A study of a nitrogen heat pipe
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Scholars' Mine Scholars' Mine
Masters Theses Student Theses and Dissertations
1971
A study of a nitrogen heat pipe A study of a nitrogen heat pipe
Jay Dudheker
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A STUDY OF A NITROGEN HEAT PIPE
by
JAY DUDHEKER, 1932-
A THESIS
submitted to the faculty of the
UNIVERSITY OF MISSOURI-ROLLA
in partial fulfillment of the requirements for the
Degree of
MASTER OF SCIENCE IN MECHANICAL ENGINEERING
Rolla, Missouri
1971
1.9426.0
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ii
ABSTRACT
An experimental heat pipe, 33.25 inches long and 0.75
inches in diameter, with modacrylic fiber wick and liquid
nitrogen as its working fluid was constructed to study the
operating characteristics of a cryogenic heat pipe. The
effective thermal conductivity and the axial temperature
distribution were determined for various levels of power
input. The effect of inclination angle on the above
parameters was also measured.
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iii
ACKNOWLEDGEMENTS
The author wishes to express his sincere appreciation
to his advisor, Dr. B.F. Armaly, for his guidance, counsel,
and constant encouragement throughout the project and pre
paration of this thesis.
The constructive criticism of Dr. H.J. Sauer and
Dr. J.B. Prater is appreciated.
Sincere thanks and appreciation go to my wife Georgia
for her patience and understanding throughout the past year.
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TABLE OF CONTENTS
ABSTRACT • . • • •
ACKNOWLEDGEMENTS .
LIST OF FIGURES
LIST OF TABLES
NOMENCLATURE •
I. INTRODUCTION
II. REVIEW OF LITERATURE
iv
Page
ii
iii
v
vii
• • viii
1
5
III. DESCRIPTION OF THE EXPERIMENTAL APPARATUS • 15
IV. EXPERIMENTAL PROCEDURE AND DATA REDUCTION • 24
v. RESULTS AND DISCUSSION
VI. CONCLUSIONS AND RECOMMENDATIONS •
BIBLIOGRAPHY •
VITA ••
APPENDIX A.
APPENDIX B.
. . . . . . . . . . . . . . . .
EXPERIMENTAL DATA AND RESULTS
THERMOPHYSICAL PROPERTIES OF NITROGEN • • • . • • • • • . .
34
43
45
47
48
53
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Figure
1.
2.
3.
4.
5.
6.
7.
8.
9 .
10.
11.
12.
13.
14.
15.
LIST OF FIGURES
Heat Pipe • • • • • . . . . . . . Radial Temperatures at Evaporator and Condenser Sections . • • • . • •
Cryogenic Tank . . . . . . . . . Heat Pipe with Vacuum Jacket
Calibration Curve for Nitrogen Storage Tank. . . . ·· . . . . . . . . . . . . .
Thermocouple Locations
Experim~ntal Set-Up of Nitrogen Heat Pipe Assembly . • • • • • • • • •
Experimental Results for the Case of Horizontal Level . . • . • • . • .
Steady State Temperature Distribution, Horizontal Level • • • . • • • • • • • •
Experimental Results for 1 Degree Inclination Angle with Evaporator Above Condenser • . . • . . • • • •
Experimental Results for 1.75 Degree Inclination Angle with Evaporator Above Condenser • • . . . • • • . • .
Experimental Results for 5.25 Degree Inclination Angle with Condenser Above Evaporator • • . . • • • • . • . . .
Steady State Temperature Distribution for 5.25 Degree Inclination Angle with Condenser Above Evaporator . . •
Effective Thermal Conductivity for the Heat Pipe at Different Angles of Inclination • • • • • • • •
Effective Thermal Conductivity of Saturated Wick in Evaporator and Condenser Sections Based on Chi's Analysis
v
Page
3
10
16
17
20
21
23
27
28
30
31
32
33
35
38
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Figure
16.
B-1
B-2
B-3
B-4
B-5
B-6
B-7
B-8
B-9
B-10
LIST OF FIGURES (continued)
Comparison of the Total Experimental Temperature Difference • . • . • • •
Vapor Pressure of Nitrogen • • • • • • •
Density of Saturated Liquid Nitrogen
Density of Gaseous Nitrogen (Saturated Vapor) • . • . • . . • • . . • .
Dynamic Viscosity of Liquid Nitrogen •
Dynamic Viscosity of Gaseous Nitrogen at Atmospheric Pressure • . . • • • • •
Heat of Vaporization of Nitrogen •
Surface Tension of Saturated Liquid Nitrogen . • • • . . •.•••..
Specific Heat of Saturated Liquid Nitrogen . . . . • • • • •..
Thermal Conductivity of Saturated Liquid Nitrogen . . . . . . . . . . . . . . . .
Thermal Conductivity of Gaseous Nitro-gen at Atmospheric Pressure • • . . . • • .
vi
Page
41
54
55
56
57
58
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Table
A-1
A-2
A-3
A-4
A-5
LIST OF TABLES
Experimental Results, Heat Pipe Operating in the Horizontal Position
Experimental Results, Heat Pipe Operating at 1 Degree Angle, Evaporator Above Condenser • • • • • • • • . • .
Experimental Results, Heat Pipe
. . .
Operating at 1.75 Degree Angle, Evaporator Above Condenser • • • • • • • • • • • • •
Experimental Results, Heat Pipe Operating at 5.75 Degree Angle, Condenser Above Eva para tor . . . . . . . . . . . . . . . . .
Deduced Experimental Results • • . • • • • •
vii
Page
49
so
so
51
52
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Symbol
a
A
b
c p
e
g
K ss
L
NOMENCLATURE
Quantity
Sonic velocity
Vapor passage cross-sectional area
Property of wick, due to tortuous path taken by fluid through pores
Specific heat at constant pressure
Wick porosity
Acceleration due to gravity
Constant factor of propotionality
Latent heat of vaporization
Thermal conductivity of liquid saturated wick
Thermal conductivity of stainless steel tube
Effective thermal conductivity of heat pipe
Total length of heat pipe
Length of adiabatic section
Length of evaporator section
Length of condenser section
viii
Units
ft/sec
dimensionless
dimensionless
ft/sec 2
lbm-ft/lbf-sec2
Btu/lbm
Btu/hr-ft-0 R
ft
ft
ft
ft
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Symbol
• m
Q
t
T . Cl.
T co
T . el.
