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Experimental study of a nitrogen-charged cryogenic loop heat pipe
Cryogenics 52 (2012) 557–563
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Experimental study of a nitrogen-charged cryogenic loop heat pipe Lizhan Bai a,⇑, Guiping Lin a, Hongxing Zhang b, Jianyin Miao b, Dongsheng Wen c a
School of Aeronautic Science and Engineering, Beihang University, Beijing 100191, PR China Beijing Institute of Spacecraft System Engineering, Beijing 100086, PR China c School of Engineering and Materials Science, Queen Mary, University of London, London E1 4NS, UK b
a r t i c l e
i n f o
Article history: Received 26 May 2012 Received in revised form 26 July 2012 Accepted 26 July 2012 Available online 8 August 2012 Keywords: Loop heat pipe Cryogenic Operating characteristics Supercritical startup Experiment
a b s t r a c t Cryogenic loop heat pipes (CLHPs) are effective and efficient cryogenic heat transport devices suitable for many space applications. In this work, a miniature cryogenic loop heat pipe (CLHP) with nitrogen as the working fluid was designed and experimentally investigated. An auxiliary loop was employed to assist the supercritical startup of the primary evaporator. The operational characteristics of the CLHP and the matching characteristics of heat loads applied to the primary and secondary evaporators were investigated experimentally. The results show that the CLHP can achieve reliably the supercritical startup when the heat load applied to the secondary evaporator is no less than 3 W; when the heat load applied to the primary evaporator is no less than 2.5 W, the primary evaporator can operate independently, otherwise a proper selection of the heat load applied to the secondary evaporator should be considered to overcome the parasitic heat load from the ambient. The CLHP is working at the variable conductance mode and can achieve smooth operational transition subject to a large step change of the heat load applied to the primary evaporator. Ó 2012 Elsevier Ltd. All rights reserved.
1. Introduction As an effective and efficient two-phase heat transfer device, loop heat pipe (LHP) utilizes the evaporation and condensation of a working fluid to transfer heat, and the capillary force developed in fine porous wicks to circulate the working fluid [1]. Their high pumping capability and excellent heat transport performance have been traditionally utilized to address the thermal control problems of spacecraft, and successfully applied in many space missions [2,3]. Recently, its application has been extended to terrestrial surroundings such as in electronics cooling [4–6] and thermal-management system of aircraft [7,8]. Their relative long distance heat transport capability and flexibility in design could offer many advantages compared with traditional heat pipes, especially under the situation of antigravity operation. Currently, most LHPs investigated or in use are operating at the ambient temperature range (0–60 °C), i.e. so-called the ambient loop heat pipes (ALHPs), and the working fluids charged are typically ammonia, water and acetone, etc. However, for the low temperature applications, such as in a space exploration system where the space infrared sensors/detectors have to be maintained at 80– 100 K or even lower temperature, ALHPs are not applicable. It is necessary to develop LHPs operating in the low temperature range, i.e. cryogenic loop heat pipes (CLHPs). Inheriting the advantage of ⇑ Corresponding author. Tel.: +86 10 8233 8469; fax: +86 10 8233 8600. E-mail address: [email protected] (L. Bai). 0011-2275/$ - see front matter Ó 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.cryogenics.2012.07.005
long distance and flexible heat transport features of ALHPs, CLHPs can realize the separation of the infrared optical instruments from the cryocoolers with effective long distance cryogenic heat transport between them. The advantages of the application of CLHPs onto space optical instrument are evident: first, the pointing agility of the optical instruments can be improved considerably; secondly, the vibration induced by the cryocoolers can be isolated from the optical instruments, which can provide jitter-free observations of the space telescope at a target that may prove invaluable for most space missions. Since 2000, many investigations have been performed on the design and functional study of CLHPs, which are briefly reviewed here. Pereira et al. [9] designed and investigated experimentally a CLHP with different working fluids, which utilized the gravity to realize the temperature drop and liquid saturation of the evaporator wick. The CLHP was able to transfer up to 20 W when filled with argon, 25 W when filled with krypton and 30 W when filled with propane in a gravity assisted orientation. The limitation of heat transfer in each case was only due to the limited cooling power of the cryocooler used to perform the tests. Khrustalev et al. [10–12] experimentally investigated an oxygen-charged CLHP, which employed a secondary evaporator to realize the temperature drop and liquid saturation of the primary evaporator wick. The CLHP can operate at the temperature range of 65– 140 K, and the experimental results showed that the CLHP could startup and operate reliably when the heat load applied to the main evaporator varied from 0.5 W to 9 W with zero power on
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the secondary evaporator. The maximum 9 W could be obtained under anti-gravity condition, i.e. with the main evaporator elevated 5 cm above the condenser. Mo et al. [13–15] designed and experimentally investigated a nitrogen-charged CLHP, which also employed a secondary evaporator to realize the temperature drop and liquid saturation of the primary evaporator wick, and the effects of gravity, volume of the gas reservoir, effective pore diameter of the wick and different working fluids on the operating performance of the CLHP were reported. It is worth noted that the structure of the secondary evaporator in Refs. [10–12] is similar to that of the primary evaporator, while the structure of the secondary evaporator in Refs. [13–15] is like a traditional grooved heat pipe. James et al. [16] developed a CLHP and conducted the tests in a thermal vacuum chamber for passive optical bench cooling applications. Ethane was selected as the working fluid to provide an operating temperature range of 215–218 K. The experimental results clearly demonstrated the capability of the CLHP, i.e. it could startup reliably from a supercritical temperature of 335 K to achieve a normal operating temperature of 215 K by switching on the secondary evaporator. With a heat load of 5 W applied to the secondary evaporator, the CLHP could achieve a 50 W heat transport capability at 215 K. Hoang and O’Connell [17] developed a nitrogen-charged CLHP, which could realize the supercritical startup and operate at the temperature range of 80–110 K. The CLHP showed good performance in power cycling and long duration low heat load tests, and its maximum heat transport capability was found to be 5 W with a transport distance of 4.3 m. In Ref. [18], a hydrogen-charged CLHP was experimentally investigated, which could realize the supercritical startup and operated at the temperature range of 20–30 K. A maximum heat transport capability of 5 W with a transport distance of 2.5 m was obtained. An optimization of the CLHP to minimize its mass and volume was