[American Institute of Aeronautics and Astronautics 35th AIAA Thermophysics Conference -...

(c)2001 American Institute of Aeronautics & Astronautics or Published with Permission of Author(s) and/or Author(s)' Sponsoring Organization.

AM A A A01-31449

AIAA-2001-3079Design of the Brazilian Cryogenic HeatPipesP. Couto and M. B. H. MantelliSatellite Thermal Control LaboratoryFederal University of Santa CatarinaBrazil

35th AIAA ThermophysicsConference

11-14 June 2001 /Anaheim, CAFor permission to copy or to republish, contact the American Institute of Aeronautics and Astronautics,1801 Alexander Bell Drive, Suite 500, Reston, VA, 20191-4344.


AIAA-2001-3079

Design of the Brazilian Cryogenic Heat Pipes

Paulo Couto*Marcia B. H. Mantelli"''

Satellite Thermal Control LaboratoryDepartment of Mechanical Engineering

Federal University of Santa CatarinaP. O. Box: 476 - Florianopolis SC Brazil - 88040-900

AbstractCryogenic heat pipe is a satellite thermal control

device that is used mainly for the thermal control ofoptical surfaces, infrared scanning systems, or largesuperconducting magnets in the space environment.Actually, investigations are being conducted to includeother applications of cryogenic heat pipes, such as thecooling of electronic devices, particularly in micrograv-ity environments. The most recent theoretical develop-ment on cryogenic heat pipe technology, carried out atthe Satellite Thermal Control Laboratory (NCTS), atthe Federal University of Santa Catarina (UFSC), ispresented. This research project is funded by the Brazil-ian Space Agency, as part of the Uniespa£o Program.The heat transfer limitations of these heat pipes arefocused here. Four configurations of stainless-steel/nitrogen cryogenic heat pipes are proposed forfuture experimental ground testing. An experimentalsetup is under construction at NCTS, and a descriptionof the apparatus is given at the conclusion. Descriptionsof the researches that are still under development atNCTS are also presented. This research is part of a Ph.D. thesis, currently under development at NCTS, and itwill provide a significant insight on the startup phe-nomena at cryogenic temperatures.

NomenclatureA area [m2]Cp specific heat [J/kg K]d mesh wire diameter [m]F friction parameter [s/m4]/ friction factorL length [m]Leff effective length of the heat pipe [m]g Earth's gravitational acceleration [9.81 m/s2]% latent heat [J/kg]K permeability [m"2]* Research Assistant, Department of Mechanical Engineering, Stu-

dent Member AIAA, [email protected]* Professor, Department of Mathematics, Member AIAA,

[email protected] @ 2001 by P. Couto and M. B. H. Mantelli. Published by

the American Institute for Aeronautics and Astronautics, Inc., withpermission.

k thermal conductivity [W/mK]TV mesh numberP pressure [kPa]Q heat transfer limitation [W]Re Reynolds numberRg gas constant [J/kg K]Rhyd hydraulic radius [m]RI wick structure outer radius (or liquid layer

outer radius) [m]Rv vapor region radius [m]rc capillary radius [m]T temperature [K]V volume [mj]W mesh wire spacing [m]

Acronyms:AEB Brazilian Space AgencyNCTS Satellite Thermal Control LaboratoryPNAE Brazilian Police for Space ActivitiesUFSC Federal University of Santa Catarina

Greek Symbols:a thermal diffusivity [m2/s]8 thickness [m]<f> heat pipe tilt angle [degrees]<p porosity// dynamic viscosity [N.s/m2]p density [kg/m3]a surface tension [N/m]

Subscripts:a adiabatic sectionc condenser sectione evaporator section£ liquid layers solid layerv vapor layerw wick structureboi boiling limitationcap capillary limitationent entrainment limitationson sonic limitationvis viscous limitation

1American Institute of Aeronautics and Astronautics


IntroductionIn 1996, the Satellite Thermal Control Laboratory

(NCTS/UFSC) began the development of a PassiveCryogenic Radiator1, in the frame of the UniespacoProgram, funded by the Brazilian Space Agency (AEB).Passive Cryogenic Radiators are used to cool downequipments, such as infrared sensors and CCD cameras,to the cryogenic temperature levels required for theiroptimum operation. In many satellite designs, theequipment to be cooled down cannot be placed near thecryogenic radiator (or any other cryogenic heat sink).Usually, cryogenic heat pipes are used to transfer theheat from these equipments to the cryogenic heat sink2">A5. In addition, cryogenic heat pipes are used on thethermal control of focal plans of infrared sensors6 andX-ray telescopes7, on the cooling of superconductingmagnets8, among others.

