- Email: [email protected]
Experimental study on the operating characteristics of a capillary pumped loop with a flat evaporator
International Journal of Heat and Mass Transfer 53 (2010) 268–275
Contents lists available at ScienceDirect
International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
Experimental study on the operating characteristics of a capillary pumped loop with a flat evaporator Wukchul Joung, Hyungjin Hwang, Jinho Lee * School of Mechanical Engineering, Yonsei University, 134 Shinchon-dong, Seodaemoon-gu, Seoul 120-749, Republic of Korea
a r t i c l e
i n f o
Article history: Received 6 March 2009 Received in revised form 11 June 2009 Accepted 2 September 2009 Available online 23 October 2009 Keywords: Capillary pumped loop Flat evaporator Operating temperature control Reservoir temperature control
a b s t r a c t Recent advances in electronics technology have led to growing demand for highly efficient heat transfer devices with operating temperature controllability. Although capillary pumped loops are excellent candidates for this requirement and have been successful in various applications, their typical cylindrical evaporator shape has been a major drawback for applications to flat heat sources. Moreover, operating characteristics and the corresponding physical understanding of the capillary pumped loops are still uncertain, and lack of data for their operation makes their application unreliable. Thus, in this work, a capillary pumped loop with a flat evaporator was devised, and its operating characteristics were investigated. Ó 2009 Elsevier Ltd. All rights reserved.
1. Introduction A capillary pumped loop (CPL) is a highly effective two-phase heat transfer device. Compared to a similar two-phase heat transfer device, i.e., the loop heat pipe, the CPL has the distinctive feature of operating temperature controllability. This characteristic is due to a separate two-phase working fluid accumulator called a reservoir. The reservoir, which contains two-phase working fluid that remains in a saturation state, is directly connected to the liquid transport line between the evaporator and the condenser. By controlling its temperature, the operating temperature of a CPL (i.e., the saturation temperature on the evaporating surface of the wick) is controlled [1–4]. From its early development, due to this operating temperature controllability, the CPL has found applications in the thermal management of high precision equipment, such as avionics and optics in spacecraft. More recently, electronics cooling and thermal control of renewable energy systems are CPL application trends [5–10]. Despite its operating temperature controllability, additional thermal resistance is introduced by its cylindrical evaporator structure, which requires a supplementary intermediary called a saddle. This additional thermal resistance necessitates a flat evaporator CPL. Moreover, since most heat sources have flat thermo-contact surfaces and are closely packed, the development of a CPL with a flat evaporator and symmetric heat absorbing surfaces is of primary concern [10–12]. In addition to the requirement for struc-
* Corresponding author. Tel.: +82 2 2123 2816; fax: +82 2 312 2159. E-mail address: [email protected] (J. Lee). 0017-9310/$ - see front matter Ó 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.ijheatmasstransfer.2009.09.032
tural improvement, the lack of data for steady state operations of the CPL and physical understanding of the mechanism by which the operating temperature is controlled makes its applications unreliable. Therefore, in this work, we first focused on creating a flat evaporator CPL (hereafter the FECPL) with symmetric heat absorbing surfaces. Its operating characteristics, including both steady state and transient state responses, were investigated. In particular, steady state operations with various reservoir temperatures were analyzed to understand the role of the reservoir in the operating temperature control mode. The transient state responses under the changing heat load were examined to ascertain whether the developed CPL experienced wick dry-out when subjected to an abrupt heat load decrease, which is known to be the most probable reason for CPL wick dry-out [13–15]. 2. Operating principles and experimental parameters Fig. 1(a) and (b) shows the schematic of the CPL and the corresponding thermodynamic diagram for its operation. As shown in this figure, the saturation states of the evaporating surface of the wick and the reservoir are closely related by the following pressure balance equation.
DPsat: ¼ Psat:1 Psat:8 ¼ ql gðh1 þ h2 Þ þ DPf ;1!5
ð1Þ
where Psat.1 and Psat.8 are the saturation pressures on the evaporating surface of the wick and in the reservoir, respectively, and h1 + h2 is the net height difference between the two-phase interfaces in the reservoir and the condenser. DPf,1?5 is the frictional pressure drop
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269
Nomenclature g h P R Q_ T
gravity acceleration height pressure thermal resistance heat load temperature
g l ll ll, 1 ll, 2 loss res sat sat. 1 sat. 8 th vg vl wick
Greek symbols u phase q density Subscripts cd condenser cond. out condenser outlet evap evaporator f frictional
along the working fluid flow path from the evaporating surface of the wick to the bifurcation, and is given as follows.
DPf ;1!5 ¼ DP v g þ DPv l þ DP cd þ DPll;1
ð2Þ
where DPll,1 is the liquid phase frictional pressure drop from the outlet of the condenser to the bifurcation. As shown in Eq. (1), the
gravitational liquid phase liquid transport line liquid transport line from the condenser outlet to the bifurcation liquid transport line from the bifurcation to the end of the liquid transport line loss reservoir saturation saturation state of the evaporating surface of the wick saturation state of the reservoir thermal vapor removal grooves vapor transport line filtration
operating temperature of the CPL (i.e., the saturation temperature on the evaporating surface of the wick) is controlled by the reservoir (saturation) temperature which acts as a reference point for the thermodynamic states of the rest of the CPL. The saturation pressure difference described in Eq. (1) is the driving force of the CPL, and it circulates the working fluid through the loop. The capillary pressure difference across the wick evaporating surface provides a compensating force for the saturation pressure difference between the evaporating surface of the wick and the reservoir. This pressure difference makes the working fluid flow along the loop without penetrating the wick, and is given as follows.
DPcapillary ¼ DPsat: ql gh2 ql gh3 þ DPll;2 þ DPwick ¼ ql gðh1 h3 Þ þ DPf þ DPwick
ð3Þ
where DPll,2 is the liquid phase frictional pressure drop from the bifurcation to the end of the liquid transport line, DPwick is the filtration pressure drop through the wick, and DPf is the frictional pressure drop along the working fluid path including the whole liquid transport line. The frictional pressure drop, DPf, is given as follows.
