Experimental studies on a high performance compact loop heat pipe with a square flat evaporator

Applied Thermal Engineering 30 (2010) 741–752

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Applied Thermal Engineering

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Experimental studies on a high performance compact loop heat pipewith a square flat evaporator

Ji Li a,*, Daming Wang b, G.P. Peterson c

a Laboratory of Electronics Thermal Management, College of Physics, Graduate University of Chinese Academy of Sciences, 19A Yu-quan-lu Road,Shijingshan District, Beijing 100049, PR Chinab Beijing Technology Research Center, Asia Vital Components Co., LTD., Haidian District, Beijing 100085, PR Chinac Woodruff School of Mechanical Engineering, Georgia Institute of Technology, Atlanta, GA 30318, United States

a r t i c l e i n f o

Article history:Received 18 August 2009Accepted 7 December 2009Available online 11 December 2009

Keywords:Electronics coolingLoop heat pipeFlat evaporatorStartupSteady-stateHigh performance

1359-4311/$ - see front matter � 2009 Elsevier Ltd. Adoi:10.1016/j.applthermaleng.2009.12.004

* Corresponding author. Tel.: +86 13522278866; faE-mail address: [email protected] (J. Li).

a b s t r a c t

A thorough experimental investigation was carried out on a copper–water compact loop heat pipe (LHP)with a unique flat, square evaporator with dimension of 30 mm (L)�30 mm (W)�15 mm (H) and a con-necting tube having an inner diameter of 5 mm. Using a carefully designed experimental system, thestartup process of the LHP when subjected to different heat loads was studied and the possible mecha-nisms behind the observed phenomena were explored. Two main modes, boiling trigger startup andevaporation trigger startup, were proposed to explain the varying startup behavior for different heatloads. In addition, an expression was developed to describe the radius of the receding meniscus insidethe wick, to balance the increased pressure drop along the LHP with increasing heat loads. Finally, insightinto how the compact LHP can transfer heat loads of more than 600 W (with a heat flux in excess of100 W/cm2) with no occurrence of evaporator dry-out was provided.

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1. Introduction

The thermal load resulting from high power electronics insupercomputers, avionic/military equipment, telecommunicationunits, and photoelectric devices can exceed 100 W/cm2. There existdifferent cooling methods ready for the thermal management ofthose high power electronic chips [1,2]. An optimum choice de-pends on a number of factors, such as thermal performance, reli-ability index, acoustic issues, cost of manufacturing, andpotential for miniaturization [3].

As highly efficient heat transfer devices, loop heat pipes (LHP)present several unique cooling opportunities and hold significantpromise for cooling electronics [4]. Some of the advantages of theimplementation of loop heat pipes in electronic cooling applica-tions are summarized by Maydanik et al. [4] as: (i) much highercapacity at comparable dimensions; (ii) efficient operation at anyorientation in the gravitational field; (iii) lower thermal resistance;(iv) flexibility in packaging; (v) high heat transfer loads over con-siderable distances, etc. For a loop heat pipe with a connectingpipeline of 6–8 mm in diameter, the thermal resistance of LHPsdoes not exceed 0.05 �C/W [4]. Due to the above mentioned meritsfor LHPs in electronic cooling, they have achieved extensive inter-ests worldwide from academics and industry to pursuit miniatur-ized high performance LHPs. As indicated by Maydanik [4], the

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increased heat dissipation requirements in different electroniccooling applications, coupled with the miniaturization of LHPsmake it possible to implement these devices as a promising mech-anism in the thermal management of high power electronics.

The operating mechanisms and performance characteristics ofloop heat pipes have been described in detail in a number of re-views or textbooks [5–8]. Fundamentally, the application of loopheat pipes can be roughly divided into five categories, based onthe size of the connecting pipelines between the evaporator andcondenser. Due to the importance of these connections on the per-formance and implementation of LHPs, they can be categorized as:(i) microscale LHPs, the hydraulic diameters dh,lines of connectingpipelines, dh,lines 6 1 mm; (ii) miniature LHPs, 1 mm < dh,lines 6

3 mm; (iii) compact LHPs, 3 mm < dh,lines 6 5 mm; (iv) mediumLHPs, 5 mm < dh,lines 6 10 mm; and (v) large/very large LHPs,dh,lines > 10 mm.

Several early investigations of loop heat pipe miniaturizationfocused on the cooling of high power electronic devices with cylin-drical evaporators and saddle interfaces [9–11]. For these minia-ture loop heat pipe (mLHP) prototypes, the very small diametersof the evaporator and the connecting pipelines limit the capacityto the range of 25–30 W and thermal resistances in the range of0.3–1.2 �C/W. For applications that utilize air cooling at the con-denser, the total thermal resistance can be as high as 1.7–4 �C/Wdepending on the air flow rate [9–11]. Riehl and Siqueira [12]reported a thermal resistance of 0.09 �C/W at 20 W thermal loadsfor stainless steel, acetone LHP with a 19 mm outer diameter

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Nomenclature

A cross-section area, m2

dh hydraulic diameter, mhfg latent heat, J/kgI electric current, AK permeability, m2

P pressure, PaQ heat load, Wrm mean radius of pore in the wick, mT temperature, �C/KV electric voltage, Volt

Greek symbolse porosityk thermal conductivity, W/m K

h contact angler surface tension, N/mq density, kg/m3

Subscriptsb bubblec capillary/critical bubble sizeg gravityl liquidp powders saturation statev vaporw wick

742 J. Li et al. / Applied Thermal Engineering 30 (2010) 741–752

(OD) � 67 mm (length) cylindrical evaporator with liquid/vaporlines of 2.85 mm in inner diameter (ID) and 1000 mm in length.The maximum capacity for their LHP was approximately 80 W.Vasiliev et al. [13] presented a compact loop heat pipe with an ad-vanced manufacturing method for evaporator wick preparation.The reported LHP has a cylindrical evaporator of 334 mm(L) � 18 mm (OD) and can maintain a 900 W heat load at a temper-ature level below 100 �C.