Cryogenic heat pipes are under investigation at theNCTS/UFSC9. The main objective is to develop a pas-sive cryogenic satellite thermal control device for thepayload of the Brazilian satellites, described at thePNAE10 (Brazilian Policy for Space Activities, 1996).Cryogenic heat pipes usually operates at temperaturesbelow 200 K, which makes this device very sensitive toheat loads from the surrounding environment (satellitestructure, for example). Small heat loads applied to theheat pipe evaporator can produce large temperaturevariations11. This is because cryogenic working fluidspresent a significant change of the thermophysicalproperties for small temperature variations12. Therefore,the operational temperature range of cryogenic heatpipes is quite narrow, and usually lies between the criti-cal point and the triple point1 J. Figure 1 (adapted fromGilmore14) presents the operating temperature rangesfor some cryogenic working fluids.

CriticalPoint

(at 0.10 mm Hg) Point Point (at 1 atm)

0 40 80 120 160 200 240 380 420Temperature [K]

Figure 1 - Operating temperature ranges for cryogenicworking fluids (Gilmore ).

For the development and ground testing of cryo-genic heat pipes, the heat transfer limitations and the

transient response of the temperature must be wellknown. The heat transfer process associated with thestartup process must be consistently modeled includingthe effects of the supercritical working fluid. Also, thethermophysical properties and wick structure parame-ters must be accurately determined.

Literature ReviewThe most challenging problem in cryogenic heat

pipes is modeling the startup phenomena. Differentlyfrom low and medium temperatures heat pipes'^16,cryogenic heat pipes starts up from a supercritical sta-te . The entire heat pipe must be cooled down beforenominal operation begins. Cowell17'18, Brennan19,Ochterbeck et al.20, Rosenfeld et al.21 and Van &Ochterbeck22 discussed the start-up process of cryo-genic heat pipes.

1718Cowell ' presented a theoretical transient behav-ior model of a nitrogen/stainless steel cryogenic heatpipe with circumferential wick and composite centralslab. His three-dimensional model assumes constantproperties and does not account for the working fluiddynamics. Although provisions for simulating a super-critical start-up were included, this process was mod-eled from an isothermal heat pipe, with a temperaturevery close to the critical temperature of the workingfluid (130 K for nitrogen).

Brennan presented a microgravity experiment fortwo different aluminum/ oxygen axially grooved heatpipes. The experiment, called CRYOHP, flew aboardthe STS-53 space shuttle mission, in December 1992.Reliable start-ups in flight of the two heat pipes wereperformed, but slower than in ground tests. This showsthat ground tests for transient start-up tend to presentbelter results relative to the microgravity start-up be-havior. The maximum transport capability of the twotested heat pipes were 5 W @ 88 K and 22 W @ 95 K.

Ochterbeck et al. analyzed the depriming and re-wetting process of two high-capacity external arteryheat pipes subjected to externally induced accelerations.They used previously existing analytical expressions todevelop a combined analytical/numerical model. Thismodel was not developed for predicting the startup ofthe heat pipe, but it gives an important insight of thedepriming/rewetting process, that can occur during thestartup of heat pipes, subjected to external forces.

Rosenfeld presented a study on the supercriticalstart-up of a porous metal wick titanium/nitrogen heatpipe. The test was performed during mission STS-62 ofthe Space Shuttle (March, 1994). This heat pipe reacheda non-operational steady state thermal condition duringmicrogravity tests. Only 30 % of the heat pipe length(condenser) cooled below the nitrogen critical point. Inground tests, the titanium/nitrogen heat pipe started-upsuccessfully. The authors concluded that the thermal

American Institute of Aeronautics and Astronautics


conduction of the titanium/nitrogen heat pipe was insuf-ficient to reduce its internal pressure below the criticalpressure of the nitrogen. The successful start-up duringground tests was due to enhanced thermal transport ofthe gravity-assisted convection/liquid collection effects.

Yan and Ochterbeck presented an one-dimensional transient model. They considered that thestartup process can be summarized in two stages, asfollows:

1st Stage: In the first stage, the heat pipe is cooledby pure heat conduction. The vapor temperature andpressure at the condenser are greater than the criticaltemperature and pressure (Tconj > Tcrih or P > Pcril). Thecooling effect resulting from the condenser heat rejec-tion is not immediately propagated through the heatpipe, but it is confined to a region extending from thecondenser to some penetration depth & Beyond 8, thetemperature gradient is zero. When penetration depthequals the heat pipe length, the cooling effect of thecondenser has propagated over the entire heat pipe.Then the temperature of the entire heat pipe decreases.