DPf ¼ DPv g þ DPv l þ DPcd þ DP ll
ð4Þ
As explained above, the operating temperature of the CPL is determined by the reservoir temperature. In this work, to see the effect of the reservoir temperature on the steady state responses of the FECPL, we measured the steady state operating temperature variation with heat load at fixed reservoir temperatures, and the responses to reservoir temperature change at a fixed heat load. Another important factor to be investigated regarding the reliable operation of the CPL is its tendency of wick dry-out when subjected to an abrupt heat load decrease. It is generally known that the CPL wick dries out when the heat load is suddenly decreased due to the vapor generation and accumulation in the evaporator core, which blocks the working fluid supply to the wick [13,14]. Thus, we measured transient responses of the FECPL under abrupt heat load changes to see whether or not the wick dries out. 3. FECPL design
Fig. 1. Schematic and corresponding thermodynamic operation curve of the CPL.
The assembly process and structural view of the evaporator, the wick and the reservoir are shown in Fig. 2. As shown in the figure, employing two flat porous plates and a side rim, the wick had planar heat absorbing (or evaporating) surfaces to form the flat evaporator. The wick of the CPL plays an important role in CPL
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Fig. 2. Exploded view of the evaporator and the reservoir.
operation, which is a hydraulic barrier to the saturation pressure difference between the wick evaporating surface and the reservoir [16,17]. For the wick of the CPL to be a successful hydraulic barrier, it is essential to separate the two phases from each other as completely as possible. Otherwise, the required saturation pressure difference cannot be developed properly. To this end, a flange was added at the opening of the wick to provide a better contact surface with the evaporator housing. We used custom-made sintered porous plates (AISI 304) with a mean pore size and porosity of 5 lm and 0.5, respectively, as obtained from the manufacturer’s specifications. Dimensions of the flat porous plate were 52.77 mm 52.77 mm 3 mm (width height thickness). The side rim and the flange were made of stainless steel (AISI 316L), and all elements, including the porous plates, were laser welded for complete sealing. The wick was assembled with the evaporator housing, as shown in Fig. 2. All contact lines between the evaporator housing and the wick were also laser welded for the complete separation of vapor and liquid phases. The evaporator housing was made of stainless steel (AISI 316L), and had 13 vapor removal grooves on each inner surface. The reservoir, which had a hexahedral shape, was made of copper plates, and was designed to have around half of the total loop volume. The working fluid transport lines for the vapor and liquid phase working fluids were made of 0.6 and 0.46 m-long 1=4 in. stainless steel tubes. The transport lines were connected to the evaporator (vapor outlet and liquid inlet) and the reservoir. The condenser was made of a shell and tube type (one shell pass and two tube passes with no baffles) heat exchanger which consisted of a 0.8 m long copper tube with a diameter of 1=4 in. The whole assembly of the FECPL is presented in Fig. 3.
4. Test set-up The performance of the FECPL was evaluated with its operating temperatures and thermal resistance. For this purpose, 18 T-type thermocouples were attached along the working fluid flow path. Two absolute pressure transducers were placed in the reservoir and the vapor outlet to examine the saturation pressure difference
Fig. 3. External view of the FECPL, and temperature and pressure measurement locations.
between the evaporator and the reservoir. The errors associated with these sensors were a 0.1 °C deviation at 100 °C and 0.05% FS, respectively. Fig. 3 shows the locations of the temperature readings and the pressure measurement. The thermal performance of the FECPL was evaluated in terms of the vapor outlet temperature, which showed the working fluid’s temperature at the vapor outlet, and the evaporator temperature, which was defined as the average of the temperature readings over the evaporator surface (1)–(4). Another factor used in the performance evaluation was the thermal resistance, which was defined as the ratio of the temperature difference between the evaporator and the condenser outlet to the supplied heat load, and is given as
Rth ¼
T ev ap T cond:out ; Q_ ¼ 1
ð5Þ
C th
where Cth is the thermal conductance and is the reciprocal of the thermal resistance.
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271
Fig. 4. Schematic of the test set-up.
The test set-up is shown in Fig. 4. The reservoir temperature was controlled using an isothermal water circulation bath with a volume flow rate of 1 L/min, and the range of temperature change was from 30 °C to 80 °C with a 10 °C increment. The volume flow rate of the coolant supplied to the condenser was 2 L/min. The ambient temperature was maintained at 25 °C throughout the experiment. Water was used as the working fluid because it is the most easily accessible moderate temperature working fluid. However, due to its rather uncertain compatibility with stainless steel [19,20], we measured the inside pressure of the loop before the heat load was supplied during the whole experiment period (60 days). The fluid inventory in the FECPL was fixed at 70% of the total volume. The heat load was supplied by two electric loaders through flat ceramic heaters (50 mm 50 mm, 5.6 X), which were attached to the outer surface of the evaporator (Figs. 2 and 3), and the heat input range was from 100 W to 220 W. Heat loss to the ambient was assumed to be negligible throughout this work, as the heat loss in the hottest case was estimated to be less than 1% of the total heat input (from the vapor transport line, with Tevap = 81.9 °C at 220 W). The tests were repeated at least three times to obtain a reliable steady state data set, and the uncertainty analysis was done at a 95% confidence level. 5. Results and discussion
Fig. 5. Reservoir pressure variation with time before heat input.
fluid. As shown in the figure, the internal pressure of the reservoir gradually increased with time, though daily variations were small. This was attributed to the non-condensable gas generation inside the loop, since such an internal pressure rise had never been experienced using other combinations of working fluids and stainless steel by authors with similar configurations (e.g., LHPs).
5.1. Compatibility 5.2. Start-up Compatibility is an important factor in determining a working fluid and material combination, as an incompatible combination can cause non-condensable gas generation and an unnecessary internal pressure rise. Water, despite its wide availability and easy accessibility, is known to cause hydrogen generation when used with stainless steel [19,20]. Thus, before discussing the test results, it is appropriate to mention the compatibility test results obtained for the test period of 60 days. To check the non-condensable gas generation, we measured the internal pressure in the reservoir before heat input and compared it to the saturation pressure of water at 25 °C. Fig. 5 shows the reservoir pressure variation with time before heat input (i.e., at the ambient temperature of 25 °C). Compared to the saturation pressure of water at 25 °C (3.17 kPa), the initial reservoir pressure of 11.72 kPa was due to the initially injected non-condensable gases when filling the FECPL with the working
Fig. 6 shows the temperature variations of the evaporator, vapor outlet, liquid (transport) line and reservoir during the start-up (Q_ ¼ 100 W) of the FECPL when the reservoir temperature was set at 50 °C. As shown in the figure, the reservoir temperature was set at 50 °C prior to heat input, and then a heat load was applied (around 1400 s). After heat input, due to vapor generation on the evaporating surface of the wick and also due to the density difference between the vapor and liquid phases, the temperature of the vapor outlet rapidly increased, representing vapor discharge from the vapor outlet, and was followed by a decrease in the liquid line temperature, which showed successful working fluid circulation from the vapor outlet through the condenser to the liquid inlet. After the start-up, the FECPL reached a steady state operation (around 400 s after heat input). Start-ups of other reservoir temperatures showed similar trends, and no failure was observed.