conducted for future space applications, and further results were reported in Ref. [19]. To increase the heat transport capacity, Zhao et al. [20] experimentally investigated a nitrogen-charged CLHP with a parallel condenser, which could greatly reduce the flow resistance and increase the cooling capability of the condenser. The results showed that the CLHP could operate reliably for a maximum heat load of 41 W with a small temperature drop of 6 K across a 0.48 m transport distance. Gully et al. [21] designed and experimentally investigated a prototype of CLHP working around 80 K with nitrogen as the working fluid. Experimental results were analyzed and discussed both in the transient phase of cooling from room temperature and in stationary conditions. The effects of transferred power, filling pressure and radiation heat load for two basic configurations of cold reservoir of the secondary circuit were studied in stationary conditions, and a maximum cold power of 19 W with a corresponding limited temperature difference of 5 K was achieved across a 0.5 m distance. Of particular note that all CLHPs introduced in Refs. [16–21] employed an auxiliary loop to realize the large temperature drop of the primary evaporator during the supercritical startup process. Above short review shows that CLHPs are effective and efficient cryogenic heat transport devices developed in recent years, and the study is just at the beginning mainly focusing on experimental aspects. Within these studies, experimental investigation on its working principles and operating characteristics are still inadequate. Furthermore, to push CLHP into space applications, besides each component should be optimized to reduce its volume and weight, appropriate design of the structure of the evaporator casing and the condenser is also required to satisfy the interface requirements between the CLHP and the heat source (infrared sensors/detectors) and heat sink (cryocoolers). Aiming at future space applications, a miniature CLHP with nitrogen as the working fluid, which employed an auxiliary loop to realize the temperature drop of the primary evaporator during the supercritical startup, was
designed and experimentally investigated in this paper. The outer diameters of the evaporator and the transport line were 13 mm and 2 mm respectively, and the cylindrical condenser design can provide convenient interface with the cold finger of the cryocooler, making the CLHP with considerable application potential. The experiments were mainly focused on the supercritical startup capability and matching characteristics of the heat loads applied to the primary and secondary evaporators, and its heat transport capacity and thermal resistance variation were also tested. The experimental results are presented and analyzed in detail in this paper. 2. Experimental setup Fig. 1 shows the schematic of the CLHP designed in this work, and Table 1 gives the basic parameters of the CLHP, where CC represents the compensation chamber and OD and ID represent the outer and inner diameters respectively. As shown in Fig. 1, because the CLHP employs an auxiliary loop to realize the supercritical startup, the system is composed of a main loop and an auxiliary loop apart from the gas reservoir. The main loop includes a primary evaporator, a primary CC, a primary condenser and the primary vapor and liquid lines, and the auxiliary loop consists of a secondary evaporator, a secondary CC, a secondary condenser and the secondary loop line. The gas reservoir with a comparatively large volume is utilized to reduce the system pressure at ambient state, and it is connected with the inlet of the primary condenser. All the components of the CLHP were made of stainless steel except that the primary and secondary evaporator wicks were made of sintered nickel powder. The CLHP was operating at the temperature range of 80–120 K, and nitrogen whose purity was greater than 99.9995% was used as the working fluid. In the experiments, heat loads applied to the primary and secondary evaporators were provided by two thin-film electric resistance heaters, and they were attached directly to the casings of the primary and secondary evaporators symmetrically. The heat load can be adjusted from 0 W to 30 W by altering the DC power output voltage imposed on the heaters. The primary and secondary condenser lines were coiled and soldered onto a solid cylindrical cold block (see Fig. 1) to reduce the contact thermal resistance. The cryogenic heat sink was simulated by liquid nitrogen circulating through another hollow cylindrical cold block (see Fig. 2), and the two cylindrical cold blocks were connected tightly through the end faces. For practical application, the hollow cylindrical cold block would be replaced by the cryocooler. To reduce the parasitic heat loss from the ambient, the CLHP was placed in a thermal-vacuum chamber, and all the components except the gas reservoir were covered with multilayer insulation materials. Note that, the pressure in the thermal vacuum chamber can be maintained at below 2 102 Pa, and the parasitic heat loss by convective heat transfer becomes negligible in the experiments. Type T thermal couples (TCs) were used to monitor the temperature variations of the characteristic points along the loop, the measuring uncertainties of the thermal couples are ±1.0 K by calibration, and the TC locations are shown in Fig. 1. To reduce the influence of gravity, the primary evaporator and the primary condenser were placed in a horizontal plane in the experiments. 3. Experimental results and analysis 3.1. Supercritical startup 3.1.1. Temperature variations of characteristic points along the loop Fig. 3 shows the temperature variations of the characteristic points along the loop during the supercritical startup process. In
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Fig. 1. Schematic of the CLHP and thermal couple locations.
Table 1 Basic parameters of the CLHP. Component
Parameter
Primary evaporator
Casing OD/ID length Wick OD/ID length
U13/11 35 mm U11/4 30 mm
Secondary evaporator
Casing OD/ID length Wick OD/ID length
U13/11 35 mm U11/4 30 mm
Main loop
Vapor line OD/ID length Liquid line OD/ID length Condenser line OD/ID length
U2/1 640 mm U2/1 560 mm U2/1 450 mm
Secondary loop
Transport line OD/ID length Condenser line OD/ID length
U2/1 600 mm U2/1 150 mm
Volume of primary CC/secondary CC
2.01/3.62
Volume of gas reservoir
555 ml
Wick
1.0 lm 55% >5 1014 m2
Maximum capillary radius Porosity Permeability
300
Qse= 3W Qpe= 2W
Temperature/K
250 TC5
TC8
200
TC1 TC2
TC12
150
TC3
TC11
100
50 Fig. 2. Schematic of the hollow cylindrical cold block.
Fig. 3, Qpe and Qse represent the heat loads applied to the primary and secondary evaporators respectively, and the charged pressure of the working fluid was 1.2 MPa. As shown in Fig. 3, at the initial state, all the characteristic points along the loop were at the ambient temperature of 291.6 K. About 10 min later, liquid nitrogen began to circulate through the hollow cylindrical cold block, and the temperature of the cold block (TC13) dropped rapidly. At the same time, the temperatures of the inlet (TC3 and TC8) and outlet (TC4 and TC9) of the primary and secondary condenser lines dropped fast due to their close contact with the cold block. The secondary evaporator was cooled by the cold block through the secondary
TC13
0
25
50
TC4 TC9
75
100
125
150
Time/min Fig. 3. Supercritical startup process of the CLHP (Qse = 3 W, Qpe = 2 W).
