- Email: [email protected]
An integrated heat pipe coupling the vapor chamber and two cylindrical heat pipes with high anti-gravity thermal performance
Applied Thermal Engineering 159 (2019) 113816
Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Research Paper
An integrated heat pipe coupling the vapor chamber and two cylindrical heat pipes with high anti-gravity thermal performance
T
⁎
Huawei Wanga, Pengfei Baia,b, , Honglin Zhoua, Reinder Coehoornb, Nan Lic, Hua Liaoe, Guofu Zhoua,b,c,d a Guangdong Provincial Key Laboratory of Optical Information Materials and Technology & Institute of Electronic Paper Displays, South China Academy of Advanced Optoelectronics, South China Normal University, Guangzhou 510006, PR China b National Center for International Research on Green Optoelectronics, South China Normal University, Guangzhou 510006, PR China c Shenzhen Guohua Optoelectronics Tech. Co. Ltd., Shenzhen 518110, PR China d Academy of Shenzhen Guohua Optoelectronics, Shenzhen 518110, PR China e Institute of Solar Energy, Yunnan Normal University, Kunming 650500, PR China
H I GH L IG H T S
integrated heat pipe is designed, built and tested in this study. • An performance of the heat pipe under different conditions was tested. • The lowest thermal resistance under anti-gravity condition is 0.157 K/W. • The critical heating load is affected by the filling ratio significantly. • The • The heat pipe has good performance and can adapt to many inclination angles.
A R T I C LE I N FO
A B S T R A C T
Keywords: Integrated Heat pipe Wick structure Anti-gravity
An integrated heat pipe (IHP) coupling the vapor chamber and two cylindrical heat pipes is proposed in this study. Cylindrical heat pipes were connected with the top surface of the vapor chamber to extend the area of the condenser. The copper powder rings and graded pore-size wick are used to shorten the circulation path of working fluid and strengthen the capillary pressure and wick permeability of the wick structure. The effect of various filling ratios (30%, 40%, 60%, and 80%) and different inclination angles (0°, 90°, and 180°) have been investigated experimentally. The results showed that the filling ratio has a significant impact on the heat transfer performance of IHP, and the high filling ratio IHP shows better thermal properties. Specifically, the maximum temperature is lower than 83 °C at the heat load of 150 W when the filling ratio is 60%. Moreover, the integrated heat pipe shows better anti-gravity thermal performance than a conventional heat pipe. The lowest anti-gravity thermal resistance is 0.157 K/W with a heat load of 80 W. As compared with other heat pipes reported in this literature, the IHP has a good thermal performance and can adapt to many inclination angles.
1. Introduction In recent years, the integration and power density of electronic devices is continuously becoming intensive. This trend results in a high heat flux problem and influences the lifetime and reliability of electronic significantly [1–3]. Therefore, novel components with excellent heat dissipation ability have been developed to address the problem. Heat pipes, as one of the most efficient two-phase heat transfer
technologies available [4]. They have the advantages of high heat dissipation, high thermal uniformity, and low heat transfer temperature difference. The working fluid in the heat pipe is heated by the external heat source and then evaporates and carries heat from heating source to cold source driven by the pressure difference. After vapor reaches to the cold source and condenses into liquid, the working fluid enters into wick structure and flows back to the heating section [5], completing a whole heat transfer cycle.
⁎ Corresponding author at: Guangdong Provincial Key Laboratory of Optical Information Materials and Technology & Institute of Electronic Paper Displays, South China Academy of Advanced Optoelectronics, South China Normal University, Guangzhou 510006, PR China. E-mail address: [email protected] (P. Bai).
https://doi.org/10.1016/j.applthermaleng.2019.113816 Received 11 November 2018; Received in revised form 17 May 2019; Accepted 23 May 2019 Available online 24 May 2019 1359-4311/ © 2019 Published by Elsevier Ltd.
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
φ ρ
Nomenclature I q Q R T U V
current [A] heat flux [W/cm2] heat load [W] thermal resistance [K/W] temperature [K] voltage [V] volume [L]
Subscripts c e f p w E,c hl
Greek symbols ε
filling ratio density [kg/m3]
condenser evaporator working fluid pore wick center of evaporator heat loss
porosity
of the evaporator is below 84 °C under a heat load of 90 W. Moreover, Li et al. [18] studied the effects of copper powder size on heat transfer performance of heat pipe. The results show that, for single-powder structure, 75–100 μm is the optimal size, and 25–50 μm is the optimal powder size for continuous step-graded structure. Some analytical expressions were derived to allow the designers to determine the amount of working fluid necessary for a given set or range of operating conditions, tube geometry, and working fluid properties [19–22]. In some studies the researchers manufactured the heat pipes with a series of filling ratios, they selected the optimum filling ratio according to the experimental results [14–16,23]. In this study, we first calculate the pore volume (2.94 mL) in the wick structure. Then we choose a series of filling ratios to study the effect of the filling amount of working fluid on the heat transfer performance in antigravity condition. Based on the above studies, the present work designs an integrated heat pipe (IHP), which couples a vapor chamber and two cylindrical heat pipes. In terms of the wick structure, sintered copper powder rings are introduced to shorten the flow path. In addition, this work uses sintered copper powders with different powder sizes, the composite wick within the IHP provides enough capillary force and high wick permeability to make the heat pipe operate in many inclination conditions. In the next few sections, we experimentally investigated the effects of heating rates, filling ratios and inclination angles on the thermal performance of IHP.
In heat pipes, wick structure is the core component determining the heat performance of the heat pipe. Sintered powder wick is the most effective wick structure [6]. Xuan et al. [7] made a layer of copper powder sintered on the heated surface of a flat plate heat pipe and investigated the transient behaviors. Results indicated the copper powder layer strengthens the evaporation process and enhance the response speed significantly. The sintered copper powder with high density can markedly enhance the evaporation process, but the small pore diameter and surface tension diminish the vapor flow rate in evaporator. Zhang et al. [8] experimentally verified compressed metal foam has multi-scale pores, enabling it to pump a significant amount of liquid to a long distance. Wong et al. [9] experimentally studied loosely-sintered copper-powder evaporators and measured the evaporation resistance. A glass plate was used to make the evaporation process visualized. The evaporation resistance decreases with the increase of heat flux until partial dryout occurred. The minimum evaporation resistance measured was about 0.08–0.09 W cm2/K. Tang et al. [10] proposed a multi-artery vapor chamber, copper powder rings were introduced and provide short paths for the working fluid to flow back to the evaporator. This method made the vapor chamber have a fast response speed, a low thermal resistance of 0.033 K/W. Ji et al. [11] put forward a vapor chamber with many copper foam bars inserted into the channels of the condenser. This strategy eliminates the contact thermal resistance between the surface of the vapor chamber and fins. The minimum thermal resistance is 0.03 K/W and the maximum heat flux is above 445 W/cm2 without dryout phenomenon. Gravity also influences the thermal performance of heat pipe, in the terrestrial application of the heat pipe, it is normal to have the heat source above the cold source [12,13]. Local dryout under anti-gravity condition usually occurs for conventional heat pipes. Many researchers have studied the effects of wick structure and the filling amount of working fluid on the heat transfer performance in anti-gravity condition. Trijo et al. [14] analyzed the thermal performance of cylindrical and flattened heat pipes at five inclination angles. The lowest thermal resistance of 0.46 K/W was achieved when the inclination angle is −45°. The thermal performance of a pulsating heat pipe was studied by Deng et al. [15]. The heat pipe was used for waste heat recovery from the high-temperature exhaust gas under anti-gravity condition. The start-up process of the heat pipe takes a long time and the thermal performance heavily depends on filling ratio. When the filling ratio increases under the same heat load for the anti-gravity heat pipes, the thermal resistance decreases firstly (from 50% to 70%) and then increases (from 70% to 80%), suggesting that the optimal filling ratio for the heat transfer is 70%. Yao et al. [16] designed and tested an antigravity vapor chamber, which has a tree-shaped groove evaporator surface and an 80/200 mesh wick to strengthen capillary pressure. The minimum anti-gravity thermal resistance is 0.215 K/W when the filling ratio is 60%. Tang et al. [17] developed an anti-gravity loop heat pipe with a continuous graded pore-size wick structure. The overall thermal resistance is nearly stable at 0.15 K/W, and the maximum temperature
2. IHP description The IHP primarily includes a vapor chamber, two cylindrical heat pipes, and an exhaust tube (Fig. 1). Compared with the conventional heat pipe/vapor chamber, the IHP has the following significant structural differences. Firstly, it has a similar three-dimensional shape. Two cylindrical heat pipes were carefully welded on the top plate of the vapor chamber. The copper powder layers of the two cylindrical heat pipes and the sintered copper powder layers of the vapor chamber are connected to each other, therefore, there is no contact thermal resistance between them. Second, the copper powder rings are introduced to shorten the circulation path of the working fluid. The sintered copper powder rings were mounted around the copper columns for recycling the liquid from the condenser to evaporator wick. Third, the graded pore-size wick consists of copper powders with two kinds of powder size. The two kinds of copper powders with different powder size were marked by blue color1 and coppery color in Fig. 1(b). The size of copper powder particle varies from 124 µm to 200 µm, which were sintered on the inner surface of the two cylindrical heat pipes and the top plate of the vapor chamber. The diameters of copper powder particle, sintered on the copper powder rings and the bottom surface of the vapor 1 For interpretation of color in Fig. 1, the reader is referred to the web version of this article.