T v
!J.P c
!J.P g
NOMENCLATURE (continued)
Quantity
Mass of liquid
Mass flow rate
Heat transfer rate
Capillary wick pore radius
Outside radius of containing stainless steel tube
Inside radius of containing stainless steel tube
Radius of vapor space
Thickness of wick
Temperature at the inner wall of the containing tube at the condenser
Temperature at the outer wall of the containing tube at the condenser
Temperature at the inner wall of the containing tube at the evaporator
Temperature at the outer wall of the containing tube at the evaporator
Vapor temperature
Capillary pumping pressure drop
Pressure drop due to gravity forces
ix
Units
lbm •
lbm/hr
Btu/hr
ft
ft
ft
ft
ft
_oR
lbf/ft2
lbf/ft2
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Symbol
b.P v
9
lll
NOMENCLATURE (continued)
Quantity
Pressure drop in vapor phase
Pressure drop in liquid phase
Contact angle
Viscosity of liquid
Viscosity of vapor
Density of liquid
Density of vapor
Surface tension of liquid
Inclination angle of heat pipe
X
Units
lbf/ft2
lbf/ft2
degrees
lbm/ft-hr
lbm/ft-hr
lbm/ft3
lbm/ft3
lbf/ft
degrees
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1
I. INTRODUCTION
The heat pipe is a self-contained structure made
out of a thin-wall sealed tube lined on the inside with
a wick which is saturated by a working fluid. The main
attractive feature is its capacity to transfer heat
better than the best metallic conductors, such as silver
or copper. For instance, thermal power of 11,000 watts
[1]* was transferred by a one inch heat pipe over a
distance of 27 inches with practically no temperature
drop. By way of comparison a copper rod of the same
length and nine feet in diameter weighing about 40 tons
would be required to produce the same results. The opera
tion of a heat pipe is accomplished by continuously evapor-
ating and condensing the working fluid, and transferring
heat by mass flow from evaporator to condenser utilizing
the latent heat of vaporization.
Since its invention by Gauler in the year 1942,
and a subsequent follow up by Grover [2] at Los Almos
Scientific Laboratory, more and more educational institu-
tions and industries from various fields have become
interested in its development and applications. It has
received considerable attention in space technology due to
its light weight and its ability to perform satisfactorily
under zero gravity. Other industries are also making use
*Numbers in brackets indicate references listed in Bibliography
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2
of this principle to service their needs and requirements.
For example, the railroad [3] industry are experimenting
with large capacity heat pipes (i.e., 10 to 20ft. long),
with water as working fluid, for use in transit to main
tain some commodities, such as tar, at high temperatures.
Most of the research and development has been restricted
to heat pipes operating at relatively high temperatures,
above normal ambient temperature. The design techniques
and operating characteristics of these heat pipes have
been well defined by a considerable number of investiga
tors. Review of the technical literature has indicated,
however, that only two experimental studies have been per
formed on heat pipes operating in the cryogenic temperature
range with nitrogen [4,5] (139.3°R) as the working fluid
and only two theoretical analyses [6,7] on the design of
cryogenic heat pipes has been published.
The heat pipe, shown in Figure 1, is a sealed slender
long thin-wall tube which is lined on the inside with a
saturated wick. Heat added from a source to a section of
the tube, called the evaporator, causes the working fluid
to evaporate. This action causes a pressure gradient to
exist between the hot and the cold end of the tube which
forces the vapor to flow through a central passage towards
the cooler section of the tube known as the condenser.
As the vapor enters the condenser, condensation takes place
and energy equivalent to the latent heat of vaporization
is released. Due to surface tension forces present in the
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CONDENSER SECTION
(HEAT SINK)
ADIABATIC SECTION
FIGURE 1. Heat Pipe
EVAPORATOR SECTION
(HEAT SOURCE)
w
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4
wicking material, the condensate returns along the wick
from the condenser to the evaporator section, and the
above fluid cycle is repeated.
The purpose of this study is to investigate
experimentally the operating characteristics of a cryo
genic heat pipe operating with nitrogen as a working fluid
in order to evaluate existing analytical models and experi
mental results.
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5
II. REVIEW OF LITERATURE
Since the initial work of Grover several heat pipe
analyses have appeared in the literature. Cotter [8] was
the first to develop the equations governing the dynamics
of heat transfer in a heat pipe. The formulation was
based on a balance of pressure drops due to the various
significant mechanisms promoting the energy transfer, and
is given as
( 1)
In order that the heat pipe operate satisfactorily,
the capillary pumping head developed in the wick due to
surface tension forces, ~Pc' must remain larger or equal
to the sum of all other pressure drops, for example the
pressure drop due to viscous forces in the liquid, ~P 1 ,
the vapor pressure drop, ~Pv' and the pressure drop due
to gravity forces ~P • g Maximum heat transfer capability
of the heat pipe is achieved when the capillary pumping
head is equal to the sum of all other pressure drops.
Mathematical expressions relating the pressure drop
to other system properties have been obtained by several
investigators. For example, the capillary pumping head
which can develop in a wick structure is given by
(2)
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6
Experimental evidence indicates [9] that the effective
capillary wick pore radius in equation(2) can be taken as
half the width of the opening between two parallel threads
plus half the thread or wire thickness, and the value of
contact angle e, which the liquid meniscus makes with the
capillary channel formed by parallel threads of the fiberp
can be assumed zero. An expression for the liquid pressure
drop in the wick was derived by Cotter using Hagen Poise-·
uille law, and assuming laminar flow through the porous
media. The final expression is given by
= b~l Q(Le+Lc)
2n(r~-r~)p 1er~hfg {3)
where (L +L ) is the total length of the heat pipe. The e c
expression for the pressure drop in the vapor phase was
obtained also by Cotter. The vapor flow was treated as
a laminar, incompressible flow in a circular duct with
constant injection along the evaporator section and suction
along the condenser section. This expression is given by
4~VQ(Le+Lc) =
npvr!hfg (4)
In most cases this pressure drop is negligible relative
to other pressure drops encountered in the heat pipe. The
gravitational pressure drop can be simply calculated from
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~p = + P' ~ L cos ~ g 1 gc
(5)
7
This gravitational effect could enhance or reduce the
heat transfer capability of a heat pipe. If the condenser
is at a level above the boiler, the gravity effect in-
creases the heat transfer while in the reverse case it
decreases the heat transfer. When the heat pipe is in a
horizontal position the gravity effects are nil.
Haskin modified Cotter's results, equations (3) and
(4), for application to a heat pipe with an adiabatic
transfer section between the evaporator and condenser. His
results indicate that the above equations are still appli-
cable to this new geometry if the length parameters (Le+Lc)
appearing in the above expressions is changed to(L +La),
where La is the length of the adiabatic section and L is
the total length of the heat pipe.