CC, and its temperature (TC11) also dropped fast. About 18 min later, the temperature of the cold block (TC13) began to drop slowly and maintained at about 83 K. About 26 min later, the temperature of the secondary evaporator (TC11) began to drop slowly approaching 104 K, indicating that nitrogen in the two-phase state began to exist in the secondary evaporator, which would keep the secondary evaporator temperature around the saturation temperature of nitrogen corresponding to local pressure. As time went by, the inventory of liquid nitrogen in the secondary
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evaporator would increase gradually, and it would saturate the wick of the secondary evaporator in the end. After the temperature of the secondary evaporator maintaining at 104 K for about 35 min, a heat load of 3 W was applied to the secondary evaporator (Qse), and the temperature of the secondary evaporator (TC11) rose slightly, then maintained at about 110 K. The temperature at the outlet of the secondary evaporator (TC12) dropped sharply, indicating that the liquid nitrogen in the secondary evaporator began to evaporate, and the generated cryogenic nitrogen vapor flowed out of the secondary evaporator. At the same time, the temperatures at the inlet of the primary and secondary condenser lines (TC3 and TC8) rose rapidly, which was caused by the displacement of the ambient nitrogen gas in the transport lines into the primary and secondary condensers. Meanwhile, the temperatures at the outlet of the primary and secondary condenser lines (TC4 and TC9) dropped rapidly caused by the condensate in the primary and secondary condensers being pushed out, which confirmed the normal circulation of the working fluid in the loop. The operation of the secondary evaporator could realize the transfer of the condensed liquid nitrogen from the primary condenser to the primary evaporator and CC through the primary liquid transport line, and so the temperature at the outlet of the primary liquid transport line (TC5) dropped rapidly. Meanwhile, the temperature of the primary evaporator (TC1) dropped gradually. With the operation of the secondary evaporator for about 22 min, the temperature of the primary evaporator dropped to about 109 K and then remained unchanged. The temperature at the outlet of the primary evaporator (TC2) began to drop rapidly, indicating that nitrogen in the twophase state began to exist in the wick of the primary evaporator, the wick began to be saturated with liquid nitrogen, and the generated cryogenic nitrogen vapor flowed out of the primary evaporator. As time went by, the inventory of liquid nitrogen in the primary evaporator would increase gradually and fully saturate the wick of the primary evaporator. With the temperature of the primary evaporator remaining at 108 K for about 28 min, a heat load of 2 W was applied to the primary evaporator (Qpe), and the temperature of the primary evaporator rose slightly, then maintained at about 115 K. With the normal operation of the primary evaporator, the CLHP reliably realized the supercritical startup. 3.1.2. Effect of heat load applied to the secondary evaporator Figs. 4 and 5 show the temperature variations of the characteristic points along the loop during the supercritical startup process. Note that, In Figs. 4–12, the charged pressure of the working fluid was all 1.6 MPa. The heat loads applied to the secondary evaporator were 5 W and 3 W in Figs. 4 and 5 respectively. The tempera300
Qse= 3W
300
Qpe= 2W
250
Temperature/K
560
200 TC1
150
TC11
100
50
TC13
0
20
40
60
80
100
120
Time/min Fig. 5. Supercritical startup process of the CLHP (Qse = 3 W, Qpe = 2 W).
ture variation process of the characteristic points along the loop in Fig. 5 was similar to that in Fig. 4 but taking longer duration in the startup. As the heat load applied to the secondary evaporator became smaller, the transferred cryocooling from the primary condenser to the primary evaporator per unit time decreased. At the same time, as the mass flowrate of the working fluid became smaller, the temperature rise of the working fluid flowing along the primary liquid transport line increased due to the parasitic heat load from the ambient, and the temperature at the inlet of the primary CC became higher. All these factors mentioned above were adverse to the fast temperature drop of the primary evaporator. Therefore it took 40 min for the primary evaporator to drop to 104 K from ambient temperature in Fig. 5, whereas only 15 min was required in Fig. 4. Although the temperature drop of the primary evaporator became slower, the CLHP also achieved supercritical startup at last, as shown in Fig. 5. Further reducing the heat load applied to the secondary evaporator to 2 W during the supercritical startup process, it took nearly two hours for the temperature of the primary evaporator (TC1) to drop to about 197 K, as shown in Fig. 6. Beyond that, it dropped rather slightly, indicating that the transferred cryocooling from the primary condenser was balanced by the parasitic heat load from the ambient to the primary evaporator. As this temperature was obviously higher than the critical temperature of nitrogen (126 K), no nitrogen in the two-phase state would exist in the primary evaporator, and the wick of the primary evaporator could not be saturated by liquid nitrogen. Under such a situation, the 300
Qse= 5W
Qse= 2W
Qpe= 2W
250
Temperature/K
Temperature/K
250
200 TC1
TC11
150
100
TC1
200
150
TC11
100 TC13
50 0
15
30
45
60
75
90
Time/min Fig. 4. Supercritical startup process of the CLHP (Qse = 5 W, Qpe = 2 W).
50
TC13
0
30
60
90
120
150
Time/min Fig. 6. Temperature drop process of the primary evaporator (Qse = 2 W).
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180
Qse= 0W Qpe= 4W
Qpe= 5W
180
Qpe= 2.5W
Qpe= 3W
TC3
TC1 TC4
80 60
Temperature/K
Temperature/K
140
100
Qpe= 14W
140 120
TC2
100
TC1
TC5
TC13
TC3
TC13
60 0
25
50
75
100
125
0
30
180
Qse= 1W Qpe= 2W
Qpe= 4W
Thermal resistance/(K/W)
TC3
TC11 TC5
TC4
80
150
12
TC1
100
120
Fig. 10. Heat transport capacity test of the primary evaporator (Qse = 0 W).
Qpe= 1.5W Qpe= 1.0W
140 TC2
90
14
160
120
60
Time/min
Fig. 7. Temperature variation with heat load applied to the primary evaporator (Qse = 0 W).
Temperature/K
TC4
80
Time/min
10 8 6 4 2
TC13
0
60 0
10
20
30
40
50
60
70
80
Fig. 8. Temperature variation with heat load applied to the primary evaporator (Qpe = 1 W).
180
Qse= 2W Qpe= 1W
Qpe= 2W
Qpe= 0.5W
160 140
TC11
100 80
TC2 TC5
TC1
TC4
TC13
60 0
10
20
30
40
4
6
8
10
12
50
60
Fig. 11. Thermal resistance variation of the CLHP.
secondary evaporator must reach a certain value. According to Figs. 5 and 6, the minimum heat load applied to the secondary evaporator required should be between 2 W and 3 W to realize the supercritical startup of the CLHP developed in this work. 3.2. Matching characteristics of heat loads applied to the primary and secondary evaporators
TC3
120
2
Heat load/W
Time/min
Temperature/K
Qpe= 6W Qpe= 8W Qpe= 10W Qpe= 12W
160
Qpe= 2W
TC2
Qse= 0W
Qpe= 4W
160
120
Qse= 2W
70
80
Time/min Fig. 9. Temperature variation with heat load applied to the primary evaporator (Qse = 2 W).