In addition to cylindrical evaporators with a saddle, mentionedabove, several different evaporator configurations have beendeveloped, some of which can be configured to contact directlyto the cooled chip using a clipping fixture. Boo and Chung [14] ob-tained a thermal resistance of 0.65 �C/W in a horizontal position fora maximum heat load of 87 W and a flat evaporator with an activearea of 40 � 50 mm2 and a thickness of 30 mm. Delil et al. [15] re-ported a maximum heat load of 120 W with a thermal resistancefrom 0.62 to 1.32 �C/W at different orientations for a loop heat pipewith a flat disk type evaporator, a diameter of 44 mm and a thick-ness of 22 mm. Kiseev et al. [16] observed a maximum heat load of160 W while the thermal resistance was within the range of 0.5–0.8 �C/W for an mLHP with a flat disk type evaporator, 31.8 mmin diameter and 15 mm thick. Recently, Singh et al. [3,17] reporteda LHP thermal resistance of 0.17 �C/W at a maximum heat load of70 W for a mLHP with a flat disk type evaporator having a diameterof 30 mm and a characteristic thickness of 10 mm. Zimbeck et al.[18] developed an air cooled compact LHP with a flat disk typeevaporator using a new ceramic wick material for cooling com-puter servers. The total thermal resistance from chip surface toambient air was approximately 0.46 �C/W at a heat load of100 W. Most recently, Li et al. [19,20] reported a copper–watercompact LHP with a unique flat type square evaporator of 30 mm(L)�30 mm (W)�15 mm (H), which has an extremely low totalthermal resistance of 0.152 �C/W from the heating source surfaceto the ambient air at a heat load of 628 W.

Given the number and significance of the experimental devel-opments in the field of electronics cooling with loop heat pipes,it appears that LHP’s will soon be ready for commercial implemen-tation. However, before large scale installations of LHPs in com-mercial electronic equipment are implemented, a number ofreliability issues should be considered and evaluated. Most impor-tantly, but not limited to, are long term life tests, power cyclingtests, thermal shock tests and others. Among these required tests,the thermal shock test (transient startup test at different heatloads) and the power cycling test are two of the most important.There have been a number of investigations of the operationalcharacteristics based on these two types of tests. Furthermore,for heat pipe based products, there exists some additional issues

that should be considered and investigated in order to evaluatethe performance characteristics and operational limits, and iden-tify areas for improvement, like tilt effect in gravitational environ-ment, non-condensable gas, charging mass, etc. There were somesystematic analyses from experimental or theoretical viewpointson these effects on the startup process or operational characteris-tics during different power cycles [3,6,8,18,21,22]. The commonphenomena behind these investigations are summarized brieflyhere: (1) necessary minimum heat load for LHP startup; (2) over-shooting of temperature before LHP startup; (3) temperature peri-odic fluctuations during steady-state operations at comparativelylow heat loads; (4) temperature oscillations during startup andsteady-state operation; and (5) temperature hysteresis for differ-ent power cycle tests. There are a number of suggestions madeabout the mechanisms behind these observed phenomena. How-ever, concrete evidence is rare and detailed physical explanationsthat address the measured startup temperature and the tempera-ture overshoot are almost completely absent. In addition, there isa lack of a well-accepted physical model that describes the basicmechanisms behind the transient and steady-state operations ofLHPs. The present available models are almost all limited to stea-dy-state operation, and are based on the analysis of pressure dropsalong the loop, balanced with the capillary force, as well as thethermal balance in the loop [23–27]. There do not appear to beany transient models based on the imbalance of inner pressuresalong the loop to predict the temperature oscillations that occurduring transient LHP operation.

Given the information presented above and reports on newlydeveloped high performance compact LHP’s as reported in Refs.[19,20], systematic experiments were conducted using visualiza-tion techniques. The goal of this investigation was to reveal the de-tailed characteristics of the LHP during startup and normaloperation, using very fine time intervals to provide benchmarkdata for future theoretical and numerical analyses. From the mea-surements and observed phenomena for startup at different tran-sient heat loads and for the operation during different kinds ofpower cycles, some simple theoretical models are developed to ex-plain the possible startup temperature for LHP operation. Further-more, a physical explanation is proposed to provide insight into themechanisms that control the maximum thermal capacity of a loopheat pipe as reported in this work.