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Fig. 7 shows the temperature variation of the vapor outlet with heat load at fixed reservoir temperatures. The vapor outlet temperature represents the temperature of the working fluid discharged from the evaporator, which is close to the saturation temperature on the wick evaporating surface. Thus, it is appropriate to use the vapor outlet temperature as an indicator of the saturation temperature of the evaporator. As shown clearly in the figure, the vapor outlet temperature increased as the reservoir temperature increased, indicating that the operating temperature of the FECPL was successfully controlled by the reservoir temperature. This was because the reservoir was in a saturation state. By setting the reservoir temperature, the saturation pressure of the reservoir was set, affecting the rest of the loop. This operating temperature controllability demonstrated the role of the reservoir as a reference point for thermodynamic states of other components. However, the variation with heat load showed that the vapor outlet temperature increased with heat load despite the constant reservoir temperature, and this was attributed to the effect of non-condensable gases in the reservoir. The reservoir, which was at the lowest pressure among CPL components, contained non-condensable gases thus having a mixture pressure of the saturation pressure of the working fluid and the partial pressure of the non-condensable gases. The saturation pressure of the working fluid in the reservoir was constant with heat load due to the constant reservoir temper-
ature. However, since the amount of liquid phase working fluid discharged from the condenser increased with heat load [18], and this liquid phase working fluid compressed the non-condensable gases in the reservoir, the partial pressure of the non-condensable gases increased with heat load, making the mixture pressure of the reservoir increase with heat load. Fig. 8 shows the reservoir pressure variation with heat load at various reservoir temperatures. As explained above, the internal pressure of the reservoir increased with heat load at constant reservoir temperatures due to the compression of the non-condensable gases. Fig. 9 shows the vapor outlet pressure variation with heat load, and due to the fact that the saturation pressure of the evaporator ( vapor outlet pressure) is the sum of the reservoir (saturation) pressure and the pressure drop along the working fluid flow path (frictional and hydrostatic), the tendency was similar to the reservoir pressure variation. In Fig. 7, at a reservoir temperature of 30 °C, the vapor outlet temperature suddenly increased at a heat load of 200 W, and this was attributed to the hard filled reservoir (i.e., loss of saturation state of the working fluid) which was confirmed by the rapid pressure rise of the reservoir at the same heat load of 200 W in Fig. 8. Fig. 7 also shows the decreasing temperature differences between the vapor outlet and the reservoir with reservoir temperature, which were, for example at 100 W, 24.8 °C, 18.6 °C, 13.1 °C, 8.1 °C, 3.9 °C and 0.8 °C for reservoir temperatures from 30 °C to 80 °C. These uneven temperature differences were due to the nonlinearly increasing nature of the water saturation curve. Thus the saturation pressure difference between the evaporator and the reservoir should have remained constant while the reservoir temperature changed. This was confirmed by the nearly constant pressure differences between the vapor outlet and the reservoir at a fixed heat load under the reservoir temperature change from Figs. 8 and 9 (pressure differences (minimum to maximum): 2.06–2.37 kPa at 100 W, 2.17–2.31 kPa at 120 W, 2.13–2.26 at 140 W, 2.15–2.29 kPa at 160 W, 2.26– 2.35 kPa at 180 W, 2.29–2.50 kPa at 200 W, 2.27–2.44 kPa at 220 W). Figs. 10 and 11 show the evaporator and condenser outlet temperature variations with heat load at fixed reservoir temperatures. Evaporator temperature showed a similar trend to the vapor outlet temperature, except for small deviations due to conduction through the housing material. For condenser outlet temperature variation, starting from the heat load of 100 W, sudden decreases in temperature at increasing heat loads with reservoir temperature were observed. This was thought to be the result of the interaction between the increasing vapor temperature with reservoir temperature and the increasing mass flow rate of the working fluid with
Fig. 7. Vapor outlet temperature variation with heat load at fixed reservoir temperatures.
Fig. 8. Reservoir pressure variation with heat load at fixed reservoir temperatures.
Fig. 6. Start-up of the FECPL at a fixed reservoir temperature of 50 °C.
5.3. Steady state responses
W. Joung et al. / International Journal of Heat and Mass Transfer 53 (2010) 268–275
Fig. 9. Vapor outlet pressure variation with heat load at fixed reservoir temperatures.
Fig. 10. Evaporator temperature variation with heat load at fixed reservoir temperatures.
Fig. 12. Thermal resistance variation with heat load temperatures.
273
at fixed reservoir
shown in the figure, most of the regimes on which the FECPL operated were variable conductance regimes where the thermal resistance (or the thermal conductance) varies with heat load. The reason for the appearance of the variable conductance regime is explained by the evaporator and condenser outlet temperature variations in Figs. 10 and 11. In Figs. 10 and 11, the increase rate of temperature difference between the evaporator and the condenser outlet was less than the increase rate of heat load, thus based on Eq. (5), resulting in a gradual decrease of thermal resistance. Fig. 13 shows the evaporator temperature variation with step increase and decrease in the reservoir temperature at a heat load of 100 W, when the reservoir temperature increment and decrement were 10 °C (i.e., 30 °C ? 40 °C ? 50 °C ? 60 °C ? 70 °C ? 80 °C ? 70 °C ? 60 °C ? 50 °C ? 40 °C? 30 °C). As shown in the figure, there was negligibly small hysteresis for the step variations in the reservoir temperature, and the results were nearly identical to the cases of the fixed reservoir temperature (deviations were less than 1 °C). Fig. 14 shows the evaporator temperature variation under abrupt reservoir temperature change at a heat load of 100 W. The reservoir temperature was changed in the following sequence.
30 C ! 80 C ! 30 C ! 70 C ! 40 C ! 70 C ! 50 C ! 60 C ! 50 C:
Fig. 11. Condenser outlet temperature variation with heat load at fixed reservoir temperatures.
heat load. Based on these temperatures and Eq. (5), thermal resistance was calculated, and the result is presented in Fig. 12. As
Fig. 13. Evaporator temperature variation with step increase and decrease of the reservoir temperature at a fixed heat load (100 W).