primary evaporator could not be started, i.e. the CLHP could not realize supercritical startup with a heat load of 2 W applied to the secondary evaporator. Such an experimental result was also consistent with the theoretical analysis in Ref. [22]. To realize the supercritical startup of the CLHP, the heat load applied to the
For CLHPs, as they operate in the cryogenic temperature range, the large temperature difference between the ambient and the working fluid in the CLHP would produce comparatively large parasitic heat loss. Most of the experimental studies on CLHPs are generally conducted in the thermal-vacuum chamber, and the whole CLHP except the gas reservoir is covered with multilayer insulation materials (MLI) to reduce the heat transfer between the ambient and the CLHP. Although such strict thermal insulation measures are employed, it is still possible that the primary evaporator fails to operate independently due to the effect of the parasitic heat load from the ambient at a relatively small heat load [23]. Under this situation, the secondary evaporator must be kept in operation to assist the operation of the primary evaporator. For a specific heat load applied to the secondary evaporator, there shall exist a minimum heat load applied to the primary evaporator to maintain the normal operation of the primary evaporator. Fig. 7 shows the temperature variations of the characteristic points along the loop when the heat load applied to the secondary
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Qse= 0W
Qse= 5W
120
Qpe= 4W
Qpe= 8W
100
TC1
TC3
90
TC4
80
Qse= 0W Qpe= 4W
Qpe= 12W
110
TC2
Temperature/K
Temperature/K
110
70
Qse= 5W
120
Qpe= 12W Qpe= 8W
TC2
100
TC3
TC1
90
TC4
80
TC13
TC13
70 0
25
50
75
100
125
0
10
20
30
40
50
Time/min
Time/min
(a) Variation amplitude of 4W
(b) Variation amplitude of 8W
60
70
Fig. 12. Response to step change of heat load applied to the primary evaporator.
evaporator was 0 W, i.e. the secondary evaporator was not in operation, and the heat load applied to the primary evaporator was decreased gradually. Under such conditions, the primary evaporator could operate normally as the heat load applied to the primary evaporator is in the regime of 2.5–5 W. However, when the heat load was further reduced to 2.0 W, the primary evaporator temperature (TC1) rose quickly, and the primary evaporator failed to operate. That is because when the heat load applied to the primary evaporator is very small, i.e. <2.5 W in this paper, it cannot overcome the effect of the parasitic heat loss from the ambient, as analyzed in detail in Ref. [23]. Under such a situation, the secondary evaporator must be kept in operation to assist the operation of the primary evaporator. Fig. 8 shows the temperature variations of the characteristic points along the loop when the heat load applied to the secondary evaporator was 1 W and the heat load applied to the primary evaporator was decreased gradually. Different from the situation in Fig. 7, the CLHP can still operate normally as the heat load applied to the primary evaporator was decreased from 4 W to 1.5 W. However when the heat load applied to the primary evaporator was further reduced to 1.0 W, the primary evaporator temperature (TC1) rose quickly, and the primary evaporator failed to operate. As the heat load applied to the secondary evaporator was increased further to 2 W in Fig. 9, the primary evaporator could operate normally at a heat load as low as 0.5 W. It is known that the mass flowrate of the working fluid in the primary liquid transport line, which is in proportional to the sum of the heat loads applied to the primary and secondary evaporators, determines the ability of the CLHP to overcome the effect of the parasitic heat load from the ambient. The results in Figs. 7–9 show clearly that when the sum of the heat loads applied to the primary and secondary evaporators was no less than 2.5 W, the CLHP could overcome the effect of the parasitic heat load and the primary evaporator could operate normally; however, when the sum of the heat loads was smaller than 2.5 W, the primary evaporator could not operate normally due to the effect of parasitic heat load, and the secondary evaporator must be kept in operation to assist the operation of the primary evaporator. 3.3. Heat transport capacity of the primary evaporator The heat transport capacity is one of the most important aspects to assess the CLHP performance, which determines the
maximum heat load that the primary evaporator can sustain. Fig. 10 shows the test results of the heat transport capacity of the primary evaporator. In the experiments, the heat load applied to the secondary evaporator was firstly set at 2 W, then it was reduced to 0 W where the primary evaporator could operate independently with a heat load of 4 W. As shown in Fig. 10, when the heat load applied to the primary evaporator was increased from 4 W to 12 W, the primary evaporator could operate normally. However, when the heat load was further increased to 14 W, the temperature of the primary evaporator rose quickly, and the primary evaporator failed to operate, indicating that the primary evaporator reached the capillary limit, i.e. 12 W 0.56 m. 3.4. Thermal resistance variation In this paper, the thermal resistance of the CLHP is defined as follows:
R¼
T w;pe T if;cb Q pe
ð1Þ
In Eq. (1), R is the thermal resistance of the CLHP; T w;pe and T if;cb are the temperatures of the primary evaporator wall and the cold block interface respectively and Qpe is the heat load applied to the primary evaporator. In this work, the temperature of the primary evaporator wall is represented by TC1, and the temperature of the cold block interface is represented by TC13. Fig. 11 shows the heat-load dependence of the thermal resistance of the CLHP. As the heat load applied to the primary evaporator increased, the thermal resistance of the CLHP decreased monotonically, i.e. from 12.1 K/W at the heat load of 2 W to 2.1 K/W at the heat load of 12 W, indicating that the CLHP was always operating in the variable conductance mode. In this mode, the two-phase zone of the primary condenser kept increasing with the increase of the heat load applied to the primary evaporator. Assuming the thermal conductivity of copper is 400 W/(m K), the thermal resistance of the solid copper rod with the same volume of the CLHP is about 160 K/W, and the thermal resistance of the CLHP is therefore much smaller than that of the equivalent copper rod, indicating that the CLHP has an excellent cryogenic heat transport performance.
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3.5. Response to step change of heat load applied to the primary evaporator Fig. 12 shows the temperature variations of the characteristic points along the loop when the heat load applied to the primary evaporator was subject to a step change with a comparatively large amplitude. In this experiment, the heat load applied to the secondary evaporator was 5 W at first, then it was reduced to 0 W, and the primary evaporator could operate independently with a heat load of 4 W. According to Fig. 12, when the heat load applied to the primary evaporator was subject to a step change of 4 W or 8 W, as shown in Fig. 12a and b respectively, the primary evaporator temperature (TC1) changed smoothly, the primary evaporator could operate normally, and the CLHP manifested good response characteristic and thermal control performance. 4. Conclusions A CLHP using nitrogen as the working fluid was designed and experimentally investigated in this work, and specific conclusions below can be drawn: (1) With a heat load of 3 W applied to the secondary evaporator, the CLHP can reliably realize the supercritical startup, and the larger the heat load applied to the secondary evaporator, the sooner the temperature drop process of the primary evaporator. (2) When the heat load applied to the secondary evaporator is 62 W, the primary evaporator cannot be cooled below the critical temperature of nitrogen due to the effect of parasitic heat load from the ambient, and the CLHP cannot realize the supercritical startup under this situation. (3) When the heat load applied to the primary evaporator is no less than 2.5 W, the primary evaporator can operate independently, i.e. the operation of the secondary evaporator is not required; however, when it is smaller than 2.5 W, the secondary evaporator must be kept in operation to assist the operation of the primary evaporator. (4) The CLHP has a heat transport capacity of 12 W 0.56 m, and the thermal resistance of the CLHP decreases with the increase of the heat load applied to the primary evaporator showing a variable conductance working mode. (5) The CLHP has the ability to operate with large step change of the heat load applied to the primary evaporator, which manifests good response characteristic and thermal control performance.