2. Development of LHP

The operating principle and the typical structure of a loop heatpipe are described in detail in Refs. [5,7]. A schematic of the LHP

(a) Inner structure of the LHP evaporator

(b) Photo of sintered wick structure inside the evaporator

J. Li et al. / Applied Thermal Engineering 30 (2010) 741–752 743

investigated here is shown in Fig. 1. This LHP has a flat squareevaporator with a dimension of 30 mm (L)�30 mm (W)�15 mm(H), a vapor line (ID = 5 mm and length = 120 mm), a cross-flowcondenser and a liquid line(ID = 5 mm and length = 120 mm). Forhigh power applications, an axial DC fan will be mounted ontothe condenser. The volume of fins encompassing the serpentinepipeline in the condenser is 120 � 120 � 20 mm3, which can easilybe incorporated with a standard 120 mm axial DC fan. The workingfluid will evaporate in the wick, and then flow automatically to thecondenser due to the pressure difference between the evaporatorand the condenser. Later the condensed working fluid is drivenby capillary force provided by the wick inside the evaporator andflows back to the compensation chamber to continually replenishit and allow continuous operation. The detailed information aboutthe design issues and manufacturing issues can be found in ourprevious work [19,20]. It should be noted that the LHP evaporatorwick structure tested here has been modified slightly from the onereported in [19,20] on the basis of heat transfer analysis for perfor-mance improvement with the theoretical model proposed by Cher-nnysheva et al. [26]. The charging volume in this work was 100%based on the calculation of the volume of the porosity of the wick,the volume of the compensation chamber, and the volume of theliquid line. The tested LHP is vertically configured, which workswith gravity. This implies that the compensation chamber wouldbe mostly filled and no vapor–liquid interface exists in the com-pensation chamber during normal operation. It is therefore ex-pected that this loop heat pipe will have an operating mode ofnearly constant conductance.

Figs. 2a and b illustrate the inner structure of the evaporator.The wick structure is sintered directly on the substrate to providethe capillary force. Copper powder with a mean pore radius of65 lm was directly sintered on the evaporator substrate under1030 �C with a shielding gas mixture of nitrogen and hydrogengases, of which one SEM inspection is shown in Fig. 2c.

The working fluid in this research is de-ionized distilled waterwhich is the most environmental-friendly and compatible withcopper wick.

Fig. 1. Schematic of the tested loop heat pipe in this work.

(c) SEM photo of wick porous surface

Fig. 2. Configuration of the LHP evaporator: (a) inner structure of the LHPevaporator; (b) photo of sintered wick structure inside the evaporator; and (c)SEM photo of wick porous surface.

3. Experimental system

The test section and the locations of the thermocouples areshown in Fig. 3. Six microscale T-type Omega thermocouples werefirmly attached onto the loop heat pipe by a film adhesive to mea-sure the following temperatures: the temperature at the bottomsurface of the evaporator (103), the temperature at the exit ofthe evaporator (104), the temperature at the entrance of the con-denser (105), the temperature at the exit of the condenser (106),the temperature at the entrance of the evaporator (107), and thetemperature at the top surface of the compensation chamber(108), respectively. It should be noted that a double-check proce-dure has been performed by detecting the electro-conductance be-tween the thermocouples and the loop heat pipe to guarantee theeffective attachment and avoid any possible failure or error in themeasurements. The tested loop heat pipe sample was fastenedonto a heating copper block with a square surface (active heatingarea) of 25 mm � 25 mm by a standard CPU clip. Ten commercialcylindrical cartridge heaters were soldered in the copper block.Two microscale T-type Omega thermocouples (101# and 102#)were inserted in the copper block and soldered, to monitor theheating power. Additionally, the top one (102) was used to

Fig. 3. Schematic of the test section and the locations of the thermocouples.


represent the case temperature as being conducted in standardthermal tests for electronic components. There was a very thinlayer of thermal grease used to fill the possible air gap betweenthe copper heating block and the loop heat pipe. The thermal con-ductivity of the grease used in the test is 3.86 W/mK and the thick-ness of the grease is typically less than 0.02 mm with a standardclip. There was another thermocouple (109) for measuring theambient temperature.

The experimental system is shown in Fig. 4. This setup includesan infrared camera, a data acquisition unit, a DC power supply and

Fig. 4. Experimenta

a tachometer, and the test section. A DC power supplier was usedto provide heating power to the cartridge heaters. The data acqui-sition unit includes an Agilent data acquisition unit and a computerto monitor and record the temperatures from the thermal couples.The infrared camera was used for the monitoring of the tempera-ture distribution in the loop heat pipe under different heatingloads. In the present tests, forced air cooling was adopted by a typ-ical 120 mm DC axial flow fan to release the heat from the con-denser to the environment. This approach of heat dissipation iswidely used in electronic cooling. The tachometer was used to re-cord the rotation speed of the fan. Combining the wind tunnel, theair mass flow rate (or air flow velocity) can be accurately measuredwith the recorded fan speed. All of the measured temperatureevery 1 s from the thermocouples (9 totally) located in the test sec-tion were sent to the data acquisition system and stored in thecomputer. Before the formal tests, several different kinds ofburn-in tests were run in order to achieve stable and repeatableoperation in the LHP.