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Fig. 14. Evaporator temperature variation under abrupt reservoir temperature change at a fixed heat load (100 W).
For abrupt reservoir temperature changes, hysteresis was observed when the reservoir temperature dropped to 30 °C from 80 °C (1.44 °C). In other cases (70 °C ? 40 °C ? 70 °C and 50 °C ? 60 °C ? 50 °C) there was no appreciable temperature hysteresis. The differences between the results of the varying reservoir temperature tests and those of the fixed reservoir temperature tests were small except for a reservoir temperature of 50 °C (deviations as large as 1.35 °C were observed). From these results, it was clear that the FECPL operated stably with reservoir temperature change, and stable operating temperature control was guaranteed. 5.4. Transient responses Transient responses of the CPL, especially to the abrupt heat load decrease, are important since it can cause CPL wick dry-out. Thus, in this work, transient responses of the FECPL to abrupt heat load decreases (200 W ? 100 W, 200 W ? 80 W, 200 W ? 60 W, 200 W ? 40 W, 200 W ? 20 W) were examined. Fig. 15a and b shows the temperature variations of the evaporator, reservoir line, vapor line and liquid line under the illustrated heat load change. As shown in the figure, no CPL wick dry-outs were observed, which feature an ever increasing evaporator temperature and transport line temperatures at rest, but the operating temperature oscillations at the evaporator and the working fluid transport lines were observed. The oscillating temperatures, particularly of vapor and liquid lines, were thought to be the result of unstable condenser operation (more specifically, an unstable two-phase front in the condenser). When the heat load is decreased, due to the increased subcooling length of the condenser (or decreased two-phase region length), the working fluid is transferred from the reservoir to the condenser, forcing its two-phase front to regress toward the condenser inlet. After settlement of the two-phase front, a stable operation begins. However, in the case of abrupt heat load change, though the working fluid discharge from the reservoir was successfully completed (the reservoir line temperature decreased when heat load decreased, and no severe oscillation was observed), the two-phase front in the condenser seemed not to be stably settled, resulting in intermittent vapor discharge to the liquid line, which was seen as intermittent temperature spikes in the liquid line. This oscillating but still operating feature of the FECPL continued until the heat load decreased to 40 W. When the heat load decreased from 200 W to 20 W, temperature rises at the evaporator, liquid line and reservoir were observed, while the vapor line temperature drastically decreased, indicating a reverse flow of the working fluid from the liquid inlet of the evaporator to the vapor outlet (i.e., the thermosyphon mode of operation). This was due to the low heat
Fig. 15. Transient responses of the FECPL under abrupt heat load change (Tres = 60 °C).
load, which was insufficient to generate the required saturation pressure difference between the evaporating surface of the wick and the reservoir. After heat load increased to 200 W again, the FECPL operated normally as it did in steady state operations. 6. Conclusion The flat evaporator capillary pumped loop (FECPL) developed in this study operated satisfactorily and showed excellent operating temperature controllability in a tested heat load range. The following conclusions were drawn from this study: 1. The operating temperature of the FECPL was controlled by setting the reservoir temperature (i.e., its saturation pressure). This demonstrated the role of the reservoir as a reference point for thermodynamic states of other components. 2. The operating temperature of the FECPL increased with heat load at a fixed reservoir temperature. This was due to the non-condensable gases in the reservoir, which were compressed by liquid phase working fluid as the heat load increased, increasing the mixture pressure of the reservoir. 3. The temperature difference between the vapor outlet ( the saturation temperature on the wick evaporating surface) and the reservoir decreased as the reservoir temperature increased. This was due to the nonlinearly increasing saturation curve of water. 4. The FECPL showed negligible evaporator temperature hysteresis with reservoir temperature during step increases and decreases of the reservoir temperature. However, when the
W. Joung et al. / International Journal of Heat and Mass Transfer 53 (2010) 268–275
reservoir temperature was abruptly changed from 80 °C to 30 °C, there was slight temperature hysteresis. In both cases and at most reservoir temperatures, differences from the cases of the fixed reservoir temperatures were small. 5. The FECPL showed no wick dry-out when heat load abruptly decreased, except for oscillating temperatures in the vapor and liquid transport lines. This was attributed to the unstable two-phase front in the condenser, and intermittent vapor flow through the liquid line was detected. When heat load decreased from 200 W to 20 W, a thermosyphon mode of operation was observed, indicating a failure in capillary pumping due to the low level of heat load. References [1] F.J. Stenger, Experimental feasibility study of water filled capillary pumped heat transfer loops, NASA TM X-1310, (1996). [2] J. Ku, Thermodynamic aspects of capillary pumped loop operation, AIAA 942059, (1994). [3] M. Nikitkin, B. Cullimore, CPL and LHP technologies: what are the differences, what are the similarities? SAE Paper 981387, (1998). [4] D. Butler, J. Ku, T. Swanson, Loop heat pipes and capillary pumped loops – an application perspective, Space Technology and Applications International Forum (2002) 49–56. [5] Yu. Maidanik, et al., Thermoregulation of loops with capillary pumping for space use, SAE Paper 921169, (1992). [6] T.D. Swanson, G.C. Birur, NASA thermal control technologies for robotic spacecraft, Appl. Therm. Eng. 23 (2003) 1055–1065.