2nd Stage: In the second stage, the vapor tempera-ture and pressure are lower than the critical temperatureand pressure, and the heat conducted to the advancingliquid front cools the heat pipe (Tcon(J < Tcrih or P <Pent)- When the condenser temperature is lower than thecritical temperature and the internal pressure is lowerthan the critical point, the vapor begins to condense inthe condenser section. The advancing liquid layer issubjected to a capillary driving force that is induced bysurface tension and opposed by the wall shear stress. Itadvances with an average velocity that will vary withrespect to the length of the liquid layer. If the liquidlayer length is less than the condenser length, the rewet-ting process is restricted to the condenser. The heattransfer in the remaining sections of the heat pipe is stillby pure heat conduction. When the liquid front reachesthe interface between the condenser and the adiabaticsection, there are two possible cases:

Ixl case: If the liquid average velocity in the con-denser is not enough to provide cooling to the dry re-gion, the liquid front will stagnate and will not advanceimmediately. Thus, the wall temperature in the dryregion remains independent of the rewetting.

2nd case\ With the increasing time, the liquid aver-age velocity in the condenser raises, because the liquidtemperature continues to decrease with time. This re-sults in the augment of the surface tension, intensifyingthe capillary driving force. Additionally, the heat fluxfrom the adiabatic section to the condenser decreases,and the latent heat of vaporization increases. Thus, theliquid front eventually will advance again, until the heatpipe reaches its operational steady state.

This model was compared favorably with the mi-crogravity experimental data presented by Brennan et

19al. . This model did not include the effects of the hy-drostatic height, which helps the start-up process in \-genvironment. The models described in the literature didnot account for the fluid dynamics of the working fluid.However, the author believes that these effects areimportant for the determination of the heat transfercoefficient in the condenser and in the evaporator, andfor the correct prediction of the device capillary andboiling limits.

Cryogenic Heat Pipe Design Features and HeatTransfer Limitations

The objective of this project is to design a stainless-steel/nitrogen cryogenic heat pipe for operation at 100K. The porous media selected for the heat pipes is thestainless-steel screen wick, with Meshes number 62,160, 166, and 254. Two different stainless steel tubes of0.8 meters long are used as the container: one with 3/8",and another with 1/2" of external diameter. The tubewall thickness is 1.5 mm.

On the following paragraphs, the design procedureof different heat pipe configurations, and its heat trans-fer limitations, is shown, according to the followingsteps:1. Determination of the working fluid mass, given an

operational temperature (100 K);2. Determination of the pressure profile in which the

heat pipe operates during the start up (enthalpy xpressure diagram);

3. Determination of the heat transfer limitations;4. Determination of the optimum configuration;

Determination of the Working Fluid MassThe mass of the working fluid for each combina-

tion of tube/screen wick structure can be obtained bycalculating the void volume of the screen wick and thevolume of the vapor region, and multiplying these val-ues by the density of the saturated working fluid (liquidand vapor) at the operational temperature (100 K).Tables 1 and 2 show the mass of working fluid for thedifferent combinations of container and wick structure.

The void volume of an n layer screen wick struc-ture can be obtained from the following relation:

Vt = 9n Vw (1)

where (pn is the porosity of an n layer screen wick struc-ture, obtained from Imura et al23:

(2)

where n is the number of layers of the screen wickstructure, S\ and Sn are the thickness of one and n layers



of screen wick (experimentally measured with a 0 - 25(im micrometer), and q>\ is the porosity of one layer ofthe screen wick, given by:

V. V-V...,

[(d + W)(d 2 /4)(d + W)}(3)

Table 1: Working fluid mass (in kg x 10°) for the 1/2"external diameter stainless-steel container

Mesh62160166254

1

7.524.544.483.86

10.245.605.524.56

15.127.607.475.85

19.519.479.287.07

Table 2: Working fluid mass (in kg x 10°) for the 3/8"external diameter stainless-steel container

^^^Layers

Mesh ^^\62160166254

1

4.882.752.712.26

2

6.793.513.452.76

3

10.174.934.843.69

4

13.116.256.124.56

where Vf: is the void volume of the wick structure, Vw isthe total volume of the wick structure, Fm>c, is the wirevolume, d is the wire diameter of the screen wick, andW is the mesh wire spacing.