To determine the maximum capability of the heat pipe,
the equality in equation (1) is used. By substituting the
various expressions for the pressure drops, for the case of
a heat pipe with an adiabatic transfer section, a relation
for the maximum heat transfer capability of the heat pipe
can be obtained in terms of wick, vapor and liquid proper-
ties and geometric parameters.
2crcose rc
+ P1 }- L cos {I ( 6) c
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8
Haskin sol_ved the above equation, neglecting the gravita-
tiona! term, and deduced the following relationships for
optimum capillary pore radius, optimum wick thickness
ratio, and maximum heat transport as
r =[ b].llpvr! r/2 (7) c 2 2 811 p1e(r -r ) v w v
where
= (8)
and
Rearranging and using the optimum value described in
equation (8), the optimum value for heat transport can
also be expressed by
= 2'TTr~hf9crcos9 3 (L+La)
(10)
(9)
Joy [7] included the gravitational term in equation
(6), and derived the relationship for optimum capillary
pore radius and maximum heat transport. His final results
are given by
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9
b~ 1 (L+L )gLcos~ 2 [( a ) +
21Thf e(r2-r2 )g g w v c
(11)
and
2crr0-r~p 1 ....2. gc Lcosfl
0max = (12) b~l (L+La) 4r 2~ (L+L )
+ c v a
21Thfge(r~-r!>P 1 4 1Tpvhfgrv
The total temperature drop across the length of the
heat pipe, between evaporator and condenser, is equal to
the sum of the individual radial temperature drops in the
evaporator and the condenser. This drop in temperature
has been treated by Chi [6] and the results are given by
(Figure 2) •
Teo-Teo= (Teo-Tei)+(Tei-Tv)+(Tv-Tci)+(Tci-Tco> (lJ)
The radial temperature drop due to conduction through the
stainless steel container is given by
Q r T - T = ln(ro) eo ei 21TK L ss e w
(14)
and Q
r T ci - Teo =
2iTKSSLC ln(r o)
w (15)
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Tv TCI
Teo
B-B
STAINLESS STEEL TUBING
B CONDENSER
SECTION
WICK
VAPOR
ADIABATIC SECTION
A EVAPORATOR
SECTION
Tv TEl
TEO
FIGURE 2. Radial Temperatures at Evaporator and Condenser Sections
A-A
..... 0
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11
for the evaporator and condenser, respectively. The
radial temperature drop across the saturated wick thick
ness is given by
T I - T = e~ v
Q 27Tr L K we m
t
t exp
for the evaporator section, and
Tv - T . = Cl.
Q 27Tr L K w c m
m cP exp(- 47Tr L tK
w c m
{16)
for the condenser section. The mass flow rate appearing in
these expressions can be calculated using the following
expression.
xn = o hfg
( 18)
The possible use of the theoretical equations for predicting
heat pipe operating conditions depends heavily on how well
the properties of the saturated wick are known,such as ther-
mal conductivity, porosity, permeability and pore radius.
This kind of information, however, is not available and
more research is needed to specify more accurately these
properties.
Several limiting conditions on the operating character-
istics of a heat pipe have been reported in the literature.
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12
These limitations prevent the heat pipe from achieving
its. optimum operating conditions. For example, the vapor
in the core could reach sonic velocity resulting in a choked
flow condition in the heat pipe. The heat flux attained
under such a choked condition represents the maximum pos
sible heat flux which could be smaller in magnitude than
the optimum value predicted from equations (10) and (12).
This limiting criteria [10] is specified by
{19)
For cryogenic heat pipes, this limitation can be ignored
due to the low magnitude of the optimum heat fluxes.
Another significant limiting condition could be caused
by the entrainment of liquid in the vapor. Vapor and
liquid in the heat pipe normally travel in opposite direc
tions and frictional forces under this limiting condition
could cause the liquid to leave the wick structure and
become entrained in the vapor. This effect, under severe
conditions, could prevent the liquid from returning to the
condenser under the action of capillary forces and cause
the heat pipe to fail. Experimental studies conducted by
Kemme [11] have indicated that this limitation could be
controlled by selecting a fine pore wicking material. If
it is necessary to use coarse pore materials as a wick, the
inner surface which contacts the vapor should be covered
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13
with finer wicking material. The magnitude of the heat
flux under which this limiting condition is reached could
be calculated by
~P crhf Q = v g A Z
(20)
where Z is a dimension associated with wick surface and
nearly equal to wick thread diameter.
Heat transferred by the heat pipe can also be
limited by condenser and evaporator parameters. A high
energy flux at the evaporator section could cause a
phenomenon similar to film boiling where a film of vapor
can become trapped between the wick and the tube. This
film in effect increases the resistance to radial heat
flow and thus results in a corresponding increase in
evaporator temperature which leads to heat pipe failure.
In addition the working fluid must be free of all impurities
for proper heat pipe operation. Impurities which are more
volatile than the working fluid would collect in the con-
denser section and those which are less volatile would
collect in the evaporator section. These impurities in
effect reduce the effective areas of the condenser and the
evaporator resulting in a reduced heat transfer capability
of the heat pipe. At any instance during the heat pipe
operation, the capillary forces must exceed all other active
and opposing forces. When such a criterion fails to exist,
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14
for example due to an increase in gravity forces, the
heat pipe fails. This limitation is known as wick limi
tation.
The above review indicates the lack of sufficie-nt
experimental investigations of heat pipes operating in
the cryogenic temperature range. The prediction of heat
pipe operating conditions depends on how well the wick
properties are known. This type of information presently
does not exist, and thus experimental work is needed to
further study the heat pipe operating characteristics in
this low temperature range.
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15
III. DESCRIPTION OF THE EXPERIMENTAL APPARATUS
The experimental apparatus consisted of a cryogenic
tank, which provided an environment at liquid nitrogen
temperature (139.3°R); a heat pipe, on which measurements
were made; a vacuum jacket, which provided thermal pro
tection to the heat pipe; and a precalibrated tank
acting as a nitrogen gas reservoir.
The cryogenic tank, shown in Figure 3, consists of a
double jacketed rectangular container 3.5 x 0.58 x 1.5 feet
in dimensions. It was fabricated out of 16 gauge stain
less steel sheet metal and strengthened along its length
by several beams, 0.25 x l inches, to prevent buckling
when vacuum is applied to the surrounding jacket. The
jacket was maintained under a vacuum, approximately 12
microns, during the experiment in order to reduce the
evaporation rate of the liquid nitrogen contained within
the cryogenic tank.
The experimental heat pipe, shown in Figure 4, was
33.25 inches long and made of 22 gauge stainless steel,
type 304, tubing and has an outside diameter of 0.75 inches.
