB-SSF28T, Hanil Co.,Ltd., Korea) at the end of the air duct. The air-duct details, dimensions, and MPCL are shown in Fig. 3.
Herein, a miniature MPCL (shown in Fig. 3b) was established to assess the feasibility of the PMC-fabricated heat sink application. Due to its flexibility, strength, corrosion, fouling resistance, and high thermal conductivity, copper was chosen for piping. The water inlet was set constant using a water-bath (Refrigerated/Heating Circulators RCH3015, Ltd CP Ti, Korea) and the temperature was measured at four points: the inlet and outlet of the heat sink for both the water and air sides, using resistance temperature detectors (RTDs). The water mass flow was measured by an electromagnetic flow meter (ModMAG M2000, Czech Republic) placed between the water bath and heat sink water-inlet.
2.1.3. Air duct (wind tunnel) The air duct is vital for supplying a steady and uniform flow over the fins to approach the PHS. Owing to ease in the manufacturing and the suitability for fitting various heat sinks and RTDs, acrylic sheets of 1 cm 3
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̇ = ṁ air Cp, air (Tair , out − Tair , in ) Qair
(2)
Q̇ water = ṁ water Cp, water (Twater , out − Twater , in)
(3)
̇ ) and overall heat Further, the average heat transfer rate (Qave transfer coefficient (U) based on the inlet and outlet temperatures and mass flow rates of both water and air sides were calculated, using the logarithmic mean temperature difference (LMTD) method [27], as outlined by the following equations:
̇ + Q̇ water Qair 2
̇ = Qave
(4)
̇ = UAF ΔTlm, cf Qave
(5)
In which
ΔTlm, cf =
(Twater . out − Tair , in ) − (Twater , in − Tair , out ) ln [(Twater ,out − Tair , in )/(Twater , in − Tair ,out )]
(6)
where ΔTlm, cf is the LMTD for counter flow arrangement, and F is the non-dimensional correction factor. The corresponding correction factors were determined to be approximately one (F ≈ 1) based on the correction factor chart for the cross-flow heat exchanger [27]. Gnielinki’s recommended equation was chosen corresponding to the experimental condition (turbulent and transition region) for the Nusselt number (Eq.7)
(f /2)(Reb−1000) Prb
Nub =
2
1
1 + 12.7(f /2) 2 (Prb3 − 1)
(7)
where Pr is the Prandtl number, and f is the Fanning friction factor that can be obtained by equations below:
f = (1.58LnReb−3.28)−2
Reb =
Rea =
cp μ K
(9)
ρa umax dae μa
(10)
-dae : Hydraulic diameter of single flow channel Because the modified Wilson plot method constitutes a suitable technique to estimate the convection coefficients in a variety of convective heat transfer processes without requiring direct measurement of the surface temperature, this method was applied for each heat sink to obtain the air-side heat transfer resistance from the overall conductance [28]. Data were obtained for water flow rates between 0.005 and 0.015 kg/s inside the copper pipe as the hot-side fluid, and constant air flow rate and temperature of 0.096 kg/s and 21 °C as the cold-side fluid. The total thermal resistance (Rtotal) of the heat exchanger, as shown in Eq. (11), is consists of the inside thermal resistance (Rin), tube wall thermal resistance (Rwall), thermal contact resistance (Rcontact), and outside thermal resistance (Rout).
2.2. Experimental uncertainty Table 1 shows the summary of uncertainties of the measured and calculated data. The measured uncertainties were obtained from the instrument manufactures and for the calculated type, Engineering Equation Solver (EES) software was used. The method for determining the experimental uncertainties can be summarized in Eq. (1). More details can be found in Ref. [26].
∂R ∂R ∂R σx1)2 + ( σx2 )2 +⋯+( σx n )2 ∂x1 ∂x2 ∂x n
Prb =
Likewise, the Reynolds numbers for the airside was calculated from Eq. (10):
Fig. 2. (a) Geometric dimension of the fin and (b) illustration of the actual fabricated PHS (Cu/MWCNT-HDPE).
σR = ± (
4ṁ , πdi μ
(8)
(1)
where
R = f (x1, x2 , ⋯x n )
Rtotal = Rin + Rwall + R contact + Rout
In the equation above, the resultant R (which can be any parameter) is function of x1, x2, …, xn, and σR represents the uncertainty of the resultant R.
(11)
The total thermal resistance of the heat exchanger is related to the overall conductance through the following expression:
Rtotal = 2.3. Data reduction
Rtotal
Since the thermal performance is the primary objectives of this research, parameters such as thermal resistance, total heat transfer rate, and overall heat transfer coefficient were calculated for all the heat sinks. The air and water heat transfer rates were calculated using Eqs. (2) and (3) respectively.
1 F ΔTLMCF = ̇ UA Qave
1 1 = + Rwall + R contact + hi Ai ηo ho Ao
(12)
(13)
Since the thermal resistance of tube due to copper high thermal conductivity and its thickness are much smaller than the inside and outside resistance of the pipe, they are negligible. Therefore the overall 4
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H. Rohani, et al.
Fig. 3. Schematic illustration of experimental setup for (a) Air duct, and (b) mechanically pumped cooling loop. Table 1 Uncertainties of experimental data. Measured variable
Calculated variable
Variable
Uncertainty
Variable
Relative Uncertainty
Temperature
± 0.1 °C
± 1.8%
Water mass flow rate Air mass flow rate Length tr , sr Diameter
± 0.1 L/min
Reynolds number (air) Reynolds number (water)
± 0.4%
3
± 0.3 M /min ± 0.001 m ± 0.00001 m ± 0.00001 m
Fig. 5. Fin and heater installation for thermography process.
thermal resistance is expressed as Eq. (14).
1 1 1 ≈ + R contact + UA hi Ai ηo ho Ao Fig. 4. Plot used to obtain the air-side heat resistance (modified Wilson plot).