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Contents lists available at SciVerse ScienceDirect
Cryogenics journal homepage: www.elsevier.com/locate/cryogenics
Experimental study of a nitrogen-charged cryogenic loop heat pipe Lizhan Bai a,⇑, Guiping Lin a, Hongxing Zhang b, Jianyin Miao b, Dongsheng Wen c a
School of Aeronautic Science and Engineering, Beihang University, Beijing 100191, PR China Beijing Institute of Spacecraft System Engineering, Beijing 100086, PR China c School of Engineering and Materials Science, Queen Mary, University of London, London E1 4NS, UK b
a r t i c l e
i n f o
Article history: Received 26 May 2012 Received in revised form 26 July 2012 Accepted 26 July 2012 Available online 8 August 2012 Keywords: Loop heat pipe Cryogenic Operating characteristics Supercritical startup Experiment
a b s t r a c t Cryogenic loop heat pipes (CLHPs) are effective and efficient cryogenic heat transport devices suitable for many space applications. In this work, a miniature cryogenic loop heat pipe (CLHP) with nitrogen as the working fluid was designed and experimentally investigated. An auxiliary loop was employed to assist the supercritical startup of the primary evaporator. The operational characteristics of the CLHP and the matching characteristics of heat loads applied to the primary and secondary evaporators were investigated experimentally. The results show that the CLHP can achieve reliably the supercritical startup when the heat load applied to the secondary evaporator is no less than 3 W; when the heat load applied to the primary evaporator is no less than 2.5 W, the primary evaporator can operate independently, otherwise a proper selection of the heat load applied to the secondary evaporator should be considered to overcome the parasitic heat load from the ambient. The CLHP is working at the variable conductance mode and can achieve smooth operational transition subject to a large step change of the heat load applied to the primary evaporator. Ó 2012 Elsevier Ltd. All rights reserved.
1. Introduction As an effective and efficient two-phase heat transfer device, loop heat pipe (LHP) utilizes the evaporation and condensation of a working fluid to transfer heat, and the capillary force developed in fine porous wicks to circulate the working fluid [1]. Their high pumping capability and excellent heat transport performance have been traditionally utilized to address the thermal control problems of spacecraft, and successfully applied in many space missions [2,3]. Recently, its application has been extended to terrestrial surroundings such as in electronics cooling [4–6] and thermal-management system of aircraft [7,8]. Their relative long distance heat transport capability and flexibility in design could offer many advantages compared with traditional heat pipes, especially under the situation of antigravity operation. Currently, most LHPs investigated or in use are operating at the ambient temperature range (0–60 °C), i.e. so-called the ambient loop heat pipes (ALHPs), and the working fluids charged are typically ammonia, water and acetone, etc. However, for the low temperature applications, such as in a space exploration system where the space infrared sensors/detectors have to be maintained at 80– 100 K or even lower temperature, ALHPs are not applicable. It is necessary to develop LHPs operating in the low temperature range, i.e. cryogenic loop heat pipes (CLHPs). Inheriting the advantage of ⇑ Corresponding author. Tel.: +86 10 8233 8469; fax: +86 10 8233 8600. E-mail address: [email protected] (L. Bai). 0011-2275/$ - see front matter Ó 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.cryogenics.2012.07.005
long distance and flexible heat transport features of ALHPs, CLHPs can realize the separation of the infrared optical instruments from the cryocoolers with effective long distance cryogenic heat transport between them. The advantages of the application of CLHPs onto space optical instrument are evident: first, the pointing agility of the optical instruments can be improved considerably; secondly, the vibration induced by the cryocoolers can be isolated from the optical instruments, which can provide jitter-free observations of the space telescope at a target that may prove invaluable for most space missions. Since 2000, many investigations have been performed on the design and functional study of CLHPs, which are briefly reviewed here. Pereira et al. [9] designed and investigated experimentally a CLHP with different working fluids, which utilized the gravity to realize the temperature drop and liquid saturation of the evaporator wick. The CLHP was able to transfer up to 20 W when filled with argon, 25 W when filled with krypton and 30 W when filled with propane in a gravity assisted orientation. The limitation of heat transfer in each case was only due to the limited cooling power of the cryocooler used to perform the tests. Khrustalev et al. [10–12] experimentally investigated an oxygen-charged CLHP, which employed a secondary evaporator to realize the temperature drop and liquid saturation of the primary evaporator wick. The CLHP can operate at the temperature range of 65– 140 K, and the experimental results showed that the CLHP could startup and operate reliably when the heat load applied to the main evaporator varied from 0.5 W to 9 W with zero power on
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the secondary evaporator. The maximum 9 W could be obtained under anti-gravity condition, i.e. with the main evaporator elevated 5 cm above the condenser. Mo et al. [13–15] designed and experimentally investigated a nitrogen-charged CLHP, which also employed a secondary evaporator to realize the temperature drop and liquid saturation of the primary evaporator wick, and the effects of gravity, volume of the gas reservoir, effective pore diameter of the wick and different working fluids on the operating performance of the CLHP were reported. It is worth noted that the structure of the secondary evaporator in Refs. [10–12] is similar to that of the primary evaporator, while the structure of the secondary evaporator in Refs. [13–15] is like a traditional grooved heat pipe. James et al. [16] developed a CLHP and conducted the tests in a thermal vacuum chamber for passive optical bench cooling applications. Ethane was selected as the working fluid to provide an operating temperature range of 215–218 K. The experimental results clearly demonstrated the capability of the CLHP, i.e. it could startup reliably from a supercritical temperature of 335 K to achieve a normal operating temperature of 215 K by switching on the secondary evaporator. With a heat load of 5 W applied to the secondary evaporator, the CLHP could achieve a 50 W heat transport capability at 215 K. Hoang and O’Connell [17] developed a nitrogen-charged CLHP, which could realize the supercritical startup and operate at the temperature range of 80–110 K. The CLHP showed good performance in power cycling and long duration low heat load tests, and its maximum heat transport capability was found to be 5 W with a transport distance of 4.3 m. In Ref. [18], a hydrogen-charged CLHP was experimentally investigated, which could realize the supercritical startup and operated at the temperature range of 20–30 K. A maximum heat transport capability of 5 W with a transport distance of 2.5 m was obtained. An optimization of the CLHP to minimize its mass and volume was conducted for