In order to form a closed tube, two stainless steel flanges
were silver soldered at the two ends. A 0.125 inches stain
less steel tube was welded to one of the end caps to serve
as a fill and evacuating line. The heat pipe was divided
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16
. M
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FILL LINE
HEAT PIPE
PRESSURE GAUGE
TO VACUUM LINf
INSTRUMENT FEED THROUGH
THERMOCOUPLE LEADS
CONDENSER SECTION
ADIABATIC SECTION
' FIGURE 4. Heat Pipe with Vacuum Jacket
EVAPORATOR SECTION
VACUUM JACKET
~ ~
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18
into evaporator, condenser and adiabatic sections. The
evaporator section was 10 inches long. Energy was added
to this section through a nichrome wire resistance heater,
15 feet long which was wrapped uniformly around the
peripheral surface area of the evaporator. Scotch tape
and general electric adhesive, No. 7031, were used to
electrically insulate and at the same time provide good
thermal contact between the heater wire and the heat pipe
surface. The condenser section was 10 inches long at the
other end of the tube, and served as a heat sink while the
middle section between the evaporator and the condenser
served as the adiabatic section.
In order to isolate the evaporator and the adiabatic
section of the heat pipe from direct contact with the
cryogenic environment, they were enclosed inside the
vacuum jacket as shown in Figure 4. A cap was silver
soldered to the heat pipe on the line separating the con
denser from the adiabatic section to provide a seal when
connected to the vacuum jacket. A tube was welded verti
cally to the jacket and was used as an evacuating line and
a housing for the instrument wires. The open end of this
connecting tube was then closed by an instrument feed
through flange and a connection was made to the vacuum
pump line.
A stainless steel cylindrical bottle, 0.318 cubic
feet in capacity, was used as a storage tank for nitrogen
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19
gas. This bottle was equipped with a pressure gauge
which was calibrated as a function of mass contained within
the bottle. For any given pressure and room temperature,
the mass of nitrogen within the bottle could be read
directly from Figure 5. This bottle was connected to the
heat pipe through the feed line and was used to charge
and determine the amount of nitrogen charged into the heat
pipe. The feed line was also equipped with a pressure
gauge to measure the pressure within the heat pipe during
operating conditions.
The axial temperature distribution along the external
surface of the heat pipe was measured by thermocouples
and a potentiometer. The thermocouples were copper
constantan, 5 mils in diameter. Six thermocouples were
distributed as shown in Figure 6 (one on the evaporator
section, four on the adiabatic section, and one on the
condenser section) • Each thermocouple was tempered by
rapping the thermocouple leads around the tube for a length
of six inches to insure accurate temperature measurements.
Electrical energy was supplied to the evaporator section
through the heater from a 110 volt - 60 cycle AC outlet.
A powerstat was used to vary the energy input and an AC
wattmeter was used to measure the energy added to the heat
pipe.
The wick was made out of modacrylic fiber cloth which
is a synthetic fiber woven from 18/1 cc. yarn. The fiber
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-.. ~
~ '-'
1'1'"\ c::> r-i
X (/) (/) <( ~
z UJ C) 0 0::: .,_ -z
20
160r---------------------------------
40
30
20
10
PRESSURE (PSI)
FIGURE 5. Calibration Curve for Nitrogen Storage Tank
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CONTAINING TUBE
CONDENSER SECTION
ADIABATIC SECTION
EVAPORATOR SECTION
I~ 8,25" •I
1... 11'' ~
1~ 14 .5'' ... 1
1... 18" -----------~~ 1~ 22.5'L--------+i
~ 25.5''"----------+t
FIGURE 6. Thermocouple Locations N ~
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22
had 64 ends and 43 picks of yarn per square inch and was
a plain weave. The weight of the fiber was 3.6 ounces
per square yard and its thickness was 0.015 inches.
This material was selected because of its fast wicking
properties and its high capillary pumping head. The
capillary pumping head was measured for four types of
wicking materials, 100 mesh stainless steel screen and
rnodacrylic acrilon, and polyester fibers by dipping them
in distilled water and observing the speed and the maximum
height of wicking achieved by capillary pumping. The
modacrylic fiber was superior to the other three materials.
Three layers of this material were wrapped and inserted
into the heat pipe to serve as a wick. The wick was tightly
held by a helical spring inside the heat pipe tube.
The heat pipe with its protective vacuum jacket was
suspended inside the cryogenic tank through a rotating
device in such a way that the entire assembly could be
rotated and locked in position in either direction (Figure
7) . This assembly was used to investigate the operating
characteristics of the heat pipe at different angles of
inclination. For evacuating the cryogenic tank, the heat
pipe, and the vacuum jacket, each was connected to the
vacuum pump through appropriate swagekck fittings and
valves as shown in Figure 7.
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PRESSURE GAUGE IIPREs-··
VACUUM GAUGE
0
POTENTIOMETER
c
0 0
TO ELECTRICAL
SUPPLY
POWER~TAT
SURE GAUGE LIQUID NITROGEN LEVEL
II- -~--=- ~ ~:- -p - = - --- ----- -- -
N2 SUPPLY n ... L-nun vnvuun runr \...LN2 CONTAINER '1"\1'\""r'T'I r-
FIGURE 7. Experimental Set~Up of Nitrogen Heat Pipe Assembly
t.J w
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24
IV. EXPERIMENTAL PROCEDURE AND DATA REDUCTION
The successful operation of the heat pipe relies
heavily on purity of the working fluid and the non
presence of any foreign destructive material in the wick
structure. To insure this state, the heat pipe and all
the connecting tubings were evacuated at room temperature
to approximately 5 microns and then flushed by nitrogen
gas. This,procedure was repeated at least six times to
dilute and remove any non-condensable foreign gas from
the heat pipe. A leak check was performed to insure the
isolation of the heat pipe from the surrounding environ
ment. The heat pipe structure was then adjusted to a
horizontal level by a liquid level indicator situated
temporarily on the condenser section. The degree of
accuracy to fix the entire heat pipe in horizontal
position by this method could not be ascertained.
As a final step in the preparation of the heat pipe,
the whole assembly·, consisting of the cryogenic tank, pro
tective jacket and the heat pipe tube, was evacuated for a
period of 48 hours. The cryogenic tank was partially
filled, to cover the whole heat pipe, with liquid nitrogen
and acted as a heat sink. Every three hours, throughout
the experiment, liquid nitrogen was added to the cryogenic
tank to maintain the liquid nitrogen level above the heat
pipe approximately constant.
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25
To determine experimentally the amount of energy
loss by conduction through the containing tube wall and
instrument wires and by radiation to the protecting
jacket, energy was added to the evaporator section while
the heat pipe was evacuated. It was observed that for
the temperature drop range, between the evaporator and
condenser, covered during the experimental run, this
energy loss varied between 0.025 and 0.045 watts which is
insignificant relative to the energy transferred by the
heat pipe during the experiment.