Determination of the Pressure ProfileThe ratio between of the overall mass and the inter-

nal volume of the heat pipe gives the working fluiddensity, which remains constant during the startupprocess. Therefore, knowing the temperature range inwhich the heat pipe will operates (from startup tosteady-state) and the working fluid density, it is possi-ble to determine the pressure profile that the heat pipe issubmitted during the start up process. The highest pres-sure occurs for the highest density value at the initialstart up temperature (-300 K). This case corresponds tothe wick structure composed by 4 layers of screen Mesh64, for both tube diameters. At the initial start up tem-perature, the lowest initial pressure occurs for the low-est density value, which corresponds to the wick struc-ture composed by one layer of screen Mesh 254, forboth diameters. Figures 2 and 3 shows the enthalpy x

pressure chart of the proposed cryogenic heat pipes startup process. These figures were obtained using the ther-modynamic properties for the nitrogen presented byReynolds26.

On Figure 2, the line labeled (1) corresponds to thehighest density wick configuration, while the line la-beled (2) corresponds to the lowest density wick con-figuration. It can be seen that, at the highest densitywick configuration, the initial pressure (or, also, thestorage pressure of the heat pipe) is 23.3 MPa, whichcorresponds to almost 230 atmospheres, while at thelowest density wick configuration the pressure laysaround 4.1 MPa (40.5 atmospheres). Also, at the high-est density wick configuration, the working fluid entersthe saturation region of the enthalpy x pressure diagramnear the critical point, while at the lowest density wickconfiguration the vapor begins to condensate at 105.5K, which is close to the operational temperature (100K). The same effects occur for the 3/8" heat pipe, butwith higher magnitudes. At the highest density wickconfiguration, the working fluid enters the saturationregion at the critical point (T = 126.2 K and P = 3.4MPa).

0 50 100 150 200 250 300 350 400 450 500

0,10 50 100 150 200 250 300 350 400 450 500

Enthalpy [kJ/kg]

Figure 2 - Thermodynamic chart for the 1/2"cryogenic heat pipe start up.

500

50 100 150 200 250 300 350 400 450 50Enthalpy [kJ/kg]

Figure 3 - Thermodynamic chart for the 3/8"cryogenic heat pipe start up.



Determination of the heat transfer limitationsBefore determining the heat transfer limitations of

the heat pipe configurations, it is necessary to estimatethe length of the condenser and of the evaporator. Thevalue of 0.4 m for the evaporator length, and the valueof 0.2 m for the condenser length are used, as shown inTab. 3. The total length of the heat pipe is fixed in 0.8m.

Table 3: Geometric parameters of the proposedstainless-steel/nitrogen cryogenic heat pipes.

ParameterExternal radius, Rc

Internal radius, R?

Evaporator length, LcCondenser length, LcAdiabatic length, LaEffective length, Leff

Heat Pipe 14,7625 mm(0 3/8")

3,2625 mm(e* = 1,5 mm)

300mm300 mm200 mm500mm

Heat Pipe 26,35 mm(0 1/2")4.85 mm

(e* = 1,5 mm)300mm300mm200mm500mm

* Commercial dimension.

Faghri24 and Peterson25 present the expressionsused to determine the heat transfer limitations. Theyare:

Capillary limitation: The capacity of pumping liquidfrom the condenser to the evaporator is limited for agiven combination of the working fluid/wick structure.This is called the capillary limitation. The expressionsused to determine this limit are:

(4)

where a is the surface tension of the working fluid, r^-is the effective radius of the pores of the wick structure,PC is the density of the saturated liquid, g is the gravita-tional acceleration, Lt is the total length of the heat pipe,Dv is the vapor region diameter, </) is the tilt angle of theheat pipe, LCff is the effective length of the heat pipe,and the liquid and vapor friction factors (F^ and Fv,respectively) are given by:

(5)

F =•*• •» (/Rev)//v (6)

where //f and ju{: are the saturated liquid and vaporviscosity, Aw is the wick cross sectional area, Av is the

vapor region cross sectional area, Rv is the vapor regionradius, hj-K is the latent heat of the working fluid, and Kis the permeability. For laminar flow in a circular pipe,the factor (/"Rev) is equal to 16.

etalThe permeability in Eq. (5) is obtained from Imura

<*v (7)

where d is the wire diameter, cp is the porosity [given byEq. (2)], and the friction factor/was experimentallycorrelated with Re (/Re/ = 122).

