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Investigation of a compact copper–water loop heap pipe with a flat evaporator
Applied Thermal Engineering 31 (2011) 3533e3541
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Investigation of a compact copperewater loop heap pipe with a flat evaporator Yu. Maydanik*, S. Vershinin, M. Chernysheva, S. Yushakova Institute of Thermal Physics, Ural Branch of the Russian Academy of Sciences, Amundsen St. 106, Ekaterinburg 620016, Russia
a r t i c l e i n f o
a b s t r a c t
Article history: Received 10 November 2010 Accepted 1 July 2011 Available online 18 July 2011
A compact copperewater loop heat pipe (LHP) with an effective length of 310 mm equipped with a flat eoval evaporator measuring 80 (L) 42 (W) 7 (H) has been tested. The vapor line and the condenser had the same internal diameter of 5.4 mm. The internal diameter of the liquid line was 3.4 mm. Tests were conducted with a heat source which had a heating surface of 30 mm 30 mm. The condenser was cooled by running water with a temperature of 20 C. In the horizontal position the device has exhibited serviceability in the heat load range from 5 W to 1200 W at vapor temperatures from 26.5 C to 103.4 C. The maximum capacity was achieved at a heat source temperature of 143.5 C, when the LHP thermal resistance was equal to 0.044 C/W. The corresponding values of thermal resistance for the evaporator and the condenser were at a level of 0.006 C/W and 0.038 C/W. A minimum thermal resistance of 0.097 C/W for the “heat sourceeLHPecooling water” system was obtained at a heat load of about 700 W, at which the temperature of the heat source was 87 C. Ó 2011 Elsevier Ltd. All rights reserved.
Keywords: Electronics cooling Loop heat pipe Flateoval evaporator Heat load Thermal resistance Start-up
1. Introduction For quite a long time there existed a stereotype idea that the “copperewater” combination is unsuitable for use in loop heat pipes operating in temperature range below 100 C. For this there was quite a logical substantiation. The fact is that the high thermal conductivity of copper on the one hand and the low value of the derivative dP/dT, which characterizes the slope of the saturation line of the working fluid in the temperature range mentioned, on the other impede the creation of the required drop of the vapor temperature DT and the corresponding pressure drop DP between the evaporation zone and the compensation chamber. As is well known [1,2], such a drop is necessary for a start-up and maintenance of the circulation of a working fluid in an LHP. At the same time, copperewater LHPs are an extremely attractive object for development and practical application because this combination is well compatible chemically and safe from an ecological and a technological point of view. Besides, it allows manufacture of sufficiently cheap devices during their mass production. An important stimulus for the development of copperewater LHPs is also the fact that the thermal properties of copper and water with a correct approach open up possibilities for the achievement of exceptionally high thermal characteristics, including the maximum heat-transfer capacity, the heat flux and the low thermal resistance. At present these factors are in great demand for the creation of efficient and reliable cooling systems of powerful * Corresponding author. Tel.: þ7 (343) 267 87 91; fax: þ7 (343) 267 87 99. E-mail address: [email protected] (Yu. Maydanik). 1359-4311/$ e see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2011.07.008
electronics and photoelectronics used in various fields of technology. It should be expected that such a tendency will remain intact in the foreseeable future. Owing to purposeful efforts in this direction, the first miniature all-copper LHPs with water as a working fluid were created and tested as long ago as the early 2000s [3]. These devices were equipped with cylindrical evaporators 6 mm in diameter or flateoval evaporators 3.2 mm thick with thermal resistances at a level of 0.09 C/W and 0.05 C/W, respectively. The evaporators were connected to the condenser by tubes with an internal diameter of 2 mm. Under forced air cooling of the condenser LHPs exhibited serviceability in the vapor temperature range from 50 to 100 C at heat loads from 10 to 160 W. In this case the ability to operate efficiently at different orientations in the gravity field, which is characteristics of loop heat pipes, was retained in full measure. About 20 such devices were made and tested in searching for optimum designs in the period from 2000 to 2003. Quite an important advantage of water as a working fluid for LHPs is also the fact that its vapor pressure at operating temperatures below 100 C allows making flat evaporators with a relatively thin wall. Besides, the flat form of the thermocontact surface in such evaporators in many cases makes it possible to avoid the use of thermal interfaces (thermal adapters) with which cylindrical evaporators are equipped for ensuring a good thermal contact with the heat sources. Three main types of flat copper evaporators have been developed thus far: disk-shaped, rectangular and flateoval evaporators. Disk-shaped evaporators [4,5] have a cylindrical side surface and two flat butt-end surfaces, one of which serves for the heat load
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Nomenclature F H k L P Q R T W DP DT SPi
surface, m2 height, mm heat-transfer coefficient, W/m2 C length, mm pressure, Pa heat load, W thermal resistance, W/ C temperature, C width, mm pressure drop, Pa temperature drop, C total amount of pressure losses, Pa
supply. There are also designs where both flat surfaces of the evaporator may be used for supplying heat loads [6]. The main drawback of disk-shaped evaporators is the fact that their thickness, which includes the compensation chamber, cannot be sufficiently small. Usually it varies within 10e20 mm. Singh et al. [4], for instance, presented the results of testing a miniature copperewater LHP equipped with a disk-shaped evaporator 30 mm in diameter and 10 mm in thickness with a nickel wick. This device 150 mm long with 2-mm transportation lines was capable of transferring from 5 to 70 W at evaporator temperatures from 57 to 99.6 C. The minimum value of thermal resistance of the evaporator was achieved at a maximum heat load of 70 W and was equal to 0.06 C/W. The heat-transfer coefficient that corresponded to it was equal to 22,600 W/m2 C. Zimbeck et al. [7] reported the results of testing two copperewater LHPs with disk-shaped evaporators having a thickness less than 25 mm provided with a ceramic wick. On trials with a CPU thermal simulator, which had a heating surface of 20 mm 20 mm, the heat load varied in the range from 50 to 200 W. The condenser was cooled by means of forced air convection with a temperature of 25 C. The evaporator temperature in this case increased within 48e65 C. At a maximum heat load of 200 W the thermal resistance of the evaporator was equal to 0.05 C/W, and the heat-transfer coefficient corresponding to it reached a value of about 50,000 W/m2 C. Rectangular flat evaporators may have both one and two thermocontact surfaces. The compensation chamber is located here on any of the sides depending on the configuration of the evaporator, whose thickness may be considerably lower as compared with disk-shaped evaporators. Thus, for instance, Singh et al. [8] reported the results of testing a miniature copperewater LHP with a flat rectangular evaporator which measured 60 mm (L) 50 mm (W) 5 mm (H) and was also equipped with a nickel wick. The active zone measuring 22 mm 22 mm was on one of the evaporator sides, and the compensation chamber was located in the same plane as the wick and enveloped it on three sides. The device exhibited serviceability in the range of heat loads from 10 to 50 W, at which the evaporator temperature varied in the range from 75 to 100 C. A minimum value of thermal resistance of the evaporator equal to 0.25 C/W was achieved at a maximum heat load of 50 W. Li et al. [9] presented the results of testing a copperewater LHP with a flat square evaporator with dimensions 30 mm (L) 30 mm (W) 15 mm (H). The copper wick structure was sintered directly on the evaporator substrate, and the compensation chamber was located above the wick. The vapor and liquid lines with an internal diameter of 5 mm had the same length equal to 120 mm. The condenser was a serpentine-shaped one, was provided with finning and was cooled by a 120-mm axial fan. The device was tested at a vertical orientation in a gravity assisted mode when the evaporator was lower than the condenser. A heater with
Subscripts s condenser s-cool condenser-cooling medium cc compensation chamber e evaporator hs heat source hs-e heat source-evaporator l liquid ll liquid line s system v vapor vl vapor line
a heating surface 25 mm 25 mm was used as a heat source. Tests were conducted at an ambient temperature 23.7 2.1 C. The heat load increased stepwise from 10 to 600 W. The evaporator temperature in this case varied from 50 to 98 C. The minimum thermal resistance of the LHP achieved at a heat load of 450 W was 0.033 C/W. These are fairly high results. But they cannot characterize in full measure the efficiency of the device under discussion as they have been obtained in “too favorable” conditions for LHP, which are gravity assisted modes. Flateoval evaporators may also have both one and two thermocontact surfaces located on the opposite flat sides. In both cases the compensation chamber is in the same plane as the wick, and therefore the thickness of the evaporators can be easily varied within the limits from 3 mm and more. The length and the width of such evaporators are determined by the dimensions of the heat source and the volume of the compensation chamber. The authors of Ref. [10] have recently presented the results of testing a copperewater LHP equipped with a flateoval evaporator 7 mm thick. The dimensions of the active zone of the evaporator were 32 mm 32 mm. The vapor and liquid lines, and also the condenser had the same internal diameter of 3.4 mm and length of 120 mm, 180 mm and 335 mm, respectively. The devise was tested in the horizontal position with heat source that had heating surfaces of 9 cm2 and 1 cm2. The condenser was cooled by thermostatted running water with a temperature 21 1 C. The heat load varied in the range from 20 to 900 W. A minimum value of thermal resistance of the evaporator equal to 0.014 C/W was achieved at the maximum heat load with a heat source which had a heating surface of 9 cm2 when the corresponding heat-transfer coefficient was equal to 79,000 W/m2 C. It should be mentioned that for many components of electronics a limiting temperature of the heat source equal to 100 C was achieved here at a heat loads of about 550 W, which corresponds to a heat flux of 61 W/cm2. The presented results obtained by various authors strongly suggest that copperewater LHPs make it possible to ensure sufficiently high values of heat loads and heat fluxes in the heating zone. However, in most cases these characteristics are achieved only at a relatively high level of the LHP operating temperature, which is unacceptable for most electronic components. Therefore the task of decreasing the operating temperature with retention of high thermal characteristics is quite topical. It is at the solution of this problem that the present work was aimed. 2. Analysis of the main factors that affect the LHP operating temperature The fundamental parameter that characterizes the LHP operation in the most impressive way is the vapor temperature. Therefore it is conveniently used for analyzing the thermal and
Yu. Maydanik et al. / Applied Thermal Engineering 31 (2011) 3533e3541
n X i¼1
DPi ¼
dP DT; dT
(1)
where SDPi is the total amount of pressure losses in the segments of the working fluid circulation from the evaporating surface of the wick to its absorbing surface on the side of the compensation chamber, dP/dT is the derivative that characterizes the slope of the saturation line of the working fluid at a given temperature, DT is the vapor temperature drop between the evaporating surface of the wick and the vaporeliquid interface in the compensation chamber. In this turn, the product of the quantities in the right-hand side of this equation is the pressure drop determined by the difference of the vapor temperatures between the evaporation zone and the compensation chamber. Thus, for ensuring an LHP start-up and operation it is necessary that the pressure losses during the circulation of the working fluid between the evaporating and absorbing surfaces of the wick SDPi should not exceed the difference of the pressures of vapor DP connected with the difference of its temperatures DT between the surfaces mentioned above. Evidently the low value of dP/dT, which is characteristic of water at low temperatures, may be compensated either at the cost of increasing DT or at the cost of decreasing SDPi, or at the cost of both simultaneously. As for the value of DT, its increase is considerably inhibited by the high thermal conductivity of the copper body and the wick, which contributes to the increase of parasitic heat flows into the compensation chamber. The use of a wick made of less heat-conducting materials for decreasing these flows does not solve the whole problem as in this case one can observe a considerable reduction in the efficiency of heat transfer in the evaporation zone. The problem of increasing DT may be solved at the cost of an increase in the thickness of the “barrier” layer of the wick, which separates the evaporation zone and the condensation chamber. This way is fairly effective when the working fluid is water because the high capillary pressure that it is capable of creating compensates easily an increase in the hydraulic resistance of the wick. In this case for evaporators which have a butt-end compensation chamber an increase in the wick thickness leads to an increase in the thickness of the evaporator itself, and for evaporators with a compensation chamber located in the same plane as the wick this results in increasing evaporator length. The choice of one or the other variant is determined by the conditions of the LHP use.