2
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
Fig. 1. The structure of IHP. (a) Photograph of IHP and (b) Sectional view.
allowed for full decomposition of the impurities. Then, the sintering temperature was raised to 1000 °C at a heating rate of 6–7 °C/min and held at this temperature for 4 h. After that, the furnace was cooled to 100 °C with a cooling rate of 7–8 °C/min. Next, the two copper tubes were properly set on the top plate and solder paste was applied to their contact surfaces. Copper powders with powder size of 124–200 µm were placed on the top plate and fixed by a graphite die, and the copper powders were sintered in the same method as above. After that, the sintered copper layers connect each other between the two cylindrical
chamber, vary from 74 µm to 124 µm. The complete parameters of the IHP are specified in Table 1. Fig. 2 shows the flow chart of the manufacturing process of wick structure. Firstly, copper powders with powder size of 124–200 µm were placed on the inner surface of the two cylindrical copper pipes fixed by graphite core rod, and then they were sintered at a temperature of 1000 °C for four hours under hydrogen reducing atmosphere. The sintering process curve for sintered copper powder wick is presented in Fig. 3. In the pre-sintering stage, a hold time of 30 min at 800 °C 3
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
through the copper tube. De-ionized water was chosen to be working fluid because of its high latent heat, large surface tension, and nontoxicity. Finally, the exhaust tube was sealed by crimping press and spot welding. The porosity of the wick structure can be defined as the ratio of the pore volume Vpor to the wick structure volume Vw or
Table 1 Specification of the IHP. parameters
Dimension/material
Vapor chamber Length Width Height Side wall thickness Top plate thickness Bottom plate thickness Wick Wick thickness Bottom plate porosity Top plate porosity
50 mm 45 mm 5 mm 0.8 mm 0.5 mm 0.8 mm Sintered copper powder 0.5 mm 65% 50%
ε=
5.4 mm 6 mm 100 mm Sintered copper powder 0.9 mm 50%
Copper columns Height Diameter
2.8 mm 1.2 mm
Copper powder rings Inner diameter Outer diameter Wick Porosity
1.2 mm 3.6 mm Sintered copper powder 65%
Vw
=1−
m ρVw
(1)
where m is mass, ρ is the density of the used copper powder. The filling ratio is defined as the volume percentage of the pore space of the wick structure shared by the working fluid, the filling ratio is calculated as
φ=
Cylindrical heat pipe Inner diameter Outer diameter Length Wick Wick thickness Porosity
Vpor
Vf Vpor
=
Vf εVw
(2)
where Vl is the volume of working fluid. And a series of filling ratios of 30%, 40%, 60%, and 80% were selected for the experimental test. For the 80% filling ratio, the dosage of water was calculated as 2.39 g. In turn, 1.78 g, 1.21 g, and 0.87 g water were used in the IHP with 60%, 40%, and 30% filling ratios respectively. 3. Experimental methods 3.1. Experimental setup A test rig was designed to investigate the transient and stable behavior of the IHP under the effect of various parameters such as inclination angle and filling ratio at different heat load. The schematic diagram of the experimental system is illustrated in Fig. 5. The testing system consists of three modules: heating module, cooling module, and temperature data collecting module. The heating module includes a DC power supplier, a copper heating block and two cartridge heaters. The heating block contacts the bottom surface of IHP, whose active heating area is 12 mm × 12 mm. To minimize environmental heat losses, thermal insulation blocks were used. In addition, a layer of thermal grease fills the gap between IHP and the heating block to decrease the contact thermal resistance. The cooling module includes a thermostatic water tank (Julabo, F25-ME), a flowmeter, and two symmetrical copper cooling blocks. The cooling blocks contact with IHP closely and thermal grease was used on the contact surface to decrease the contact thermal resistance. The cooling water flow rate is 2.2 L/min with a temperature of 30 °C. The temperature data collecting module consists of ten T-type thermocouples, an Agilent 34970A Data
heat pipes and the top plate. Then, seven powder rings and the sintered powder layer on the bottom plate were sintered separately in the same method as above. Finally, seven copper powder rings were set around the copper columns. The top plate along with the bottom plate were welded by solder paste. The SEM images of the sintered copper powder wick in the bottom surface of the vapor chamber was shown in Fig. 4. There are many large pores and small pores inside the wick. The wick within the IHP is composite wick and it can increase the overall permeability and capillary forces of the wick [24–26]. The working fluid charging procedure is as follows: A hole in the side wall of the vapor chamber was drilled out and an exhaust tube was inserted into this hole. The air inside the IHP was pumped away by a vacuum pump to make the inside pressure of IHP at about 0.06 Torr (7.99 Pa). Working fluid was filled into the IHP at negative pressure
Fig. 2. Flow diagram of the manufacturing process of wick structure. 4
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
Fig. 3. Temperature-time curve for sintering process.
Fig. 4. SEM images of the pores inside the bottom surface wick.
power is switched off. The same procedure is repeated for every heat load, filling ratio, inclination angle.
Acquisition. Fig. 5(b) shows the different inclination configurations, the IHP was mounted to a plastic block that allowed tilt configuration. The temperature measurement points are shown in Fig. 6. Temperatures are recorded automatically by the data collector at 1 s intervals. The heat load started at 10 W and increased 10 W every time. The DC power supplier provides electrical power to a steel block with two cartridge heaters as the heating source. To minimize the environmental heat losses, the adiabatic area was wrapped by a layer of 5 mm-thick adiabatic foam made of styrene butadiene rubber (thermal conductivity is 0.19 W/(m·K)). After the test rig was set up, it is ready to perform the experiment. Firstly, the water bath is set at a desired cooling temperature. Then, the rotameter is adjusted to 2.2 L/min. Before the power is supplied to the heater, enough time is provided to ensure that all thermocouples readings reach approximately the value of adjusted cooling temperature which is another proof of thermocouples accuracy in temperature measurement. Then, the power is supplied to the electrical heater by adjusting the variable transformer to a certain value, equivalent to the desired heat input needed to apply on the evaporator as a heat source. This heat input can be obtained by multiplying the voltage times the current as well as the reading of the wattmeter. The next step is to monitor the variation of the temperature readings with time. After all temperature readings reach the steady state, the data is saved and the
3.2. Data reduction and uncertainty analysis The thermal resistance RI of the IHP is defined as the ratio of the temperature difference between evaporator and the condensing resistance of this system. It is expressed as
RI =
Te − Tc Q
Te =
1 5
∑ Ti
1 6
∑ Ti
Tc =
(3)
5 1
(4)
16 11
(5)
where Q is the value of the heat load, Te is the average temperature of the evaporator, Tc is the average temperature of the condenser respectively. Based on the measurements of different variables. The uncertainties in the measurement of the heat load, heat flux, and thermal resistance were calculated as: 5
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
Fig. 5. The experimental apparatus (a) and different inclination configurations (b).
ΔQ = Q Δq = q
ΔR = R
The thermal resistance of IHP with different filling ratios is shown in Fig. 8. It is clear that the high filling ratio gives the low evaporator temperature. For the same heat load, the heat pipe with φ = 30% almost has the highest thermal resistance except for under low heat load. This is because there is not enough liquid supply to the evaporator. When heat load is 10 W, the thermal resistance is 0.537 K/W for φ = 30% which is lower than 0.555 K/W for φ = 80%. This is due to the thick water film in evaporator, decreasing the evaporation rate at a low heat load. When filling ratio increased from 60% to 80%, the evaporator temperature increased at the same heat load. This can be explained that a slight excess charge will typically result in the formation of a small puddle in the lower portion of the heat pipe and result in a small reduction in performance. In this study, the filling ratio of 60% is necessary for the IHP, and when filling ratio increases to 80%, there exists a slight overcharge in the IHP, and some working liquid may accumulate in the evaporator and cause a partial blockage. This will lead to a high evaporator temperature and thermal resistance. The results can be cross-checked with the well-known conclusion made by G.P. Peterson [20]. Most curves of thermal resistance first go down and then go up. The minimum thermal resistance of 0.142 K/W was reached when the heat load is 130 W. The main reason is the thickness of the liquid film in evaporator varies with filling ratio and heat load. When heat load is low, the evaporation rate is low in the evaporator, and the liquid film is relatively thick, which lead to a large evaporation resistance. With the increase of heat load, the evaporation rate will increase, resulting in a thinner film in evaporator and a lower evaporation resistance. Moreover, a large evaporation rate will increase the
2
ΔU 2 ΔI 2 ΔQhl ⎞ ⎛ ⎞ +⎛ ⎞ +⎛ U ⎝ ⎠ ⎝ I ⎠ ⎝ Qhl ⎠ ⎜
2
⎟
2
⎛ ΔQ ⎞ + ⎛ ΔA ⎞ ⎝ A ⎠ ⎝ Q ⎠
⎜
⎟
2
(6b) 2
⎛ ΔQ ⎞ + ⎛ Δ(ΔT ) ⎞ ⎝ ΔT ⎠ ⎝ Q ⎠
⎜
(6a)
⎟
(6c)
The accuracy of the thermocouple was ± 0.5 °C, the accuracy of the flowmeter was ± 0.088 L/min, the input power was supplied by a DC power supplier with 0.1% voltage accuracy and 0.5% current accuracy. The uncertainty of the heat loss was found to be less than 3% and the uncertainty of the heating area was 0.67%. Therefore, the maximum uncertainties of the heat load, heat flux, and the thermal resistance were 3.04%, 3.11%, and 6.75%. 4. Results and discussion 4.1. Evaporator temperature and thermal resistance of IHP Fig. 7 shows the center temperature of the evaporator with the applied heat load. The value of Te,c increases with the increase of heat load, but the rate of temperature rise varies with the filling ratio. When the filling ratio is low, the temperature rising speed becomes faster and faster with the increase of heat load, while when the filling ratio is high, the rising speed of temperature is relatively stable. 6
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
Fig. 6. The measurement points of the IHP and heater.