future space applications, and further results were reported in Ref. [19]. To increase the heat transport capacity, Zhao et al. [20] experimentally investigated a nitrogen-charged CLHP with a parallel condenser, which could greatly reduce the flow resistance and increase the cooling capability of the condenser. The results showed that the CLHP could operate reliably for a maximum heat load of 41 W with a small temperature drop of 6 K across a 0.48 m transport distance. Gully et al. [21] designed and experimentally investigated a prototype of CLHP working around 80 K with nitrogen as the working fluid. Experimental results were analyzed and discussed both in the transient phase of cooling from room temperature and in stationary conditions. The effects of transferred power, filling pressure and radiation heat load for two basic configurations of cold reservoir of the secondary circuit were studied in stationary conditions, and a maximum cold power of 19 W with a corresponding limited temperature difference of 5 K was achieved across a 0.5 m distance. Of particular note that all CLHPs introduced in Refs. [16–21] employed an auxiliary loop to realize the large temperature drop of the primary evaporator during the supercritical startup process. Above short review shows that CLHPs are effective and efficient cryogenic heat transport devices developed in recent years, and the study is just at the beginning mainly focusing on experimental aspects. Within these studies, experimental investigation on its working principles and operating characteristics are still inadequate. Furthermore, to push CLHP into space applications, besides each component should be optimized to reduce its volume and weight, appropriate design of the structure of the evaporator casing and the condenser is also required to satisfy the interface requirements between the CLHP and the heat source (infrared sensors/detectors) and heat sink (cryocoolers). Aiming at future space applications, a miniature CLHP with nitrogen as the working fluid, which employed an auxiliary loop to realize the temperature drop of the primary evaporator during the supercritical startup, was
designed and experimentally investigated in this paper. The outer diameters of the evaporator and the transport line were 13 mm and 2 mm respectively, and the cylindrical condenser design can provide convenient interface with the cold finger of the cryocooler, making the CLHP with considerable application potential. The experiments were mainly focused on the supercritical startup capability and matching characteristics of the heat loads applied to the primary and secondary evaporators, and its heat transport capacity and thermal resistance variation were also tested. The experimental results are presented and analyzed in detail in this paper. 2. Experimental setup Fig. 1 shows the schematic of the CLHP designed in this work, and Table 1 gives the basic parameters of the CLHP, where CC represents the compensation chamber and OD and ID represent the outer and inner diameters respectively. As shown in Fig. 1, because the CLHP employs an auxiliary loop to realize the supercritical startup, the system is composed of a main loop and an auxiliary loop apart from the gas reservoir. The main loop includes a primary evaporator, a primary CC, a primary condenser and the primary vapor and liquid lines, and the auxiliary loop consists of a secondary evaporator, a secondary CC, a secondary condenser and the secondary loop line. The gas reservoir with a comparatively large volume is utilized to reduce the system pressure at ambient state, and it is connected with the inlet of the primary condenser. All the components of the CLHP were made of stainless steel except that the primary and secondary evaporator wicks were made of sintered nickel powder. The CLHP was operating at the temperature range of 80–120 K, and nitrogen whose purity was greater than 99.9995% was used as the working fluid. In the experiments, heat loads applied to the primary and secondary evaporators were provided by two thin-film electric resistance heaters, and they were attached directly to the casings of the primary and secondary evaporators symmetrically. The heat load can be adjusted from 0 W to 30 W by altering the DC power output voltage imposed on the heaters. The primary and secondary condenser lines were coiled and soldered onto a solid cylindrical cold block (see Fig. 1) to reduce the contact thermal resistance. The cryogenic heat sink was simulated by liquid nitrogen circulating through another hollow cylindrical cold block (see Fig. 2), and the two cylindrical cold blocks were connected tightly through the end faces. For practical application, the hollow cylindrical cold block would be replaced by the cryocooler. To reduce the parasitic heat loss from the ambient, the CLHP was placed in a thermal-vacuum chamber, and all the components except the gas reservoir were covered with multilayer insulation materials. Note that, the pressure in the thermal vacuum chamber can be maintained at below 2 102 Pa, and the parasitic heat loss by convective heat transfer becomes negligible in the experiments. Type T thermal couples (TCs) were used to monitor the temperature variations of the characteristic points along the loop, the measuring uncertainties of the thermal couples are ±1.0 K by calibration, and the TC locations are shown in Fig. 1. To reduce the influence of gravity, the primary evaporator and the primary condenser were placed in a horizontal plane in the experiments. 3. Experimental results and analysis 3.1. Supercritical startup 3.1.1. Temperature variations of characteristic points along the loop Fig. 3 shows the temperature variations of the characteristic points along the loop during the supercritical startup process. In
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Fig. 1. Schematic of the CLHP and thermal couple locations.
Table 1 Basic parameters of the CLHP. Component
Parameter
Primary evaporator
Casing OD/ID length Wick OD/ID length
U13/11 35 mm U11/4 30 mm
Secondary evaporator
Casing OD/ID length Wick OD/ID length
U13/11 35 mm U11/4 30 mm
Main loop
Vapor line OD/ID length Liquid line OD/ID length Condenser line OD/ID length
U2/1 640 mm U2/1 560 mm U2/1 450 mm
Secondary loop
Transport line OD/ID length Condenser line OD/ID length
U2/1 600 mm U2/1 150 mm
Volume of primary CC/secondary CC
2.01/3.62
Volume of gas reservoir
555 ml
Wick
1.0 lm 55% >5 1014 m2
Maximum capillary radius Porosity Permeability
300
Qse= 3W Qpe= 2W
Temperature/K
250 TC5
TC8
200
TC1 TC2
TC12
150
TC3
TC11
100
50 Fig. 2. Schematic of the hollow cylindrical cold block.
Fig. 3, Qpe and Qse represent the heat loads applied to the primary and secondary evaporators respectively, and the charged pressure of the working fluid was 1.2 MPa. As shown in Fig. 3, at the initial state, all the characteristic points along the loop were at the ambient temperature of 291.6 K. About 10 min later, liquid nitrogen began to circulate through the hollow cylindrical cold block, and the temperature of the cold block (TC13) dropped rapidly. At the same time, the temperatures of the inlet (TC3 and TC8) and outlet (TC4 and TC9) of the primary and secondary condenser lines dropped fast due to their close contact with the cold block. The secondary evaporator was cooled by the cold block through the secondary
TC13
0
25
50
TC4 TC9
75
100
125
150
Time/min Fig. 3. Supercritical startup process of the CLHP (Qse = 3 W, Qpe = 2 W).