The infrared camera has an inaccuracy of ±2.0 �C. The standardmeasurement error of the thermocouples is approximately ±1.0 �Cfrom the datasheet of the products (or ±0.75%). All thermocoupleswere calibrated based on the output when placed in a temperaturechamber to ensure the measurements from the data acquisitionsystem for all nine thermocouples were within ±0.1 �C. The princi-pal uncertainty in the experiments is the heat loss to the environ-ment from the copper block even it was carefully covered withthermal insulation. The loss from the copper heating block wasdetermined to be less than 10% by comparing the calculations ofthe heat load Q subjected to the loop heat pipe from Q = V � I andfrom Q ¼ kcuA T101�T102

d . This heat loss will be further discussed andpresented in the next section.

4. Results and discussion

In the present tests, the fan speed was constant for all cases.With the aid of the wind tunnel, the mean air velocity flowingthrough the fins in the condenser was determined to be 1.55 m/scorresponding to an air volume flow rate of approximately2.23 � 10�2 m3/s (or 47.21 cubic feet per minute). The ambienttemperature was 23.7 ± 2.1 �C. The sampling rate set for the dataacquisition system was one set, i.e., nine temperature data pointsper second for the startup tests, or one set, i.e., nine temperaturedata points every 3 s for the power cycle tests.

l system setup.

Fig. 5. LHP startup process with a 10 W heating power.


4.1. Startup tests

Fig. 5 shows the startup process with a 10 W input power. After75 min, the loop heat pipe was approaching steady-state condi-tions. It can be deduced that at this low heat load the LHP didnot start up from the very small temperature difference betweenthe condenser inlet (105#) and the condenser outlet (106#). Thismeasurement indicates that there is no fluid circulation insidethe LHP. Most of the input heat energy at the evaporator bottomsurface is dissipated by the natural convection from the exposedexternal surface of the evaporator to the environment. Thishypothesis is also confirmed by the negligible temperature differ-ence between the evaporator bottom surface (103#) and the com-pensation chamber top surface (108#), indicating no two-phasecirculation. The small temperature difference between the evapo-rator outlet (104#) and the evaporator inlet (107#) is attributedto the fact that: (i) the fluid at evaporator outlet is directly in con-tact with the hottest evaporator bottom surface (103#), however,the fluid at evaporator inlet (107#) is connecting to the compensa-tion chamber; (ii) even this is no massive working fluid circulationin the loop, occasionally there will be little amount of vapor bubblereleased from the evaporator and a local minute circulation maytake place in the distance range between the evaporator outletand the condenser inlet (in the vapor line) if taking into accountof the gravitational effect. Further visualization work is recom-mended to study this kind of local circulation in LHPs at verylow heat load.

Fig. 6. LHP startup process for a 30 W heat load.

Fig. 6 illustrates the transient temperature oscillations duringthe startup process at a heat load of 30 W. From Fig. 6, it is appar-ent that when the evaporator temperature was higher than 52 �C, asudden temperature drop occurred at the evaporator bottom sur-face (103#) with a magnitude of approximately 10�. In addition,the temperature curves exhibited chaotic oscillations for all mea-surements. One possible reason is that irregular massive vaporbubble release causes rapidly changing temperature variations.This phenomenon will be discussed in more detail later.

Fig. 7 presents the transient temperature oscillations during thestartup process at a heat load of 50 W. Similar to the 30 W casestartup process shown in Fig. 6, from Fig. 7, when the evaporatortemperature is higher than 48 �C, a sudden temperature drop oc-curred at the evaporator bottom surface (103#) with a magnitudeof approximately 5�. Again the temperature curves exhibited irreg-ular oscillations for all measurements for a short time (approxi-mately 8 min). Later the evaporator temperature increased in astepwise fashion to a level of more than 62 �C, at which anothersudden temperature drop occurred with a magnitude of approxi-mately 10�. Following this rapid drop, the temperature oscillationsfor all measurement appeared to occur in a regular periodic fashion– this type of periodic fluctuation has also been reported elsewhere[3,6,12,13,15]). The frequency of the fluctuations for all tempera-tures is approximately the same at different locations, but theamplitude is different. At the evaporator bottom surface, the mea-sured fluctuation amplitude is ±7.5 �C with a period of 100 s. Whenthe LHP has established this kind of quasi-steady-state with peri-odic fluctuations, basically the temperature variations at the con-denser inlet and outlet exhibit a trend that is opposite to thetemperature variations at the evaporator bottom surface, the evap-orator outlet and the compensation inlet.

Fig. 8 shows the transient temperature evaluations during thestartup process at a heat load of 100 W. From Fig. 8, it is apparentthat during the startup process of the LHP, at approximately 43 �C,the LHP functioned for a short time with the style of the chaoticoscillations, this temporary LHP operation results from intermit-tent vapor bubble release. Since boiling will in all likelihood notmaintain a reliable and sustainable heat transfer, the LHP temper-ature continues to increase. When the LHP evaporator temperaturewas higher than 65 �C as plotted in Fig. 8, similar to the one as ob-served for 50 W, a sudden temperature drop occurred again with amagnitude of approximately 15�. After that, the temperature for allmeasurement presented a quasi-steady-state and periodic fluctua-tion with a small amplitude but high frequency compared to theone observed for 50 W, e.g., in a range of ±2 �C at the evaporatorbottom surface for 100 W with a period of 25 s.

Fig. 7. LHP startup process at a heat load of 50 W.