Boiling limitation: If the heat flux or the temperature ofwall is too high, the boiling phenomena takes placeinside the wick structure, affecting severely the liquidflow and modifying the heat pipe performance. Vaporbubbles can block the liquid path in the wick structure,leading the heat pipe to a dry-out condition. The ex-pressions used to determine this limit are:

(0e/>,max

(8)

Where Lc is the evaporator length, TV is the saturatedvapor temperature, rn is the bubble nucleation radius, rwis the effective capillary radius of the screen wick, andkeff is the effective conductivity of the wick structure,given by:

(9)

where kf and kv are the saturated liquid and vapor con-ductivity, respectively.

Sonic limitation: The vapor flow for some heat pipesmay reaches sonic or supersonic values during thestartup or even during the steady state operation. Theexpression used to determine this limit is:

(10)

where p0 and TQ are the density and the temperature atthe evaporator end cap, defined at the operational con-dition in the heat pipe. Rg is the gas constant, and thecoefficient yis the ratio of specific heats.

Viscous limitation: If the viscous forces dominate thevapor flow, the vapor pressure at the condenser may



reduces to zero. Under this condition, the heat transfercapacity of the heat pipe may be limited. The expres-sion used to determine this limit is:

(11)

Where Pv is the saturated vapor pressure, and/Re = 16for laminar flow in a circular pipe.

Entrainment limitation: When the vapor flow velocity issufficiently high, the shear force at the liquid-vaporinterface can drag liquid into the vapor flow. This phe-nomenon reduces the liquid flow to the evaporator,leading to a partial or total heat pipe dry-out. The ex-pression used to determine this limit is:

(12)

where rhyd^, is the hydraulic radius of the pores of thewick structure.

During the calculations of the heat transfer limita-tions, it was found that the strongest constraints are dueto the capillary and the boiling limits, with the order ofmagnitude of 1 W. These mechanisms will be discussedhere. The orders of magnitude of the other limitationsare: 104 W for the sonic limitation, 102 W for the en-trainment limitation, and 1010 W for the viscous limita-tion. These will not be discussed.

Table 4 and 5 shows the lowest heat transfer limita-tions for the various configurations under investigation,at the temperature of 100 K. These values were ob-tained by using the nitrogen thermodynamics propertiespresented by Reynolds26. The index at the right side ofeach value indicates the corresponding heat transferlimitation mechanism.

The best wick structure configuration for the 1/2"heat pipe (Table 4), and for the geometrical parametersgiven in Tab. 3, is that composed by 3 layers of screenMesh 254, which is boiling limited at 2.32 W at 100 K.The second best configuration for the 1/2" heat pipe isthat composed by 2 layers of screen Mesh 166, which iscapillary limited at 1.90 W at 100 K. It can be observedthat the heat transfer limitation mechanism for thesetwo configurations is different (capillary and boiling).Actually, increasing the number of layers of the screenwick causes an increase on the liquid layer thickness,which reduces the normal hydrostatic pressure, improv-ing the capillary limitation. In the other hand, the in-crease of the numbers of the screen wick layers dimin-ishes the radius of the vapor region (Rv\ which reducesthe liquid/vapor interface area in the wick structure,decreasing the boiling limitation.

Table 4 - Heat transfer limitations (in Watts) for the1/2" heat pipe.

^^\ n

62160166254

1

-18.41cap

1.63cap

1.86cap

1 3 ] c a p

2

-11.91cap

1.65cap

! 90 cap

1 3 , cap

3

-8.66 cap

1.47boi

1.49 boi

2.32 boi

4

0.32 b0'1 .02 boi

, 04boi

1.63boi

cap Capillary limitation; b01 Boiling limitation.

Table 5 - Heat transfer limitations (in Watts) for the3/8" heat pipe.

^^\ n

62160166254

1

0.80 b01

1.68cap

L86cap

1.06cap

2

0.57 b01

1.66cap

1.73 boi

1.06cap

3

0.29 b01

0.95 boi

0.97 boi

1.53boi

4

0.17b°l

0.65 boi

0.67 boi

1.06boi

:ap Capillary limitation; b01 Boiling limitation.

It is important to observe that the configurationcomposed by 1, 2 or 3 layers of screen Mesh 62 for the1/2" heat pipe cannot supply the required capillarypressure to pump the saturated liquid from the con-denser to the evaporator. For these configurations, thecapillary limitation values are below zero.