dP/dT, kPa/K
4 3 2 1 0
20
30
40
50
60
70
80
90 100 110
o
Tv, C Fig. 1. Temperature dependence dP/dT for water.
Another way of realizing condition (1) is connected with a decrease in the value of SDPi. This may be first of all achieved by decreasing pressure losses in the LHP vapor line. For water, which has a low vapor density in a low-temperature range, even a small increase in the diameter of the vapor line makes it possible to considerably decrease these losses. Figs. 1 and 2 present as examples, respectively, temperature dependences of the value of dP/dT and the amount of pressure losses in a vapor line 300 mm long with different diameters at a heat load of 50 W. Here it can be seen, in particular, that at a vapor temperature of 30 C pressure losses in a vapor line 3 mm in diameter are approximately equal to 880 Pa, and the value of dP/dT at this temperature is 230 Pa/K. Evaluation by formula (1) gives a value of DT equal to 3.8 C. It is difficult to attain such a value in a copper evaporator at a heat load of 50 W. At the same time at a temperature of 60 C a value of DT equal to 0.25 C may be quite sufficient. Approximately the same result can also be obtained at a vapor temperature of 30 C by increasing the diameter of the vapor line to 6 mm. 3. Description of an experimental LHP A copperewater loop heat pipe with a flateoval evaporator was specially made for experiments in the framework of this work. Such evaporators are easy to make, and besides they have shown themselves to good advantage in testing various versions of copperewater LHPs [3,10e13]. In this case the LHP had a flateoval evaporator measuring 80 mm (L) 42 mm (W) 7 mm (H). The active zone of the evaporator (the zone to which a heat load may be applied) measured 32 mm 32 mm. The evaporator was provided with a copper wick with a porosity of about 67% and a break-down pore radius of 21 microns. Twelve longitudinal channels 1, 8 mm in diameter were made in the wick. They formed the evaporating
0,9 0,8 0,7
ΔPvl, kPa
hydrodynamic processes that take place in these devices. It is quite evident that the lower limiter of the LHP operating temperature is the temperature of the medium which cools the condenser and the upper one the temperature of the heat source. The closer the vapor temperature to the temperature of one and the other, the more efficient the LHP operation, and the lower the thermal resistance of the “heat sourceecooling medium” system. The problem here consists in ensuring the operation of a copperewater LHP at a lower temperature level with respect to the ambient temperature retaining at the same time an acceptable efficiency of the devise operation. An analysis of the results obtained by various authors on trials of copperewater LHPs [3e5,7e10] provided with both copper and some other, less heat-conducting, wicks shows that the minimum value of the vapor temperature was always at a level above 40 C even at minimum heat loads. The operating vapor temperature in an LHP depends on many factors, among which the thermal properties of the working fluid occupy an important place. The thermal properties of water in temperature ranges below 100 C, such as the density and viscosity of vapor, and also the value of dP/dT, make it not a very “convenient” working fluid for LHPs. To understand this, one should turn to the second, the so-called thermodynamic condition of the LHP serviceability, which may be written as follows:
3535
d=3 mm
0,6 0,5 0,4
d=4 mm d=5 mm d=6 mm
0,3 0,2 0,1 0,0
20
30
40
50
60 70 Tv, oC
80
90
100 110
Fig. 2. Temperature dependence of pressure losses in a vapor part 300 mm long with different diameters.
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surface and at the same time served for removing vapor from the evaporation zone. On one of the flat sides of the evaporator, near which the vapor-removal channels were located, there was a special thermal interface, i.e. a copper plate 1 mm thick measuring 30 mm 30 mm, which was soldered to the surface of the active zone. Despite the fact that such an interface creates an additional thermal resistance, its use was dictated by the necessity of locating thermocouples for temperature measurements right on the evaporator wall. For this purpose the plate had a through cut 1 mm wide, where a thermocouple was placed. The vapor line 120 mm long and the condenser with a length of 335 mm had the same internal diameter equal to 5.4 mm. The liquid line 180 mm in length had an internal diameter of 3.4 mm. The length of the vapor path of the device with allowance for the length of the vaporremoval channels in the evaporation zone, the vapor line and the part of the condenser which rids itself of the liquid during the operation was about 370 mm. The condenser was provided with a spiral fin located in its outer surface and a plastic jacket for pumping the cooling water. The general view and the scheme of the copperewater loop heat pipe are presented in Figs. 3 and 4, respectively. 4. Testing procedure All tests in the framework of these investigations were conducted at a horizontal orientation of an LHP, when the evaporator and the condenser were located in one and the same horizontal plane. The heat load was delivered to the evaporator from below. The heat load source was a copper cylindrical block, in which 6 heating cartridges of 250 W each were positioned parallel to the longitudinal axis. One of the butt-ends of the copper block had a square ledge, whose heating surface measuring 30 mm 30 mm was in thermal contact with the thermal interface of the evaporator. The whole copper block was in thermal insulation, limited from the outside by a thin-walled metal shell. Three thermocouples were located at the shell surface, and their readings made it possible to evaluate the heat losses from the heat source into the outside ambient and to determine the magnitude of the actual heat load supplied to the evaporator. The scheme of the heat source is shown in Fig. 5. The electric power supplied to the heaters was measured by a wattmeter and changed stepwise with a step of 20 W to a value of 100 W, and then with a step of 100 W to the maximum magnitude. The magnitude of the maximum power was limited by the vapor temperature in the LHP equal to 105 C, at which the internal pressure reached 120 kPa. A further increase in the pressure was liable to cause deformation of the thin-walled body of the evaporator. The condenser was cooled by running thermostatted water with a temperature of 20 0.1 C, which was pumped through the
Fig. 3. General view of a copperewater LHP.
Fig. 4. Scheme of a copperewater LHP.
cooling jacket with a flow rate equal to 0.5 103 m3/min. The LHP operating temperature was measured at several points with the help of coppereconstantan thermocouples “OMEGA” TT-T-30. The temperature of the evaporator wall at the center of the heating zone was measured by a thermocouple which was located in the above-mentioned special cut of the thermal interface. To measure the vapor temperature, use was made of a thermocouple positioned in a metal capillary which was inserted into the vapor line. The sufficiently large diameter of that line allowed doing it without any
Fig. 5. Scheme of a heat source.
5.1. LHP start-up In the opinion of some authors in Refs. [4,9], which is based on their own experimental experience, a start-up at low heat loads is a problematic operating mode for LHPs. The problem here is connected with the fact that at low heat loads a start-up does not take place (there is no circulation of the working fluid), or a start-up is accompanied by strong temperature pulsation, whose amplitude may reach 10 C. In this case it is assumed that the conditions leading to such phenomena are inherent properties for all LHPs regardless of what manufacturing process is utilized. Special tests were conducted to check how critical the start-up process is for the given LHP. Here it would be expedient to mention that the notion of “low heat load” is relative as the essential factor is the magnitude of the surface over which it is distributed. It would be more adequate to use the term “heat flux”, which with allowance for the specific character of LHPs with water as a working fluid may be considered low if its value does not exceed 5 W/cm2. For ammonia LHPs, for instance, under adequate conditions a low heat flux may be 1 W/cm2 and lower. Fig. 6 presents start-up processes at heat loads of 5 W, 20 W, 100 W, 200 W, which approximately correspond to heat fluxes of 0.5 W/cm2, 2.2 W/cm2, 11.1 W/cm2 and 22.2 W/cm2. The first two of them may be referred to low heat fluxes, the third and the fourth to moderate ones. It is easy to see that common to all of them is the presence of the so-called temperature “overshoot”, after which one can observe a rapid temperature drop and an actual LHP start-up. The higher the heat load, the faster the heating and the subsequent start-up. Thus, for instance, at a heat load of 5 W an actual LHP start-up takes place in about 60 min, and at 200 W in 2.2 min after a heat load delivery. In this case the time of going into a steady state level is noted, respectively, in about 100 min and 7 min. At the same time the absolute ceiling of the temperature “overshoot” at all the above-mentioned heat loads remains approximately the same and is within 42e44 C for the evaporator and 31e33 C for vapor. As is shown in Ref. [14], such a picture corresponds entirely to one possible scenario of an LHP start-up, when in the initial state vaporremoval channels in the evaporation zone are flooded with a working fluid and there is no completed “vaporeliquid” interface here. For such an interface to appear the boiling-up of a working fluid is required, which is accompanied by some superheating corresponding in value to the temperature “overshoot”. However, such a start-up scenario is not the only possible one. When for some reasons or other there is a free “vaporeliquid” interface in the evaporation zone, a start-up may proceed “smoothly”, without
45 5W
40 35
Te
30 25
Tv2
20
Tl1
0
Te mpera ture, oC
In LHP trials a study was made of all the main thermal parameters that characterize the behavior of a “heat sourceeLHPecooling water” system under changes of the heat load when the ambient conditions remained constant. The heat load dependence of the LHP temperature at characteristic points, the heat-transfer coefficient in the evaporation zone and thermal resistance in some sections and in the system as a whole was determined. Investigations were also made of the LHP behavior during start-ups at low and moderate heat loads.