Nui =
5
hi × di ki
(14)
(15)
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Fig. 6. Schematic of filler alignment in hot-press molding process (most particles are aligned to flow direction inside the mold).
Fig. 7. Thermal conductivity for various filler volume percentage for (a) MWCNT/HDPE and (b) Cu+MWCNT/HDPE composite.
Through-plane
In-plane
Nusselts number. Therefore, the plot in Fig. 4 that shows both the water-side and air-side thermal resistances is applicable. The important aspects regarding this chart are the line extrapolation and intersection point with the vertical axes (1/U) which represent the neat air-side thermal resistance.
2.7 1.717
205 1.317
205 2.574
2.4. Thermographic analysis of materials
0.952
0.57
0.57
1.043
0.644
0.71
Table 2 Selected materials properties near room temperature. Material
Density (g/ cm3)
Aluminum (Al) Cu/MWCNT-HDPE (CMH) composite High density polyethylene (HDPE) MWCNT-HDPE (MH) composite
Thermal conductivity (W/m∙K)
Because the contact and fins thermal resistance were assumed to be constant value for all cases, therefore, by maintaining the air temperature and mass flow rate steady and changing the water mass flow, all changes in the total thermal resistance value will be caused by water-side heat transfer coefficient (hi). Likewise, the heat transfer coefficient is related to the Nusselts number (Nu) by Eq. (15). By considering Eqs. (14) and (15), Eqs. (16)–(19) can be achieved.
For a better comparison among four different materials that were used as heat sinks, the thermographic method was applied to perceive heat diffusion inside the fins. For simplifying the process, a cartridge heater of diameter 8 mm was passed through the center of the squareshaped fin of length 80 mm and thickness 2 mm as shown in Fig. 5. To reduce the radiation effect on the results and for equal conditions for all the cases, a thin graphite layer (GRAPHIT 33, CRC Industries, Deutschland) was sprayed on the fin surfaces. Then each fin was heated gradually from room temperature (26 °C) until it was stabilized at 65 °C. Thermal pictures (IR) were captured from fin surfaces at steady state using a thermal camera (FLIR T620, FLIR Systems, Inc., Sweden).
1 d 1 1 = A (( i · ) + (R contact + )) U ki Ai Nui ηo ho Ao
3. Results and discussion
α = A·
di = constant ki Ai
β = A (Rcontact +
(18), (19) →
1 ) = constant ηo ho Ao
1 α = +β U Nui
(16)
3.1. Thermal conductivity results (17) To maximize the heat rejection from tubes using molded fins, the inplane thermal conductivity is more important than that of the throughplane. From fluid mechanics, the shear stress profile for viscoelastic substance in a channel changes through the transvers axis. Thus, the stress value for the fluid layer next to the wall is zero and increases by getting far from wall until it reaches to maximum level at the channel center line. This shear stress difference between fluid layers renders the
(18)
(19)
As shown, the left side of Eq. (19) is inversely related to the water 6
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After 30 min
Steady situation
HDPE
MWCNT/HDPE
Cu+MWCNT/HDPE
Aluminum
After 15 min
(a)
(b) Fig. 8. Heat diffusion results for fins, using time lapse thermography (a) for range of 25–65 °C and (b) 26–38 °C.
filler particles rotation and alignment parallel to the flow direction. Therefore, the shear stress due to pressing process prior to the conductivity measurement causes the particles to be aligned perpendicularly to the samples measuring plane [29] (Fig. 6). Since the copper particles are larger than MWCNTs (almost 20 times larger), they were affected more than the MWCNTs and consequently aligned better to the flow path. The C-Therm TCi thermal conductivity analyzer (C-Therm Technologies Ltd., Canada) was used to thermally characterize the
conductive polymeric samples. This method is based on the difference in voltage in the modified transient plane source (MTPS) probe induced by the temperature gradient during the sampling time. Both the measured in-plane and through-plane thermal conductivity of two new composite materials for various fillers volumetric ratios can be seen in Fig. 7. An improvement of up to 22.4% and 245.5% for the in-plane and 11.03% and 88.9% for the through-plane thermal conductivity was achieved for MWCNT/HDPE and MWCNT+ Cu/HDPE, respectively.
7
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Fig. 11. Modified Wilson plot results for the heatsinks.
Fig. 9. Experimental results for average heat transfer rate under various Reynolds numbers on air-side.
Fig. 12. Numerical simulation results for fin radial temperature distribution. Fig. 10. Experimental results of overall heat transfer coefficient changes (U) for various Reynolds number.
still demonstrates the highest heat rejection rate.
The higher thermal conductivity of the CMH is attributed to better networks with larger mean free paths for the ballistic motion of phonons through the composite. Table 2 shows the final results including density and thermal conductivity of each material near the room temperature for this research.
3.3. Thermal performance The average heat transfer rate and overall heat transfer coefficient versus various Reynolds number [3,30,31] in variety of 2000 up to 100,000 for the air side were calculated and are shown in Figs. 9 and 10, respectively. For the lowest Reynolds number (around 2000), the heat transfer rate gap between the best (aluminum) and worst (HDPE) cases was approximately 0.027 kW. By increasing the air flow rate, the heat transfer rate number also raised up significantly for aluminum and with a smooth slop for three other cases (due to increasing of heat transfer coefficient). After the Reynolds number reached to around 80,000, heat transfer rate becames constant for all four heat sinks and difference between the best and worst cases was 0.177 kW. A similar trend was also observed for overall heat transfer coefficient. The convective heat transfer is always the result of mass flow rate, temperature difference between fluid and surface, and heat transfer area. Since the mass flow, fluid temperature, and area were the same for all cases, therefore the only effective parameter was the fin efficiency, which was directly related to the material thermal conductivity. An enhancement of up to 23.4% and 52.4% in average heat transfer rate, and 147% and
3.2. Thermographic analysis of materials Because thermography facilitates the visual comparison of heat diffusion through materials, square-shaped manufactured fins were heated using a cartridge heater until they reached to steady state (Fig. 5), and then the thermal pictures (IR) were taken from fin surfaces. Fig. 8a shows the significant temperature difference between the composite cases and aluminum. Therefore, another temperature range was set on the thermal camera for a better comparison among the polymeric cases (Fig. 8b). The results of HDPE and MWCNT/HDPE were extremely close to each other. At the steady state, CMH exhibits a higher heat rejection compared with other polymeric fins. This is due to the higher thermal conductivity of the CMH and better interface between the copper particles and polymeric matrix. Aluminum, however, 8
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Fig. 13. Simulation domain and boundary condition.