Fig. 9 illustrates the transient temperature evaluations duringthe startup process at a heat load of 150 W. Most phenomena for150 W are similar to those for 100 W, with the exception that dur-ing the established quasi-state state, the periodic characteristicsseem negligible. The range of temperature fluctuations at the evap-orator bottom surface was approximately ±1.0 �C with a very smallbut indistinguishable period.

Fig. 10 presents the temperature evolutions at different placesfor loop heat pipe startup test under a heat load of 300 W. It wasclearly shown that the very stable operation mode has beenachieved at this comparatively high heat load with negligible fluc-tuations in the temperature curves. The evaporator bottom surfacehas a temperature of 63 �C at steady-state. It should be noted thatsimilarly, the LHP first became operational at approximately 50 �C.

The temperature oscillations which occur at comparatively lowheat loads were recently reported and discussed by Singh et al.[3,17] and Li et al. [19,20]. Singh et al. [3,17] argued that thestart-up phenomenon involves satisfying two main conditions thatinclude: (1) clearing of liquid from the evaporator, vapor line andpart of the condenser and (2) setting up required pressure a differ-ence across the wick that in order to circulate the working fluidaround the loop. Li et al. [19,20] speculated that the possible rea-sons for this oscillation are believed to include:

(1) the draining effect of the porous wick presented in the con-denser at low heat loads. As the vapor condensed and theliquid drained back to the compensation chamber throughthe wick structure, flow oscillations occurred and



(2) at low heat loads, the vaporization process and the floodingin the porous wick took place alternatively, because the lowheating power could not maintain a stable meniscus in thewick. This phenomenon is believed to be exacerbated if thedegree of sub-cooling of the returning liquid was increased.

Combining the above information, it is hypothesized that theoscillations with different characteristics arise mainly from the fol-lowing three effects: (1) original liquid blocking along the vaporflow passage; (2) alternative turn-out of menisci and flooding inthe wick; and (3) two-phase flow instability in the condenser.The above mentioned three factors are believed to be inherentfor any kind of loop heat pipe, no matter what kind of manufactur-ing/operation process is utilized, since it was very hard (maybeimpossible) to establish a perfect distribution of working fluid inthe loop before/during the LHP operation.

After careful examination of the recorded data shown in Figs. 6and 7, and referring to our latest high-speed visualization work[28] as depicted in Fig. 11, it is believed that there exists two mainmodes associated with the LHP startup behavior that address howvapor is generated inside the evaporator, to help displace liquidfrom the evaporator, vapor line and a portion of the condenser.These are: (i) boiling trigger startup at comparatively low temper-atures and (ii) evaporation trigger startup at comparatively hightemperatures.

Summarizing the plotted experimental data shown in (Figs. 5–10) and based on the previous visualization work in this field[28] and the proposed mechanisms behind the LHP startup processas described above, the reasons for the temperature variations and

Fig. 11. Heat transfer regimes of boiling/evaporation on thin sintered mesh screenwith demonstration of high-speed visualization images [28].

(a) Temperature oscillations at the evaporator and

(b) Temperature oscillations at the condenser inlet and

(c) Temperature difference between the condenser inlet and

compensation chamber

condenser outlet

outlet as given in (b)

Fig. 12. Temperature oscillations during startup process at 50 W heat load aspresented in Fig. 7: (a) temperature oscillations at the evaporator and compensa-tion chamber; (b) temperature oscillations at the condenser inlet and condenseroutlet; and (c) temperature difference between the condenser inlet and outlet asgiven in (b).


why they exhibit different characteristics can be determined. Atvery low heat loads, e.g., 30 W and referring to Fig. 6, the temper-ature rise rate was low and at a certain evaporator bottom surfacetemperature (around 45–50 �C for the present LHP), startup canbegin in an unstable mode. One possible reason for the instabilityis that discontinuous bubble generation and collapse combinedwith the instability of the condensation in the condenser will resultin irregular oscillations in the LHP temperature. With increases inthe heat load, e.g., 50 W and referring to Fig. 7, the boiling triggeredthe LHP (around 45–50 �C). Then if the boiling process cannotmaintain the appropriate heat balance, the LHP temperature goesup. With the liquid–vapor interface retreating into the wick (theevaporated mass cannot be circulated back completely to the com-pensation chamber at the boiling stage due to lack of sufficientdriving force), the menisci will form in the wick and the capillaryforce will drive the liquid to sustain the liquid–vapor interfaceand evaporation takes the role to govern the heat transfer. How-ever, when the startup mode switches from the boiling triggermode to the evaporation trigger mode, temperature oscillationsstill exist. One possible reason can be that since the heat load atthis stage is still comparatively low, immediately after the menisciare established in the wick and circulation is generated inside theloop, the very large evaporation heat transfer will result in a sud-den cooling of the evaporator, the compensation chamber andthe entire loop heat pipe, the vapor temperature at the exit ofthe vapor removal channel inside the wick will decrease corre-spondingly. This implies that the vapor pressure at the liquid–va-por interface will turn to a lower value, which will cause thecollapse of the meniscus and result in temporary re-flooding ofthe wick. Repeating the process of meniscus formation and re-flooding of the wick is the mechanism behind the temperatureoscillations during the evaporation trigger startup mode.