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[7] C. Wang, T. Leu, T. Lai, Micro capillary pumped loop system for a cooling high power device, Exper. Therm. Fluid Sci. 32 (2008) 1090–1095. [8] E. Bazzo, M. Nogoseke, Capillary pumping system for solar heating application, Appl. Therm. Eng. 23 (2003) 1153–1165. [9] J. Jung et al., A capillary pumped loop with microcone shaped capillary structure for cooling electronic devices, J. Micromech. Microeng. 18 (2008) 017002. [10] P. Chen, W. Lin, The application of capillary pumped loop for cooling of electronic components, Appl. Therm. Eng. 21 (2001) 1739–1754. [11] H. Lin, W. Lin, Thermal analysis of reservoir structure versus capillary pumped loop, Appl. Therm. Eng. 29 (2009) 186–194. [12] Wukchul Joung, Taeu Yu, Jinho Lee, Experimental study on the loop heat pipe with a planar bifacial wick structure, Int. J. Heat Mass Transfer 51 (2008) 1573–1581. [13] L. Ottenstein, J, Ku, D. Butler, Thermal vacuum testing of the capillary pumped loop flight experiment, SAE Paper 941599, (1994). [14] E. Bazzo, R.R. Riehl, Operating characteristics of a small-scale capillary pumped loop, Appl. Therm. Eng. 23 (2003) 687–705. [15] T.J. LaClair, I. Mudawar, Theoretical model for fast bubble growth in small channels with reference to startup of capillary pumped loops used in spacecraft thermal management systems, Int. J. Heat Mass Transfer 32 (2009) 716–723. [16] T. O’connel, T. Hoang, Effects of wick properties on pressure oscillations in a capillary pumped loop, SAE Paper 961434, (1996). [17] J. Yu et al., An experimental investigation on capillary pumped loop with the meshes wick, Int. J. Heat Mass Transfer (50) (2007) 4503–4507. [18] E. Pouzet et al., Dynamic response of a capillary pumped loop subjected to various heat load transients, Int. J. Heat Mass Transfer 47 (2004) 2293–2316. [19] A. Faghri, Heat Pipe Science and Technology, Taylor & Francis, New York, 1995. pp. 19–24. [20] D. Reay, P. Kew, Heat Pipes, fifth ed., Butterworth-Heinemann, 2006. pp.126– 139.
Contents lists available at ScienceDirect
International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
Experimental study on the operating characteristics of a capillary pumped loop with a flat evaporator Wukchul Joung, Hyungjin Hwang, Jinho Lee * School of Mechanical Engineering, Yonsei University, 134 Shinchon-dong, Seodaemoon-gu, Seoul 120-749, Republic of Korea
a r t i c l e
i n f o
Article history: Received 6 March 2009 Received in revised form 11 June 2009 Accepted 2 September 2009 Available online 23 October 2009 Keywords: Capillary pumped loop Flat evaporator Operating temperature control Reservoir temperature control
a b s t r a c t Recent advances in electronics technology have led to growing demand for highly efficient heat transfer devices with operating temperature controllability. Although capillary pumped loops are excellent candidates for this requirement and have been successful in various applications, their typical cylindrical evaporator shape has been a major drawback for applications to flat heat sources. Moreover, operating characteristics and the corresponding physical understanding of the capillary pumped loops are still uncertain, and lack of data for their operation makes their application unreliable. Thus, in this work, a capillary pumped loop with a flat evaporator was devised, and its operating characteristics were investigated. Ó 2009 Elsevier Ltd. All rights reserved.
1. Introduction A capillary pumped loop (CPL) is a highly effective two-phase heat transfer device. Compared to a similar two-phase heat transfer device, i.e., the loop heat pipe, the CPL has the distinctive feature of operating temperature controllability. This characteristic is due to a separate two-phase working fluid accumulator called a reservoir. The reservoir, which contains two-phase working fluid that remains in a saturation state, is directly connected to the liquid transport line between the evaporator and the condenser. By controlling its temperature, the operating temperature of a CPL (i.e., the saturation temperature on the evaporating surface of the wick) is controlled [1–4]. From its early development, due to this operating temperature controllability, the CPL has found applications in the thermal management of high precision equipment, such as avionics and optics in spacecraft. More recently, electronics cooling and thermal control of renewable energy systems are CPL application trends [5–10]. Despite its operating temperature controllability, additional thermal resistance is introduced by its cylindrical evaporator structure, which requires a supplementary intermediary called a saddle. This additional thermal resistance necessitates a flat evaporator CPL. Moreover, since most heat sources have flat thermo-contact surfaces and are closely packed, the development of a CPL with a flat evaporator and symmetric heat absorbing surfaces is of primary concern [10–12]. In addition to the requirement for struc-
* Corresponding author. Tel.: +82 2 2123 2816; fax: +82 2 312 2159. E-mail address: [email protected] (J. Lee). 0017-9310/$ - see front matter Ó 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.ijheatmasstransfer.2009.09.032
tural improvement, the lack of data for steady state operations of the CPL and physical understanding of the mechanism by which the operating temperature is controlled makes its applications unreliable. Therefore, in this work, we first focused on creating a flat evaporator CPL (hereafter the FECPL) with symmetric heat absorbing surfaces. Its operating characteristics, including both steady state and transient state responses, were investigated. In particular, steady state operations with various reservoir temperatures were analyzed to understand the role of the reservoir in the operating temperature control mode. The transient state responses under the changing heat load were examined to ascertain whether the developed CPL experienced wick dry-out when subjected to an abrupt heat load decrease, which is known to be the most probable reason for CPL wick dry-out [13–15]. 2. Operating principles and experimental parameters Fig. 1(a) and (b) shows the schematic of the CPL and the corresponding thermodynamic diagram for its operation. As shown in this figure, the saturation states of the evaporating surface of the wick and the reservoir are closely related by the following pressure balance equation.
DPsat: ¼ Psat:1 Psat:8 ¼ ql gðh1 þ h2 Þ þ DPf ;1!5
ð1Þ
where Psat.1 and Psat.8 are the saturation pressures on the evaporating surface of the wick and in the reservoir, respectively, and h1 + h2 is the net height difference between the two-phase interfaces in the reservoir and the condenser. DPf,1?5 is the frictional pressure drop
W. Joung et al. / International Journal of Heat and Mass Transfer 53 (2010) 268–275
269
Nomenclature g h P R Q_ T
gravity acceleration height pressure thermal resistance heat load temperature
g l ll ll, 1 ll, 2 loss res sat sat. 1 sat. 8 th vg vl wick
Greek symbols u phase q density Subscripts cd condenser cond. out condenser outlet evap evaporator f frictional
along the working fluid flow path from the evaporating surface of the wick to the bifurcation, and is given as follows.
DPf ;1!5 ¼ DP v g þ DPv l þ DP cd þ DPll;1
ð2Þ
where DPll,1 is the liquid phase frictional pressure drop from the outlet of the condenser to the bifurcation. As shown in Eq. (1), the
gravitational liquid phase liquid transport line liquid transport line from the condenser outlet to the bifurcation liquid transport line from the bifurcation to the end of the liquid transport line loss reservoir saturation saturation state of the evaporating surface of the wick saturation state of the reservoir thermal vapor removal grooves vapor transport line filtration
operating temperature of the CPL (i.e., the saturation temperature on the evaporating surface of the wick) is controlled by the reservoir (saturation) temperature which acts as a reference point for the thermodynamic states of the rest of the CPL. The saturation pressure difference described in Eq. (1) is the driving force of the CPL, and it circulates the working fluid through the loop. The capillary pressure difference across the wick evaporating surface provides a compensating force for the saturation pressure difference between the evaporating surface of the wick and the reservoir. This pressure difference makes the working fluid flow along the loop without penetrating the wick, and is given as follows.