Fig. 14. Single fin simulation results for (a) HDPE, (b) MWCNT/HDPE, (c) Cu+MWCNT/HDPE, and d) aluminum.
between the through-plane and in-plane thermal conductivities, simulation was performed in a three-dimensional (3D) domain. The steadystate solver and energy model were chosen for natural conduction heat transfer. Following the experimental conditions, the heater and room temperature were set at 65 °C and 26 °C respectively. The radial heat diffusion and fins average temperature results are shown in Fig. 12, showing similarity with the experimental results shown in Fig. 8, and represent the best and worst heat diffusion for Aluminum and pure HDPE respectively. Owing to the complex geometry of the heat sink, the simulation was done for a single fin for each case. The mesh properties for the 3D domain was 1,661,992 nodes and the maximum aspect ratio of 74.6. The air flow with constant thermal properties were assumed to be a steady, incompressible and non-viscous dissipation throughout the computational domain. The continuity, momentum and energy equations for the RNG κ − ε model were set for the water and air turbulent conditions. The simulation domain and final results are shown in
287% in overall heat transfer coefficient were achieved for MWCNT/ HDPE and MWCNT+Cu/HDPE, respectively, compared to pure HDPE. As mentioned earlier, the modified Wilson plot method was considered to determine the air-side overall heat resistance. Hence, the airside mass flow and temperature were maintained constant while the water mass flow was varied. Based on the modified Wilson plot method, the intersection point of trend lines in Fig. 11 with the vertical axis represents the total thermal resistance for the air-side. Because all the conditions such as heat sink geometry and thermal contact resistance (between pipes and fins) were the same for all cases, the heat sinks material thermal conductivity would be regarded as the only impact parameter on the overall air side thermal resistance. 3.4. Simulation The ANSYS Fluent software was used to obtain more details regarding the heat dispersion inside the fins. Due to the difference 9
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Figs. 13 and 14.
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Newman, Composite Materials Technology: Processes and Properties, Hanser Puplishers, Munich Vienna New York, 2000. [30] M. Saeed, M.H. Kim, Heat transfer enhancement using nanofluids (Al2O3-H2O) in mini-channel heatsinks, Int. J. Heat Mass Transf. 120 (2018) 671–682, https://doi. org/10.1016/j.ijheatmasstransfer.2017.12.075. [31] Y. Cormier, P. Dupuis, A. Farjam, A. Corbeil, B. Jodoin, Additive manufacturing of pyramidal pin fins: height and fin density effects under forced convection, Int. J. Heat Mass Transf. 75 (2014) 235–244, https://doi.org/10.1016/j. ijheatmasstransfer.2014.03.053.
4. Conclusions 1. Experimental results indicated that improvement in both in-plane and through-plane thermal conductivity could be achieved by employing high thermally conductive filler in polymer matrix. Enhancements of up to of 22.4% and 245.5% for in-plane and 11.03% and 88.9% for through-plane thermal conductivities, as well as 23.4% and 52.4% increase in heat transfer rate were achieved for MWCNT/HDPE and MWCNT+Cu/HDPE respectively. 2. Heat transfer rate difference for all four cases at small air Reynolds numbers (close to natural convection) was much smaller than that of forced convection. In other words, for low air Reynolds numbers, the overall heat transfer coefficient ratio of aluminum over other cases were 4.07, 3.97, and 1.9 for CMH, MH, and HDPE respectively, and 5, 3.06, and, 2.27 respectively for high Reynolds numbers. 3. Notwithstanding the MWCNTs shows higher thermal conductivity than copper (Cu), a better interface between copper particles and polymer chain resulted in a superior path for heat to disperse inside the composite. 4. As copper particles are approximately 20 times larger than the MWCNTs, fluid flow in the molding process affected the copper particles alignment significantly. Consequently, the in-plane over through-plane thermal conductivity ratios were calculated as 1.95 and 1.1 for CMP and MH composites respectively. Acknowledgment This work was supported by the Ministry of Trade, Industry, and Energy (MOTIE) of the Republic of Korea (P0002131) and by the National Research Foundation of Korea funded by the Korean Government (grant number 2019R1A2C2010607). Appendix A. Supplementary material Supplementary data to this article can be found online at https:// doi.org/10.1016/j.applthermaleng.2019.113823. References [1] K.C. Otiaba, N.N. Ekere, E.H. Amalu, R.S. Bhatti, S. Mallik, Thermal Management materials for electronic control unit: trends, processing technology and R and D challenges, Adv. Mater. Res. 367 (2011) 301–307, https://doi.org/10.4028/www. scientific.net/AMR.367.301. [2] X. Chen, Y. Su, D. Reay, S. Riffat, Recent research developments in polymer heat exchangers – a review, Renew. Sustain. Energy Rev. 60 (2016) 1367–1386, https:// doi.org/10.1016/j.rser.2016.03.024. [3] M.A. Arie, A.H. Shooshtari, R. Tiwari, S.V. Dessiatoun, M.M. Ohadi, J.M. Pearce, Experimental characterization of heat transfer in an additively manufactured polymer heat exchanger, f Eng. 113 (2017) 575–584, https://doi.org/10.1016/j. applthermaleng.2016.11.030. [4] R. Trojanowski, T. Butcher, M. Worek, G. Wei, Polymer heat exchanger design for condensing boiler applications, Appl. Therm. Eng. 103 (2016) 150–158, https:// doi.org/10.1016/j.applthermaleng.2016.03.004. [5] I. Astrouski, M. Raudensky, I. Krásny, Polymeric hollow fiber heat exchanger as an automotive radiator, Appl. Therm. Eng. 108 (2016) 798–803, https://doi.org/10. 1016/j.applthermaleng.2016.07.181. [6] A. Roja, J. Hussain, A.A. Alahyari, S.A. Eastman, C. Thibaud-erkey, S. Johnston, M.J. Sobkowicz, Review of polymers for heat exchanger applications: factors concerning thermal conductivity, Appl. Therm. Eng. 113 (2017) 1118–1127, https:// doi.org/10.1016/j.applthermaleng.2016.11.041. [7] I. Astrouski, M. Dohnal, M. Raudensky, Intensification of heat transfer of polymeric hollow fiber heat exchangers by chaotisation, Appl. Therm. Eng. 113 (2017) 632–638, https://doi.org/10.1016/j.applthermaleng.2016.11.038. [8] J.G. Cevallos, A.E. Bergles, A. Bar-Cohen, P. Rodgers, S.K. Gupta, Polymer heat exchangers—history, opportunities, and challenges, Heat Transf. Eng. 33 (2012) 1075–1093, https://doi.org/10.1080/01457632.2012.663654.