To operate the heat pipe, nitrogen gas was injected
to the heat pipe through the fill line from the calibrated
plenum bottle. At this time electrical energy was not
being added to the evaporator section. Due to the fact
that the heat pipe's condenser section was at nitrogen
temperature (139°R), the gas added was immediately conden
sed. This condensate then was carried by the capillary
action of the wick towards the evaporator section. The
axial temperatures along the heat pipe were monitored
throughout the condensation process. As the condensate
travelled through the wick towards the warmer end, the
temperature at each cross section dropped rapidly. The
entire heat pipe was at nitrogen temperature (139°R) within
ten minutes from the start of the filling process. This
indicated that the wick was operating well and being
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26
saturated with liquid nitrogen. The magnitude of the mass
that was charged into the heat pipe was 0.1 pounds and
corresponds to a pressure drop from 100 psi to 36 psi in
the calibrated plenum bottle (Figure 5) • This amount was
chosen as sufficient to saturate the wick based on the
calculation from
(21)
This amount of mass was kept constant throughout the
experiment.
The experimental results consisted of the axial tern-
perature distribution for various power level input to
the evaporator section and for different angles of inclina-
tions. The operating pressure for every case was also
measured. For the horizontal level the power input was
varied from 5 to 110 watts and the steady state operating
pressure and temperature distribution for each of these
runs are reported in Table A-I and Figures 8 and 9. The
angle of inclination of the heat pipe while the condenser
was below the evaporator was fixed at 1.0 and 1.75 degrees
and while the condenser was above the evaporator at 5.25
degrees, to study the effects of inclination angles on the
heat pipe operating characteristics. In each case the
steady state operating pressure and the axial temperature
distribution were recorded for various power level input
to the evaporator section. The results of these runs are
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I 901- I
,-...
7+ I
0::
I 0 .......,
1-<l I UJ u
I z UJ 0:: 50 UJ
I u. u. -t:l I UJ 0:: :::J I 1- 30 <( 0::
I UJ Q.. 0 6T total ~ UJ b. 6T condenser 1-
10J 0 !J.T evaporator
I -- 6 T copper rod
10 30 50 70 :JU J.J.V 1\)
~
POWER INPUT (WATTS) FIGURE 8. Experimental Resultsfor the Case of Horizontal Level
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-Lt: 0 -
UJ Lt: :J t< Lt: UJ a.. ~ UJ t-UJ u < u.. Lt: :J (/)
240 . EVAPORATOR'
230. SECTION ADIABATIC SECTION
50 WATTS
30 WATTS
20 WATTS 10 WATTS
DISTANCE FROM EVAPORATOR END (INCHES)
CONDENSER SECTION
FIGURE 9. Steady State Temperature Distribution, Horizontal Level
"' 00
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presented in Tables.II,III, IV and Figures 10, 11, 12
and 13.
29
In each experimental run the effective thermal con
ductivity of the heat pipe, based on a solid rod area of
0.75 inches in diameter equivalent to the diameter of
the heat pipe and a length of 17.25 inches equivalent to
the distance between thermocouples number 1 and 6, was
calculated using the Fourier conduction law.
(22)
where ~x and ~T are the spacing and the temperature differ
ence between thermocouples 1 and 6 respectively, and q is
the heat flux through the heat pipe. This effective thermal
conductivity was then compared with that of copper [12] at
the same average temperature.
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-. 0:::
0 -I-<J
UJ u z UJ 0::: UJ u. u. -Q
UJ 0::: :J I-<( 0::: UJ ~ ~ UJ I-
40 I
I 30 I
I 20 I
I I 0 ~T total
10 6. ~T evaporator
I 0 ~T condenser
-- ~ T copper rod
15 20 25 30 POWER INPUT (WATTS)
FIGURE 10. Experimental Results for 1° Degree Inclination Angle with Evaporator Above Condenser
w 0
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,..... 0::
0 -1-<I
LLl u z LLl 0:: LLl LL LL -Q
LLl 0:: ::> 1-<( 0:: LLl c.. ~ LLl 1-
50~--------------------------------~
40
30
201 l/ 0 fiT total
-- fiT copper rod
10 5 10 . 15 20 25 30 35
POWER INPUT (WATTS)
FIGURE 11. Experimental Results for 1.75° Degrees Inclination Angle With Evaporator Above Condenser
w ~
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-. ~
0
"""' 1-<J
LU u z LU ~ LU u.. u.. ..... Q
LU ~ ::l 1-<C ~ LU a.. ~ LU 1-
100....----------------------~
80
60
40
I
20 I-/ I
I
I I I
I I
I I
I
t.T condenser
0 t.T total
-- t.T copper rod -_L
4 60--· 80 100 120
POWER INPUT (WATTS) FIGURE 12. Experimental Results for 5.25° Degree Inclination Angle With
Condenser Above Evaporator
w ~
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-0:: 0 .........
UJ 0:: :::J .... < 0:: UJ 0.. ~ UJ .... UJ u < u.. 0::: :::J en
. 230 ..
EVAPORATOR 210 ~sECTION
190
170
160
150
5
ADIABATIC SECTION
10 15 20
DISTANCE FROM EVAPORATOR END (INCHES)
CONDENSER SECTION
25
FIGURE 13. Steady State Temperature Distribution for 5.25 Degrees Inclination Angle with Condenser Above Evaporator
w w
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34
V. RESULTS AND DISCUSSION
The results of experimental measurements for the
nitrogen heat pipe, summarized in Appendix A, indicate
that it is possible to achieve relatively high effective
thermal conductivity in the operating temperature range.
For example, for horizontal operating conditions at low
heat fluxes (20 watts) , the effective thermal conductivity
of the heat pipe was eight times that of copper at the
same average temperature (Figure 14). The effective
thermal conductivity decreased with an increase in heat
flux, and with an increase in the angle of inclination with
evaporator above the condenser section.