3537
15 2000
4000
6000
45 20 W
40 35
Te
30
Tv2
25 20
Tl1
15 0
Tem pe rature, oC
5. Test results and discussion
400
800
1200
45 100 W
40
Te
35 30
Tv2
25
Tl1
20 15 0
Tem pe rature, oC
essential changes in the local hydraulic resistance to the vapor flow. Besides, two thermocouples were located on the wall of the vapor line. One thermocouple was also used for measuring temperature near the thermocontact surface of the heat source. The readings of the thermocouples were recorded by a data acquisition unit Agilent 34970A. The scheme of location of the thermocouples is given in Fig. 4.
Tempe rature, oC
Yu. Maydanik et al. / Applied Thermal Engineering 31 (2011) 3533e3541
200
400
600
800
45 Te
200 W
40
Tv2
35 30 25
Tl1
20 15 0
100
200
300
400
500
Time, s Fig. 6. Process of an LHP start-up at different heat loads.
being accompanied by a temperature “overshoot”, as is shown, for instance, in Fig. 7. The presence or the absence of a completed “vaporeliquid” interface in the evaporation zone depends on the initial distribution of a working fluid in an LHP connected with its “thermal prehistory”. As for studying the start-up process at the minimum heat load at which a stable circulation of the working fluid in an LHP begins, it should be mentioned that it has an academic rather than a practical meaning. It is rather difficult to imagine some exotic case when an LHP intended for operation at a nominal heat load of, for instance, 150 W will start up in the process of operation at a heat load of 5 or even 10 W. As a rule, an actual range of operating heat loads is relatively small. Thus, in particular, a CPU dissipating 100 W at a maximum load in the standby mode, when the heat load is
Yu. Maydanik et al. / Applied Thermal Engineering 31 (2011) 3533e3541
Te mpe rature , oC
3538
40 5W
35
Te
30 25
Tv2
20
Tl1
15 0
2000
4000 Time, s
6000
Fig. 7. Process of a “smooth” LHP start-up at a heat load of 5 W.
minimum, dissipates 40e50 W. Therefore, if it is necessary to remove heat from an object dissipating 5 or 10 W, an appropriate LHP has to be made, for which such heat loads are nominal. It is also significant that in the process of LHP tests at the above-mentioned heat loads the start-up was not accompanied by considerable temperature pulsations. The maximum amplitude of vapor temperature pulsation did not exceed 0.6 C, and the amplitude of temperature pulsation of the evaporator and the heat source was still lower. The main reason and kinds of LHP temperature pulsations are analyzed at great length in Refs. [15e17]. Here it may be added that the operation of LHPs without temperature pulsations is impossible in principle, as well as of any other closed system where joined processes of evaporation/boiling and condensation take place and where there are movable interfaces between the vapor and liquid phases of the working fluid. Difference may lie in the organization of these processes and the parameters of pulsations. If the processes of heat and mass transfer in an LHP are organized optimally, the amplitude of pulsations is small and they are of little practical importance. 5.2. LHP operating characteristics The LHP operating characteristics are presented in Fig. 8 as heat load dependences of the temperature of the heat source Ths, evaporator Te, vapor Tv2 and compensation chamber Tcc. The type of the curve Te ¼ f(Q), which has a near-linear form, not quite typical of LHPs, may point to the absence of a variable-conductance mode. Such a mode is usually observed as long as there is a redistribution of a liquid between the condenser and the compensation chamber. However, the vapor temperature, which is more sensitive to the situation inside the LHP, and which is presented by the curve Tv2 ¼ f(Q), shows that such a mode exists up to heat loads of about 300 W, though in this case it is not strongly pronounced. Of particular importance is the fact that the vapor temperature in the LHP remains lower than 40 C up to heat loads close to 250 W. In this case the minimum value of the vapor temperature obtained in
Temperature, oC
160
5.3. Thermal resistance The thermal resistance is one of the most important thermal parameters that characterize the LHP efficiency. The lower is the thermal resistance of a heat-transfer device, the closer to the temperature of the medium cooling the condenser may be the temperature of the heat source. The thermal resistance of an LHP is mainly determined by the thermal resistance of the evaporator and the condenser as the thermal resistance of the vapor and liquid lines in most cases may be neglected. To analyze thermal resistance, it is convenient to use the scheme presented in Fig. 9. In this scheme thermal resistances are determined in the following way: 1) The thermal resistance from the heat source to the evaporator
Rhse ¼
Thse Te Q
2) The thermal resistance of the evaporator is usually determined by the readings of thermocouples located on the evaporator wall in the heating zone and on the vapor line as
Re ¼
Te Tv1 : Q
In this case, however, a more precise value of Re can be determined by using the readings of the thermocouple Tv2 located inside the vapor line:
Te Tv2 : Q
Ths
40
(4)
Te Rhs-e
0 200
(3)
The thermal resistance of the evaporator may also be presented in a different form, which reflects more precisely the physical meaning of this quantity:
80
0
(2)
is determined by the quality of finish of the conjugate surfaces, the force of pressing, and the thermal conductivity of the grease used for improving the thermal contact. Its value is practically independent of the heat load, but it may make a considerable contribution to the thermal resistance of the whole system. In this case its value was 0.027 20%.
Re ¼
Ths Te Tv2 Tcc
120
LHP trials was at a level of about 26.5 C. In accordance with the vapor temperature, the temperature of the evaporator and the temperature of the heat source were also at a lower level. In particular, a temperature of the heat source Ths equal to 80 C, which is an admissible temperature for many components of electronics, was achieved at a heat load of about 600 W, the evaporator temperature being equal to 67 C, and the vapor temperature to 56 C. If we considered a CPU capable of dissipating 130 W at a maximum heat load as a heat source, its temperature with such an LHP would be approximately 43 C. It should also be mentioned that the limiting heat load for this LHP was not achieved owing to the limitations connected with the vapor pressure. At an actual maximum heat load of 1174 W, which corresponds to a heat flux in the heating zone equal to 130.5 W/m2, no signs of heat-transfer crisis in the evaporator were observed.
400
600
800
Heat load, W Fig. 8. LHP operating characteristics.
1000
1200
Tv1 Re
Tv3 Rvl
Tl2
Tl1
Tc Rc
Tcool Rc-cool
Rll Fig. 9. Scheme of thermal resistances of the “heat sourceeLHPecooling medium” system.
Re ¼
1 ; ke Fe
Thermal resistance, oC/W
Yu. Maydanik et al. / Applied Thermal Engineering 31 (2011) 3533e3541
(5)
where ke is the heat-transfer coefficient in the evaporation zone, Fe is the area of the heating surface. At a fixed value of Fe the value of Re depends on the thickness and thermal conductivity of the evaporator wall and the heat-transfer intensity in the evaporation zone. 3) The thermal resistance of the condenser, by analogy with that of the evaporator, may be determined as
T Tc Rc ¼ v2 : Q
Tc ¼
(7)
1 ; kc Fc
(8)
where kc is the heat-transfer coefficient in the condenser, Fc is the area of the condenser surface. 4) The thermal resistance of an LHP is usually presented by the formula:
RLHP ¼
Te Tc ; Q
(9)
With allowance for (5) and (8) it may also be written as:
RLHP ¼
1 1 þ : ke Fe kc Fc
(10)
5) The thermal resistance of the heat transfer from the condenser to the cooling medium may also have a considerable effect on the resistance of the whole system. It may be determined as:
Rccool ¼
Tc Tcool : Q
(11)
6) The total thermal resistance of the system
Ths Tcool Q
0,20 0,15 0,10 0,05 0,00 0
200
400 600 800 Heat load, W
1000
1200
Fig. 11. Heat load dependence of the condenser thermal resistance.
may also be presented as the sum of the resistances:
(12)
0,08 0,06 0,04 0,02
(13)
Now that all the determinations of thermal resistances are given, we may return to the results of LHP trials. Figs. 10e14 present heat load dependence of the main thermal resistances of the system. It can be seen (Fig. 10) that the lowest thermal resistance is that of evaporator. A minimum value of 0.006 C/W was achieved here at a heat flux of 130.4 W/cm2, which corresponded to a maximum actual heat load of 1174 W. The heat source temperature in this case was equal to 143.5 C, and the evaporator and vapor temperatures had values of 109.8 and 103.2 C, respectively. These are too high temperatures for modern electronic components. However, even at a temperature of 80 C, which is admissible for many of them, the thermal resistance of the evaporator remained sufficiently low and was equal to 0.02 C/W. Despite the relatively small heating surface Fe, so low values of the evaporator thermal resistance were achieved owing to the high values of the heat-transfer coefficient ke in the evaporation zone, which was calculated by the formula:
ke ¼
Q : Fe ðTe Tv2 Þ
(14)
In the whole range of heat loads the value of ke increased from 18,000 W/m2 C to 198,000 W/m2 C. If instead of Tv2 in formula (14) we put Tv1, which is used by many authors as the vapor temperature for calculating ke, the values of the latter will decrease on the average by 22%. In their turn, such values of ke were achieved owing to the high thermal conductivity of the copper used for making the body and the wick of the evaporator, a good thermal contact between them, and also optimum organization of the evaporation zone. It can also be seen (Fig. 11) that a “weak” point of the system is the thermal resistance of the condenser, whose minimum value of Thermal resistance, oC/W
Thermal resistance, oC/W
Rs ¼
0,25
Rs ¼ Rhse þ RLHP þ Rccool :
Tv3 þ Tll : 2
As with the evaporator, the thermal resistance of the condenser may be written in the following way:
Rc ¼
0,30
(6)
Since in the conditions of the present experiment it is very difficult to precisely determine the average temperature of the condenser by means of direct measurements, it was calculated as follows:
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0,16 0,12 0,08 0,04 0,00 0
0,00 0
200
400 600 800 Heat load, W
1000
1200
Fig. 10. Heat load dependence of the evaporator thermal resistance.