DPcapillary ¼ DPsat: ql gh2 ql gh3 þ DPll;2 þ DPwick ¼ ql gðh1 h3 Þ þ DPf þ DPwick
ð3Þ
where DPll,2 is the liquid phase frictional pressure drop from the bifurcation to the end of the liquid transport line, DPwick is the filtration pressure drop through the wick, and DPf is the frictional pressure drop along the working fluid path including the whole liquid transport line. The frictional pressure drop, DPf, is given as follows.
DPf ¼ DPv g þ DPv l þ DPcd þ DP ll
ð4Þ
As explained above, the operating temperature of the CPL is determined by the reservoir temperature. In this work, to see the effect of the reservoir temperature on the steady state responses of the FECPL, we measured the steady state operating temperature variation with heat load at fixed reservoir temperatures, and the responses to reservoir temperature change at a fixed heat load. Another important factor to be investigated regarding the reliable operation of the CPL is its tendency of wick dry-out when subjected to an abrupt heat load decrease. It is generally known that the CPL wick dries out when the heat load is suddenly decreased due to the vapor generation and accumulation in the evaporator core, which blocks the working fluid supply to the wick [13,14]. Thus, we measured transient responses of the FECPL under abrupt heat load changes to see whether or not the wick dries out. 3. FECPL design
Fig. 1. Schematic and corresponding thermodynamic operation curve of the CPL.
The assembly process and structural view of the evaporator, the wick and the reservoir are shown in Fig. 2. As shown in the figure, employing two flat porous plates and a side rim, the wick had planar heat absorbing (or evaporating) surfaces to form the flat evaporator. The wick of the CPL plays an important role in CPL
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Fig. 2. Exploded view of the evaporator and the reservoir.
operation, which is a hydraulic barrier to the saturation pressure difference between the wick evaporating surface and the reservoir [16,17]. For the wick of the CPL to be a successful hydraulic barrier, it is essential to separate the two phases from each other as completely as possible. Otherwise, the required saturation pressure difference cannot be developed properly. To this end, a flange was added at the opening of the wick to provide a better contact surface with the evaporator housing. We used custom-made sintered porous plates (AISI 304) with a mean pore size and porosity of 5 lm and 0.5, respectively, as obtained from the manufacturer’s specifications. Dimensions of the flat porous plate were 52.77 mm 52.77 mm 3 mm (width height thickness). The side rim and the flange were made of stainless steel (AISI 316L), and all elements, including the porous plates, were laser welded for complete sealing. The wick was assembled with the evaporator housing, as shown in Fig. 2. All contact lines between the evaporator housing and the wick were also laser welded for the complete separation of vapor and liquid phases. The evaporator housing was made of stainless steel (AISI 316L), and had 13 vapor removal grooves on each inner surface. The reservoir, which had a hexahedral shape, was made of copper plates, and was designed to have around half of the total loop volume. The working fluid transport lines for the vapor and liquid phase working fluids were made of 0.6 and 0.46 m-long 1=4 in. stainless steel tubes. The transport lines were connected to the evaporator (vapor outlet and liquid inlet) and the reservoir. The condenser was made of a shell and tube type (one shell pass and two tube passes with no baffles) heat exchanger which consisted of a 0.8 m long copper tube with a diameter of 1=4 in. The whole assembly of the FECPL is presented in Fig. 3.
4. Test set-up The performance of the FECPL was evaluated with its operating temperatures and thermal resistance. For this purpose, 18 T-type thermocouples were attached along the working fluid flow path. Two absolute pressure transducers were placed in the reservoir and the vapor outlet to examine the saturation pressure difference
Fig. 3. External view of the FECPL, and temperature and pressure measurement locations.
between the evaporator and the reservoir. The errors associated with these sensors were a 0.1 °C deviation at 100 °C and 0.05% FS, respectively. Fig. 3 shows the locations of the temperature readings and the pressure measurement. The thermal performance of the FECPL was evaluated in terms of the vapor outlet temperature, which showed the working fluid’s temperature at the vapor outlet, and the evaporator temperature, which was defined as the average of the temperature readings over the evaporator surface (1)–(4). Another factor used in the performance evaluation was the thermal resistance, which was defined as the ratio of the temperature difference between the evaporator and the condenser outlet to the supplied heat load, and is given as
Rth ¼
T ev ap T cond:out ; Q_ ¼ 1
ð5Þ
C th
where Cth is the thermal conductance and is the reciprocal of the thermal resistance.
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Fig. 4. Schematic of the test set-up.
The test set-up is shown in Fig. 4. The reservoir temperature was controlled using an isothermal water circulation bath with a volume flow rate of 1 L/min, and the range of temperature change was from 30 °C to 80 °C with a 10 °C increment. The volume flow rate of the coolant supplied to the condenser was 2 L/min. The ambient temperature was maintained at 25 °C throughout the experiment. Water was used as the working fluid because it is the most easily accessible moderate temperature working fluid. However, due to its rather uncertain compatibility with stainless steel [19,20], we measured the inside pressure of the loop before the heat load was supplied during the whole experiment period (60 days). The fluid inventory in the FECPL was fixed at 70% of the total volume. The heat load was supplied by two electric loaders through flat ceramic heaters (50 mm 50 mm, 5.6 X), which were attached to the outer surface of the evaporator (Figs. 2 and 3), and the heat input range was from 100 W to 220 W. Heat loss to the ambient was assumed to be negligible throughout this work, as the heat loss in the hottest case was estimated to be less than 1% of the total heat input (from the vapor transport line, with Tevap = 81.9 °C at 220 W). The tests were repeated at least three times to obtain a reliable steady state data set, and the uncertainty analysis was done at a 95% confidence level. 5. Results and discussion
Fig. 5. Reservoir pressure variation with time before heat input.
fluid. As shown in the figure, the internal pressure of the reservoir gradually increased with time, though daily variations were small. This was attributed to the non-condensable gas generation inside the loop, since such an internal pressure rise had never been experienced using other combinations of working fluids and stainless steel by authors with similar configurations (e.g., LHPs).