Fig. 7. Center temperature of the evaporator for different filling ratios.
Fig. 8. Thermal resistance of IHP with different filling ratios.
condensation rate simultaneously. When the heat load exceeds a critical value, the thermal resistance will increase due to partial dryout in the evaporator. The lower the filling ratio, the more obvious the increase of thermal resistance.
heated, nucleate boiling firstly occurs in the evaporator, then the nucleate boiling becomes more and more intense with the increase of heat load. When the heat load reaches Qc, the boiling regime in the evaporator changes from nucleate boiling to transition boiling [27]. In this case, many vapor blankets cover on boiling surfaces, the dried surfaces have no more nucleate boiling chances. Thus, the thermal resistance RI keeps increasing as the curves in Fig. 8 for φ = 30%. Table 2 shows Qc for four different filling ratios under three inclination angles.
4.2. The critical heat load of IHP As shown in Fig. 8, the thermal resistance decreases with the increase of heat load when the heat load is low. Especially when the heat load was applied from 10 W to 20 W, there is a steep decrease in the thermal resistance. The phenomena may be explained as the startup of the IHP. After that, the thermal resistance continues to decrease until the critical heat load Qc was achieved. This is because when IHP is
4.3. Anti-gravity thermal performance 4.3.1. Mechanism of anti-gravity thermal performance Gravity has a great impact on the vapor-liquid distribution and flow. 7
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
Table 2 Critical heat load of IHP. Charging ratio
Qc for θ = 0° (W)
Qc for θ = 90° (W)
Qc for θ = 180° (W)
30% 40% 60% 80%
30 60 130 130
20 30 130 > 140
10 40 60 80
Therefore, it is necessary to study the effect of gravity on the performance of IHP. Fig. 9 shows the center temperature of evaporator varies with the heat load for three inclination angles. The IHP achieved the maximum heat transfer capability of 100 W with the peak temperature of 82.8 °C at the outlet of the evaporator. When the filling ratio is low, e.g., φ = 40%, the IHP has a higher evaporator temperature when it is under anti-gravity condition than that of other inclination angles. When the filling ratio is high, the IHP has a lower evaporator temperature under anti-gravity condition than under gravity-assisted condition. The evaporator temperature increases quickly when the heat load is above 90 W. This phenomenon may attribute to the following reasons: (1) the graded pore-size wick provides a strong capillary force and high permeability to supply enough water to the evaporation section even in anti-gravity condition; (2) in anti-gravity condition, a high filling ratio will allow the working fluid to contact with the heating zone and form a thin liquid film in the evaporator. This will lead to reduced evaporation resistance of the working fluid and promote the evaporation. Fig. 10 shows the thermal resistance of IHP with high filling ratios under three inclination angles. When heat load is 80 W, the minimum thermal
Fig. 10. Thermal resistance of IHP under three inclination angles with high filling ratios.
resistance of 0.157 K/W is obtained. The thermal resistance of IHP with φ = 60% is almost lower than that with φ = 80% when under antigravity situation. The reason is that when the filling ratio is high, excess working fluid would accumulate on the surface of the condenser, restraining the condensation of vapor. When IHP operated at gravity-assisted orientation, the evaporator is located below the condenser. The gravity helps the liquid to move down to the evaporator, and the heat pipe exhibits best heat transfer
Fig. 9. Center temperature of the evaporator under different inclination angles. Filling ratio of (a) 30%, (b) 40%, (c) 60%, (d) 80%. 8
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
performance when the filling ratio is low (30% and 40%). The heat transfer capability of the IHP is not the best when the filling ratio is high (60% and 80%). This is due to sufficient working fluid accumulate in evaporator, the evaporation resistance and the thermal resistance increased. Therefore, gravity-assisted orientation creates a favorable condition for the IHP when the filling ratio is low. The influence of gravity has weakened when the filling ratio is high. This can be crosschecked with the conclusion made by Amir Faghri [19] that providing a sufficient amount of working fluid can avoid dryout limit at satrtup. When IHP at horizontal orientation, the evaporator parallel to the condenser. The IHP has a higher evaporator temperature than the IHP at gravity-assisted orientation because of the dryout limit. When the filling ratio is high, the IHP has a lower evaporator temperature at horizontal orientation than the IHP at gravity-assisted orientation. This is due to the working fluid accumulate in the lower part of the vapor chamber, the evaporator is in the center of the bottom surface of the vapor chamber, the excess liquid did not accumulate in the evaporator. Therefore, the IHP at horizontal orientation has a better thermal performance than the IHP at gravity-assisted orientation.
it has a similar three-dimensional shape as the heat-pipe cooling device studied by Xiao et al. [28]. The difference is that they assembled several heat pipes with heat sink and there must be contact thermal resistance between the heat sink and heat pipes. For the IHP, the sintered copper powder layers of the vapor chamber and the copper powder layers of the two cylindrical heat pipes are connected to each other and there is no contact thermal resistance. The thermal resistance of the IHP is 0.142 K/W when the heat load is 130 W. The thermal resistance of their heat pipe is 0.373 K/W when the heat load is 12 W. For the practicality of IHP, from Fig. 9(c), we can easily find the Te,c is below 83 °C when the heat load is 150 W. Fig. 12 shows some representative temperatures of the vapor chamber. The results present the great thermal uniformity of the IHP. This can meet the heat dissipation requirements of most electronic devices such as high-performance central processing unit and light emitting diode. Moreover, verified in Section 4.3, the IHP can work in anti-gravity condition and adapt to many inclination angles. 5. Conclusion An IHP was proposed and tested. The IHP couples a vapor chamber and two cylindrical heat pipes and there is no contact thermal resistance between them. In addition, sintered copper powder rings were used to shorten the flow path of working fluid and graded pore-size wick was introduced to strengthen the capillary force and wick permeability of the wick structure. The conclusions are summarized as follows.
4.3.2. Comparison of anti-gravity heat transfer performance In this study, the heating area and the cooling area is 1.44 cm2 and 17.6 cm2 respectively. In order to identify the anti-gravity heat transfer performance of IHP, a comparison between the anti-gravity thermal resistance of this IHP and that of other heat pipes in literature was presented in Fig. 11. Ji et al. [11] designed and tested a copper foam based VC with integrated structure. The heating area is 0.785 cm2 and the cooling area is above 78 cm2. The minimum thermal resistance of 0.185 K/W is reached when the heat flux is 80 W/cm2, their anti-gravity thermal resistance values are close to each other except some data points in the high heat load region. Their heat pipe has a high heat transfer limit due to the large cooling area. Li et al. [18] designed and tested an antigravity loop-shaped heat pipe, which has a graded capillary wick structure. With a heating area of 2.51 cm2 and cooling area of 19.95 cm2, the minimum thermal resistance of 0.153 K/W is reached when the heat flux is 16 W/cm2, beyond that the thermal resistance sharply increased owing to dryout or the boiling limit. Yao et al. [16] studied the performance of a novel anti-gravity vapor chamber, which has a tree-shaped groove evaporator surface and a hybrid mesh wick. With a heating area of 2.54 cm2 and cooling area of 22.90 cm2, the minimum thermal resistance of 0.215 K/W is reached at the maximum heat flux of 25.5 W/cm2, and the capillary limit or boiling limit was not reached due to the high porosity of hybrid mesh wick. For the IHP presented in this study, the minimum anti-gravity thermal resistance of 0.157 K/W is reached at the heat load of 80 W corresponds to a heat flux of 55.56 W/cm2. The anti-gravity thermal resistance in this study is much lower than most if not all of the results from the literature. For the dryout heat flux, only the VC of Ji et al. [11] can work at a high heat flux which exceeds 150 W/cm2 without apparent dryout. But the heating area of 0.785 cm2 is smaller than our heating area and the cooling area of that study (78 cm2) is much larger than this study. The dryout heat flux would decrease if the cooling area decreased and the thermal resistance would increase if the heating area increased, so comparisons with their VC need further study. The IHP has lowest anti-gravity thermal resistance when heat flux ranges from 25 W/cm2 to 63 W/cm2. In general, the IHP designed and tested in this study has good anti-gravity performance compared with the former works. This is mainly because of the excellent capillary performance and high permeability of the graded pore-size wick and the short liquid flow path from the condenser to the evaporator through the copper powder rings.