All of the above discussions can be verified indirectly throughthe examination of the temperature measurements as shown inFig. 12a–c. It is well known that from the temperature differencebetween the evaporator and the compensation chamber and thetemperature difference between the condenser inlet and the con-denser outlet, one can tell whether the LHP is operating normallyor not. Fig. 12a shows the temperature variations at the evaporatorbottom surface and the top surface of the compensation chamber,respectively. Fig. 12b shows the temperature variations at the con-denser inlet and the condenser outlet, respectively. Fig. 12c givesthe temperature difference between the condenser inlet and thecondenser outlet as calculated from Fig. 12b. It is obvious that fromthe different LHP startup trigger mode, the variations of the LHPtemperature at difference locations will exhibit different charac-teristics: compared to the evaporation trigger mode, the tempera-ture is more chaotic for the boiling trigger mode.

In order to further verify the proposed mechanisms for startup,a simple visualization experimental system was established, with acapillary wick covering an artificial compensation chamber whichmaintains a constant system pressure as the same as the atmo-spheric pressure. As illustrated in Fig. 13a and b, the proposedmechanism was fully demonstrated: (1) at the beginning of a100 W heating process (steady-state was not established), severalvapor bubbles were generated and released from the vapor re-moval channels into the environment, e.g., Fig. 13a and (2) Afterthe bubble generation process, in the wick all the vapor removalchannels are clear of bubbles and the heat transfer was conductedonly through evaporation in the wick. e.g., Fig. 13b. Fig. 13c illus-trates infrared photos during evaporation corresponding to thestate as pictured in Fig. 13b.

When the heating power was incremented beyond a point(150 W for the present LHP), steady-state of LHP is quickly estab-lished and the evaporation process becomes very stable. Further,the increased pressure difference between the vapor pressure in

the vapor removal channels and the vapor pressure in the con-denser when coupled with the increase in the thermal load willaid in the flow stability in the liquid and vapor pipelines.

(a) Vapor bubbles generated and released from the vapor removal channels

(b) Clearance of vapor bubbles (c) infrared photos during evaporation corresponding to (b)

Fig. 13. Visualization study on the transient phase change phenomena during a heating process at 100 W.


For the bubble nucleation temperature in a fully flooded wick,the reader is referred to Fig. 14a. Hsu and Graham [29] proposedthat the vapor bubble in a cavity will begin to grow if Tl(yb) P Tb,where Tl(yb) is the liquid temperature at the bubble top surfaceand Tb is the vapor temperature inside the bubble, which can becalculated from the following equation:

Tb � Ts ¼2c1rTs

c2hfgqvybð1Þ

with

c1 ¼ 1þ cos h; c2 ¼1

sin h; yb ¼ c1rc; rc ¼ c2rm

where h is the contact angle for liquid on a solid surface, yb the pro-jection height of vapor bubbles over the solid surface and rm is thegiven surface equivalent radius of the cavity. From Eq. (1), it is quitestraightforward to calculate the level of superheat necessary todrive a bubble embryo entrapped in the wick structure, however,it is very hard to define the saturation temperature inside the LHPbefore it commences operation with the boiling trigger mode, sincethe two-phase thermodynamic status inside the loop is very non-uniform. One way to begin is to assume that the saturation statusin the compensation chamber is identical to the original status inthe wick, by substituting measured Tcc into Eq. (1) to replace Ts.From a simple derivation, it is clear that Tb � 1.0028 � Tcc. The aboveestimation agrees well with the measured data as shown in (Figs. 6and 7).

To determine the startup temperature for the evaporation con-trol process, we must first evaluate the possible pressure losses forthe working fluid along the loop under a certain heat load and thenusing the capillary pressure obtained from the force balance,

DPc ¼ DPw þ DPv þ DPl þ DPg ð2Þ

Then, the required meniscus radius inside the porous wick can beobtained from the Young–Laplace equation

DPc ¼2rrc

ð3Þ

here, rc is the radius of the meniscus. Referring to Fig. 14b, as pro-posed by Li et al. [30], for a transient heating process, the local boil-ing/evaporation process on a uniformly sintered wick surface hasfive typical configurations: (a) initial state; (b) convection; (c)nucleate boiling; (d) thin liquid film evaporation; and (e) dry-out.For the evaporation control process, the meniscus radius in the wickwill decrease in order to provide the capillary force to pump theneeded water into the heating area.

Referring to Fig. 15a, as the liquid–vapor interface recedes intothe wick the capillary pressure increases and the pressure balanceis maintained. At a certain position of the meniscus in the pore ofthe wick, the capillary pressure is equal to the value obtained fromEq. (3), and the necessary circulation of the working fluid can beachieved.

Through an analysis of the vapor pressure at the pore, we canestimate the vapor temperature through the Clausius–Claperonequation. This analysis has been reported previously and is wellunderstood [6,23,26]. However, nearly all of the existing analyticalmodels obtain the operating temperature of the LHPs at steady-state. If we want to implement those models in the startup temper-ature analysis, we must evaluate the temperature inside the com-pensation chamber before the circulation starts since thetemperature in the compensation chamber plays a critical role inthe evaporator temperature during both normal operation and

(a) Model of the meniscus radius at different location

(b) Calculated results from Eqs.(4-5)

along the wick powder/wire

Fig. 15. Meniscus radius with the variations of the meniscus position in the wick.