5.1. Compatibility 5.2. Start-up Compatibility is an important factor in determining a working fluid and material combination, as an incompatible combination can cause non-condensable gas generation and an unnecessary internal pressure rise. Water, despite its wide availability and easy accessibility, is known to cause hydrogen generation when used with stainless steel [19,20]. Thus, before discussing the test results, it is appropriate to mention the compatibility test results obtained for the test period of 60 days. To check the non-condensable gas generation, we measured the internal pressure in the reservoir before heat input and compared it to the saturation pressure of water at 25 °C. Fig. 5 shows the reservoir pressure variation with time before heat input (i.e., at the ambient temperature of 25 °C). Compared to the saturation pressure of water at 25 °C (3.17 kPa), the initial reservoir pressure of 11.72 kPa was due to the initially injected non-condensable gases when filling the FECPL with the working
Fig. 6 shows the temperature variations of the evaporator, vapor outlet, liquid (transport) line and reservoir during the start-up (Q_ ¼ 100 W) of the FECPL when the reservoir temperature was set at 50 °C. As shown in the figure, the reservoir temperature was set at 50 °C prior to heat input, and then a heat load was applied (around 1400 s). After heat input, due to vapor generation on the evaporating surface of the wick and also due to the density difference between the vapor and liquid phases, the temperature of the vapor outlet rapidly increased, representing vapor discharge from the vapor outlet, and was followed by a decrease in the liquid line temperature, which showed successful working fluid circulation from the vapor outlet through the condenser to the liquid inlet. After the start-up, the FECPL reached a steady state operation (around 400 s after heat input). Start-ups of other reservoir temperatures showed similar trends, and no failure was observed.
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Fig. 7 shows the temperature variation of the vapor outlet with heat load at fixed reservoir temperatures. The vapor outlet temperature represents the temperature of the working fluid discharged from the evaporator, which is close to the saturation temperature on the wick evaporating surface. Thus, it is appropriate to use the vapor outlet temperature as an indicator of the saturation temperature of the evaporator. As shown clearly in the figure, the vapor outlet temperature increased as the reservoir temperature increased, indicating that the operating temperature of the FECPL was successfully controlled by the reservoir temperature. This was because the reservoir was in a saturation state. By setting the reservoir temperature, the saturation pressure of the reservoir was set, affecting the rest of the loop. This operating temperature controllability demonstrated the role of the reservoir as a reference point for thermodynamic states of other components. However, the variation with heat load showed that the vapor outlet temperature increased with heat load despite the constant reservoir temperature, and this was attributed to the effect of non-condensable gases in the reservoir. The reservoir, which was at the lowest pressure among CPL components, contained non-condensable gases thus having a mixture pressure of the saturation pressure of the working fluid and the partial pressure of the non-condensable gases. The saturation pressure of the working fluid in the reservoir was constant with heat load due to the constant reservoir temper-
ature. However, since the amount of liquid phase working fluid discharged from the condenser increased with heat load [18], and this liquid phase working fluid compressed the non-condensable gases in the reservoir, the partial pressure of the non-condensable gases increased with heat load, making the mixture pressure of the reservoir increase with heat load. Fig. 8 shows the reservoir pressure variation with heat load at various reservoir temperatures. As explained above, the internal pressure of the reservoir increased with heat load at constant reservoir temperatures due to the compression of the non-condensable gases. Fig. 9 shows the vapor outlet pressure variation with heat load, and due to the fact that the saturation pressure of the evaporator ( vapor outlet pressure) is the sum of the reservoir (saturation) pressure and the pressure drop along the working fluid flow path (frictional and hydrostatic), the tendency was similar to the reservoir pressure variation. In Fig. 7, at a reservoir temperature of 30 °C, the vapor outlet temperature suddenly increased at a heat load of 200 W, and this was attributed to the hard filled reservoir (i.e., loss of saturation state of the working fluid) which was confirmed by the rapid pressure rise of the reservoir at the same heat load of 200 W in Fig. 8. Fig. 7 also shows the decreasing temperature differences between the vapor outlet and the reservoir with reservoir temperature, which were, for example at 100 W, 24.8 °C, 18.6 °C, 13.1 °C, 8.1 °C, 3.9 °C and 0.8 °C for reservoir temperatures from 30 °C to 80 °C. These uneven temperature differences were due to the nonlinearly increasing nature of the water saturation curve. Thus the saturation pressure difference between the evaporator and the reservoir should have remained constant while the reservoir temperature changed. This was confirmed by the nearly constant pressure differences between the vapor outlet and the reservoir at a fixed heat load under the reservoir temperature change from Figs. 8 and 9 (pressure differences (minimum to maximum): 2.06–2.37 kPa at 100 W, 2.17–2.31 kPa at 120 W, 2.13–2.26 at 140 W, 2.15–2.29 kPa at 160 W, 2.26– 2.35 kPa at 180 W, 2.29–2.50 kPa at 200 W, 2.27–2.44 kPa at 220 W). Figs. 10 and 11 show the evaporator and condenser outlet temperature variations with heat load at fixed reservoir temperatures. Evaporator temperature showed a similar trend to the vapor outlet temperature, except for small deviations due to conduction through the housing material. For condenser outlet temperature variation, starting from the heat load of 100 W, sudden decreases in temperature at increasing heat loads with reservoir temperature were observed. This was thought to be the result of the interaction between the increasing vapor temperature with reservoir temperature and the increasing mass flow rate of the working fluid with
Fig. 7. Vapor outlet temperature variation with heat load at fixed reservoir temperatures.
Fig. 8. Reservoir pressure variation with heat load at fixed reservoir temperatures.
Fig. 6. Start-up of the FECPL at a fixed reservoir temperature of 50 °C.
5.3. Steady state responses
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Fig. 9. Vapor outlet pressure variation with heat load at fixed reservoir temperatures.
Fig. 10. Evaporator temperature variation with heat load at fixed reservoir temperatures.
Fig. 12. Thermal resistance variation with heat load temperatures.