(1) The IHP has good heat transfer performance, the center temperature of the evaporator is below 83 °C when the heat load is 150 W. The temperature increases with increasing heat load. When the filling ratio is low, the temperature rises more quickly and the local dryout phenomenon of the IHP is more likely to occur. When the filling ratio is high, the temperature rises steadily. (2) For the same heat load, the thermal resistance varies with filling ratios and the IHP with a filling ratio of 60% has the lowest thermal resistance. Most curves of thermal resistance first go down and then go up. The minimum thermal resistance of 0.142 K/W was reached when the heat load is 130 W. (3) When heat load is low, the boiling regime in the evaporator is nucleate boiling. When heat load reaches critical heat load Qc, the boiling regime changes from nucleate boiling to transition boiling. Therefore, the heat transfer performance decreases with increasing heat load when the heat load is above Qc. (4) The IHP realize stable heat transfer at the anti-gravity condition. In addition, the IHP achieves the maximum heat transfer capability of
4.4. Structural advantages of IHP Fig. 11. Comparisons of anti-gravity thermal resistances of the present work and those reported in the literature.
The IHP coupling the vapor chamber and two cylindrical heat pipes, 9
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
Thermomechanical Phenomena in Electronic Systems, 2002. [2] V.V. Zhirnov, R.K. Cavin, J.A. Hutchby, G.I. Bourianoff, Limits to binary logic switch scaling - a gedanken model, Proc. IEEE 91 (2003) 1934–1939. [3] L. Zhu, H. Tan, J. Yu, Analysis on optimal heat exchanger size of thermoelectric cooler for electronic cooling applications, Energy Convers. Manage. 76 (2013) 685–690. [4] H. Jouhara, A. Chauhan, T. Nannou, S. Almahmoud, B. Delpech, L.C. Wrobel, Heat pipe based systems - Advances and applications, Energy 128 (2017) 729–754. [5] S. Motahar, R. Khodabandeh, Experimental study on the melting and solidification of a phase change material enhanced by heat pipe, Int. Commun. Heat Mass Transfer 73 (2016) 1–6. [6] L.L. Vasiliev, Micro and miniature heat pipes – Electronic component coolers, Appl. Therm. Eng. 28 (2008) 266–273. [7] Y. Xuan, Y. Hong, Q. Li, Investigation on transient behaviors of flat plate heat pipes, Exp. Therm Fluid Sci. 28 (2004) 249–255. [8] H. Zhang, Q. Pan, H. Zhang, Multi-scale porous copper foams as wick structures, Mater. Lett. 106 (2013) 360–362. [9] S.-C. Wong, J.-H. Liou, C.-W. Chang, Evaporation resistance measurement with visualization for sintered copper-powder evaporator in operating flat-plate heat pipes, Int. J. Heat Mass Transf. 53 (2010) 3792–3798. [10] Y. Tang, D. Yuan, L. Lu, Z. Wang, A multi-artery vapor chamber and its performance, Appl. Therm. Eng. 60 (2013) 15–23. [11] X. Ji, H. Li, J. Xu, Y. Huang, Integrated flat heat pipe with a porous network wick for high-heat-flux electronic devices, Exp. Therm Fluid Sci. 85 (2017) 119–131. [12] L. Yang, R. Zhou, X. Jin, X. Ling, H. Peng, Experimental investigate on thermal properties of a novel high temperature flat heat pipe receiver in solar power tower plant, Appl. Therm. Eng. 109 (2016) 610–618. [13] T. Yousefi, S.A. Mousavi, B. Farahbakhsh, M.Z. Saghir, Experimental investigation on the performance of CPU coolers: effect of heat pipe inclination angle and the use of nanofluids, Microelectron. Reliab. 53 (2013) 1954–1961. [14] T. Tharayil, L.G. Asirvatham, C.F.M. Cassie, S. Wongwises, Performance of cylindrical and flattened heat pipes at various inclinations including repeatability in antigravity – A comparative study, Appl. Therm. Eng. 122 (2017) 685–696. [15] Z. Deng, Y. Zheng, X. Liu, B. Zhu, Y. Chen, Experimental study on thermal performance of an anti-gravity pulsating heat pipe and its application on heat recovery utilization, Appl. Therm. Eng. 125 (2017) 1368–1378. [16] F. Yao, S. Miao, M. Zhang, Y. Chen, An experimental study of an anti-gravity vapor chamber with a tree-shaped evaporator, Appl. Therm. Eng. 141 (2018) 1000–1008. [17] Y. Tang, R. Zhou, L. Lu, Z. Xie, Anti-Gravity Loop-shaped heat pipe with graded pore-size wick, Appl. Therm. Eng. 36 (2012) 78–86. [18] H. Li, X. Wang, Z. Liu, Y. Tang, W. Yuan, R. Zhou, Y. Li, Experimental investigation on the sintered wick of the anti-gravity loop-shaped heat pipe, Exp. Therm Fluid Sci. 68 (2015) 689–696. [19] A. Faghri, Heat Pipe Science and Technology, Global Digital Press, 1995. [20] G.P. Peterson, An introduction to heat pipes: modeling, testing, and applications, (1994). [21] A. Strel’tsov, Theoretical and experimental investigation of optimum filling for heat pipes, Heat Transfer-Sov. Res. 7 (1975) 23–27. [22] G. Bartsch, J. Unk, A contribution to calculating the optimum quantity for filling a closed two-phase thermosyphon, Proceedings of 6th IHPC, 1987, pp. 641–646. [23] W. Liu, J. Gou, Y. Luo, M. Zhang, The experimental investigation of a vapor chamber with compound columns under the influence of gravity, Appl. Therm. Eng. 140 (2018) 131–138. [24] Y. Tang, D. Deng, L. Lu, M. Pan, Q. Wang, Experimental investigation on capillary force of composite wick structure by IR thermal imaging camera, Exp. Therm Fluid Sci. 34 (2010) 190–196. [25] L. Jiang, J. Ling, L. Jiang, Y. Tang, Y. Li, W. Zhou, J. Gao, Thermal performance of a novel porous crack composite wick heat pipe, Energy Convers. Manage. 81 (2014) 10–18. [26] Y. Wang, J. Cen, F. Jiang, W. Cao, J. Guo, LHP heat transfer performance: a comparison study about sintered copper powder wick and copper mesh wick, Appl. Therm. Eng. 92 (2016) 104–110. [27] S.H. Kim, I.C. Chu, M.H. Choi, D.J. Euh, Mechanism study of departure of nucleate boiling on forced convective channel flow boiling, Int. J. Heat Mass Transf. 126 (2018) 1049–1058. [28] C. Xiao, H. Liao, Y. Wang, J. Li, W. Zhu, A novel automated heat-pipe cooling device for high-power LEDs, Appl. Therm. Eng. 111 (2017) 1320–1329.
Fig. 12. Some representative temperatures of the vapor chamber.
100 W with the peak temperature of 82.8 °C at the outlet of the evaporator. The minimum of thermal resistance 0.157 K/W is reached when the heat load is 80 W under anti-gravity condition. Acknowledgements This work was supported by National Key Research and Development Program of China (2016YFB0401502), Natural Science Foundation of China under Grant (No. 61771204); Program for Chang Jiang Scholars and Innovative Research Teams in Universities (No. IRT_17R40); Guangdong Innovative Research Team Program under Grant (No. 2013C102); Science and technology project of Guangdong Province (NO. 2018A050501013, 2017B090903008); Science and Technology Project of Shenzhen Municipal Science and Technology Innovation Committee (GQYCZZ20150721150 406) Longhua District Technological SMEs Technological Innovation Project (20171228A1300902); Guangdong Provincial Key Laboratory of Optical Information Materials and Technology under Grant (No. 2017B030301007); MOE International Laboratory for Optical Information Technologies; Guangzhou Key Laboratory of Electronic Paper Displays Materials and Devices (201705030007); Leading talents of Guangdong province Program (No. 00201504); The 111 Project and Yunnan expert workstation (2017IC011). Appendix A. Supplementary material Supplementary data to this article can be found online at https:// doi.org/10.1016/j.applthermaleng.2019.113816. References [1] M. Arik, J. Petroski, S. Weaver, Thermal challanges in the future generation solid state lighting applications: Light emitting diodes, in: Conference on Thermal &
10
Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Research Paper
An integrated heat pipe coupling the vapor chamber and two cylindrical heat pipes with high anti-gravity thermal performance
T
⁎
Huawei Wanga, Pengfei Baia,b, , Honglin Zhoua, Reinder Coehoornb, Nan Lic, Hua Liaoe, Guofu Zhoua,b,c,d a Guangdong Provincial Key Laboratory of Optical Information Materials and Technology & Institute of Electronic Paper Displays, South China Academy of Advanced Optoelectronics, South China Normal University, Guangzhou 510006, PR China b National Center for International Research on Green Optoelectronics, South China Normal University, Guangzhou 510006, PR China c Shenzhen Guohua Optoelectronics Tech. Co. Ltd., Shenzhen 518110, PR China d Academy of Shenzhen Guohua Optoelectronics, Shenzhen 518110, PR China e Institute of Solar Energy, Yunnan Normal University, Kunming 650500, PR China
H I GH L IG H T S
integrated heat pipe is designed, built and tested in this study. • An performance of the heat pipe under different conditions was tested. • The lowest thermal resistance under anti-gravity condition is 0.157 K/W. • The critical heating load is affected by the filling ratio significantly. • The • The heat pipe has good performance and can adapt to many inclination angles.