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Research Paper
Thermal performance of modified polymeric heatsinks as an alternative for aluminum in heat rejection systems Hossein Rohania, Arash Badakhsha,b, Chan Woo Parka, a b
T
⁎
Division of Mechanical Design Engineering, Chonbuk National University, Jeonju, Jeonbuk 54896, Republic of Korea Multifunctional Carbon Materials Research Laboratory, Korea Institute of Carbon Convergence Technology KCTECH, Jeonju, Jeonbuk 54853, Republic of Korea
H I GH L IG H T S :
process for the two new polymeric composite material. • Manufacture effect of different filler ratios on the products thermal conductivity. • The the thermal performance of composite materials with the aluminum case. • Comparison • Heat diffusion observation inside the materials.
A R T I C LE I N FO
A B S T R A C T
Keywords: Polymer-matrix composite (PMC) Heat sink Infrared thermography Heat transfer
Many studies have been conducted to obtain a viable alternative for metallic compartments in heat exchangers. The present research has been conducted to distinctly reveal the performance characteristics of polymer based composite materials as heat transfer medium. Various prototypes of polymeric fins with different volume fractions for multiwalled-carbon nanotubes, copper as components, and high density polyethylene (HDPE) as a matrix, were fabricated and stacked around copper pipes as an air/water finned-pipe heat exchanger. The thermal performance of each case was tested and compared with those of aluminum fins under the same conditions in a miniature mechanically pumped cooling loop (MPCL). Results indicate that for the airside, the difference between total heat transfer rates of the aluminum heat sink increased more rapidly than that of polymeric cases at the beginning, and became constant when a certain Reynolds number is reached. Further, for the best case of new composite heat sinks, a 451% and 52% enhancement in thermal conductivity and heat rejection were achieved, respectively, compared with the pure HDPE. The thermographic method also was used for better visual comparison among the materials. Finally, numerical simulations using ANSYS Fluent, indicated consistency with the experimental results of heat distribution for each case.
1. Introduction The thermal management of sensitive units such as electronic control units (ECUs) require sensitive measures in system design [1] to achieve high reliability and safety. Weight, size, efficiency, and heat rejection mechanism are few significant features that must be considered. The chosen material is critical for this purpose. For long, polymers have been under the spotlight as the base material for many engineering applications [2] such as heat recovery, solar water heating, and cooling/refrigeration systems, [3–7]. However, metals are still the material of choice for heat exchangers hitherto; the feasibility of polymer-based heat rejection systems in an actual cooling application is deemed possible and viable [8,9]. Nonetheless, insufficient information
⁎
on polymeric stacked-fin heat sink deprive thermal engineers from a better and broader conclusion on the actual application of such significant units and their potential use in any other applications in which polymer characteristics may be required. The present study was conducted to reveal this actual applicability and to better understand the operational characteristics of such units. The closest study in terms of geometry was performed by Chen et al. [10]. They experimentally studied the application of thermally modified polypropylene (PP) in a finned-tube heat exchanger, and the results showed that a threshold value of material thermal conductivity existed that divided the performance curve into two parts. Below this value, improving the thermal conductivity can improve the heat exchanger performance considerably; meanwhile, over this value, the
Corresponding author. E-mail address: [email protected] (C.W. Park).
https://doi.org/10.1016/j.applthermaleng.2019.113823 Received 16 January 2019; Received in revised form 24 April 2019; Accepted 24 May 2019 Available online 25 May 2019 1359-4311/ © 2019 Published by Elsevier Ltd.