200
400
600
800
1000
1200
Heat load, W Fig. 12. Heat load dependence of the thermal resistance of heat transfer from the condenser to the cooling medium.
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Thermal resistance, oC/W
0,40 0,30 0,20 0,10 0,00 0
200
400
600
800
1000
1200
Heat load, W
itself (Fig. 13), its minimum value of 0.044 C/W was obtained in the same way as the minimum thermal resistance of the evaporator at a maximum heat load. In this case an LHP thermal resistance corresponding to a heat-source temperature of 80 C and equal to 0.055 C/W was achieved at a heat load of about 600 W, the vapor temperature being at a level of about 56 C. It can also be noticed that the thermal resistance of the system Rs varied only slightly in a range of heat loads above 500 W (Fig. 14). Its minimum value of 0.097 C/W was obtained at a heat load of about 700 W, which corresponded to the minimum thermal resistance of the condenser. 6. Conclusion
Fig. 13. Heat load dependence of the LHP thermal resistance. C/W
Thermal resistance, oC/W
0.035 was achieved at a heat load of about 700 W, when the vapor temperature in the LHP was equal to 62.5 C. This is twice as large as the thermal resistance of the evaporator at the same heat load. The relatively high thermal resistance of the condenser is caused here by at least two circumstances. First, under film condensation, which is observed in the present case, the heattransfer intensity is lower than in the evaporation zone. Second, not the whole condenser participates in the condensation process, but only its part that is free of liquid. In this case, depending on the heat load value, the length of the condensation zone and the magnitude of the condensing surface corresponding to it vary. The other, internal heat-transfer part of the condenser, which is filled with liquid, is not involved in the condensation process. It serves only to supercool the condensed working fluid, and the heattransfer intensity here, as in a one-phase flow, decreases still further. It is quite evident that the efforts aimed at a further reduction in the LHP thermal resistance should be concentrated on decreasing the thermal resistance of the condenser. Work is now underway in this direction [18e21]. On this way, however, there are problems connected with the fact that for the time being it is impossible to achieve as high heat-transfer coefficients as in the evaporation zone. At the same time the enhancement of the heattransfer surface in the condenser is very often limited by its dimensions, which in view of the well-known reasons cannot be sufficiently large. Very often another “weak” point of the cooling system is the thermal resistance of heat transfer from the condenser to the cooling medium Rc-cool. In this case, however, owing to the intense cooling of the condenser, by means of forced water convection and the condenser relatively large length, this thermal resistance (Fig. 12) in the heat load range above 200 W has proved to be practically equal to the thermal resistance Rhs-e. Therefore, from the view point of thermal resistance this system has proved to be sufficiently well balanced. As for the thermal resistance of the LHP
0,6 0,5 0,4 0,3 0,2 0,1 0,0 0
200
400 600 800 Heat load, W
1000
1200
Fig. 14. Heat load dependence of the thermal resistance of the “heat sourceeLHPecooling medium” system.
Possibilities of a considerable decrease in the operating temperature of vapor in a copperewater loop heat pipe made entirely of copper have been determined and demonstrated for the first time. In particular, at heat loads from 5 W to 1200 W the vapor operating temperature in a compact copperewater LHP with a flateoval evaporator, whose condenser was cooled by running water with a temperature of 20 C, was in the range from 26.5 C to 103.4 C. In this case the vapor temperature remained lower than 40 C up to a heat load close to 250 W. This, in its turn, made it possible to maintain the temperature of the heat source at a sufficiently low level equal to 55.6 C. A minimum thermal resistance of the LHP equal to 0.044 C/W was achieved at a maximum heat load which was not limiting and was restricted only by the value of the excess vapor pressure in the LHP. The corresponding values of thermal resistance of the evaporator and the condenser were in this case 0.006 C/W and 0.038 C/W. It is also noted that the “weakest” point of the “heat sourceeLHPecooling water” system is the thermal resistance of the condenser, whose minimum value of 0.035 C/W was achieved at a heat load of about 700 W. The thermal resistance of the system in this case also had a minimum value and was equal to 0.097 C/W. It was obtained owing to the good thermal contact of the heat source with the evaporator, the LHP low thermal resistance and the sufficient enough cooling of the condenser. References [1] Yu.F. Maydanik, Yu.G. Fershtater, Theoretical basis and classification of loop heat pipes and capillary pumped loops, Preprint of the 10th Int. Heat Pipe Conf., Stuttgart, Germany, 1997. [2] Yu.F. Maydanik, Loop heat pipes (review), Applied Thermal Engineering 25 (2005) 635e657. [3] Yu.F. Maydanik, Miniature loop heat pipes, in: Proceedings of the 13th Int. Heat Pipe Conf, Shanghai, China, 2004, pp. 24e37. [4] R. Singh, A. Akbarzadech, C. Dixon, M. Mochizuki, R.R. Riehl, Miniature loop heat pipe with flat evaporator for cooling computer CPU, IEEE Transactions on Computers and Packaging Technologies 30 (1) (2007) 42e49. [5] R. Singh, A. Akbarzadech, C. Dixon, M. Mochizuki, T. Nguen, R.R. Riehl, Miniature loop heat pipe with different evaporator configuration for cooling compact electronics, in: Proceedings of the 14th Int. Heat Pipe Conf., Florianopolis, Brazil, 2007, pp. 176e181. [6] Yu. Maydanik, S. Vershinin, M. Chernysheva, Evaporation chamber of a loop heat pipe, Russian Patent 2170401, 2001. [7] W. Zimbek, G. Slavik, J. Cennamo, S. Kang, J. Yun, E. Kroliczek, Loop heat pipe technology for cooling computer servers, in: 11th ITHERM Conf., Orlando, USA, 2008, pp. 19e25. [8] R. Singh, A. Akbarzadech, C. Dixon, M. Mochizuki, T. Nguen, Novel design of miniature loop heat pipe and their future potential for cooling high power laptops, in: Proceedings of the 9th Int. Heat Pipe Symp., Kuala Lumpur, Malaysia, 2008, pp. 244e253. [9] J. Li, D. Wang, G.P. Peterson, Experimental studies on a high performance loop heat pipe with a square flat evaporator, Applied Thermal Engineering 30 (2010) 741e752. [10] Yu.F. Maydanik, S.V. Vershinin, Development and investigation of copper water LHP with high operating characteristics, Preprint of the 15th Int. Heat Pipe Conference, Clemson, USA, 2010. [11] V.G. Pastukhov, Yu.F. Maydanik, Active coolers based on the copperewater LHPs for a desktop PC, Applied Thermal Engineering 29 (2009) 3140e3143.
Yu. Maydanik et al. / Applied Thermal Engineering 31 (2011) 3533e3541 [12] V.G. Pastukhov, Yu.F. Maydanik, LHP-based coolers for graphic cards, Preprint of the 15th Int. Heat Pipe Conf., Clemson, USA, 2010. [13] V.G. Pastukhov, Yu.F. Maydanik, Development and results of testing coolers on the basis of copperewater LHPs for a desktop PC, in: Proceedings of the 14th Int. Heat Pipe Conference, Florianopolis, Brazil, 2007, pp. 157e162. [14] Yu. F. Maydanik, N.N. Solodovnik, Yu.G. Fershtater, Investigation of dynamic and stationary characteristics of a loop heat pipe, in: Proceedings of the 9th Int. Heat Pipe Conference, Albuquerque, USA, 1995, pp. 1002e1006. [15] J. Ku, High frequency low amplitude temperature oscillations in loop heat pipe operation, in: Proceedings of the Int. Two-Phase Thermal Control Technology Workshop, Noordwijk, Netherlands, 2003. [16] J. Ku, J. Rodriguez, Low frequency high amplitude temperature oscillations in loop heat pipe operation, in: Proceedings of the Int. Two-Phase Thermal Control Technology Workshop, Noordwijk, Netherlands, 2003.
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[17] S.V. Vershinin, Yu.F. Maydanik, Investigation of pulsation of the operating temperature in a miniature loop heat pipe, International Journal of Heat and Mass Transfer 50 (2007) 5232e5240. [18] M. Chernysheva, S. Vershinin, Yu. Maydanik, Heat transfer during condensation of moving steam in a narrow channel, International Journal of Heat and Mass Transfer 50 (2009) 2437e2443. [19] M. Goncharov, V. Buz, A. Kochetkov, Modeling and experimental researches of condensation in LHP, Preprint of the 15th Int. Heat Pipe Conference, Clemson, USA, 2010. [20] C. Kovael, T. Peters, H. Kariya, J. Allison, M. McCarthy, J. Brisson, E. Wang, Design of parallel plate condensers with sintered wick for a capillary-pumped loop heat pipe, Preprint of the 15th Int. Heat Pipe Conference, Clemson, USA, 2010. [21] M. Chernysheva, E. Bartuli, Yu. Maydanik, Heat-exchange processes in a splittype condenser of a copperewater loop heat pipe, Thermal Processes in Engineering 8 (2) (2010) 354e363.
Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Investigation of a compact copperewater loop heap pipe with a flat evaporator Yu. Maydanik*, S. Vershinin, M. Chernysheva, S. Yushakova Institute of Thermal Physics, Ural Branch of the Russian Academy of Sciences, Amundsen St. 106, Ekaterinburg 620016, Russia
a r t i c l e i n f o
a b s t r a c t
Article history: Received 10 November 2010 Accepted 1 July 2011 Available online 18 July 2011
A compact copperewater loop heat pipe (LHP) with an effective length of 310 mm equipped with a flat eoval evaporator measuring 80 (L) 42 (W) 7 (H) has been tested. The vapor line and the condenser had the same internal diameter of 5.4 mm. The internal diameter of the liquid line was 3.4 mm. Tests were conducted with a heat source which had a heating surface of 30 mm 30 mm. The condenser was cooled by running water with a temperature of 20 C. In the horizontal position the device has exhibited serviceability in the heat load range from 5 W to 1200 W at vapor temperatures from 26.5 C to 103.4 C. The maximum capacity was achieved at a heat source temperature of 143.5 C, when the LHP thermal resistance was equal to 0.044 C/W. The corresponding values of thermal resistance for the evaporator and the condenser were at a level of 0.006 C/W and 0.038 C/W. A minimum thermal resistance of 0.097 C/W for the “heat sourceeLHPecooling water” system was obtained at a heat load of about 700 W, at which the temperature of the heat source was 87 C. Ó 2011 Elsevier Ltd. All rights reserved.
Keywords: Electronics cooling Loop heat pipe Flateoval evaporator Heat load Thermal resistance Start-up
1. Introduction For quite a long time there existed a stereotype idea that the “copperewater” combination is unsuitable for use in loop heat pipes operating in temperature range below 100 C. For this there was quite a logical substantiation. The fact is that the high thermal conductivity of copper on the one hand and the low value of the derivative dP/dT, which characterizes the slope of the saturation line of the working fluid in the temperature range mentioned, on the other impede the creation of the required drop of the vapor temperature DT and the corresponding pressure drop DP between the evaporation zone and the compensation chamber. As is well known [1,2], such a drop is necessary for a start-up and maintenance of the circulation of a working fluid in an LHP. At the same time, copperewater LHPs are an extremely attractive object for development and practical application because this combination is well compatible chemically and safe from an ecological and a technological point of view. Besides, it allows manufacture of sufficiently cheap devices during their mass production. An important stimulus for the development of copperewater LHPs is also the fact that the thermal properties of copper and water with a correct approach open up possibilities for the achievement of exceptionally high thermal characteristics, including the maximum heat-transfer capacity, the heat flux and the low thermal resistance. At present these factors are in great demand for the creation of efficient and reliable cooling systems of powerful * Corresponding author. Tel.: þ7 (343) 267 87 91; fax: þ7 (343) 267 87 99. E-mail address: [email protected] (Yu. Maydanik). 1359-4311/$ e see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2011.07.008
electronics and photoelectronics used in various fields of technology. It should be expected that such a tendency will remain intact in the foreseeable future. Owing to purposeful efforts in this direction, the first miniature all-copper LHPs with water as a working fluid were created and tested as long ago as the early 2000s [3]. These devices were equipped with cylindrical evaporators 6 mm in diameter or flateoval evaporators 3.2 mm thick with thermal resistances at a level of 0.09 C/W and 0.05 C/W, respectively. The evaporators were connected to the condenser by tubes with an internal diameter of 2 mm. Under forced air cooling of the condenser LHPs exhibited serviceability in the vapor temperature range from 50 to 100 C at heat loads from 10 to 160 W. In this case the ability to operate efficiently at different orientations in the gravity field, which is characteristics of loop heat pipes, was retained in full measure. About 20 such devices were made and tested in searching for optimum designs in the period from 2000 to 2003. Quite an important advantage of water as a working fluid for LHPs is also the fact that its vapor pressure at operating temperatures below 100 C allows making flat evaporators with a relatively thin wall. Besides, the flat form of the thermocontact surface in such evaporators in many cases makes it possible to avoid the use of thermal interfaces (thermal adapters) with which cylindrical evaporators are equipped for ensuring a good thermal contact with the heat sources. Three main types of flat copper evaporators have been developed thus far: disk-shaped, rectangular and flateoval evaporators. Disk-shaped evaporators [4,5] have a cylindrical side surface and two flat butt-end surfaces, one of which serves for the heat load
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Nomenclature F H k L P Q R T W DP DT SPi
surface, m2 height, mm heat-transfer coefficient, W/m2 C length, mm pressure, Pa heat load, W thermal resistance, W/ C temperature, C width, mm pressure drop, Pa temperature drop, C total amount of pressure losses, Pa
supply. There are also designs where both flat surfaces of the evaporator may be used for supplying heat loads [6]. The main drawback of disk-shaped evaporators is the fact that their thickness, which includes the compensation chamber, cannot be sufficiently small. Usually it varies within 10e20 mm. Singh et al. [4], for instance, presented the results of testing a miniature copperewater LHP equipped with a disk-shaped evaporator 30 mm in diameter and 10 mm in thickness with a nickel wick. This device 150 mm long with 2-mm transportation lines was capable of transferring from 5 to 70 W at evaporator temperatures from 57 to 99.6 C. The minimum value of thermal resistance of the evaporator was achieved at a maximum heat load of 70 W and was equal to 0.06 C/W. The heat-transfer coefficient that corresponded to it was equal to 22,600 W/m2 C. Zimbeck et al. [7] reported the results of testing two copperewater LHPs with disk-shaped evaporators having a thickness less than 25 mm provided with a ceramic wick. On trials with a CPU thermal simulator, which had a heating surface of 20 mm 20 mm, the heat load varied in the range from 50 to 200 W. The condenser was cooled by means of forced air convection with a temperature of 25 C. The evaporator temperature in this case increased within 48e65 C. At a maximum heat load of 200 W the thermal resistance of the evaporator was equal to 0.05 C/W, and the heat-transfer coefficient corresponding to it reached a value of about 50,000 W/m2 C. Rectangular flat evaporators may have both one and two thermocontact surfaces. The compensation chamber is located here on any of the sides depending on the configuration of the evaporator, whose thickness may be considerably lower as compared with disk-shaped evaporators. Thus, for instance, Singh et al. [8] reported the results of testing a miniature copperewater LHP with a flat rectangular evaporator which measured 60 mm (L) 50 mm (W) 5 mm (H) and was also equipped with a nickel wick. The active zone measuring 22 mm 22 mm was on one of the evaporator sides, and the compensation chamber was located in the same plane as the wick and enveloped it on three sides. The device exhibited serviceability in the range of heat loads from 10 to 50 W, at which the evaporator temperature varied in the range from 75 to 100 C. A minimum value of thermal resistance of the evaporator equal to 0.25 C/W was achieved at a maximum heat load of 50 W. Li et al. [9] presented the results of testing a copperewater LHP with a flat square evaporator with dimensions 30 mm (L) 30 mm (W) 15 mm (H). The copper wick structure was sintered directly on the evaporator substrate, and the compensation chamber was located above the wick. The vapor and liquid lines with an internal diameter of 5 mm had the same length equal to 120 mm. The condenser was a serpentine-shaped one, was provided with finning and was cooled by a 120-mm axial fan. The device was tested at a vertical orientation in a gravity assisted mode when the evaporator was lower than the condenser. A heater with
Subscripts s condenser s-cool condenser-cooling medium cc compensation chamber e evaporator hs heat source hs-e heat source-evaporator l liquid ll liquid line s system v vapor vl vapor line
a heating surface 25 mm 25 mm was used as a heat source. Tests were conducted at an ambient temperature 23.7 2.1 C. The heat load increased stepwise from 10 to 600 W. The evaporator temperature in this case varied from 50 to 98 C. The minimum thermal resistance of the LHP achieved at a heat load of 450 W was 0.033 C/W. These are fairly high results. But they cannot characterize in full measure the efficiency of the device under discussion as they have been obtained in “too favorable” conditions for LHP, which are gravity assisted modes. Flateoval evaporators may also have both one and two thermocontact surfaces located on the opposite flat sides. In both cases the compensation chamber is in the same plane as the wick, and therefore the thickness of the evaporators can be easily varied within the limits from 3 mm and more. The length and the width of such evaporators are determined by the dimensions of the heat source and the volume of the compensation chamber. The authors of Ref. [10] have recently presented the results of testing a copperewater LHP equipped with a flateoval evaporator 7 mm thick. The dimensions of the active zone of the evaporator were 32 mm 32 mm. The vapor and liquid lines, and also the condenser had the same internal diameter of 3.4 mm and length of 120 mm, 180 mm and 335 mm, respectively. The devise was tested in the horizontal position with heat source that had heating surfaces of 9 cm2 and 1 cm2. The condenser was cooled by thermostatted running water with a temperature 21 1 C. The heat load varied in the range from 20 to 900 W. A minimum value of thermal resistance of the evaporator equal to 0.014 C/W was achieved at the maximum heat load with a heat source which had a heating surface of 9 cm2 when the corresponding heat-transfer coefficient was equal to 79,000 W/m2 C. It should be mentioned that for many components of electronics a limiting temperature of the heat source equal to 100 C was achieved here at a heat loads of about 550 W, which corresponds to a heat flux of 61 W/cm2. The presented results obtained by various authors strongly suggest that copperewater LHPs make it possible to ensure sufficiently high values of heat loads and heat fluxes in the heating zone. However, in most cases these characteristics are achieved only at a relatively high level of the LHP operating temperature, which is unacceptable for most electronic components. Therefore the task of decreasing the operating temperature with retention of high thermal characteristics is quite topical. It is at the solution of this problem that the present work was aimed. 2. Analysis of the main factors that affect the LHP operating temperature The fundamental parameter that characterizes the LHP operation in the most impressive way is the vapor temperature. Therefore it is conveniently used for analyzing the thermal and
Yu. Maydanik et al. / Applied Thermal Engineering 31 (2011) 3533e3541
n X i¼1
DPi ¼
dP DT; dT
(1)
where SDPi is the total amount of pressure losses in the segments of the working fluid circulation from the evaporating surface of the wick to its absorbing surface on the side of the compensation chamber, dP/dT is the derivative that characterizes the slope of the saturation line of the working fluid at a given temperature, DT is the vapor temperature drop between the evaporating surface of the wick and the vaporeliquid interface in the compensation chamber. In this turn, the product of the quantities in the right-hand side of this equation is the pressure drop determined by the difference of the vapor temperatures between the evaporation zone and the compensation chamber. Thus, for ensuring an LHP start-up and operation it is necessary that the pressure losses during the circulation of the working fluid between the evaporating and absorbing surfaces of the wick SDPi should not exceed the difference of the pressures of vapor DP connected with the difference of its temperatures DT between the surfaces mentioned above. Evidently the low value of dP/dT, which is characteristic of water at low temperatures, may be compensated either at the cost of increasing DT or at the cost of decreasing SDPi, or at the cost of both simultaneously. As for the value of DT, its increase is considerably inhibited by the high thermal conductivity of the copper body and the wick, which contributes to the increase of parasitic heat flows into the compensation chamber. The use of a wick made of less heat-conducting materials for decreasing these flows does not solve the whole problem as in this case one can observe a considerable reduction in the efficiency of heat transfer in the evaporation zone. The problem of increasing DT may be solved at the cost of an increase in the thickness of the “barrier” layer of the wick, which separates the evaporation zone and the condensation chamber. This way is fairly effective when the working fluid is water because the high capillary pressure that it is capable of creating compensates easily an increase in the hydraulic resistance of the wick. In this case for evaporators which have a butt-end compensation chamber an increase in the wick thickness leads to an increase in the thickness of the evaporator itself, and for evaporators with a compensation chamber located in the same plane as the wick this results in increasing evaporator length. The choice of one or the other variant is determined by the conditions of the LHP use.