A R T I C LE I N FO
A B S T R A C T
Keywords: Integrated Heat pipe Wick structure Anti-gravity
An integrated heat pipe (IHP) coupling the vapor chamber and two cylindrical heat pipes is proposed in this study. Cylindrical heat pipes were connected with the top surface of the vapor chamber to extend the area of the condenser. The copper powder rings and graded pore-size wick are used to shorten the circulation path of working fluid and strengthen the capillary pressure and wick permeability of the wick structure. The effect of various filling ratios (30%, 40%, 60%, and 80%) and different inclination angles (0°, 90°, and 180°) have been investigated experimentally. The results showed that the filling ratio has a significant impact on the heat transfer performance of IHP, and the high filling ratio IHP shows better thermal properties. Specifically, the maximum temperature is lower than 83 °C at the heat load of 150 W when the filling ratio is 60%. Moreover, the integrated heat pipe shows better anti-gravity thermal performance than a conventional heat pipe. The lowest anti-gravity thermal resistance is 0.157 K/W with a heat load of 80 W. As compared with other heat pipes reported in this literature, the IHP has a good thermal performance and can adapt to many inclination angles.
1. Introduction In recent years, the integration and power density of electronic devices is continuously becoming intensive. This trend results in a high heat flux problem and influences the lifetime and reliability of electronic significantly [1–3]. Therefore, novel components with excellent heat dissipation ability have been developed to address the problem. Heat pipes, as one of the most efficient two-phase heat transfer
technologies available [4]. They have the advantages of high heat dissipation, high thermal uniformity, and low heat transfer temperature difference. The working fluid in the heat pipe is heated by the external heat source and then evaporates and carries heat from heating source to cold source driven by the pressure difference. After vapor reaches to the cold source and condenses into liquid, the working fluid enters into wick structure and flows back to the heating section [5], completing a whole heat transfer cycle.
⁎ Corresponding author at: Guangdong Provincial Key Laboratory of Optical Information Materials and Technology & Institute of Electronic Paper Displays, South China Academy of Advanced Optoelectronics, South China Normal University, Guangzhou 510006, PR China. E-mail address: [email protected] (P. Bai).
https://doi.org/10.1016/j.applthermaleng.2019.113816 Received 11 November 2018; Received in revised form 17 May 2019; Accepted 23 May 2019 Available online 24 May 2019 1359-4311/ © 2019 Published by Elsevier Ltd.
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
φ ρ
Nomenclature I q Q R T U V
current [A] heat flux [W/cm2] heat load [W] thermal resistance [K/W] temperature [K] voltage [V] volume [L]
Subscripts c e f p w E,c hl
Greek symbols ε
filling ratio density [kg/m3]
condenser evaporator working fluid pore wick center of evaporator heat loss
porosity
of the evaporator is below 84 °C under a heat load of 90 W. Moreover, Li et al. [18] studied the effects of copper powder size on heat transfer performance of heat pipe. The results show that, for single-powder structure, 75–100 μm is the optimal size, and 25–50 μm is the optimal powder size for continuous step-graded structure. Some analytical expressions were derived to allow the designers to determine the amount of working fluid necessary for a given set or range of operating conditions, tube geometry, and working fluid properties [19–22]. In some studies the researchers manufactured the heat pipes with a series of filling ratios, they selected the optimum filling ratio according to the experimental results [14–16,23]. In this study, we first calculate the pore volume (2.94 mL) in the wick structure. Then we choose a series of filling ratios to study the effect of the filling amount of working fluid on the heat transfer performance in antigravity condition. Based on the above studies, the present work designs an integrated heat pipe (IHP), which couples a vapor chamber and two cylindrical heat pipes. In terms of the wick structure, sintered copper powder rings are introduced to shorten the flow path. In addition, this work uses sintered copper powders with different powder sizes, the composite wick within the IHP provides enough capillary force and high wick permeability to make the heat pipe operate in many inclination conditions. In the next few sections, we experimentally investigated the effects of heating rates, filling ratios and inclination angles on the thermal performance of IHP.
In heat pipes, wick structure is the core component determining the heat performance of the heat pipe. Sintered powder wick is the most effective wick structure [6]. Xuan et al. [7] made a layer of copper powder sintered on the heated surface of a flat plate heat pipe and investigated the transient behaviors. Results indicated the copper powder layer strengthens the evaporation process and enhance the response speed significantly. The sintered copper powder with high density can markedly enhance the evaporation process, but the small pore diameter and surface tension diminish the vapor flow rate in evaporator. Zhang et al. [8] experimentally verified compressed metal foam has multi-scale pores, enabling it to pump a significant amount of liquid to a long distance. Wong et al. [9] experimentally studied loosely-sintered copper-powder evaporators and measured the evaporation resistance. A glass plate was used to make the evaporation process visualized. The evaporation resistance decreases with the increase of heat flux until partial dryout occurred. The minimum evaporation resistance measured was about 0.08–0.09 W cm2/K. Tang et al. [10] proposed a multi-artery vapor chamber, copper powder rings were introduced and provide short paths for the working fluid to flow back to the evaporator. This method made the vapor chamber have a fast response speed, a low thermal resistance of 0.033 K/W. Ji et al. [11] put forward a vapor chamber with many copper foam bars inserted into the channels of the condenser. This strategy eliminates the contact thermal resistance between the surface of the vapor chamber and fins. The minimum thermal resistance is 0.03 K/W and the maximum heat flux is above 445 W/cm2 without dryout phenomenon. Gravity also influences the thermal performance of heat pipe, in the terrestrial application of the heat pipe, it is normal to have the heat source above the cold source [12,13]. Local dryout under anti-gravity condition usually occurs for conventional heat pipes. Many researchers have studied the effects of wick structure and the filling amount of working fluid on the heat transfer performance in anti-gravity condition. Trijo et al. [14] analyzed the thermal performance of cylindrical and flattened heat pipes at five inclination angles. The lowest thermal resistance of 0.46 K/W was achieved when the inclination angle is −45°. The thermal performance of a pulsating heat pipe was studied by Deng et al. [15]. The heat pipe was used for waste heat recovery from the high-temperature exhaust gas under anti-gravity condition. The start-up process of the heat pipe takes a long time and the thermal performance heavily depends on filling ratio. When the filling ratio increases under the same heat load for the anti-gravity heat pipes, the thermal resistance decreases firstly (from 50% to 70%) and then increases (from 70% to 80%), suggesting that the optimal filling ratio for the heat transfer is 70%. Yao et al. [16] designed and tested an antigravity vapor chamber, which has a tree-shaped groove evaporator surface and an 80/200 mesh wick to strengthen capillary pressure. The minimum anti-gravity thermal resistance is 0.215 K/W when the filling ratio is 60%. Tang et al. [17] developed an anti-gravity loop heat pipe with a continuous graded pore-size wick structure. The overall thermal resistance is nearly stable at 0.15 K/W, and the maximum temperature
2. IHP description The IHP primarily includes a vapor chamber, two cylindrical heat pipes, and an exhaust tube (Fig. 1). Compared with the conventional heat pipe/vapor chamber, the IHP has the following significant structural differences. Firstly, it has a similar three-dimensional shape. Two cylindrical heat pipes were carefully welded on the top plate of the vapor chamber. The copper powder layers of the two cylindrical heat pipes and the sintered copper powder layers of the vapor chamber are connected to each other, therefore, there is no contact thermal resistance between them. Second, the copper powder rings are introduced to shorten the circulation path of the working fluid. The sintered copper powder rings were mounted around the copper columns for recycling the liquid from the condenser to evaporator wick. Third, the graded pore-size wick consists of copper powders with two kinds of powder size. The two kinds of copper powders with different powder size were marked by blue color1 and coppery color in Fig. 1(b). The size of copper powder particle varies from 124 µm to 200 µm, which were sintered on the inner surface of the two cylindrical heat pipes and the top plate of the vapor chamber. The diameters of copper powder particle, sintered on the copper powder rings and the bottom surface of the vapor 1 For interpretation of color in Fig. 1, the reader is referred to the web version of this article.