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Nomenclature A Atot Cp d dae F f K ṁ Nu Pr Q̇ Re SR T tl tq
tR U umax
area (m2) total surface area per meter (m) specific heat at constant pressure (j/kg.K) diameter (m) hydraulic diameter for single flow channel (m) LMTD Correction factor fanning friction factor thermal conductivity (W/mK) mass flow rate (kg/s) Nusselt number Prandtl number heat transfer rate (W) Reynolds number fin thickness (m) temperature (°C) longitudinal pitch (m) transvers pitch (m)
fin pitch (m) overall heat transfer coefficient (W/m2.K) maximum air speed (m/s)
Subscribes a b c h i Lm,cf o
air side bulk cold hot inside logarithmic mean temperature for cross-counter flow outside
Greek letters µ ρ
dynamic viscosity (m2/s) fluid density (kg/m3)
fillers were chosen and the best result (in case of highest thermal conductivity) was selected for the heat sink material [15,16]. Finally, in the MWCNT/HDPE case, the filler (MWCNT) volume fraction was chosen to be 6% and for composite reinforced with hybrid filler (Cu +MWCNT/HDPE), this number was 10% (5 vol% for Cu and 5 vol% for MWCNT). All composite materials and pristine HDPE (LUTENE-HPE0235, LG Chemical Co., Ltd., Korea) were molded into a plate-shaped fin, and dimensionally corresponding aluminum fins were purchased from ThermoLab Co., Ltd., Korea. The melt blending method [14] was employed to prepare the granular pristine polymers for the shaping process. Polymer granules were placed in a heated chamber (180 °C) and after complete melting and minimization of air voids, the polymer was stirred for 30 min by two rotating shafts. Then, fillers were added gently to the melted HDPE and allowed to mix for 60 min. The pasty material removed from the mixer was formed into the desired shape by a hot press machine. Due to the small thickness of the fin, a male-and-female paired mold with high dimensional accuracy was specifically designed for this purpose. The composite material preparation, molding process, developed pair molds view, and manufactured fin batch are shown in Fig. 1. Narrow channels connecting all six casting sections and allowed a uniform and continues flow and distribution of the molten material. The ultimate application of the miniature heat rejection system is in small-scale thermally sensitive units. To achieve an efficiently large heat transfer area, the polymeric heat sink was schemed as an aircooled heat exchanger formed by six parallel multi-pass copper tubes (t = 0.65 mm, dout = 6.35 mm) dressed with 0.43 mm-thick plateshaped fins. Fig. 2a illustrates the geometrical specifications of a single fin. To ensure a firm attachment of the fins to the tube walls, and an equal fin spacing, the fins have 2 mm protruded edges around the holes. Moreover, a high thermal conductivity paste (Omegatherm 201, Omega Engineering, Inc., USA) was applied on the contact areas in all the cases to ensure an equal contact condition. The final fabricated heat sink consists of 40 fins with 2.43 mm for the fin pitch, as shown in Fig. 2b.
thermal performance enhancement is little. Moreover, thermally conductive polyphenylene sulphide (PPS), which is among the typical polymer materials for heat exchange applications [11], has been used in some doubly-finned plate liquid-liquid heat exchangers [12] and pin fin heat sinks [13], and was also examined under the same conditions. Consequently, the PPS polymer heat exchangers were found to provide comparable thermal properties to those achieved with aluminum heat sinks. Several commercially available polymeric materials were tested to investigate the performance of polymers in harsh chemical environment [14]. The results indicated their potential for use in gas-fired condensing heat exchangers with no evidence of degradation after 10,000 cycles. Likewise, in the author’s previous study, the thermal and mechanical properties of polymer-based composites were studied for the mixture of high density polyethylene (HDPE) with multiwalled-carbon nanotubes (MWCNTs) and copper powder under different volume fractions and ball milling time [15–17]. The result indicated enhancement in thermal conductivity for all cases especially for the Cu +MWCNT/HDPE hybrid, which was almost thrice higher than that of pure HDPE. In this study, an air/water finned-tube heat exchanger was selected to compare the thermal performances of several novel polymer-matrix composites (PMCs): MWCNT/HDPE (MH) and Cu+MWCNT/HDPE (CMH) with pure HDPE, and aluminum heat sink. The comparison results of all the polymeric cases with typically used aluminum heat sink clarify the operational possibility of the fabricated composites in heat rejection systems. Lastly, the experimental results were verified versus the numerical simulation of the setup.
2. Experimental procedure 2.1. Experimental apparatus 2.1.1. Design and fabrication of the polymeric heat sink (PHS) In this research, high density polyethylene (HDPE) was chosen for the matrix material as it possesses the highest thermal conductivity (0.57 W/mK) among the widely used polymers. Likewise, 60-min ball milled multi-walled-carbon nanotubes “MWCNTs” (Carbon Nano-material Technology) and copper “Cu” powder (Chang Sung Corporation, ACU-325) were used as fillers owing to their excellent thermal conductivity. Because filler volume fraction and particle size are the most important and effective factors in composite thermal conductivity and crucial in composite design [18–25], different volumetric ratios for
2.1.2. Mechanically pumped cooling loop (MPCL) With the rapid progress of electronics in numerous electromechanical systems and the consequent ongoing demand for size and weight reduction, a compatible thermal management solution is inevitably a challenge. A mechanically pumped cooling loop (MPCL) is a key theme in developing a reliable alternative for conventional aluminum and/or copper cooling units. 2
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Fig. 1. (a) Composites material preparation, molding steps and (b) Developed pair molds illustrations.
thickness were used to create a square tunnel of 15 cm internal length. Two U-shaped-finned-pipe-heaters were installed at the air-inlet hatch and were connected to a digital-controller (OKE-2002, im Tech) to set the inlet-air temperature constant. To achieve a uniform airstream, the air passed through a honeycomb straightener before approaching the fins. After the heat sink, the air-duct cross-section was changed to a circular shape with an inner diameter of 15.4 cm (polyethylene pipe), using an adaptor, such that a high accuracy thermal mass flow meter (Fastflo 620G, Sierra Instruments, Inc. Netherland) could be used. Since the mass flow meter measurement could be influenced by the air stream profile, adequate distances from the adaptor and suction were adhered to avoid a non-uniform stream. The air flow was controlled by a digital inverter (Hyundai N700E Inventor, Ltd., Korea) connected to a suction pump (HSB-SSF28T, Hanil Co.,Ltd., Korea) at the end of the air duct. The air-duct details, dimensions, and MPCL are shown in Fig. 3.
Herein, a miniature MPCL (shown in Fig. 3b) was established to assess the feasibility of the PMC-fabricated heat sink application. Due to its flexibility, strength, corrosion, fouling resistance, and high thermal conductivity, copper was chosen for piping. The water inlet was set constant using a water-bath (Refrigerated/Heating Circulators RCH3015, Ltd CP Ti, Korea) and the temperature was measured at four points: the inlet and outlet of the heat sink for both the water and air sides, using resistance temperature detectors (RTDs). The water mass flow was measured by an electromagnetic flow meter (ModMAG M2000, Czech Republic) placed between the water bath and heat sink water-inlet.