(a) Bubble nucleation model proposed by Hsu [29]

(b) Local evaporation process and meniscus receding

from mesh wire surface [30]

Fig. 14. Two mechanisms governing the startup process of LHPs at different heatloads.


the startup process. This is especially true for startup, where thecompensation chamber has a higher temperature than during nor-mal operation, negatively influencing the startup process. For thisreason, it is necessary to conduct the analysis of the temperaturein the evaporator using a detailed unsteady-state 3D numericalsimulation for a specific LHP and a certain heat load.

Because of the complexity, the transient meniscus and the pro-cess whereby it recedes into the pore of the wick is very hard to de-scribe mathematically. Li et al. [28] has evaluated this process usinga high-speed visualization technique to study the evaporation heattransfer and meniscus development on a thin layer of sinteredscreen mesh. Wang and Peterson [31] provided a simple mathemat-ical model to calculate the curvature of the meniscus and the thinfilm thickness around the mesh wires, and also analyzed the tem-perature distribution around the mesh wire through a simple 2Dsteady-state model. Nevertheless, during startup, (see Fig. 15b),the meniscus continues to recede along the wick powder or wireto establish the curvature needed to provide the necessary capillaryforce to circulate the working fluid. A simple mathematical model isproposed here to describe the meniscus radius with the variationsof the meniscus position along the powder or wire. It should benoted here that the meniscus position (here an angle a is used torepresent the position to simplify the problem) is an independentvariable, but actually the meniscus position (x, y) is closely depen-dent on the transient temperature distribution and the non-equilib-rium heat transfer process. This is a relatively simple method forestimating the meniscus radius for the capillary force calculations.For a certain meniscus position a, refer to Fig. 15b,

if h � a � p=2;rc ¼ ðrp þ rm � rp sinaÞ= cosðhþ p=2� aÞ ð4Þif p=2 � a � p;

rc ¼ ½rp þ rm � rp cosða� p=2Þ�= cosða� h� p=2Þ ð5Þ

The calculated results are plotted in Fig. 15c, here a recedingcontact angle of h = 33� is selected for a water–copper combination[32,33]. As shown, when the meniscus recedes to the position ofa = 100�, the radius has a minimum value which corresponds tothe maximum capillary force, not at a = 90� as traditionallyassumed.

4.2. Power cycle tests

Tests using different heat load cycles were carried out to vali-date the operational reliability and transient response to changesin the heat load. There were two kinds of power cycles selected:power up ? power down and power down ? power up.

Fig. 16 shows the test results and operating characteristics ofthe LHP for increasing step increments from 10 W to 600 W inincrements of 40 or 50 W. From Fig. 16, it is apparent that the tem-perature oscillations fade when the heat load is increased beyond150 W for the compact loop heat pipe evaluated here. The recordedmaximum temperature oscillation at the evaporator (thermocou-ple 103) was approximately ±7.5 �C at low heat loads (50 W),which is similar to what was observed in the startup tests.

Fig. 17 shows the test results and operating characteristics ofthe LHP for decreasing step increments from 600 W to 10 W inincrements of 40 or 50 W. From Fig. 17, it is apparent that the

Fig. 16. Operation characteristics of the LHP with increase of heat loads from 10 Wto 600 W.

Fig. 17. Operation characteristics of the LHP with decrease of heat loads from600 W to 10 W.


temperature oscillations occur when the heat load is decreased to150 W. The recorded maximum temperature oscillation at theevaporator was the same as in Fig. 16. Comparing Figs. 16 and17, a distinct operation difference occurs at a heat load of 10 W,where the LHP does not operate for loads of 10 W or less regardlessof a power up or power down process.

Fig. 18 presents the test results for another power cycle: powerdown ? power up from 300 W to 10 W and from 10 W to 300 Wwith a step of 90 or 100 W. Similarly, the temperature oscillationsbecome significant when the heat load is less than 150 W. From

Fig. 18. Test results for power cycle from 300 W to 10 W and from 10 W to 300 W.

Fig. 18, it should be noted that at 100 W one test indicated a tem-perature pulse during the power decreasing process, which impliesa sudden flooding of the wick following meniscus re-forming as ex-plained previously. The other operational characteristics are simi-lar to those observed during the power up ? power down cycle.

As discussed previously, the principal uncertainty in the exper-iments is the heat loss from the copper heating block to the envi-ronment, which can be determined by examining the heatconduction in the copper block. This value was typically less than10% of the calculated heat load, Q, subjected to the loop heat pipebetween Q = V � I and Q ¼ kcuA T101�T102

d . Fig. 19 presents the compar-ison of the heat loads between the two different methods. It can befound that at high heat loads, e.g., 600 W calculated from Q = V � I,the actual heat load through the copper block is 540 W. This im-plies that there will be a 10% heat loss from this comparison. Atthe comparatively low heat load, since the measured temperaturefrom thermocouple 102# is representing the case temperaturewhich is influenced by the loop heat pipe transient operating char-acteristics and shows some fluctuations, the calculated values fromQ ¼ kcuA T101�T102

d are not constant, but the averaged values are inrelatively good agreement with the data obtained from the rela-tionship Q = V � I.