dP/dT, kPa/K
4 3 2 1 0
20
30
40
50
60
70
80
90 100 110
o
Tv, C Fig. 1. Temperature dependence dP/dT for water.
Another way of realizing condition (1) is connected with a decrease in the value of SDPi. This may be first of all achieved by decreasing pressure losses in the LHP vapor line. For water, which has a low vapor density in a low-temperature range, even a small increase in the diameter of the vapor line makes it possible to considerably decrease these losses. Figs. 1 and 2 present as examples, respectively, temperature dependences of the value of dP/dT and the amount of pressure losses in a vapor line 300 mm long with different diameters at a heat load of 50 W. Here it can be seen, in particular, that at a vapor temperature of 30 C pressure losses in a vapor line 3 mm in diameter are approximately equal to 880 Pa, and the value of dP/dT at this temperature is 230 Pa/K. Evaluation by formula (1) gives a value of DT equal to 3.8 C. It is difficult to attain such a value in a copper evaporator at a heat load of 50 W. At the same time at a temperature of 60 C a value of DT equal to 0.25 C may be quite sufficient. Approximately the same result can also be obtained at a vapor temperature of 30 C by increasing the diameter of the vapor line to 6 mm. 3. Description of an experimental LHP A copperewater loop heat pipe with a flateoval evaporator was specially made for experiments in the framework of this work. Such evaporators are easy to make, and besides they have shown themselves to good advantage in testing various versions of copperewater LHPs [3,10e13]. In this case the LHP had a flateoval evaporator measuring 80 mm (L) 42 mm (W) 7 mm (H). The active zone of the evaporator (the zone to which a heat load may be applied) measured 32 mm 32 mm. The evaporator was provided with a copper wick with a porosity of about 67% and a break-down pore radius of 21 microns. Twelve longitudinal channels 1, 8 mm in diameter were made in the wick. They formed the evaporating
0,9 0,8 0,7
ΔPvl, kPa
hydrodynamic processes that take place in these devices. It is quite evident that the lower limiter of the LHP operating temperature is the temperature of the medium which cools the condenser and the upper one the temperature of the heat source. The closer the vapor temperature to the temperature of one and the other, the more efficient the LHP operation, and the lower the thermal resistance of the “heat sourceecooling medium” system. The problem here consists in ensuring the operation of a copperewater LHP at a lower temperature level with respect to the ambient temperature retaining at the same time an acceptable efficiency of the devise operation. An analysis of the results obtained by various authors on trials of copperewater LHPs [3e5,7e10] provided with both copper and some other, less heat-conducting, wicks shows that the minimum value of the vapor temperature was always at a level above 40 C even at minimum heat loads. The operating vapor temperature in an LHP depends on many factors, among which the thermal properties of the working fluid occupy an important place. The thermal properties of water in temperature ranges below 100 C, such as the density and viscosity of vapor, and also the value of dP/dT, make it not a very “convenient” working fluid for LHPs. To understand this, one should turn to the second, the so-called thermodynamic condition of the LHP serviceability, which may be written as follows:
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d=3 mm
0,6 0,5 0,4
d=4 mm d=5 mm d=6 mm
0,3 0,2 0,1 0,0
20
30
40
50
60 70 Tv, oC
80
90
100 110
Fig. 2. Temperature dependence of pressure losses in a vapor part 300 mm long with different diameters.
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surface and at the same time served for removing vapor from the evaporation zone. On one of the flat sides of the evaporator, near which the vapor-removal channels were located, there was a special thermal interface, i.e. a copper plate 1 mm thick measuring 30 mm 30 mm, which was soldered to the surface of the active zone. Despite the fact that such an interface creates an additional thermal resistance, its use was dictated by the necessity of locating thermocouples for temperature measurements right on the evaporator wall. For this purpose the plate had a through cut 1 mm wide, where a thermocouple was placed. The vapor line 120 mm long and the condenser with a length of 335 mm had the same internal diameter equal to 5.4 mm. The liquid line 180 mm in length had an internal diameter of 3.4 mm. The length of the vapor path of the device with allowance for the length of the vaporremoval channels in the evaporation zone, the vapor line and the part of the condenser which rids itself of the liquid during the operation was about 370 mm. The condenser was provided with a spiral fin located in its outer surface and a plastic jacket for pumping the cooling water. The general view and the scheme of the copperewater loop heat pipe are presented in Figs. 3 and 4, respectively. 4. Testing procedure All tests in the framework of these investigations were conducted at a horizontal orientation of an LHP, when the evaporator and the condenser were located in one and the same horizontal plane. The heat load was delivered to the evaporator from below. The heat load source was a copper cylindrical block, in which 6 heating cartridges of 250 W each were positioned parallel to the longitudinal axis. One of the butt-ends of the copper block had a square ledge, whose heating surface measuring 30 mm 30 mm was in thermal contact with the thermal interface of the evaporator. The whole copper block was in thermal insulation, limited from the outside by a thin-walled metal shell. Three thermocouples were located at the shell surface, and their readings made it possible to evaluate the heat losses from the heat source into the outside ambient and to determine the magnitude of the actual heat load supplied to the evaporator. The scheme of the heat source is shown in Fig. 5. The electric power supplied to the heaters was measured by a wattmeter and changed stepwise with a step of 20 W to a value of 100 W, and then with a step of 100 W to the maximum magnitude. The magnitude of the maximum power was limited by the vapor temperature in the LHP equal to 105 C, at which the internal pressure reached 120 kPa. A further increase in the pressure was liable to cause deformation of the thin-walled body of the evaporator. The condenser was cooled by running thermostatted water with a temperature of 20 0.1 C, which was pumped through the
Fig. 3. General view of a copperewater LHP.
Fig. 4. Scheme of a copperewater LHP.
cooling jacket with a flow rate equal to 0.5 103 m3/min. The LHP operating temperature was measured at several points with the help of coppereconstantan thermocouples “OMEGA” TT-T-30. The temperature of the evaporator wall at the center of the heating zone was measured by a thermocouple which was located in the above-mentioned special cut of the thermal interface. To measure the vapor temperature, use was made of a thermocouple positioned in a metal capillary which was inserted into the vapor line. The sufficiently large diameter of that line allowed doing it without any
Fig. 5. Scheme of a heat source.
5.1. LHP start-up In the opinion of some authors in Refs. [4,9], which is based on their own experimental experience, a start-up at low heat loads is a problematic operating mode for LHPs. The problem here is connected with the fact that at low heat loads a start-up does not take place (there is no circulation of the working fluid), or a start-up is accompanied by strong temperature pulsation, whose amplitude may reach 10 C. In this case it is assumed that the conditions leading to such phenomena are inherent properties for all LHPs regardless of what manufacturing process is utilized. Special tests were conducted to check how critical the start-up process is for the given LHP. Here it would be expedient to mention that the notion of “low heat load” is relative as the essential factor is the magnitude of the surface over which it is distributed. It would be more adequate to use the term “heat flux”, which with allowance for the specific character of LHPs with water as a working fluid may be considered low if its value does not exceed 5 W/cm2. For ammonia LHPs, for instance, under adequate conditions a low heat flux may be 1 W/cm2 and lower. Fig. 6 presents start-up processes at heat loads of 5 W, 20 W, 100 W, 200 W, which approximately correspond to heat fluxes of 0.5 W/cm2, 2.2 W/cm2, 11.1 W/cm2 and 22.2 W/cm2. The first two of them may be referred to low heat fluxes, the third and the fourth to moderate ones. It is easy to see that common to all of them is the presence of the so-called temperature “overshoot”, after which one can observe a rapid temperature drop and an actual LHP start-up. The higher the heat load, the faster the heating and the subsequent start-up. Thus, for instance, at a heat load of 5 W an actual LHP start-up takes place in about 60 min, and at 200 W in 2.2 min after a heat load delivery. In this case the time of going into a steady state level is noted, respectively, in about 100 min and 7 min. At the same time the absolute ceiling of the temperature “overshoot” at all the above-mentioned heat loads remains approximately the same and is within 42e44 C for the evaporator and 31e33 C for vapor. As is shown in Ref. [14], such a picture corresponds entirely to one possible scenario of an LHP start-up, when in the initial state vaporremoval channels in the evaporation zone are flooded with a working fluid and there is no completed “vaporeliquid” interface here. For such an interface to appear the boiling-up of a working fluid is required, which is accompanied by some superheating corresponding in value to the temperature “overshoot”. However, such a start-up scenario is not the only possible one. When for some reasons or other there is a free “vaporeliquid” interface in the evaporation zone, a start-up may proceed “smoothly”, without
45 5W
40 35
Te
30 25
Tv2
20
Tl1
0
Te mpera ture, oC
In LHP trials a study was made of all the main thermal parameters that characterize the behavior of a “heat sourceeLHPecooling water” system under changes of the heat load when the ambient conditions remained constant. The heat load dependence of the LHP temperature at characteristic points, the heat-transfer coefficient in the evaporation zone and thermal resistance in some sections and in the system as a whole was determined. Investigations were also made of the LHP behavior during start-ups at low and moderate heat loads.