2.1.3. Air duct (wind tunnel) The air duct is vital for supplying a steady and uniform flow over the fins to approach the PHS. Owing to ease in the manufacturing and the suitability for fitting various heat sinks and RTDs, acrylic sheets of 1 cm 3
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̇ = ṁ air Cp, air (Tair , out − Tair , in ) Qair
(2)
Q̇ water = ṁ water Cp, water (Twater , out − Twater , in)
(3)
̇ ) and overall heat Further, the average heat transfer rate (Qave transfer coefficient (U) based on the inlet and outlet temperatures and mass flow rates of both water and air sides were calculated, using the logarithmic mean temperature difference (LMTD) method [27], as outlined by the following equations:
̇ + Q̇ water Qair 2
̇ = Qave
(4)
̇ = UAF ΔTlm, cf Qave
(5)
In which
ΔTlm, cf =
(Twater . out − Tair , in ) − (Twater , in − Tair , out ) ln [(Twater ,out − Tair , in )/(Twater , in − Tair ,out )]
(6)
where ΔTlm, cf is the LMTD for counter flow arrangement, and F is the non-dimensional correction factor. The corresponding correction factors were determined to be approximately one (F ≈ 1) based on the correction factor chart for the cross-flow heat exchanger [27]. Gnielinki’s recommended equation was chosen corresponding to the experimental condition (turbulent and transition region) for the Nusselt number (Eq.7)
(f /2)(Reb−1000) Prb
Nub =
2
1
1 + 12.7(f /2) 2 (Prb3 − 1)
(7)
where Pr is the Prandtl number, and f is the Fanning friction factor that can be obtained by equations below:
f = (1.58LnReb−3.28)−2
Reb =
Rea =
cp μ K
(9)
ρa umax dae μa
(10)
-dae : Hydraulic diameter of single flow channel Because the modified Wilson plot method constitutes a suitable technique to estimate the convection coefficients in a variety of convective heat transfer processes without requiring direct measurement of the surface temperature, this method was applied for each heat sink to obtain the air-side heat transfer resistance from the overall conductance [28]. Data were obtained for water flow rates between 0.005 and 0.015 kg/s inside the copper pipe as the hot-side fluid, and constant air flow rate and temperature of 0.096 kg/s and 21 °C as the cold-side fluid. The total thermal resistance (Rtotal) of the heat exchanger, as shown in Eq. (11), is consists of the inside thermal resistance (Rin), tube wall thermal resistance (Rwall), thermal contact resistance (Rcontact), and outside thermal resistance (Rout).
2.2. Experimental uncertainty Table 1 shows the summary of uncertainties of the measured and calculated data. The measured uncertainties were obtained from the instrument manufactures and for the calculated type, Engineering Equation Solver (EES) software was used. The method for determining the experimental uncertainties can be summarized in Eq. (1). More details can be found in Ref. [26].
∂R ∂R ∂R σx1)2 + ( σx2 )2 +⋯+( σx n )2 ∂x1 ∂x2 ∂x n
Prb =
Likewise, the Reynolds numbers for the airside was calculated from Eq. (10):
Fig. 2. (a) Geometric dimension of the fin and (b) illustration of the actual fabricated PHS (Cu/MWCNT-HDPE).
σR = ± (
4ṁ , πdi μ
(8)
(1)
where
R = f (x1, x2 , ⋯x n )
Rtotal = Rin + Rwall + R contact + Rout
In the equation above, the resultant R (which can be any parameter) is function of x1, x2, …, xn, and σR represents the uncertainty of the resultant R.
(11)
The total thermal resistance of the heat exchanger is related to the overall conductance through the following expression:
Rtotal = 2.3. Data reduction
Rtotal
Since the thermal performance is the primary objectives of this research, parameters such as thermal resistance, total heat transfer rate, and overall heat transfer coefficient were calculated for all the heat sinks. The air and water heat transfer rates were calculated using Eqs. (2) and (3) respectively.
1 F ΔTLMCF = ̇ UA Qave
1 1 = + Rwall + R contact + hi Ai ηo ho Ao
(12)
(13)
Since the thermal resistance of tube due to copper high thermal conductivity and its thickness are much smaller than the inside and outside resistance of the pipe, they are negligible. Therefore the overall 4
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Fig. 3. Schematic illustration of experimental setup for (a) Air duct, and (b) mechanically pumped cooling loop. Table 1 Uncertainties of experimental data. Measured variable
Calculated variable
Variable
Uncertainty
Variable
Relative Uncertainty
Temperature
± 0.1 °C
± 1.8%
Water mass flow rate Air mass flow rate Length tr , sr Diameter
± 0.1 L/min
Reynolds number (air) Reynolds number (water)
± 0.4%
3
± 0.3 M /min ± 0.001 m ± 0.00001 m ± 0.00001 m
Fig. 5. Fin and heater installation for thermography process.
thermal resistance is expressed as Eq. (14).
1 1 1 ≈ + R contact + UA hi Ai ηo ho Ao Fig. 4. Plot used to obtain the air-side heat resistance (modified Wilson plot).
Nui =
5
hi × di ki
(14)
(15)
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Fig. 6. Schematic of filler alignment in hot-press molding process (most particles are aligned to flow direction inside the mold).
Fig. 7. Thermal conductivity for various filler volume percentage for (a) MWCNT/HDPE and (b) Cu+MWCNT/HDPE composite.
Through-plane
In-plane
Nusselts number. Therefore, the plot in Fig. 4 that shows both the water-side and air-side thermal resistances is applicable. The important aspects regarding this chart are the line extrapolation and intersection point with the vertical axes (1/U) which represent the neat air-side thermal resistance.