The best wick structure configuration for the 3/8"heat pipe (Table 5), for the geometrical parametersgiven in Tab. 3, is that composed by one layer of screenMesh 166, which is capillary limited at 1.86 W at 100K. The second best configuration is that composed by 2layers of screen Mesh 166, which is boiling limited at1.73 W at 100 K. No negative values for the heat trans-fer limitations were found for any configuration of the3/8" heat pipe.

Optimum configuration featuresFigure 4 and 5 shows the lowest heat transfer limi-

tation mechanisms for the 1/2" and 3/8" heat pipe (re-spectively) with a wick structure composed of two

. layers of screen Mesh 166, highlighted in Table 4 and5, as a function of the lengths of the condenser and ofthe evaporator at the temperature of 100 K. The con-denser and the evaporator length range from 0 to 0.4meters. The maximum length of the heat pipes (Lc + La+ Lc) is kept constant, and equal to 0.8 meters. Fromthese figures, one can observe that the heat transferlimitation tends to increase when the length of both theevaporator and the condenser rises. The optimum val-ues are the highest ranges value on the scale. For futureexperimental purposes, it will be necessary to consideran adiabatic section, corresponding to the region where



the cryogenic heat pipe will be fixed at the experimentalapparatus. In this sense, a 0.1 meter adiabatic section isconsidered. Therefore, the numerical values of Fig. 4and 5 suggest that the optimum values for the evapora-tor and the condenser length are 0.4 meter and 0.3 me-ter, respectively. The heat transfer limitation for the3/8" heat pipe configuration is similar, and its optimumlengths are also 0.4 and 0.3 meters for the evaporatorand the condenser, respectively.

Figure 6 and 7 shows the heat transfer limitationsmechanisms for the 1/2" and 3/8" heat pipe (respec-tively) with a wick structure composed of two layers ofscreen Mesh 166, as a function of the temperature andof the tilt angle. The white line corresponds to the tem-perature of 100 K. Negative angles means that the con-denser is higher than the evaporator. The highest heattransfer limitation at 100 K is obtained with a tilt angleof-7° and -6° (2.63 W and 1.73 W, respectively). For

Optimum point Optimum point for a1 m adiabatic section

120^45 ̂

Figure 4 - Heat transfer limitation for the 1/2" heatpipe with a wick structure composed of two layers of

screen Mesh 166.

Figure 6 - Heat transfer limitation for the tilted 1/2"heat pipe with a wick structure composed of two layers

of screen Mesh 166.

2.5

Figure 5 - Heat transfer limitation for the 3/8" heatpipe with a wick structure composed of two layers of

screen Mesh 166.

Figure 7 - Heat transfer limitation for the tilted 3/8"heat pipe with a wick structure composed of two layers

of screen Mesh 166.



tilt angles below -7° and -6°, the heat transfer limitationremains constant. At 100 K, the heat pipes would notwork properly at tilt angles greater than 27° (1/2" heatpipe) and 40° (3/8"heat pipe), where the capillary limitbecomes negative (black region of Fig. 4 and 5). In theblack region the capillary structure cannot provide therequired capillary pressure to pump the liquid back tothe evaporator. The heat transfer limitation increases, asthe temperature decreases for a fixed tilt angle.

Table 6 summarizes the dimensions and wick struc-tures for the four heat pipe configurations under inves-tigation in this paper. The heat transfer limitation pre-sented is given at 100 K and at horizontal position (0°of tilt angle).

Table 6 - Optimum geometric parameters and wickstructure for the stainless steel/nitrogen heat pipes.

HP1 HP 2 HP 3 HP 4Externalradius (inches) \IT 3/8" 3/8"

Vaporradius (mm) 4.5 4.47 4.635 4.5

Wick structurethickness (mm) 0.3500 0.3800 0.2155 0.3500

Wick structurematerial AISI 316 Stainless steel screen

Mesh 166 254 166 166No. of layersPorosity 0.883 0.901Permeability ] 6 x 1Q_9 }^x^(m )

0.905 0.883

2.6 x 10'9 1.6x 10's

Heat transferlimitation (W)* 2.12 2.32 2.06 1.73

Evaporator length (m)Adiabatic section length (m)

0.40m0.10m

Condenser length (m) 0.30m

heat transfer limitation at 0° of tilt angle.

Figure 8 and 9 shows the heat transfer limitation asa function of the operational temperature for the 1/2"and the 3/8" heat pipe (respectively) with a wick struc-ture composed of two layers of screen Mesh 166, at thehorizontal position. The 1/2" heat pipe would not workproperly at temperatures over 111 K, where the capil-lary limit becomes negative. The temperature limitationfor the 3/8" heat pipe is 115 K. At 100 K, the capillarylimit is dominant for the 1/2" heat pipe, while, for the3/8" heat pipe, there is a composition of boiling andcapillary limitations.