For the case of horizontal operating condition the
axial temperature drop between evaporator and condenser
and the radial temperature drop in each of these sections
is presented in Figure B. The axial temperature distri
bution along the surface of the heat pipe for different
heat fluxes is presented in Figure 9. It is evident from
these figures that the temperature drop across the length
of the adiabatic section is relatively small. It is also
apparent from the plot of the radial temperature drop in
condenser and evaporator sections, that the thermal resis
tance across the condenser is smaller than that across the
evaporator for small heat fluxes and this trend reverses
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a: 0
I 1-u.. I
0:: :c
......... :::J 1-~
>-1--> -1-u :::J Cl z 0 u _J
<C ::E: 0:::: LIJ ::J: 1-
LIJ > -1-<..J LIJ LL LL LIJ
35
7000 r----.=1-----,---::::----:--------0 condenser raised s.zso
5000
Lmoo
3000
2000
600
500
400
300
0 horizontal operation
0 evaporator raised 1°
~evaporator raised 1.7so - copper rod
80 AVERAeE TEMPERATURE ( 0 R)
FIGURE 14. Effective Thermal Conductivity for the Heat Pipe at Different Angles of Inclination
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36
for heat fluxes higher than 95 watts. This behavior can
be physically explained from the fact that the boiling
phenomenon, nucleate boiling, taking place in the evaporator
at low heat fluxes has a relatively higher heat transfer
coefficient than the condensation phenomenon taking place in
the condenser. At higher heat fluxes, however, when the
radial temperature drop in the evaporator is high, the
boiling phenomenon could change from nucleate to film
boiling resulting in a rapid decrease of the heat transfer
coefficient. Under these conditions bubbles and vapor
film becomes trapped between the wick and the inner surface
of the tube and prevent the liquid from coming in contact
with the warm surface. The exact temperature difference
at which the boiling phenomenon changes from nucleate to
film in the present geometry cannot be specified due to the
presence of the wick and the lack of experimental results
on this type of a problem. Also at this stage the increase
in the temperature difference between evaporator and con
denser i~ mainly due to the temperature rise of the evapora
tor section while the temperature in the condenser section
approaches an asymptotic value.
As discussed in' section IV, Chi developed a mathe
matical model for predicting the total temperature drop,
between evaporator and condenser, as a function of heat
flux. This model assumes that the total temperature drop
is due mainly to radial temperature drop across the
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37
evaporator and condenser sections, meaning isothermal
adiabatic section. This model was used to predict the
radial temperature drop obtained experimentally across
the evaporator and the condenser. Equations (15) and
(17) were used for the condenser section. The thickness
appearing in equation (17) was taken as 0.045 inches equiva-
lent to the thickness of the wick. It was found that in
order to use this model and predict accurately the experi
mental temperature drop, the thermal conductivity must be
taken as a variable as shown in Figure 15. It was inter
esting to note that the required value was higher than the
thermal conductivity of liquid nitrogen (0.0795 Btu/hr0 R
Ft) which implies that the radial heat transfer mechanism
was at least in part due to direct condensation on the
tube wall, instead of conduction through the liquid nitro
gen layer. The upper half of the tube in the condenser
section could be exposed to direct condensation. The
gravity forces could have forced the liquid film to drop
to the lower half forming a pool, while leaving the upper
half partially exposed. This feature could explain the
reason for the high thermal conductivity obtained from the
above analysis.
To predict the radial temperature drop in the evapor
ator section, equations (14) and (16} were used. The
thickness appearing in equation (16} was taken as 0.045
inches, equivalent to the thickness of the wick. It was
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.,._ LL
I 0::::
0 I
0::: :J:: .......... => .,._ J:O
> 1-..... > ..... 1-u :::> Q z 0 u _. < ~ 0::: LIJ :::J: 1-
10 evaporator 0.81- 0 6 condenser
I I
0.7
0.6
0.5
0.4
0.3
0.2
0.1
.0
POWER INPUT (WATTS)
FIGURE 15. Effective Thermal Conductivity of Saturated Wick in Evaporator and Condenser Sections Based on Chi's
Analysis
w aa
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39
found that in order to predict accurately the experimental
results the thermal conductivity must be varied as a func
tion of heat flux as shown in Figure 15. This variation
of thermal conductivity confirms what was discussed earlier
regarding the changes in the boiling phenomenon which
takes place in the evaporator section. At high heat
fluxes the thermal conductivity decreases due to the vapor
bubbles trapped in and below the wick. As can be seen from
Figure 15, at high heat fluxes the thermal conductivity re
quired to predict the experimental results,while using
this model, in the evaporator section becomes lower than
the values at the condenser section due to this type of
boiling mechanism.
The effect of inclination angle, with reference to a
horizontal level, on t~e operating characteristics of the
heat pipe was experimentally measured for the case of 1
and 1.75 degrees with evaporator above the condenser and
5.25 degrees with evaporator below condenser. The results
are shown in Figure 10, 11, 12 and 13, respectively. It is
evident that when the evaporator is above the condenser,
gravity forces limits the capability of the heat pipe and
reduces its effective thermal conductivity. These cases
represent an example of a wick limiting heat pipe where
the capillary forces are not sufficient to overcome both
the gravity and viscous forces. The reverse is true for
the case of evaporator below the condenser where the
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40
gravity forces add to the capillary forces. The effective
thermal conductivity for different angles of inclinations
is shown in Figure 14. It reaches a high value of approxi
mately 20 times that of copper at low heat fluxes for the
case of 5.25 degrees, evaporator below condenser, and low
value of 1.25 that of copper for the case of 1.75 degrees
with evaporator above condenser.
The experimental results are compared with the results
of Haskin in Figure 16. The present total temperature dif
ference is smaller than that of Haskin's at the same heat
flux input. In comparing the two experimental results for
the horizontal case a difference of approximately 55 degrees
exist for the same heat flux with the net result of higher
effective thermal conductivity for the present experimental
heat pipe. This difference could be due to differences in
design, type of wick material used and the different loca
tions of the thermocouples. No attempt has been made to
optimize such a heat pipe in either study. A small error
in the horizontal level of the heat pipe could contribute
to some of these differences. The general trend of the
results, however, agree favorably.
The maximum heat transfer capability of this nitrogen
heat pipe w~s predicted from equation (9), using the proper
ties of liquid and gaseous nitrogen, Appendix B [12,13],
which was approximately 500 watts. The maximum heat trans
ferred during the experiment was only one fifth of the
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,...... 0:::
0 .........
!;:} UJ
f5 0::: UJ u.. u.. ..... Q
UJ 0:::
~ UJ 0...
ffi 1-
•
Haskin 6T evaporator high by 1°
() Has~in 6T evaporator low by 1° ~ Haskin 6T horizontal operation
• present exp. evaporator high by 1° 41 present exp. horizontal operation
~ present exp. condenser high by 5.25°
I'OtVER INPUT (WATTS)
FIGURE 16. Comparison of the Total Experimental Temperature Difference
~ 1-'
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42
predicted value. The radial thermal resistance at the
evaporator and condenser sections, which is not taken into
consideration in deriving equation (9) are probably the
cause of this difference.