CC, and its temperature (TC11) also dropped fast. About 18 min later, the temperature of the cold block (TC13) began to drop slowly and maintained at about 83 K. About 26 min later, the temperature of the secondary evaporator (TC11) began to drop slowly approaching 104 K, indicating that nitrogen in the two-phase state began to exist in the secondary evaporator, which would keep the secondary evaporator temperature around the saturation temperature of nitrogen corresponding to local pressure. As time went by, the inventory of liquid nitrogen in the secondary
L. Bai et al. / Cryogenics 52 (2012) 557–563
evaporator would increase gradually, and it would saturate the wick of the secondary evaporator in the end. After the temperature of the secondary evaporator maintaining at 104 K for about 35 min, a heat load of 3 W was applied to the secondary evaporator (Qse), and the temperature of the secondary evaporator (TC11) rose slightly, then maintained at about 110 K. The temperature at the outlet of the secondary evaporator (TC12) dropped sharply, indicating that the liquid nitrogen in the secondary evaporator began to evaporate, and the generated cryogenic nitrogen vapor flowed out of the secondary evaporator. At the same time, the temperatures at the inlet of the primary and secondary condenser lines (TC3 and TC8) rose rapidly, which was caused by the displacement of the ambient nitrogen gas in the transport lines into the primary and secondary condensers. Meanwhile, the temperatures at the outlet of the primary and secondary condenser lines (TC4 and TC9) dropped rapidly caused by the condensate in the primary and secondary condensers being pushed out, which confirmed the normal circulation of the working fluid in the loop. The operation of the secondary evaporator could realize the transfer of the condensed liquid nitrogen from the primary condenser to the primary evaporator and CC through the primary liquid transport line, and so the temperature at the outlet of the primary liquid transport line (TC5) dropped rapidly. Meanwhile, the temperature of the primary evaporator (TC1) dropped gradually. With the operation of the secondary evaporator for about 22 min, the temperature of the primary evaporator dropped to about 109 K and then remained unchanged. The temperature at the outlet of the primary evaporator (TC2) began to drop rapidly, indicating that nitrogen in the twophase state began to exist in the wick of the primary evaporator, the wick began to be saturated with liquid nitrogen, and the generated cryogenic nitrogen vapor flowed out of the primary evaporator. As time went by, the inventory of liquid nitrogen in the primary evaporator would increase gradually and fully saturate the wick of the primary evaporator. With the temperature of the primary evaporator remaining at 108 K for about 28 min, a heat load of 2 W was applied to the primary evaporator (Qpe), and the temperature of the primary evaporator rose slightly, then maintained at about 115 K. With the normal operation of the primary evaporator, the CLHP reliably realized the supercritical startup. 3.1.2. Effect of heat load applied to the secondary evaporator Figs. 4 and 5 show the temperature variations of the characteristic points along the loop during the supercritical startup process. Note that, In Figs. 4–12, the charged pressure of the working fluid was all 1.6 MPa. The heat loads applied to the secondary evaporator were 5 W and 3 W in Figs. 4 and 5 respectively. The tempera300
Qse= 3W
300
Qpe= 2W
250
Temperature/K
560
200 TC1
150
TC11
100
50
TC13
0
20
40
60
80
100
120
Time/min Fig. 5. Supercritical startup process of the CLHP (Qse = 3 W, Qpe = 2 W).
ture variation process of the characteristic points along the loop in Fig. 5 was similar to that in Fig. 4 but taking longer duration in the startup. As the heat load applied to the secondary evaporator became smaller, the transferred cryocooling from the primary condenser to the primary evaporator per unit time decreased. At the same time, as the mass flowrate of the working fluid became smaller, the temperature rise of the working fluid flowing along the primary liquid transport line increased due to the parasitic heat load from the ambient, and the temperature at the inlet of the primary CC became higher. All these factors mentioned above were adverse to the fast temperature drop of the primary evaporator. Therefore it took 40 min for the primary evaporator to drop to 104 K from ambient temperature in Fig. 5, whereas only 15 min was required in Fig. 4. Although the temperature drop of the primary evaporator became slower, the CLHP also achieved supercritical startup at last, as shown in Fig. 5. Further reducing the heat load applied to the secondary evaporator to 2 W during the supercritical startup process, it took nearly two hours for the temperature of the primary evaporator (TC1) to drop to about 197 K, as shown in Fig. 6. Beyond that, it dropped rather slightly, indicating that the transferred cryocooling from the primary condenser was balanced by the parasitic heat load from the ambient to the primary evaporator. As this temperature was obviously higher than the critical temperature of nitrogen (126 K), no nitrogen in the two-phase state would exist in the primary evaporator, and the wick of the primary evaporator could not be saturated by liquid nitrogen. Under such a situation, the 300
Qse= 5W
Qse= 2W
Qpe= 2W
250
Temperature/K
Temperature/K
250
200 TC1
TC11
150
100
TC1
200
150
TC11
100 TC13
50 0
15
30
45
60
75
90
Time/min Fig. 4. Supercritical startup process of the CLHP (Qse = 5 W, Qpe = 2 W).
50
TC13
0
30
60
90
120
150
Time/min Fig. 6. Temperature drop process of the primary evaporator (Qse = 2 W).
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180
Qse= 0W Qpe= 4W
Qpe= 5W
180
Qpe= 2.5W
Qpe= 3W
TC3
TC1 TC4
80 60
Temperature/K
Temperature/K
140
100
Qpe= 14W
140 120
TC2
100
TC1
TC5
TC13
TC3
TC13
60 0
25
50
75
100
125
0
30
180
Qse= 1W Qpe= 2W
Qpe= 4W
Thermal resistance/(K/W)
TC3
TC11 TC5
TC4
80
150
12
TC1
100
120
Fig. 10. Heat transport capacity test of the primary evaporator (Qse = 0 W).
Qpe= 1.5W Qpe= 1.0W
140 TC2
90
14
160
120
60
Time/min
Fig. 7. Temperature variation with heat load applied to the primary evaporator (Qse = 0 W).
Temperature/K
TC4
80
Time/min
10 8 6 4 2
TC13
0
60 0
10
20
30
40
50
60
70
80
Fig. 8. Temperature variation with heat load applied to the primary evaporator (Qpe = 1 W).
180
Qse= 2W Qpe= 1W
Qpe= 2W
Qpe= 0.5W
160 140
TC11
100 80
TC2 TC5
TC1
TC4
TC13
60 0
10
20
30
40
4
6
8
10
12
50
60
Fig. 11. Thermal resistance variation of the CLHP.
secondary evaporator must reach a certain value. According to Figs. 5 and 6, the minimum heat load applied to the secondary evaporator required should be between 2 W and 3 W to realize the supercritical startup of the CLHP developed in this work. 3.2. Matching characteristics of heat loads applied to the primary and secondary evaporators
TC3
120
2
Heat load/W
Time/min
Temperature/K
Qpe= 6W Qpe= 8W Qpe= 10W Qpe= 12W
160
Qpe= 2W
TC2
Qse= 0W
Qpe= 4W
160
120
Qse= 2W
70
80
Time/min Fig. 9. Temperature variation with heat load applied to the primary evaporator (Qse = 2 W).