2
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
Fig. 1. The structure of IHP. (a) Photograph of IHP and (b) Sectional view.
allowed for full decomposition of the impurities. Then, the sintering temperature was raised to 1000 °C at a heating rate of 6–7 °C/min and held at this temperature for 4 h. After that, the furnace was cooled to 100 °C with a cooling rate of 7–8 °C/min. Next, the two copper tubes were properly set on the top plate and solder paste was applied to their contact surfaces. Copper powders with powder size of 124–200 µm were placed on the top plate and fixed by a graphite die, and the copper powders were sintered in the same method as above. After that, the sintered copper layers connect each other between the two cylindrical
chamber, vary from 74 µm to 124 µm. The complete parameters of the IHP are specified in Table 1. Fig. 2 shows the flow chart of the manufacturing process of wick structure. Firstly, copper powders with powder size of 124–200 µm were placed on the inner surface of the two cylindrical copper pipes fixed by graphite core rod, and then they were sintered at a temperature of 1000 °C for four hours under hydrogen reducing atmosphere. The sintering process curve for sintered copper powder wick is presented in Fig. 3. In the pre-sintering stage, a hold time of 30 min at 800 °C 3
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
through the copper tube. De-ionized water was chosen to be working fluid because of its high latent heat, large surface tension, and nontoxicity. Finally, the exhaust tube was sealed by crimping press and spot welding. The porosity of the wick structure can be defined as the ratio of the pore volume Vpor to the wick structure volume Vw or
Table 1 Specification of the IHP. parameters
Dimension/material
Vapor chamber Length Width Height Side wall thickness Top plate thickness Bottom plate thickness Wick Wick thickness Bottom plate porosity Top plate porosity
50 mm 45 mm 5 mm 0.8 mm 0.5 mm 0.8 mm Sintered copper powder 0.5 mm 65% 50%
ε=
5.4 mm 6 mm 100 mm Sintered copper powder 0.9 mm 50%
Copper columns Height Diameter
2.8 mm 1.2 mm
Copper powder rings Inner diameter Outer diameter Wick Porosity
1.2 mm 3.6 mm Sintered copper powder 65%
Vw
=1−
m ρVw
(1)
where m is mass, ρ is the density of the used copper powder. The filling ratio is defined as the volume percentage of the pore space of the wick structure shared by the working fluid, the filling ratio is calculated as
φ=
Cylindrical heat pipe Inner diameter Outer diameter Length Wick Wick thickness Porosity
Vpor
Vf Vpor
=
Vf εVw
(2)
where Vl is the volume of working fluid. And a series of filling ratios of 30%, 40%, 60%, and 80% were selected for the experimental test. For the 80% filling ratio, the dosage of water was calculated as 2.39 g. In turn, 1.78 g, 1.21 g, and 0.87 g water were used in the IHP with 60%, 40%, and 30% filling ratios respectively. 3. Experimental methods 3.1. Experimental setup A test rig was designed to investigate the transient and stable behavior of the IHP under the effect of various parameters such as inclination angle and filling ratio at different heat load. The schematic diagram of the experimental system is illustrated in Fig. 5. The testing system consists of three modules: heating module, cooling module, and temperature data collecting module. The heating module includes a DC power supplier, a copper heating block and two cartridge heaters. The heating block contacts the bottom surface of IHP, whose active heating area is 12 mm × 12 mm. To minimize environmental heat losses, thermal insulation blocks were used. In addition, a layer of thermal grease fills the gap between IHP and the heating block to decrease the contact thermal resistance. The cooling module includes a thermostatic water tank (Julabo, F25-ME), a flowmeter, and two symmetrical copper cooling blocks. The cooling blocks contact with IHP closely and thermal grease was used on the contact surface to decrease the contact thermal resistance. The cooling water flow rate is 2.2 L/min with a temperature of 30 °C. The temperature data collecting module consists of ten T-type thermocouples, an Agilent 34970A Data
heat pipes and the top plate. Then, seven powder rings and the sintered powder layer on the bottom plate were sintered separately in the same method as above. Finally, seven copper powder rings were set around the copper columns. The top plate along with the bottom plate were welded by solder paste. The SEM images of the sintered copper powder wick in the bottom surface of the vapor chamber was shown in Fig. 4. There are many large pores and small pores inside the wick. The wick within the IHP is composite wick and it can increase the overall permeability and capillary forces of the wick [24–26]. The working fluid charging procedure is as follows: A hole in the side wall of the vapor chamber was drilled out and an exhaust tube was inserted into this hole. The air inside the IHP was pumped away by a vacuum pump to make the inside pressure of IHP at about 0.06 Torr (7.99 Pa). Working fluid was filled into the IHP at negative pressure
Fig. 2. Flow diagram of the manufacturing process of wick structure. 4
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
Fig. 3. Temperature-time curve for sintering process.
Fig. 4. SEM images of the pores inside the bottom surface wick.
power is switched off. The same procedure is repeated for every heat load, filling ratio, inclination angle.
Acquisition. Fig. 5(b) shows the different inclination configurations, the IHP was mounted to a plastic block that allowed tilt configuration. The temperature measurement points are shown in Fig. 6. Temperatures are recorded automatically by the data collector at 1 s intervals. The heat load started at 10 W and increased 10 W every time. The DC power supplier provides electrical power to a steel block with two cartridge heaters as the heating source. To minimize the environmental heat losses, the adiabatic area was wrapped by a layer of 5 mm-thick adiabatic foam made of styrene butadiene rubber (thermal conductivity is 0.19 W/(m·K)). After the test rig was set up, it is ready to perform the experiment. Firstly, the water bath is set at a desired cooling temperature. Then, the rotameter is adjusted to 2.2 L/min. Before the power is supplied to the heater, enough time is provided to ensure that all thermocouples readings reach approximately the value of adjusted cooling temperature which is another proof of thermocouples accuracy in temperature measurement. Then, the power is supplied to the electrical heater by adjusting the variable transformer to a certain value, equivalent to the desired heat input needed to apply on the evaporator as a heat source. This heat input can be obtained by multiplying the voltage times the current as well as the reading of the wattmeter. The next step is to monitor the variation of the temperature readings with time. After all temperature readings reach the steady state, the data is saved and the
3.2. Data reduction and uncertainty analysis The thermal resistance RI of the IHP is defined as the ratio of the temperature difference between evaporator and the condensing resistance of this system. It is expressed as
RI =
Te − Tc Q
Te =
1 5
∑ Ti
1 6
∑ Ti
Tc =
(3)
5 1
(4)
16 11
(5)
where Q is the value of the heat load, Te is the average temperature of the evaporator, Tc is the average temperature of the condenser respectively. Based on the measurements of different variables. The uncertainties in the measurement of the heat load, heat flux, and thermal resistance were calculated as: 5
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
Fig. 5. The experimental apparatus (a) and different inclination configurations (b).
ΔQ = Q Δq = q
ΔR = R
The thermal resistance of IHP with different filling ratios is shown in Fig. 8. It is clear that the high filling ratio gives the low evaporator temperature. For the same heat load, the heat pipe with φ = 30% almost has the highest thermal resistance except for under low heat load. This is because there is not enough liquid supply to the evaporator. When heat load is 10 W, the thermal resistance is 0.537 K/W for φ = 30% which is lower than 0.555 K/W for φ = 80%. This is due to the thick water film in evaporator, decreasing the evaporation rate at a low heat load. When filling ratio increased from 60% to 80%, the evaporator temperature increased at the same heat load. This can be explained that a slight excess charge will typically result in the formation of a small puddle in the lower portion of the heat pipe and result in a small reduction in performance. In this study, the filling ratio of 60% is necessary for the IHP, and when filling ratio increases to 80%, there exists a slight overcharge in the IHP, and some working liquid may accumulate in the evaporator and cause a partial blockage. This will lead to a high evaporator temperature and thermal resistance. The results can be cross-checked with the well-known conclusion made by G.P. Peterson [20]. Most curves of thermal resistance first go down and then go up. The minimum thermal resistance of 0.142 K/W was reached when the heat load is 130 W. The main reason is the thickness of the liquid film in evaporator varies with filling ratio and heat load. When heat load is low, the evaporation rate is low in the evaporator, and the liquid film is relatively thick, which lead to a large evaporation resistance. With the increase of heat load, the evaporation rate will increase, resulting in a thinner film in evaporator and a lower evaporation resistance. Moreover, a large evaporation rate will increase the
2
ΔU 2 ΔI 2 ΔQhl ⎞ ⎛ ⎞ +⎛ ⎞ +⎛ U ⎝ ⎠ ⎝ I ⎠ ⎝ Qhl ⎠ ⎜
2
⎟
2
⎛ ΔQ ⎞ + ⎛ ΔA ⎞ ⎝ A ⎠ ⎝ Q ⎠
⎜
⎟
2
(6b) 2
⎛ ΔQ ⎞ + ⎛ Δ(ΔT ) ⎞ ⎝ ΔT ⎠ ⎝ Q ⎠
⎜
(6a)
⎟
(6c)
The accuracy of the thermocouple was ± 0.5 °C, the accuracy of the flowmeter was ± 0.088 L/min, the input power was supplied by a DC power supplier with 0.1% voltage accuracy and 0.5% current accuracy. The uncertainty of the heat loss was found to be less than 3% and the uncertainty of the heating area was 0.67%. Therefore, the maximum uncertainties of the heat load, heat flux, and the thermal resistance were 3.04%, 3.11%, and 6.75%. 4. Results and discussion 4.1. Evaporator temperature and thermal resistance of IHP Fig. 7 shows the center temperature of the evaporator with the applied heat load. The value of Te,c increases with the increase of heat load, but the rate of temperature rise varies with the filling ratio. When the filling ratio is low, the temperature rising speed becomes faster and faster with the increase of heat load, while when the filling ratio is high, the rising speed of temperature is relatively stable. 6
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
Fig. 6. The measurement points of the IHP and heater.