85 95 100 105Temperature [K]

115 125

Figure 8 - Heat transfer limitations for the 1/2" heatpipes as a function of the temperature.

85 95 100 105Temperature [K]

115 125

Figure 9 - Heat transfer limitations for the 3/8" heatpipes as a function of the temperature.

Future DevelopmentsAn experimental apparatus for ground testing of

cryogenic heat pipe (showed in Fig. 10) is under con-struction, and it is supposed to be ready for testing byJuly 2001. The experimental apparatus is, basically, apivoted vacuum chamber, with feed-through for ther-mocouples, electric power (for the evaporator section ofthe heat pipe), and liquid nitrogen (for the condensersection of the heat pipe). The apparatus is able to testheat pipes up to 0.8 meters long with a external diame-ter of up to 3/4".

The one-dimensional modeling of the startup ofcryogenic heat pipes, based on the analytical model ofYan and Ochterbeck is under development. Also, thetwo-dimensional modeling of the vapor flow of cryo-genic heat pipes during startup is being considered.These results will be soon released in the literature.



Vacuum chamber||$t section)

Figure 10 — Experimental apparatus for testing of cryo-genic heat pipes under development at NCTS.

ConclusionsThe present paper presents the design procedures

for the Brazilian cryogenic heat pipes, under develop-ment at the Satellite Thermal Control Laboratory, of theFederal University of Santa Catarina. The proceduresare based on the heat transfer limitation mechanisms,and the results showed to be coherent with the theoreti-cal expressions.

Four configurations are proposed for future groundtests at the NCTS facilities in Brazil. Values for the heattransfer limitations for all these configurations werepresented based on an initial estimation of the evapora-tor and condenser lengths. Three-dimensional graphsfor two of these four configurations were presented inorder to determine the optimum values for the evapora-tor and condenser lengths. Also, three-dimensionalgraphs, for the heat transfer limitation, were showed asa function of the tilt angle and steady-state temperature.It was shown that the dominant heat transfer limitationsfor the four cryogenic heat pipes presented in this workare the capillary and the boiling limits.

Acknowledgments - The authors would like to ac-knowledge: Prof. Jay Ochterbeck (Clemson University,SC, USA) for his help during the development of thiswork; Prof. G. P. Peterson (Rensselaer PolytechnicInstitute, NY, USA) for his help during the literaturereview; and finally, the Brazilian Space Agency (AEB),the Brazilian Council for Technological Development(CNPq) and the Brazilian Coordination for the Devel-opment of Graduate Level Human Resources (CAPES)for supporting and funding this project.

ReferencesCouto, P., "Development and Design of Passives

Cryogenic Radiators for Spacecraft Applications", M.Sc. Thesis (original version in Portuguese), MechanicalEng. Dept, Federal University of Santa Catarina, Flori-anopolis, SC, Brazil, Jul 1999.

2 Brand, O. and Schlitt, R., "Low Temperature Ra-diator Design for the ABRJXAS X-Ray Satellite", Proc.6th European Symp. on Space Environmental ControlSystems, pp. 151 - 159, Noordwijk, The Netherlands,May 1997, (ESA SP-400, Aug 1997).

3 Wright, J. P., "Development of a 5 W 70 K Pas-sive Radiator", AIAA Paper No. 80-1512, Jul. 1980.

4 Wright, J. P. and Pence, W. R., "Development of aCryogenic Heat Pipe Radiator for a Detector CoolingSystem", ASME Paper No. 73-ENAs-47, Jul. 1973.

5 Zelenov, I. A., Poskonin, U. A., Timofeev, V. N.,Kostenko, V. I., Ribkin, B. I., Romanovsky, O. I., Si-dorenko, E. M., and Guskov, A. S., "Flexible LowTemperature and Cryogenic Heat Pipes for the SpaceControlled Radiator-Emitter", Proc. of the 8th Int. HeatPipe Conf, pp. E-P66/1 -4, Beijing, China, 1992.

6 Voyer, E., Moschetti, B., Briet, R., Alet, 1., andEvin, R., "Heat Pipes for Cryogenic Applications onSatellites", SAE Technical Paper No. 972450, 27th Int.Conf. on Environmental Systems, Lake Tahoe, Nevada,Jul. 1997.