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43
VI. CONCLUSIONS AND RECOMMENDATIONS
The experimental heat pipe designed with the moda
crylic fiber wick proved to operate satisfactorily in the
cryogenic temperature range. The adiabatic section for
the case of horizontal operating conditions was operating
approximately isothermal for heat fluxes below 20 watts
and with approximately 8 degrees drop for a heat flux of
100 watts. The effective thermal conductivity was approxi
mately 8 times that of copper for the horizontal case, and
increases to 20 times that of copper for the case of 5.25
inclination angle with evaporator below condenser. The
major resistance for heat flow at high heat fluxes appears
to be the radial thermal resistance at the evaporator sec
tion where bubbles and vapor film could become trapped in
and under the wick structure.
As the angle of inclination increases with evaporator
above condenser the effective thermal conductivity de
creases and the maximum heat flux capability of the heat
pipe decreases. The reverse of the above is true for the
case of evaporator below condenser. In comparing the
present results with that of Haskin's, it appears that the
present effective thermal conductivity is higher for the
same heat flux loads. This difference is probably due to
differences in design, wick material and locations of
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44
thermocouples. A small error in the angle of inclination
could contribute to some of this difference between the
two experimental results.
To improve the experimental set up and increase the
heat pipe maximum heat flux capability the present heat
pipe design should be modified to include highly conduc
tive paths such as metal fingers in the evaporator section.
These paths would decrease the radial thermal resistance
in this section. Also, a copper sleeve should be installed
over the evaporator section to make it isothermal. An
additional pressure gauge should be installed at the
evaporator end to detect pressure difference between
evaporator and condenser.
Research is needed in the area of measuring effective
thermal resistance for conditions similar to the ones
existing in the evaporator and condenser sections, caused
by boiling and condensation on the interior area of a
cylindrical surface covered by a wick. This type of infor
mation is essential, and not presently available, for
analytically predicting the operating characteristics of
the heat pipe.
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1.
2.
3.
4.
5.
6.
7.
8.
9.
10.
11.
45
BIBLIOGRAPHY
Eastman, G.Y., "The Heat Pipe", Scientific American, May 1968, p. 38-46.
Grover, G.M., Cotter, T.P. and Erickson, G.F., "Structures of Very High Thermal Conductance", Journal Applied Physics, Vol. 35, 1964.
Coyle, E.C. and Gyer, W.T. "The Development of Large Capacity Heat Pipe", ACF Industries, Inc., St. Charles, (work under progress - report unpublished), Personal Communication, December 1969.
Haskin, W.L. "Cryogenic Heat Pipe", Air Force Flight Dynamics Laboratory, Technical Report AFFDL-TR~ 66-228, June 1967.
Philip, E.C. and Aleck, s.w., "Development of Cryogenic Heat Pipes", ASME Paper No. 70-WA-ENER-1, Dec. 1970.
Chi, s.w. and Cygnavowicz, T.A. "Theoretical Analysis of Cryogenic Heat Pipes", ASME Paper No. 70 HT/SPT-6, June 1970.
Joy, Patrick, "Optimum Cryogenic Heat-Pipe Design", ASME Paper No. 70 HT/SPT-7, June 1970.
Cotter, T.P., "Theory of Heat Pipes", Los Alamos Scientific Laboratory, LA-3246-MS, March 1965.
Katzoff, s., "Notes on Heat Pipes and Vapor Chambers and Their Application to Thermal Control of Spacecraft", SC-M-66-623, October 1966.
Dzakowic, G.S., Tang, Y.S. and Arcella, F.G., "Experimental Study of Vapor Velocity Limit in a Sodium Heat Pipe", ASME Paper No. 69-HT-21, August 1969.
Kemme, J.E. "High Performance Heat Pipes", Los Alamos Scientific Laboratory, Thermionic Conversion Specialist Conference, Palo Alto, Calif., October 1967.
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46
12. Scott, R.B., "Cryogenic Engineering", D. Van Nostrand Company, Inc., Princeton, New Jersey, 1966, p. 277-286.
13. Johnson, V.J. "A Compendium of the Properties of Materials at Low Temperature (Phase I) - Part I -Properties of Fluids", United States Air Force, Wright Patterson Air Force Base, Ohio, Oct. 1960.
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47
VITA
Jay Dudheker was born on November 21, 1932, in
Patna, India, where he attended elementary and secondary
schools. He graduated from Oswal Jain High School in
June, 1949.
He entered the Seattle University in Spring of 1961
and received the Bachelor of Science degree in Mechanical
Engineering in May, 1965. The following fall he entered
the UMR Graduate Engineering Center of St. Louis.
On August 4, 1968, he married Georgia Leventogianni
of Patras, Greece. During the course of his research work,
he and his wife have been blessed with the birth of one
child, Maryann, on September 1, 1970.
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APPENDIX A
EXPERIMENTAL DATA
AND RESULTS
48
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Run Power Number Watts
1 10
2 20
3 25
4 30
5 40
6 50
7 60
8 70
9 80
10 90
11 100
12 110
TABLE A-I EXPERIMENTAL RESULTS
HEAT PIPE OPERATING IN HORIZONTAL POSITION
Evaporator Adiabatic Section Condenser
T 0 R 1 T 0 R 2 T OR
3 T 0 R 4 T 0 R 5 T 0 R 6
143.50 141.50 141.50 141.50 141.50 139.30
152 148.20 147.67 147.67 147.67 139.30
157.34 152 151.50 151.50 151 139.60
160.40 154.67 154.10 154.10 153.10 140
166.80 159.20 159.20 158.10 157.10 140
172.95 163.50 162.40 162.40 159.40 140
182.5Q 170 168 167 164.50 140
200.20 173.40 171.50 170.50 167.50 140.20
211.91 177.30 174.40 173.40 170.50 .140.20
221 180.10 177.30 176.40 174.40 140.30
223.95 183 180.91 178.20 176.40 140.40
242 186.70 183 181.51 179.10 140.40
-
Operating Pressure
PSIA
16.70
27.70
31.70
37.20
47.70
58.70
74.70
90.70
104.70
118.70
137.70
147.70
.. \D
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TABLE A-II EXPERIMENTAL RESULTS
HEAT PIPE OPERATING AT 1 DEGREE ANGLE, EVAPORATOR ABOVE CONDENSER
Run Power Evaporator Adiabatic Section ~ondenser Operating Number Watts T 0 R T 0 R T 0 R T 0 R T 0 R T 0 R
Pressure 1 2 3 4 5 6 PSIA
13 5 142 141 141 141 141 139.30 14.70 14 10 146.30 144 144 144 144 140 16.70 15 20 179.20 149.67 148.74 148.74 147.6~ 140 26.70
TABLE A-III EXPERIMENTAL RESULTS
HEAT PIPE OPERATING AT 1.75 DEGREE ANGLE, EVAPORATOR ABOVE CONDENSER
. Run Power Evaporatoi Adiabatic Section Condenser Operating
Pressure Number Watts T 0 R IT. 0 R T 0 R T 0 R T 0 R T 0 R PSIA 1 2 3 4 5 6
16 5 155.20 141.50 1~1.50 141.50 Ull.50 139.30 14.70
17 10 187.62 144 143 143 143 140 14.70 ----- -~---~ --·-
' I
I I
I
U1 0
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Run Number
18
19
20
21
22
23
24
25
TABLE A-IV EXPERIMENTAL RESULTS
HEAT PIPE OPERATING AT 5.75- DEGREE ANGLE, CONDENSER ABOVE EVAPORATOR
Power Evaporator Adiabatic Section ~ondenser Operating Watts T 0 R T 0 R T 0 R T 0 R T 0 R T 0 R
Pressure PSIA 1 . 2 3 4 5 6
5 140.50 140.50 140.50 140.50 140.50 . 139.30 14.70
10 142 141 141 141 141 139.30 14.70
20 145.34 142 142 142 142 139.30 14.70
30 151 143 143 143 143 139.30 20.70
50 163.40 153.10 153.10 153.10 153.10 140.50 34.70
70 174.40 159.40 159.40 158.34 157.80 141 46.70
90 204 163.54 162.20 161.40 161.40 140.90 59.70
120 228.34 170.50 169 167 .. 50 167.50 141.50 69.30 ---- -
'
U1 .....