2.7 1.717
205 1.317
205 2.574
2.4. Thermographic analysis of materials
0.952
0.57
0.57
1.043
0.644
0.71
Table 2 Selected materials properties near room temperature. Material
Density (g/ cm3)
Aluminum (Al) Cu/MWCNT-HDPE (CMH) composite High density polyethylene (HDPE) MWCNT-HDPE (MH) composite
Thermal conductivity (W/m∙K)
Because the contact and fins thermal resistance were assumed to be constant value for all cases, therefore, by maintaining the air temperature and mass flow rate steady and changing the water mass flow, all changes in the total thermal resistance value will be caused by water-side heat transfer coefficient (hi). Likewise, the heat transfer coefficient is related to the Nusselts number (Nu) by Eq. (15). By considering Eqs. (14) and (15), Eqs. (16)–(19) can be achieved.
For a better comparison among four different materials that were used as heat sinks, the thermographic method was applied to perceive heat diffusion inside the fins. For simplifying the process, a cartridge heater of diameter 8 mm was passed through the center of the squareshaped fin of length 80 mm and thickness 2 mm as shown in Fig. 5. To reduce the radiation effect on the results and for equal conditions for all the cases, a thin graphite layer (GRAPHIT 33, CRC Industries, Deutschland) was sprayed on the fin surfaces. Then each fin was heated gradually from room temperature (26 °C) until it was stabilized at 65 °C. Thermal pictures (IR) were captured from fin surfaces at steady state using a thermal camera (FLIR T620, FLIR Systems, Inc., Sweden).
1 d 1 1 = A (( i · ) + (R contact + )) U ki Ai Nui ηo ho Ao
3. Results and discussion
α = A·
di = constant ki Ai
β = A (Rcontact +
(18), (19) →
1 ) = constant ηo ho Ao
1 α = +β U Nui
(16)
3.1. Thermal conductivity results (17) To maximize the heat rejection from tubes using molded fins, the inplane thermal conductivity is more important than that of the throughplane. From fluid mechanics, the shear stress profile for viscoelastic substance in a channel changes through the transvers axis. Thus, the stress value for the fluid layer next to the wall is zero and increases by getting far from wall until it reaches to maximum level at the channel center line. This shear stress difference between fluid layers renders the
(18)
(19)
As shown, the left side of Eq. (19) is inversely related to the water 6
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After 30 min
Steady situation
HDPE
MWCNT/HDPE
Cu+MWCNT/HDPE
Aluminum
After 15 min
(a)
(b) Fig. 8. Heat diffusion results for fins, using time lapse thermography (a) for range of 25–65 °C and (b) 26–38 °C.
filler particles rotation and alignment parallel to the flow direction. Therefore, the shear stress due to pressing process prior to the conductivity measurement causes the particles to be aligned perpendicularly to the samples measuring plane [29] (Fig. 6). Since the copper particles are larger than MWCNTs (almost 20 times larger), they were affected more than the MWCNTs and consequently aligned better to the flow path. The C-Therm TCi thermal conductivity analyzer (C-Therm Technologies Ltd., Canada) was used to thermally characterize the
conductive polymeric samples. This method is based on the difference in voltage in the modified transient plane source (MTPS) probe induced by the temperature gradient during the sampling time. Both the measured in-plane and through-plane thermal conductivity of two new composite materials for various fillers volumetric ratios can be seen in Fig. 7. An improvement of up to 22.4% and 245.5% for the in-plane and 11.03% and 88.9% for the through-plane thermal conductivity was achieved for MWCNT/HDPE and MWCNT+ Cu/HDPE, respectively.
7
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Fig. 11. Modified Wilson plot results for the heatsinks.
Fig. 9. Experimental results for average heat transfer rate under various Reynolds numbers on air-side.
Fig. 12. Numerical simulation results for fin radial temperature distribution. Fig. 10. Experimental results of overall heat transfer coefficient changes (U) for various Reynolds number.
still demonstrates the highest heat rejection rate.
The higher thermal conductivity of the CMH is attributed to better networks with larger mean free paths for the ballistic motion of phonons through the composite. Table 2 shows the final results including density and thermal conductivity of each material near the room temperature for this research.
3.3. Thermal performance The average heat transfer rate and overall heat transfer coefficient versus various Reynolds number [3,30,31] in variety of 2000 up to 100,000 for the air side were calculated and are shown in Figs. 9 and 10, respectively. For the lowest Reynolds number (around 2000), the heat transfer rate gap between the best (aluminum) and worst (HDPE) cases was approximately 0.027 kW. By increasing the air flow rate, the heat transfer rate number also raised up significantly for aluminum and with a smooth slop for three other cases (due to increasing of heat transfer coefficient). After the Reynolds number reached to around 80,000, heat transfer rate becames constant for all four heat sinks and difference between the best and worst cases was 0.177 kW. A similar trend was also observed for overall heat transfer coefficient. The convective heat transfer is always the result of mass flow rate, temperature difference between fluid and surface, and heat transfer area. Since the mass flow, fluid temperature, and area were the same for all cases, therefore the only effective parameter was the fin efficiency, which was directly related to the material thermal conductivity. An enhancement of up to 23.4% and 52.4% in average heat transfer rate, and 147% and
3.2. Thermographic analysis of materials Because thermography facilitates the visual comparison of heat diffusion through materials, square-shaped manufactured fins were heated using a cartridge heater until they reached to steady state (Fig. 5), and then the thermal pictures (IR) were taken from fin surfaces. Fig. 8a shows the significant temperature difference between the composite cases and aluminum. Therefore, another temperature range was set on the thermal camera for a better comparison among the polymeric cases (Fig. 8b). The results of HDPE and MWCNT/HDPE were extremely close to each other. At the steady state, CMH exhibits a higher heat rejection compared with other polymeric fins. This is due to the higher thermal conductivity of the CMH and better interface between the copper particles and polymeric matrix. Aluminum, however, 8
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Fig. 13. Simulation domain and boundary condition.