3537
15 2000
4000
6000
45 20 W
40 35
Te
30
Tv2
25 20
Tl1
15 0
Tem pe rature, oC
5. Test results and discussion
400
800
1200
45 100 W
40
Te
35 30
Tv2
25
Tl1
20 15 0
Tem pe rature, oC
essential changes in the local hydraulic resistance to the vapor flow. Besides, two thermocouples were located on the wall of the vapor line. One thermocouple was also used for measuring temperature near the thermocontact surface of the heat source. The readings of the thermocouples were recorded by a data acquisition unit Agilent 34970A. The scheme of location of the thermocouples is given in Fig. 4.
Tempe rature, oC
Yu. Maydanik et al. / Applied Thermal Engineering 31 (2011) 3533e3541
200
400
600
800
45 Te
200 W
40
Tv2
35 30 25
Tl1
20 15 0
100
200
300
400
500
Time, s Fig. 6. Process of an LHP start-up at different heat loads.
being accompanied by a temperature “overshoot”, as is shown, for instance, in Fig. 7. The presence or the absence of a completed “vaporeliquid” interface in the evaporation zone depends on the initial distribution of a working fluid in an LHP connected with its “thermal prehistory”. As for studying the start-up process at the minimum heat load at which a stable circulation of the working fluid in an LHP begins, it should be mentioned that it has an academic rather than a practical meaning. It is rather difficult to imagine some exotic case when an LHP intended for operation at a nominal heat load of, for instance, 150 W will start up in the process of operation at a heat load of 5 or even 10 W. As a rule, an actual range of operating heat loads is relatively small. Thus, in particular, a CPU dissipating 100 W at a maximum load in the standby mode, when the heat load is
Yu. Maydanik et al. / Applied Thermal Engineering 31 (2011) 3533e3541
Te mpe rature , oC
3538
40 5W
35
Te
30 25
Tv2
20
Tl1
15 0
2000
4000 Time, s
6000
Fig. 7. Process of a “smooth” LHP start-up at a heat load of 5 W.
minimum, dissipates 40e50 W. Therefore, if it is necessary to remove heat from an object dissipating 5 or 10 W, an appropriate LHP has to be made, for which such heat loads are nominal. It is also significant that in the process of LHP tests at the above-mentioned heat loads the start-up was not accompanied by considerable temperature pulsations. The maximum amplitude of vapor temperature pulsation did not exceed 0.6 C, and the amplitude of temperature pulsation of the evaporator and the heat source was still lower. The main reason and kinds of LHP temperature pulsations are analyzed at great length in Refs. [15e17]. Here it may be added that the operation of LHPs without temperature pulsations is impossible in principle, as well as of any other closed system where joined processes of evaporation/boiling and condensation take place and where there are movable interfaces between the vapor and liquid phases of the working fluid. Difference may lie in the organization of these processes and the parameters of pulsations. If the processes of heat and mass transfer in an LHP are organized optimally, the amplitude of pulsations is small and they are of little practical importance. 5.2. LHP operating characteristics The LHP operating characteristics are presented in Fig. 8 as heat load dependences of the temperature of the heat source Ths, evaporator Te, vapor Tv2 and compensation chamber Tcc. The type of the curve Te ¼ f(Q), which has a near-linear form, not quite typical of LHPs, may point to the absence of a variable-conductance mode. Such a mode is usually observed as long as there is a redistribution of a liquid between the condenser and the compensation chamber. However, the vapor temperature, which is more sensitive to the situation inside the LHP, and which is presented by the curve Tv2 ¼ f(Q), shows that such a mode exists up to heat loads of about 300 W, though in this case it is not strongly pronounced. Of particular importance is the fact that the vapor temperature in the LHP remains lower than 40 C up to heat loads close to 250 W. In this case the minimum value of the vapor temperature obtained in
Temperature, oC
160
5.3. Thermal resistance The thermal resistance is one of the most important thermal parameters that characterize the LHP efficiency. The lower is the thermal resistance of a heat-transfer device, the closer to the temperature of the medium cooling the condenser may be the temperature of the heat source. The thermal resistance of an LHP is mainly determined by the thermal resistance of the evaporator and the condenser as the thermal resistance of the vapor and liquid lines in most cases may be neglected. To analyze thermal resistance, it is convenient to use the scheme presented in Fig. 9. In this scheme thermal resistances are determined in the following way: 1) The thermal resistance from the heat source to the evaporator
Rhse ¼
Thse Te Q
2) The thermal resistance of the evaporator is usually determined by the readings of thermocouples located on the evaporator wall in the heating zone and on the vapor line as
Re ¼
Te Tv1 : Q
In this case, however, a more precise value of Re can be determined by using the readings of the thermocouple Tv2 located inside the vapor line:
Te Tv2 : Q
Ths
40
(4)
Te Rhs-e
0 200
(3)
The thermal resistance of the evaporator may also be presented in a different form, which reflects more precisely the physical meaning of this quantity:
80
0
(2)
is determined by the quality of finish of the conjugate surfaces, the force of pressing, and the thermal conductivity of the grease used for improving the thermal contact. Its value is practically independent of the heat load, but it may make a considerable contribution to the thermal resistance of the whole system. In this case its value was 0.027 20%.
Re ¼
Ths Te Tv2 Tcc
120
LHP trials was at a level of about 26.5 C. In accordance with the vapor temperature, the temperature of the evaporator and the temperature of the heat source were also at a lower level. In particular, a temperature of the heat source Ths equal to 80 C, which is an admissible temperature for many components of electronics, was achieved at a heat load of about 600 W, the evaporator temperature being equal to 67 C, and the vapor temperature to 56 C. If we considered a CPU capable of dissipating 130 W at a maximum heat load as a heat source, its temperature with such an LHP would be approximately 43 C. It should also be mentioned that the limiting heat load for this LHP was not achieved owing to the limitations connected with the vapor pressure. At an actual maximum heat load of 1174 W, which corresponds to a heat flux in the heating zone equal to 130.5 W/m2, no signs of heat-transfer crisis in the evaporator were observed.
400
600
800
Heat load, W Fig. 8. LHP operating characteristics.
1000
1200
Tv1 Re
Tv3 Rvl
Tl2
Tl1
Tc Rc
Tcool Rc-cool
Rll Fig. 9. Scheme of thermal resistances of the “heat sourceeLHPecooling medium” system.
Re ¼
1 ; ke Fe
Thermal resistance, oC/W
Yu. Maydanik et al. / Applied Thermal Engineering 31 (2011) 3533e3541
(5)
where ke is the heat-transfer coefficient in the evaporation zone, Fe is the area of the heating surface. At a fixed value of Fe the value of Re depends on the thickness and thermal conductivity of the evaporator wall and the heat-transfer intensity in the evaporation zone. 3) The thermal resistance of the condenser, by analogy with that of the evaporator, may be determined as
T Tc Rc ¼ v2 : Q
Tc ¼
(7)
1 ; kc Fc
(8)
where kc is the heat-transfer coefficient in the condenser, Fc is the area of the condenser surface. 4) The thermal resistance of an LHP is usually presented by the formula:
RLHP ¼
Te Tc ; Q
(9)
With allowance for (5) and (8) it may also be written as:
RLHP ¼
1 1 þ : ke Fe kc Fc
(10)
5) The thermal resistance of the heat transfer from the condenser to the cooling medium may also have a considerable effect on the resistance of the whole system. It may be determined as:
Rccool ¼
Tc Tcool : Q
(11)
6) The total thermal resistance of the system
Ths Tcool Q
0,20 0,15 0,10 0,05 0,00 0
200
400 600 800 Heat load, W
1000
1200
Fig. 11. Heat load dependence of the condenser thermal resistance.
may also be presented as the sum of the resistances:
(12)
0,08 0,06 0,04 0,02
(13)
Now that all the determinations of thermal resistances are given, we may return to the results of LHP trials. Figs. 10e14 present heat load dependence of the main thermal resistances of the system. It can be seen (Fig. 10) that the lowest thermal resistance is that of evaporator. A minimum value of 0.006 C/W was achieved here at a heat flux of 130.4 W/cm2, which corresponded to a maximum actual heat load of 1174 W. The heat source temperature in this case was equal to 143.5 C, and the evaporator and vapor temperatures had values of 109.8 and 103.2 C, respectively. These are too high temperatures for modern electronic components. However, even at a temperature of 80 C, which is admissible for many of them, the thermal resistance of the evaporator remained sufficiently low and was equal to 0.02 C/W. Despite the relatively small heating surface Fe, so low values of the evaporator thermal resistance were achieved owing to the high values of the heat-transfer coefficient ke in the evaporation zone, which was calculated by the formula:
ke ¼
Q : Fe ðTe Tv2 Þ
(14)
In the whole range of heat loads the value of ke increased from 18,000 W/m2 C to 198,000 W/m2 C. If instead of Tv2 in formula (14) we put Tv1, which is used by many authors as the vapor temperature for calculating ke, the values of the latter will decrease on the average by 22%. In their turn, such values of ke were achieved owing to the high thermal conductivity of the copper used for making the body and the wick of the evaporator, a good thermal contact between them, and also optimum organization of the evaporation zone. It can also be seen (Fig. 11) that a “weak” point of the system is the thermal resistance of the condenser, whose minimum value of Thermal resistance, oC/W
Thermal resistance, oC/W
Rs ¼
0,25
Rs ¼ Rhse þ RLHP þ Rccool :
Tv3 þ Tll : 2
As with the evaporator, the thermal resistance of the condenser may be written in the following way:
Rc ¼
0,30
(6)
Since in the conditions of the present experiment it is very difficult to precisely determine the average temperature of the condenser by means of direct measurements, it was calculated as follows:
3539
0,16 0,12 0,08 0,04 0,00 0
0,00 0
200
400 600 800 Heat load, W
1000
1200
Fig. 10. Heat load dependence of the evaporator thermal resistance.