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at fixed reservoir
shown in the figure, most of the regimes on which the FECPL operated were variable conductance regimes where the thermal resistance (or the thermal conductance) varies with heat load. The reason for the appearance of the variable conductance regime is explained by the evaporator and condenser outlet temperature variations in Figs. 10 and 11. In Figs. 10 and 11, the increase rate of temperature difference between the evaporator and the condenser outlet was less than the increase rate of heat load, thus based on Eq. (5), resulting in a gradual decrease of thermal resistance. Fig. 13 shows the evaporator temperature variation with step increase and decrease in the reservoir temperature at a heat load of 100 W, when the reservoir temperature increment and decrement were 10 °C (i.e., 30 °C ? 40 °C ? 50 °C ? 60 °C ? 70 °C ? 80 °C ? 70 °C ? 60 °C ? 50 °C ? 40 °C? 30 °C). As shown in the figure, there was negligibly small hysteresis for the step variations in the reservoir temperature, and the results were nearly identical to the cases of the fixed reservoir temperature (deviations were less than 1 °C). Fig. 14 shows the evaporator temperature variation under abrupt reservoir temperature change at a heat load of 100 W. The reservoir temperature was changed in the following sequence.
30 C ! 80 C ! 30 C ! 70 C ! 40 C ! 70 C ! 50 C ! 60 C ! 50 C:
Fig. 11. Condenser outlet temperature variation with heat load at fixed reservoir temperatures.
heat load. Based on these temperatures and Eq. (5), thermal resistance was calculated, and the result is presented in Fig. 12. As
Fig. 13. Evaporator temperature variation with step increase and decrease of the reservoir temperature at a fixed heat load (100 W).
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Fig. 14. Evaporator temperature variation under abrupt reservoir temperature change at a fixed heat load (100 W).
For abrupt reservoir temperature changes, hysteresis was observed when the reservoir temperature dropped to 30 °C from 80 °C (1.44 °C). In other cases (70 °C ? 40 °C ? 70 °C and 50 °C ? 60 °C ? 50 °C) there was no appreciable temperature hysteresis. The differences between the results of the varying reservoir temperature tests and those of the fixed reservoir temperature tests were small except for a reservoir temperature of 50 °C (deviations as large as 1.35 °C were observed). From these results, it was clear that the FECPL operated stably with reservoir temperature change, and stable operating temperature control was guaranteed. 5.4. Transient responses Transient responses of the CPL, especially to the abrupt heat load decrease, are important since it can cause CPL wick dry-out. Thus, in this work, transient responses of the FECPL to abrupt heat load decreases (200 W ? 100 W, 200 W ? 80 W, 200 W ? 60 W, 200 W ? 40 W, 200 W ? 20 W) were examined. Fig. 15a and b shows the temperature variations of the evaporator, reservoir line, vapor line and liquid line under the illustrated heat load change. As shown in the figure, no CPL wick dry-outs were observed, which feature an ever increasing evaporator temperature and transport line temperatures at rest, but the operating temperature oscillations at the evaporator and the working fluid transport lines were observed. The oscillating temperatures, particularly of vapor and liquid lines, were thought to be the result of unstable condenser operation (more specifically, an unstable two-phase front in the condenser). When the heat load is decreased, due to the increased subcooling length of the condenser (or decreased two-phase region length), the working fluid is transferred from the reservoir to the condenser, forcing its two-phase front to regress toward the condenser inlet. After settlement of the two-phase front, a stable operation begins. However, in the case of abrupt heat load change, though the working fluid discharge from the reservoir was successfully completed (the reservoir line temperature decreased when heat load decreased, and no severe oscillation was observed), the two-phase front in the condenser seemed not to be stably settled, resulting in intermittent vapor discharge to the liquid line, which was seen as intermittent temperature spikes in the liquid line. This oscillating but still operating feature of the FECPL continued until the heat load decreased to 40 W. When the heat load decreased from 200 W to 20 W, temperature rises at the evaporator, liquid line and reservoir were observed, while the vapor line temperature drastically decreased, indicating a reverse flow of the working fluid from the liquid inlet of the evaporator to the vapor outlet (i.e., the thermosyphon mode of operation). This was due to the low heat
Fig. 15. Transient responses of the FECPL under abrupt heat load change (Tres = 60 °C).
load, which was insufficient to generate the required saturation pressure difference between the evaporating surface of the wick and the reservoir. After heat load increased to 200 W again, the FECPL operated normally as it did in steady state operations. 6. Conclusion The flat evaporator capillary pumped loop (FECPL) developed in this study operated satisfactorily and showed excellent operating temperature controllability in a tested heat load range. The following conclusions were drawn from this study: 1. The operating temperature of the FECPL was controlled by setting the reservoir temperature (i.e., its saturation pressure). This demonstrated the role of the reservoir as a reference point for thermodynamic states of other components. 2. The operating temperature of the FECPL increased with heat load at a fixed reservoir temperature. This was due to the non-condensable gases in the reservoir, which were compressed by liquid phase working fluid as the heat load increased, increasing the mixture pressure of the reservoir. 3. The temperature difference between the vapor outlet ( the saturation temperature on the wick evaporating surface) and the reservoir decreased as the reservoir temperature increased. This was due to the nonlinearly increasing saturation curve of water. 4. The FECPL showed negligible evaporator temperature hysteresis with reservoir temperature during step increases and decreases of the reservoir temperature. However, when the
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reservoir temperature was abruptly changed from 80 °C to 30 °C, there was slight temperature hysteresis. In both cases and at most reservoir temperatures, differences from the cases of the fixed reservoir temperatures were small. 5. The FECPL showed no wick dry-out when heat load abruptly decreased, except for oscillating temperatures in the vapor and liquid transport lines. This was attributed to the unstable two-phase front in the condenser, and intermittent vapor flow through the liquid line was detected. When heat load decreased from 200 W to 20 W, a thermosyphon mode of operation was observed, indicating a failure in capillary pumping due to the low level of heat load. References [1] F.J. Stenger, Experimental feasibility study of water filled capillary pumped heat transfer loops, NASA TM X-1310, (1996). [2] J. Ku, Thermodynamic aspects of capillary pumped loop operation, AIAA 942059, (1994). [3] M. Nikitkin, B. Cullimore, CPL and LHP technologies: what are the differences, what are the similarities? SAE Paper 981387, (1998). [4] D. Butler, J. Ku, T. Swanson, Loop heat pipes and capillary pumped loops – an application perspective, Space Technology and Applications International Forum (2002) 49–56. [5] Yu. Maidanik, et al., Thermoregulation of loops with capillary pumping for space use, SAE Paper 921169, (1992). [6] T.D. Swanson, G.C. Birur, NASA thermal control technologies for robotic spacecraft, Appl. Therm. Eng. 23 (2003) 1055–1065.
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