Fig. 7. Center temperature of the evaporator for different filling ratios.
Fig. 8. Thermal resistance of IHP with different filling ratios.
condensation rate simultaneously. When the heat load exceeds a critical value, the thermal resistance will increase due to partial dryout in the evaporator. The lower the filling ratio, the more obvious the increase of thermal resistance.
heated, nucleate boiling firstly occurs in the evaporator, then the nucleate boiling becomes more and more intense with the increase of heat load. When the heat load reaches Qc, the boiling regime in the evaporator changes from nucleate boiling to transition boiling [27]. In this case, many vapor blankets cover on boiling surfaces, the dried surfaces have no more nucleate boiling chances. Thus, the thermal resistance RI keeps increasing as the curves in Fig. 8 for φ = 30%. Table 2 shows Qc for four different filling ratios under three inclination angles.
4.2. The critical heat load of IHP As shown in Fig. 8, the thermal resistance decreases with the increase of heat load when the heat load is low. Especially when the heat load was applied from 10 W to 20 W, there is a steep decrease in the thermal resistance. The phenomena may be explained as the startup of the IHP. After that, the thermal resistance continues to decrease until the critical heat load Qc was achieved. This is because when IHP is
4.3. Anti-gravity thermal performance 4.3.1. Mechanism of anti-gravity thermal performance Gravity has a great impact on the vapor-liquid distribution and flow. 7
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
Table 2 Critical heat load of IHP. Charging ratio
Qc for θ = 0° (W)
Qc for θ = 90° (W)
Qc for θ = 180° (W)
30% 40% 60% 80%
30 60 130 130
20 30 130 > 140
10 40 60 80
Therefore, it is necessary to study the effect of gravity on the performance of IHP. Fig. 9 shows the center temperature of evaporator varies with the heat load for three inclination angles. The IHP achieved the maximum heat transfer capability of 100 W with the peak temperature of 82.8 °C at the outlet of the evaporator. When the filling ratio is low, e.g., φ = 40%, the IHP has a higher evaporator temperature when it is under anti-gravity condition than that of other inclination angles. When the filling ratio is high, the IHP has a lower evaporator temperature under anti-gravity condition than under gravity-assisted condition. The evaporator temperature increases quickly when the heat load is above 90 W. This phenomenon may attribute to the following reasons: (1) the graded pore-size wick provides a strong capillary force and high permeability to supply enough water to the evaporation section even in anti-gravity condition; (2) in anti-gravity condition, a high filling ratio will allow the working fluid to contact with the heating zone and form a thin liquid film in the evaporator. This will lead to reduced evaporation resistance of the working fluid and promote the evaporation. Fig. 10 shows the thermal resistance of IHP with high filling ratios under three inclination angles. When heat load is 80 W, the minimum thermal
Fig. 10. Thermal resistance of IHP under three inclination angles with high filling ratios.
resistance of 0.157 K/W is obtained. The thermal resistance of IHP with φ = 60% is almost lower than that with φ = 80% when under antigravity situation. The reason is that when the filling ratio is high, excess working fluid would accumulate on the surface of the condenser, restraining the condensation of vapor. When IHP operated at gravity-assisted orientation, the evaporator is located below the condenser. The gravity helps the liquid to move down to the evaporator, and the heat pipe exhibits best heat transfer
Fig. 9. Center temperature of the evaporator under different inclination angles. Filling ratio of (a) 30%, (b) 40%, (c) 60%, (d) 80%. 8
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
performance when the filling ratio is low (30% and 40%). The heat transfer capability of the IHP is not the best when the filling ratio is high (60% and 80%). This is due to sufficient working fluid accumulate in evaporator, the evaporation resistance and the thermal resistance increased. Therefore, gravity-assisted orientation creates a favorable condition for the IHP when the filling ratio is low. The influence of gravity has weakened when the filling ratio is high. This can be crosschecked with the conclusion made by Amir Faghri [19] that providing a sufficient amount of working fluid can avoid dryout limit at satrtup. When IHP at horizontal orientation, the evaporator parallel to the condenser. The IHP has a higher evaporator temperature than the IHP at gravity-assisted orientation because of the dryout limit. When the filling ratio is high, the IHP has a lower evaporator temperature at horizontal orientation than the IHP at gravity-assisted orientation. This is due to the working fluid accumulate in the lower part of the vapor chamber, the evaporator is in the center of the bottom surface of the vapor chamber, the excess liquid did not accumulate in the evaporator. Therefore, the IHP at horizontal orientation has a better thermal performance than the IHP at gravity-assisted orientation.
it has a similar three-dimensional shape as the heat-pipe cooling device studied by Xiao et al. [28]. The difference is that they assembled several heat pipes with heat sink and there must be contact thermal resistance between the heat sink and heat pipes. For the IHP, the sintered copper powder layers of the vapor chamber and the copper powder layers of the two cylindrical heat pipes are connected to each other and there is no contact thermal resistance. The thermal resistance of the IHP is 0.142 K/W when the heat load is 130 W. The thermal resistance of their heat pipe is 0.373 K/W when the heat load is 12 W. For the practicality of IHP, from Fig. 9(c), we can easily find the Te,c is below 83 °C when the heat load is 150 W. Fig. 12 shows some representative temperatures of the vapor chamber. The results present the great thermal uniformity of the IHP. This can meet the heat dissipation requirements of most electronic devices such as high-performance central processing unit and light emitting diode. Moreover, verified in Section 4.3, the IHP can work in anti-gravity condition and adapt to many inclination angles. 5. Conclusion An IHP was proposed and tested. The IHP couples a vapor chamber and two cylindrical heat pipes and there is no contact thermal resistance between them. In addition, sintered copper powder rings were used to shorten the flow path of working fluid and graded pore-size wick was introduced to strengthen the capillary force and wick permeability of the wick structure. The conclusions are summarized as follows.
4.3.2. Comparison of anti-gravity heat transfer performance In this study, the heating area and the cooling area is 1.44 cm2 and 17.6 cm2 respectively. In order to identify the anti-gravity heat transfer performance of IHP, a comparison between the anti-gravity thermal resistance of this IHP and that of other heat pipes in literature was presented in Fig. 11. Ji et al. [11] designed and tested a copper foam based VC with integrated structure. The heating area is 0.785 cm2 and the cooling area is above 78 cm2. The minimum thermal resistance of 0.185 K/W is reached when the heat flux is 80 W/cm2, their anti-gravity thermal resistance values are close to each other except some data points in the high heat load region. Their heat pipe has a high heat transfer limit due to the large cooling area. Li et al. [18] designed and tested an antigravity loop-shaped heat pipe, which has a graded capillary wick structure. With a heating area of 2.51 cm2 and cooling area of 19.95 cm2, the minimum thermal resistance of 0.153 K/W is reached when the heat flux is 16 W/cm2, beyond that the thermal resistance sharply increased owing to dryout or the boiling limit. Yao et al. [16] studied the performance of a novel anti-gravity vapor chamber, which has a tree-shaped groove evaporator surface and a hybrid mesh wick. With a heating area of 2.54 cm2 and cooling area of 22.90 cm2, the minimum thermal resistance of 0.215 K/W is reached at the maximum heat flux of 25.5 W/cm2, and the capillary limit or boiling limit was not reached due to the high porosity of hybrid mesh wick. For the IHP presented in this study, the minimum anti-gravity thermal resistance of 0.157 K/W is reached at the heat load of 80 W corresponds to a heat flux of 55.56 W/cm2. The anti-gravity thermal resistance in this study is much lower than most if not all of the results from the literature. For the dryout heat flux, only the VC of Ji et al. [11] can work at a high heat flux which exceeds 150 W/cm2 without apparent dryout. But the heating area of 0.785 cm2 is smaller than our heating area and the cooling area of that study (78 cm2) is much larger than this study. The dryout heat flux would decrease if the cooling area decreased and the thermal resistance would increase if the heating area increased, so comparisons with their VC need further study. The IHP has lowest anti-gravity thermal resistance when heat flux ranges from 25 W/cm2 to 63 W/cm2. In general, the IHP designed and tested in this study has good anti-gravity performance compared with the former works. This is mainly because of the excellent capillary performance and high permeability of the graded pore-size wick and the short liquid flow path from the condenser to the evaporator through the copper powder rings.