200
400
600
800
1000
1200
Heat load, W Fig. 12. Heat load dependence of the thermal resistance of heat transfer from the condenser to the cooling medium.
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Yu. Maydanik et al. / Applied Thermal Engineering 31 (2011) 3533e3541
Thermal resistance, oC/W
0,40 0,30 0,20 0,10 0,00 0
200
400
600
800
1000
1200
Heat load, W
itself (Fig. 13), its minimum value of 0.044 C/W was obtained in the same way as the minimum thermal resistance of the evaporator at a maximum heat load. In this case an LHP thermal resistance corresponding to a heat-source temperature of 80 C and equal to 0.055 C/W was achieved at a heat load of about 600 W, the vapor temperature being at a level of about 56 C. It can also be noticed that the thermal resistance of the system Rs varied only slightly in a range of heat loads above 500 W (Fig. 14). Its minimum value of 0.097 C/W was obtained at a heat load of about 700 W, which corresponded to the minimum thermal resistance of the condenser. 6. Conclusion
Fig. 13. Heat load dependence of the LHP thermal resistance. C/W
Thermal resistance, oC/W
0.035 was achieved at a heat load of about 700 W, when the vapor temperature in the LHP was equal to 62.5 C. This is twice as large as the thermal resistance of the evaporator at the same heat load. The relatively high thermal resistance of the condenser is caused here by at least two circumstances. First, under film condensation, which is observed in the present case, the heattransfer intensity is lower than in the evaporation zone. Second, not the whole condenser participates in the condensation process, but only its part that is free of liquid. In this case, depending on the heat load value, the length of the condensation zone and the magnitude of the condensing surface corresponding to it vary. The other, internal heat-transfer part of the condenser, which is filled with liquid, is not involved in the condensation process. It serves only to supercool the condensed working fluid, and the heattransfer intensity here, as in a one-phase flow, decreases still further. It is quite evident that the efforts aimed at a further reduction in the LHP thermal resistance should be concentrated on decreasing the thermal resistance of the condenser. Work is now underway in this direction [18e21]. On this way, however, there are problems connected with the fact that for the time being it is impossible to achieve as high heat-transfer coefficients as in the evaporation zone. At the same time the enhancement of the heattransfer surface in the condenser is very often limited by its dimensions, which in view of the well-known reasons cannot be sufficiently large. Very often another “weak” point of the cooling system is the thermal resistance of heat transfer from the condenser to the cooling medium Rc-cool. In this case, however, owing to the intense cooling of the condenser, by means of forced water convection and the condenser relatively large length, this thermal resistance (Fig. 12) in the heat load range above 200 W has proved to be practically equal to the thermal resistance Rhs-e. Therefore, from the view point of thermal resistance this system has proved to be sufficiently well balanced. As for the thermal resistance of the LHP
0,6 0,5 0,4 0,3 0,2 0,1 0,0 0
200
400 600 800 Heat load, W
1000
1200
Fig. 14. Heat load dependence of the thermal resistance of the “heat sourceeLHPecooling medium” system.
Possibilities of a considerable decrease in the operating temperature of vapor in a copperewater loop heat pipe made entirely of copper have been determined and demonstrated for the first time. In particular, at heat loads from 5 W to 1200 W the vapor operating temperature in a compact copperewater LHP with a flateoval evaporator, whose condenser was cooled by running water with a temperature of 20 C, was in the range from 26.5 C to 103.4 C. In this case the vapor temperature remained lower than 40 C up to a heat load close to 250 W. This, in its turn, made it possible to maintain the temperature of the heat source at a sufficiently low level equal to 55.6 C. A minimum thermal resistance of the LHP equal to 0.044 C/W was achieved at a maximum heat load which was not limiting and was restricted only by the value of the excess vapor pressure in the LHP. The corresponding values of thermal resistance of the evaporator and the condenser were in this case 0.006 C/W and 0.038 C/W. It is also noted that the “weakest” point of the “heat sourceeLHPecooling water” system is the thermal resistance of the condenser, whose minimum value of 0.035 C/W was achieved at a heat load of about 700 W. The thermal resistance of the system in this case also had a minimum value and was equal to 0.097 C/W. It was obtained owing to the good thermal contact of the heat source with the evaporator, the LHP low thermal resistance and the sufficient enough cooling of the condenser. References [1] Yu.F. Maydanik, Yu.G. Fershtater, Theoretical basis and classification of loop heat pipes and capillary pumped loops, Preprint of the 10th Int. Heat Pipe Conf., Stuttgart, Germany, 1997. [2] Yu.F. Maydanik, Loop heat pipes (review), Applied Thermal Engineering 25 (2005) 635e657. [3] Yu.F. Maydanik, Miniature loop heat pipes, in: Proceedings of the 13th Int. Heat Pipe Conf, Shanghai, China, 2004, pp. 24e37. [4] R. Singh, A. Akbarzadech, C. Dixon, M. Mochizuki, R.R. Riehl, Miniature loop heat pipe with flat evaporator for cooling computer CPU, IEEE Transactions on Computers and Packaging Technologies 30 (1) (2007) 42e49. [5] R. Singh, A. Akbarzadech, C. Dixon, M. Mochizuki, T. Nguen, R.R. Riehl, Miniature loop heat pipe with different evaporator configuration for cooling compact electronics, in: Proceedings of the 14th Int. Heat Pipe Conf., Florianopolis, Brazil, 2007, pp. 176e181. [6] Yu. Maydanik, S. Vershinin, M. Chernysheva, Evaporation chamber of a loop heat pipe, Russian Patent 2170401, 2001. [7] W. Zimbek, G. Slavik, J. Cennamo, S. Kang, J. Yun, E. Kroliczek, Loop heat pipe technology for cooling computer servers, in: 11th ITHERM Conf., Orlando, USA, 2008, pp. 19e25. [8] R. Singh, A. Akbarzadech, C. Dixon, M. Mochizuki, T. Nguen, Novel design of miniature loop heat pipe and their future potential for cooling high power laptops, in: Proceedings of the 9th Int. Heat Pipe Symp., Kuala Lumpur, Malaysia, 2008, pp. 244e253. [9] J. Li, D. Wang, G.P. Peterson, Experimental studies on a high performance loop heat pipe with a square flat evaporator, Applied Thermal Engineering 30 (2010) 741e752. [10] Yu.F. Maydanik, S.V. Vershinin, Development and investigation of copper water LHP with high operating characteristics, Preprint of the 15th Int. Heat Pipe Conference, Clemson, USA, 2010. [11] V.G. Pastukhov, Yu.F. Maydanik, Active coolers based on the copperewater LHPs for a desktop PC, Applied Thermal Engineering 29 (2009) 3140e3143.
Yu. Maydanik et al. / Applied Thermal Engineering 31 (2011) 3533e3541 [12] V.G. Pastukhov, Yu.F. Maydanik, LHP-based coolers for graphic cards, Preprint of the 15th Int. Heat Pipe Conf., Clemson, USA, 2010. [13] V.G. Pastukhov, Yu.F. Maydanik, Development and results of testing coolers on the basis of copperewater LHPs for a desktop PC, in: Proceedings of the 14th Int. Heat Pipe Conference, Florianopolis, Brazil, 2007, pp. 157e162. [14] Yu. F. Maydanik, N.N. Solodovnik, Yu.G. Fershtater, Investigation of dynamic and stationary characteristics of a loop heat pipe, in: Proceedings of the 9th Int. Heat Pipe Conference, Albuquerque, USA, 1995, pp. 1002e1006. [15] J. Ku, High frequency low amplitude temperature oscillations in loop heat pipe operation, in: Proceedings of the Int. Two-Phase Thermal Control Technology Workshop, Noordwijk, Netherlands, 2003. [16] J. Ku, J. Rodriguez, Low frequency high amplitude temperature oscillations in loop heat pipe operation, in: Proceedings of the Int. Two-Phase Thermal Control Technology Workshop, Noordwijk, Netherlands, 2003.
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[17] S.V. Vershinin, Yu.F. Maydanik, Investigation of pulsation of the operating temperature in a miniature loop heat pipe, International Journal of Heat and Mass Transfer 50 (2007) 5232e5240. [18] M. Chernysheva, S. Vershinin, Yu. Maydanik, Heat transfer during condensation of moving steam in a narrow channel, International Journal of Heat and Mass Transfer 50 (2009) 2437e2443. [19] M. Goncharov, V. Buz, A. Kochetkov, Modeling and experimental researches of condensation in LHP, Preprint of the 15th Int. Heat Pipe Conference, Clemson, USA, 2010. [20] C. Kovael, T. Peters, H. Kariya, J. Allison, M. McCarthy, J. Brisson, E. Wang, Design of parallel plate condensers with sintered wick for a capillary-pumped loop heat pipe, Preprint of the 15th Int. Heat Pipe Conference, Clemson, USA, 2010. [21] M. Chernysheva, E. Bartuli, Yu. Maydanik, Heat-exchange processes in a splittype condenser of a copperewater loop heat pipe, Thermal Processes in Engineering 8 (2) (2010) 354e363.