7 Abrosimov, A., Baryshev, O., Horonenko, V.,Kosorotov, M., Lobanov, A., and Parfentiev, M., "Di-ode Cryogenic Heat Pipe for "SODART" TelescopeSilicon Detector Cooling', Proc. of the 8th Int. HeatPipe Conf, pp. E-P54/1 -4, Beijing, China, 1992.

8 Ishigohka, T., Hirayama, Y., Ninomiya, A., and S.Maezawa, "Conduction Cooling of High-TC Supercon-ducting Magnet Using Heat Pipe", Proc. of the 11th Int.Heat Pipe Conf, pp. 139 - 144, Tokyo, Japan, Sep.1999.

9 Couto, P., and Mantelli, M. B. H., "CryogenicHeat Pipe - A Review of the State-of-the-Art", Proc.Brazilian Congress of Thermal Eng. and Sciences, Vol.CD-ROM, Porto Alegre, RS, Brazil, Oct., 2000.

10 Brazilian Space Agency, "PNAE - BrazilianPolicy for Space Activities" (original in Portuguese),Published by AEB, Brasilia, DF, Brazil, 1996.

11 Peterson, G. P., and Compagna, G. L., "Reviewof Cryogenic Heat Pipes", AIAA Journal of Spacecraft,Vol. 24, No. 2, Mar. - Apr., 1987.

12 Barron, R. F., "Cryogenic Systems", pp. 3 - 55,2nd Edition, Oxford University Press, New York, NY,USA, 1985.

13 Cm", S. W., and Cygnarowicz, T. A., "TheoreticalAnalyses of Cryogenic Heat Pipes", ASME Paper No.70-HT/SpT-6, 1970.



14 Gilmore, D. G., "Satellite Thermal ControlHandbook", 1st Edition, The Aerospace CorporationPress, El Segundo, CA, 1994.

15 Tournier, J. -ML, and El-Genk, M. S., "A HeatPipe Transient Analysis ModeF, Int. Journal of Heatand Mass transfer, Vol. 37, No. 5, pp. 753 - 762, 1994.

16 Tournier, J. -M., and El-Genk, M. S., "Segre-gated Solution Technique for Simulating the TransientOperation of Heat Pipes", Numerical Heat Transfer,Part B, Vol. 25, pp. 331 - 355, 1994.

Colwell, G. T., "Prediction of Cryogenic HeatPipe Performance", Annual Report for 1975, ReportNo. 2 (NASA Final Report NSG-2054), Feb. 1976.

18 Colwell, G. T., "Prediction of Cryogenic HeatPipe Performance", NASA Final Report NSG-2054,Mar. 1977.

19 Brennan, P. J., Thienen, L., Swanson, T., andMorgan, M., "Flight Data for the Cryogenic Heat Pipe(CRYOHP) Experiment", AIAA Paper No. 93-2735,Jul. 1993.

20 Ochterbeck, J. M., Peterson, G. P., and Ungar, E.K., "Depriming/Rewetting of Arterial Heat Pipes:Comparison with Share-11 Flight Experiment^, AIAAJournal of Thermophysics and Heat Transfer, Vol. 9,No. l,pp. 101-108, Jan-Mar. 1995.

21 Rosenfeld, J. H., Buchko, M. T., and Brennan, P.J., "A Supercritical Start-Up Limit to Cryogenic HeatPipes in Micro gravity", Proc. of the 9th Int. Heat PipeConf., Vol. II, pp. 742 - 753, Albuquerque, NM, 1995.

22 Van, Y. H. and Ochterbeck, J. M., "Analysis ofSupercritical Start-Up Behavior for Cryogenic HeatPipes", AIAA Journal of Thermophysics and HeatTransfer, Vol. 13, No. 1, pp. 140 - 145, Jan-Mar. 1999.

2j Imura, H., Kozai, H., Hayashida, S., and Taka-shima, K., "Heat Transfer Characteristics in ScreenWick Heat Pipes", JSME International Journal - SeriesII, Vol. 31, No. 1, pp. 88-97, 1988.

24 Faghri A., "Heat Pipe Science and Technol-ogy", Taylor & Francis, pp. 278 - 295, Washington,DC, 1995.

25 Peterson G. P., "An Introduction to Heat Pipes- Modeling, Testing, and Applications", John Wiley& Sons, Inc., New York, NY, USA, 1994.

26 Reynolds, W. C., "Thermodynamic Propertiesin SI", Published by the Dept. of Mechanical Engineer-ing, Stanford Univ., Stanford, CA, 1979.


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