Fig. 20 gives the heat load dependence of the evaporator aver-aged temperature. If we consider the heat loss as discussed above,it is apparent that the present loop heat pipe has an operationmode of constant conductance after stable operation has beenestablished. This implies that during normal operation, the presentconfigured LHP has a fully charged compensation chamber. How-ever, due to lack of sufficient liquid inventory at low heat loads,caused by instable boiling or evaporation process, the LHP deviatesfrom the constant conductance mode. Similar behavior has beenreported elsewhere [3].

Through careful review of the experimental results presentedhere, several possible methods are proposed for reducing the insta-bilities inherent during LHP startup and those resulting from boil-ing instabilities. These can be summarized but are not limited to:(i) elevating the system operation pressure to depress the instabil-ity, e.g., flow restriction or smaller connecting pipelines. The conse-quence of this approach is that the operating temperature will beincreased, and sometimes smaller tubes may experience deprimingof the wick by the formation of a slug in the vapor line just prior tothe condenser entrance, which should be further considered andsolved; (ii) selecting a wick with a smaller pore size to providehigher capillary forces, and at the same time increasing the flowresistance inside the wick to weaken the flooding of the wick dur-ing operation, that is, selecting a wick with a low permeability

from K ¼ r2pe3

37:5ð1�eÞ2for sintered powder wick; (iii) reducing the tem-

perature of the compensation chamber as much as possible to pro-mote the thermal performance of the LHP (reduce the temperatureovershoot during startup), especially at low heat loads, e.g., byselecting a wick material with low thermal conductivity; and (iv)modifying the wick structure and the vapor removal channels to

Fig. 19. Heat loss analysis from the heating block to the environment.

Fig. 20. Heat load dependence of the evaporator averaged temperature.

Fig. 21. Macro liquid–vapor interface evolution in the porous fins of the wick withthe increase of heat loads.


prevent the blocking of the vapor passage caused by flooding of thewick.

For the present compact loop heat pipe, even for a heat load ofup to 600 W on an active heating area of 25 mm � 25 mm (corre-sponding to a heat flux of 100 W/cm2), this LHP does not presentany sign of dry-out. Above 600 W, the solder for fixing cartridgesin the heating block was starting to melt, thus experiments at high-er heat loads were not conducted. The obvious question here ishow is it possible for this compact LHP to transport such high heatloads? From Li et al.’s experimental work [30], it was demonstratedthat, with good contact between the wick and the substrate, anevaporation CHF of 367.9 W/cm2 was obtained on an 8 cm � 8 cmmesh film with a thickness of 0.82 mm directly sintered on theheated wall (copper substrate) using water as the coolant. If thereis sufficient liquid inventory into the evaporation area, the evapo-rator would not dry-out prior to reaching this value. Furthermore,combining the previous visualization work [28], it is assumed thatthe macro liquid–vapor interface in the porous fins of the wick willevolve with the increase of heat loads as shown in Fig. 21. Thesemacro interfaces in the porous wicks will extend the effectiveevaporation surface, thus these extended interfaces will enhancethe evaporation heat transfer (please keep in mind that these inter-faces are formed by local menisci). The detailed heat transferenhancement could be further revealed by a 3D numerical analysis.

5. Concluding remarks

Temperature oscillations can occur in loop heat pipes, duringboth startup and normal operation, especially at comparatively

low heat loads. It is hypothesized here that these operational oscil-lations result principally from three factors: (1) original liquidblocking along the vapor flow passage. For this situation, the boil-ing startup mode cannot be avoided, and the lower the backgroundsaturation temperature in the compensation chamber, the lowerthe operating temperature. Thus heat leaks to the compensationchamber should be minimized or the temperature of the compen-sation chamber should be kept as low as possible; (2) alternativeturn-out of menisci and flooding in the wick. This factor can beconsidered as an inherent factor, which causes the temperatureperiodic oscillations and the operating instabilities. However, theeffect of this factor can be reduced (not eliminated) by choosingthe proper configuration of the wick i.e., proper pore size, other un-ique vapor removal channels and others; (3) two-phase flow insta-bility in the condenser and the connecting pipelines. Thisinstability should be given much attention since sometimes thisinstability may cause mal-function of the CPL/LHP, increasing theflow velocity inside the tubes or elevating the system operationpressure are two possible ways to depress this instability, e.g., add-ing flow restrictions along the loop or choosing connecting tubeswith smaller diameters. But some potential negatives should beconsidered further.

Nevertheless, this unique compact loop heat pipe can manage aheat load up to 450 W when the case temperature is below 100 �C(e.g., a CPU chip surface temperature), where with a total thermalresistance as low as 0.162 �C/W at the system level and the thermalresistance of the LHP is as low as 0.033 �C/W. Finally, the repeat-able performance validates the reliability of this LHP regardlessof the type of power cycle to which it is subjected. Further reduc-tion in the temperature oscillations during LHP normal operationsat comparatively low heat loads is being pursued from the resultsof the present research.

Acknowledgement

This work was partially supported by Asia Vital ComponentsCo., LTD. and the Presidents Fund of Graduate University of the Chi-nese Academy of Sciences (085101AM03).

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Documents

Experimental studies on a high performance compact loop heat pipe with a square flat evaporator