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52
TABLE A-V DEDUCED EXPERIMENTAL RESULTS
Run Vapor 6T(Total) 0 R 6T (Cond.) 0 R 6T (Evap.) 0 R Number Temperature T1 - T6 T - T T - T
OR v 6 1 v
1 141.31 4.20 2.01 2.19
2 149.83 12.70 10.53 2.18
3 152.20 17.74 12.60 5.14
4 155.20 20.40 15.20 5.20
5 159.80 26.80 19.80 7.00
6 164.25 32.95 24.25 8. 70·
7 169.75 42.50 29.75 12.75
8 174.40 60.00 34.20 25.80
9 178.00 71.71 37.80 33.91
10 181.30 80.70 41.00 39.70
11 185.20 93.55 44.80 48.75
12 187.00 101.60 46.60 55.00
13 139.30 2.70 0.0 2.70
14 141.31 6.30 1.31 4.99
15 149.50 39.20 9.50 29.70
16 139.30 15.90 0.0 15.90
17 139.30 47.62 0.70 48.32
18 139.30 1.20 0.0 1.20
19 139.30 2.70 o.o 2.70
20 139.30 6.04 0.0 6.07
21 145.20 11.70 5.90 5.80
22 154.00 22.90 13.50 9.40
23 159.50 33.40 18.50 14.90
24 164.75 63.10 23.85 39.25
25 168.00 86.84 26.50 60.34
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APPENDIX B
THERMOPHYSICAL PROPERTIES
OF
NITROGEN
53
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c U\
....-!
c:: C
'J ....-!
c:: C
'· ....-!
c oc
,-... <t: -cr.. a.
-· UJ
o:-::J U
:· u: u
: o:-a.
54
s:: <U O
l 0 J.t ~
..... z ~
0 <U J.t ::s U
l U
l <U J.t tl.t
J.t 0 ~
"' :> . ...... I IXl
fi1 :::> t.!) H
I'Ll
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55
50
-~'of'. .... u.
40 .......... ~ c:Q ....I -> .... -cr.
35 :z: L1.l Q
30
?.5
1. 1::>. 17S TEMPERATURE {0 R)
FIGURE B-2. Density of Saturated Liquid Nitrogen
55
215
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-r-r-. 1-u..
' ::E: c:tl _J --> 1--en :z: u.: ~
56
20.~----------------------------------------~
10.
1.0
.lf)
FIGURE B-3. Density of Gaseous Nitrogen (Saturated Vapor)
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57
0.6~------------------------------------~
0.5 LC'.
Co ....-\
X
-.. n .1~ ('.J 1-u..
.......... u ~ u:. I u.. ~ 0.3 .....1 -->-1--(/)
0 u UJ f1.? -> u ..... ~-
< :z >-~ 0.1
120 1~0 ?00 TEMPERATURE ( 0 R)
FIGURE B-4. Dynamic Viscosity of Liquid Nitrogen
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.......... c r-1
X -('.; 1-LL
........ u UJ U)
I LL JXI ...J -->· l-.... en 0 u en -> u -~ <( z >-Cl
L~
1
I 1~n
I I I 34()
I I ?.?n !nn )f)() 3~'1 42n 60
TEMPERATURE (0 R)
FIGURE B-5. Dynamic Viscosity of Gaseous Nitrogen at Atmospheric Pressure
U1 co
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59
100~----------------------------------,
BO
-~ Jll 60 ..J ........
~ ..... JXl -z: c -..... LH1 <( 1'-l -0:: 0 0... < > LL. ~n 0
..... < LU :X:
120 140 lf'O 1~0 ?.?.n
FIGURE B-6. Heat of Vaporization of Nitrogen
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L.r. c: r-f
X -..... u.. ........ u.. Q:l _J -z: 0 -en z LLl ..... UJ u <( u.. 0::: ::::> cr.
68
64
no
56
52
l.JR
llll
1 ... 0 TEMPERATURE (0 R)
FIGURE B-7. Surface Tension of Saturated Liquid Nitrogen
60
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-c::: 0
I X: ~ ..J ........ ~ I-~ -I-<( w :r: u -u.. -u UJ a.. (Jj
n.65
O.f'0
n.ss
n.sn
16'1
TEMPERATURE (0 R)
FIGURE B-8. Specific Heat of Saturated Liquid Nitrogen
61
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......... u.
0 I
1--u. I
0::: :r: ....... => 1--al _..
>-1---· > -1--u => 0 z 0 u _J <( ~ c.::: U.J :t: 1--
0.09
1),08
0.07
0.~6
(),05
0 I Q!J
0.03
100
CRITICAL TEMPERATURE (227.4)
FIGURE B-9. Thermal Conductivity of Saturated Liquid Nitrogen
62
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.::so r-1
X
-. 0::: o, t-u. I
0::: :I:
........ ::::> t-.xi -> t--> -. t-u ::;) Q z 0 u ..J <( ~ 0::: w :I: t-
gn
RO
70
IRn ?.on ??.n ;.z~n ~~n ?.RO 300 320 TEMPERATURE (0 R)
FIGURE B-10. Thermal Conductivity of Gaseous Nitrogen at Atmospheric Pressure
a\ w