primary evaporator could not be started, i.e. the CLHP could not realize supercritical startup with a heat load of 2 W applied to the secondary evaporator. Such an experimental result was also consistent with the theoretical analysis in Ref. [22]. To realize the supercritical startup of the CLHP, the heat load applied to the
For CLHPs, as they operate in the cryogenic temperature range, the large temperature difference between the ambient and the working fluid in the CLHP would produce comparatively large parasitic heat loss. Most of the experimental studies on CLHPs are generally conducted in the thermal-vacuum chamber, and the whole CLHP except the gas reservoir is covered with multilayer insulation materials (MLI) to reduce the heat transfer between the ambient and the CLHP. Although such strict thermal insulation measures are employed, it is still possible that the primary evaporator fails to operate independently due to the effect of the parasitic heat load from the ambient at a relatively small heat load [23]. Under this situation, the secondary evaporator must be kept in operation to assist the operation of the primary evaporator. For a specific heat load applied to the secondary evaporator, there shall exist a minimum heat load applied to the primary evaporator to maintain the normal operation of the primary evaporator. Fig. 7 shows the temperature variations of the characteristic points along the loop when the heat load applied to the secondary
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Qse= 0W
Qse= 5W
120
Qpe= 4W
Qpe= 8W
100
TC1
TC3
90
TC4
80
Qse= 0W Qpe= 4W
Qpe= 12W
110
TC2
Temperature/K
Temperature/K
110
70
Qse= 5W
120
Qpe= 12W Qpe= 8W
TC2
100
TC3
TC1
90
TC4
80
TC13
TC13
70 0
25
50
75
100
125
0
10
20
30
40
50
Time/min
Time/min
(a) Variation amplitude of 4W
(b) Variation amplitude of 8W
60
70
Fig. 12. Response to step change of heat load applied to the primary evaporator.
evaporator was 0 W, i.e. the secondary evaporator was not in operation, and the heat load applied to the primary evaporator was decreased gradually. Under such conditions, the primary evaporator could operate normally as the heat load applied to the primary evaporator is in the regime of 2.5–5 W. However, when the heat load was further reduced to 2.0 W, the primary evaporator temperature (TC1) rose quickly, and the primary evaporator failed to operate. That is because when the heat load applied to the primary evaporator is very small, i.e. <2.5 W in this paper, it cannot overcome the effect of the parasitic heat loss from the ambient, as analyzed in detail in Ref. [23]. Under such a situation, the secondary evaporator must be kept in operation to assist the operation of the primary evaporator. Fig. 8 shows the temperature variations of the characteristic points along the loop when the heat load applied to the secondary evaporator was 1 W and the heat load applied to the primary evaporator was decreased gradually. Different from the situation in Fig. 7, the CLHP can still operate normally as the heat load applied to the primary evaporator was decreased from 4 W to 1.5 W. However when the heat load applied to the primary evaporator was further reduced to 1.0 W, the primary evaporator temperature (TC1) rose quickly, and the primary evaporator failed to operate. As the heat load applied to the secondary evaporator was increased further to 2 W in Fig. 9, the primary evaporator could operate normally at a heat load as low as 0.5 W. It is known that the mass flowrate of the working fluid in the primary liquid transport line, which is in proportional to the sum of the heat loads applied to the primary and secondary evaporators, determines the ability of the CLHP to overcome the effect of the parasitic heat load from the ambient. The results in Figs. 7–9 show clearly that when the sum of the heat loads applied to the primary and secondary evaporators was no less than 2.5 W, the CLHP could overcome the effect of the parasitic heat load and the primary evaporator could operate normally; however, when the sum of the heat loads was smaller than 2.5 W, the primary evaporator could not operate normally due to the effect of parasitic heat load, and the secondary evaporator must be kept in operation to assist the operation of the primary evaporator. 3.3. Heat transport capacity of the primary evaporator The heat transport capacity is one of the most important aspects to assess the CLHP performance, which determines the
maximum heat load that the primary evaporator can sustain. Fig. 10 shows the test results of the heat transport capacity of the primary evaporator. In the experiments, the heat load applied to the secondary evaporator was firstly set at 2 W, then it was reduced to 0 W where the primary evaporator could operate independently with a heat load of 4 W. As shown in Fig. 10, when the heat load applied to the primary evaporator was increased from 4 W to 12 W, the primary evaporator could operate normally. However, when the heat load was further increased to 14 W, the temperature of the primary evaporator rose quickly, and the primary evaporator failed to operate, indicating that the primary evaporator reached the capillary limit, i.e. 12 W 0.56 m. 3.4. Thermal resistance variation In this paper, the thermal resistance of the CLHP is defined as follows:
R¼
T w;pe T if;cb Q pe
ð1Þ
In Eq. (1), R is the thermal resistance of the CLHP; T w;pe and T if;cb are the temperatures of the primary evaporator wall and the cold block interface respectively and Qpe is the heat load applied to the primary evaporator. In this work, the temperature of the primary evaporator wall is represented by TC1, and the temperature of the cold block interface is represented by TC13. Fig. 11 shows the heat-load dependence of the thermal resistance of the CLHP. As the heat load applied to the primary evaporator increased, the thermal resistance of the CLHP decreased monotonically, i.e. from 12.1 K/W at the heat load of 2 W to 2.1 K/W at the heat load of 12 W, indicating that the CLHP was always operating in the variable conductance mode. In this mode, the two-phase zone of the primary condenser kept increasing with the increase of the heat load applied to the primary evaporator. Assuming the thermal conductivity of copper is 400 W/(m K), the thermal resistance of the solid copper rod with the same volume of the CLHP is about 160 K/W, and the thermal resistance of the CLHP is therefore much smaller than that of the equivalent copper rod, indicating that the CLHP has an excellent cryogenic heat transport performance.
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3.5. Response to step change of heat load applied to the primary evaporator Fig. 12 shows the temperature variations of the characteristic points along the loop when the heat load applied to the primary evaporator was subject to a step change with a comparatively large amplitude. In this experiment, the heat load applied to the secondary evaporator was 5 W at first, then it was reduced to 0 W, and the primary evaporator could operate independently with a heat load of 4 W. According to Fig. 12, when the heat load applied to the primary evaporator was subject to a step change of 4 W or 8 W, as shown in Fig. 12a and b respectively, the primary evaporator temperature (TC1) changed smoothly, the primary evaporator could operate normally, and the CLHP manifested good response characteristic and thermal control performance. 4. Conclusions A CLHP using nitrogen as the working fluid was designed and experimentally investigated in this work, and specific conclusions below can be drawn: (1) With a heat load of 3 W applied to the secondary evaporator, the CLHP can reliably realize the supercritical startup, and the larger the heat load applied to the secondary evaporator, the sooner the temperature drop process of the primary evaporator. (2) When the heat load applied to the secondary evaporator is 62 W, the primary evaporator cannot be cooled below the critical temperature of nitrogen due to the effect of parasitic heat load from the ambient, and the CLHP cannot realize the supercritical startup under this situation. (3) When the heat load applied to the primary evaporator is no less than 2.5 W, the primary evaporator can operate independently, i.e. the operation of the secondary evaporator is not required; however, when it is smaller than 2.5 W, the secondary evaporator must be kept in operation to assist the operation of the primary evaporator. (4) The CLHP has a heat transport capacity of 12 W 0.56 m, and the thermal resistance of the CLHP decreases with the increase of the heat load applied to the primary evaporator showing a variable conductance working mode. (5) The CLHP has the ability to operate with large step change of the heat load applied to the primary evaporator, which manifests good response characteristic and thermal control performance.
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