(1) The IHP has good heat transfer performance, the center temperature of the evaporator is below 83 °C when the heat load is 150 W. The temperature increases with increasing heat load. When the filling ratio is low, the temperature rises more quickly and the local dryout phenomenon of the IHP is more likely to occur. When the filling ratio is high, the temperature rises steadily. (2) For the same heat load, the thermal resistance varies with filling ratios and the IHP with a filling ratio of 60% has the lowest thermal resistance. Most curves of thermal resistance first go down and then go up. The minimum thermal resistance of 0.142 K/W was reached when the heat load is 130 W. (3) When heat load is low, the boiling regime in the evaporator is nucleate boiling. When heat load reaches critical heat load Qc, the boiling regime changes from nucleate boiling to transition boiling. Therefore, the heat transfer performance decreases with increasing heat load when the heat load is above Qc. (4) The IHP realize stable heat transfer at the anti-gravity condition. In addition, the IHP achieves the maximum heat transfer capability of
4.4. Structural advantages of IHP Fig. 11. Comparisons of anti-gravity thermal resistances of the present work and those reported in the literature.
The IHP coupling the vapor chamber and two cylindrical heat pipes, 9
Applied Thermal Engineering 159 (2019) 113816
H. Wang, et al.
Thermomechanical Phenomena in Electronic Systems, 2002. [2] V.V. Zhirnov, R.K. Cavin, J.A. Hutchby, G.I. Bourianoff, Limits to binary logic switch scaling - a gedanken model, Proc. IEEE 91 (2003) 1934–1939. [3] L. Zhu, H. Tan, J. Yu, Analysis on optimal heat exchanger size of thermoelectric cooler for electronic cooling applications, Energy Convers. Manage. 76 (2013) 685–690. [4] H. Jouhara, A. Chauhan, T. Nannou, S. Almahmoud, B. Delpech, L.C. Wrobel, Heat pipe based systems - Advances and applications, Energy 128 (2017) 729–754. [5] S. Motahar, R. Khodabandeh, Experimental study on the melting and solidification of a phase change material enhanced by heat pipe, Int. Commun. Heat Mass Transfer 73 (2016) 1–6. [6] L.L. Vasiliev, Micro and miniature heat pipes – Electronic component coolers, Appl. Therm. Eng. 28 (2008) 266–273. [7] Y. Xuan, Y. Hong, Q. Li, Investigation on transient behaviors of flat plate heat pipes, Exp. Therm Fluid Sci. 28 (2004) 249–255. [8] H. Zhang, Q. Pan, H. Zhang, Multi-scale porous copper foams as wick structures, Mater. Lett. 106 (2013) 360–362. [9] S.-C. Wong, J.-H. Liou, C.-W. Chang, Evaporation resistance measurement with visualization for sintered copper-powder evaporator in operating flat-plate heat pipes, Int. J. Heat Mass Transf. 53 (2010) 3792–3798. [10] Y. Tang, D. Yuan, L. Lu, Z. Wang, A multi-artery vapor chamber and its performance, Appl. Therm. Eng. 60 (2013) 15–23. [11] X. Ji, H. Li, J. Xu, Y. Huang, Integrated flat heat pipe with a porous network wick for high-heat-flux electronic devices, Exp. Therm Fluid Sci. 85 (2017) 119–131. [12] L. Yang, R. Zhou, X. Jin, X. Ling, H. Peng, Experimental investigate on thermal properties of a novel high temperature flat heat pipe receiver in solar power tower plant, Appl. Therm. Eng. 109 (2016) 610–618. [13] T. Yousefi, S.A. Mousavi, B. Farahbakhsh, M.Z. Saghir, Experimental investigation on the performance of CPU coolers: effect of heat pipe inclination angle and the use of nanofluids, Microelectron. Reliab. 53 (2013) 1954–1961. [14] T. Tharayil, L.G. Asirvatham, C.F.M. Cassie, S. Wongwises, Performance of cylindrical and flattened heat pipes at various inclinations including repeatability in antigravity – A comparative study, Appl. Therm. Eng. 122 (2017) 685–696. [15] Z. Deng, Y. Zheng, X. Liu, B. Zhu, Y. Chen, Experimental study on thermal performance of an anti-gravity pulsating heat pipe and its application on heat recovery utilization, Appl. Therm. Eng. 125 (2017) 1368–1378. [16] F. Yao, S. Miao, M. Zhang, Y. Chen, An experimental study of an anti-gravity vapor chamber with a tree-shaped evaporator, Appl. Therm. Eng. 141 (2018) 1000–1008. [17] Y. Tang, R. Zhou, L. Lu, Z. Xie, Anti-Gravity Loop-shaped heat pipe with graded pore-size wick, Appl. Therm. Eng. 36 (2012) 78–86. [18] H. Li, X. Wang, Z. Liu, Y. Tang, W. Yuan, R. Zhou, Y. Li, Experimental investigation on the sintered wick of the anti-gravity loop-shaped heat pipe, Exp. Therm Fluid Sci. 68 (2015) 689–696. [19] A. Faghri, Heat Pipe Science and Technology, Global Digital Press, 1995. [20] G.P. Peterson, An introduction to heat pipes: modeling, testing, and applications, (1994). [21] A. Strel’tsov, Theoretical and experimental investigation of optimum filling for heat pipes, Heat Transfer-Sov. Res. 7 (1975) 23–27. [22] G. Bartsch, J. Unk, A contribution to calculating the optimum quantity for filling a closed two-phase thermosyphon, Proceedings of 6th IHPC, 1987, pp. 641–646. [23] W. Liu, J. Gou, Y. Luo, M. Zhang, The experimental investigation of a vapor chamber with compound columns under the influence of gravity, Appl. Therm. Eng. 140 (2018) 131–138. [24] Y. Tang, D. Deng, L. Lu, M. Pan, Q. Wang, Experimental investigation on capillary force of composite wick structure by IR thermal imaging camera, Exp. Therm Fluid Sci. 34 (2010) 190–196. [25] L. Jiang, J. Ling, L. Jiang, Y. Tang, Y. Li, W. Zhou, J. Gao, Thermal performance of a novel porous crack composite wick heat pipe, Energy Convers. Manage. 81 (2014) 10–18. [26] Y. Wang, J. Cen, F. Jiang, W. Cao, J. Guo, LHP heat transfer performance: a comparison study about sintered copper powder wick and copper mesh wick, Appl. Therm. Eng. 92 (2016) 104–110. [27] S.H. Kim, I.C. Chu, M.H. Choi, D.J. Euh, Mechanism study of departure of nucleate boiling on forced convective channel flow boiling, Int. J. Heat Mass Transf. 126 (2018) 1049–1058. [28] C. Xiao, H. Liao, Y. Wang, J. Li, W. Zhu, A novel automated heat-pipe cooling device for high-power LEDs, Appl. Therm. Eng. 111 (2017) 1320–1329.
Fig. 12. Some representative temperatures of the vapor chamber.
100 W with the peak temperature of 82.8 °C at the outlet of the evaporator. The minimum of thermal resistance 0.157 K/W is reached when the heat load is 80 W under anti-gravity condition. Acknowledgements This work was supported by National Key Research and Development Program of China (2016YFB0401502), Natural Science Foundation of China under Grant (No. 61771204); Program for Chang Jiang Scholars and Innovative Research Teams in Universities (No. IRT_17R40); Guangdong Innovative Research Team Program under Grant (No. 2013C102); Science and technology project of Guangdong Province (NO. 2018A050501013, 2017B090903008); Science and Technology Project of Shenzhen Municipal Science and Technology Innovation Committee (GQYCZZ20150721150 406) Longhua District Technological SMEs Technological Innovation Project (20171228A1300902); Guangdong Provincial Key Laboratory of Optical Information Materials and Technology under Grant (No. 2017B030301007); MOE International Laboratory for Optical Information Technologies; Guangzhou Key Laboratory of Electronic Paper Displays Materials and Devices (201705030007); Leading talents of Guangdong province Program (No. 00201504); The 111 Project and Yunnan expert workstation (2017IC011). Appendix A. Supplementary material Supplementary data to this article can be found online at https:// doi.org/10.1016/j.applthermaleng.2019.113816. References [1] M. Arik, J. Petroski, S. Weaver, Thermal challanges in the future generation solid state lighting applications: Light emitting diodes, in: Conference on Thermal &
10