Fig. 14. Single fin simulation results for (a) HDPE, (b) MWCNT/HDPE, (c) Cu+MWCNT/HDPE, and d) aluminum.
between the through-plane and in-plane thermal conductivities, simulation was performed in a three-dimensional (3D) domain. The steadystate solver and energy model were chosen for natural conduction heat transfer. Following the experimental conditions, the heater and room temperature were set at 65 °C and 26 °C respectively. The radial heat diffusion and fins average temperature results are shown in Fig. 12, showing similarity with the experimental results shown in Fig. 8, and represent the best and worst heat diffusion for Aluminum and pure HDPE respectively. Owing to the complex geometry of the heat sink, the simulation was done for a single fin for each case. The mesh properties for the 3D domain was 1,661,992 nodes and the maximum aspect ratio of 74.6. The air flow with constant thermal properties were assumed to be a steady, incompressible and non-viscous dissipation throughout the computational domain. The continuity, momentum and energy equations for the RNG κ − ε model were set for the water and air turbulent conditions. The simulation domain and final results are shown in
287% in overall heat transfer coefficient were achieved for MWCNT/ HDPE and MWCNT+Cu/HDPE, respectively, compared to pure HDPE. As mentioned earlier, the modified Wilson plot method was considered to determine the air-side overall heat resistance. Hence, the airside mass flow and temperature were maintained constant while the water mass flow was varied. Based on the modified Wilson plot method, the intersection point of trend lines in Fig. 11 with the vertical axis represents the total thermal resistance for the air-side. Because all the conditions such as heat sink geometry and thermal contact resistance (between pipes and fins) were the same for all cases, the heat sinks material thermal conductivity would be regarded as the only impact parameter on the overall air side thermal resistance. 3.4. Simulation The ANSYS Fluent software was used to obtain more details regarding the heat dispersion inside the fins. Due to the difference 9
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Figs. 13 and 14.
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4. Conclusions 1. Experimental results indicated that improvement in both in-plane and through-plane thermal conductivity could be achieved by employing high thermally conductive filler in polymer matrix. Enhancements of up to of 22.4% and 245.5% for in-plane and 11.03% and 88.9% for through-plane thermal conductivities, as well as 23.4% and 52.4% increase in heat transfer rate were achieved for MWCNT/HDPE and MWCNT+Cu/HDPE respectively. 2. Heat transfer rate difference for all four cases at small air Reynolds numbers (close to natural convection) was much smaller than that of forced convection. In other words, for low air Reynolds numbers, the overall heat transfer coefficient ratio of aluminum over other cases were 4.07, 3.97, and 1.9 for CMH, MH, and HDPE respectively, and 5, 3.06, and, 2.27 respectively for high Reynolds numbers. 3. Notwithstanding the MWCNTs shows higher thermal conductivity than copper (Cu), a better interface between copper particles and polymer chain resulted in a superior path for heat to disperse inside the composite. 4. As copper particles are approximately 20 times larger than the MWCNTs, fluid flow in the molding process affected the copper particles alignment significantly. Consequently, the in-plane over through-plane thermal conductivity ratios were calculated as 1.95 and 1.1 for CMP and MH composites respectively. Acknowledgment This work was supported by the Ministry of Trade, Industry, and Energy (MOTIE) of the Republic of Korea (P0002131) and by the National Research Foundation of Korea funded by the Korean Government (grant number 2019R1A2C2010607). Appendix A. Supplementary material Supplementary data to this article can be found online at https:// doi.org/10.1016/j.applthermaleng.2019.113823. References [1] K.C. Otiaba, N.N. Ekere, E.H. Amalu, R.S. Bhatti, S. Mallik, Thermal Management materials for electronic control unit: trends, processing technology and R and D challenges, Adv. Mater. Res. 367 (2011) 301–307, https://doi.org/10.4028/www. scientific.net/AMR.367.301. [2] X. Chen, Y. Su, D. Reay, S. Riffat, Recent research developments in polymer heat exchangers – a review, Renew. Sustain. Energy Rev. 60 (2016) 1367–1386, https:// doi.org/10.1016/j.rser.2016.03.024. [3] M.A. Arie, A.H. Shooshtari, R. Tiwari, S.V. Dessiatoun, M.M. Ohadi, J.M. Pearce, Experimental characterization of heat transfer in an additively manufactured polymer heat exchanger, f Eng. 113 (2017) 575–584, https://doi.org/10.1016/j. applthermaleng.2016.11.030. [4] R. Trojanowski, T. Butcher, M. Worek, G. Wei, Polymer heat exchanger design for condensing boiler applications, Appl. Therm. Eng. 103 (2016) 150–158, https:// doi.org/10.1016/j.applthermaleng.2016.03.004. [5] I. Astrouski, M. Raudensky, I. Krásny, Polymeric hollow fiber heat exchanger as an automotive radiator, Appl. Therm. Eng. 108 (2016) 798–803, https://doi.org/10. 1016/j.applthermaleng.2016.07.181. [6] A. Roja, J. Hussain, A.A. Alahyari, S.A. Eastman, C. Thibaud-erkey, S. Johnston, M.J. Sobkowicz, Review of polymers for heat exchanger applications: factors concerning thermal conductivity, Appl. Therm. Eng. 113 (2017) 1118–1127, https:// doi.org/10.1016/j.applthermaleng.2016.11.041. [7] I. Astrouski, M. Dohnal, M. Raudensky, Intensification of heat transfer of polymeric hollow fiber heat exchangers by chaotisation, Appl. Therm. Eng. 113 (2017) 632–638, https://doi.org/10.1016/j.applthermaleng.2016.11.038. [8] J.G. Cevallos, A.E. Bergles, A. Bar-Cohen, P. Rodgers, S.K. Gupta, Polymer heat exchangers—history, opportunities, and challenges, Heat Transf. Eng. 33 (2012) 1075–1093, https://doi.org/10.1080/01457632.2012.663654.
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