thesis 0510044,0510064,0510091 heat pipe 15.02.11
DESCRIPTION
heat transfer characteristics of closed loop heat pipeTRANSCRIPT
CHAPTER 1
INTRODUCTION
1.1 MOTIVATION
Meandering tube pulsating heat pipes, (PHPs) have already been found some
applications in micro-processor and power electronics applications owing to favorable
operational characteristics coupled with relatively cheaper costs. Although grouped as
a subclass of the overall family of heat pipes, the subtle complexity of thermo-fluidic
transport phenomena is quite unique justifying the need of a completely different
research outlook. Comprehensive theory of operation and reliable database or tools
for the design of PHPs according to a given micro-electronics-cooling requirement is
still an unrealized task. A closed loop pulsating or oscillating heat pipe consists of a
metallic tube of capillary dimensions wound in a serpentine manner and joined end to
end. It is first evacuated and then filled partially with a working fluid, which
distributes itself naturally in the form of liquid–vapor slugs and bubbles inside the
capillary tube. Respective tube sections thus have a different volumetric phase
distribution. One end of this tube bundle receives heat transferring it to the other end
by a pulsating action of the liquid–vapor slug-bubble system. There may exist an
optional adiabatic zone in between. This type of heat pipe is essentially a non-
equilibrium heat transfer device. The performance success primarily depends on
continuous maintenance or sustenance of these non-equilibrium conditions in the
system. The liquid and vapor slug/bubble transport is caused by the pressure
pulsations inside the device. Since these pressure pulsations are fully thermally
driven, because of the inherent construction of the device, there is no external
mechanical power source required for the fluid transport.
In a working PHP, there exist temperature gradients between the evaporator and the
condenser section. These are coupled with inherent real-time perturbations, due to:
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local non-uniform heating and cooling within the evaporator and condenser tube
sections,
unsymmetrical liquid–vapor distributions causing uneven void fraction in the
tubes, and,
presence of approximately triangular or saw-tooth alternating component of
pressure drop superimposed on the average pressure gradient in a capillary slug
flow due to the presence of vapor bubbles.
The net effect of all these temperature gradients and perturbations is to create a non-
equilibrium pressure condition which, in conjunction with the non-uniform void
fraction distribution in respective tubes, as stated earlier, is the primary driving force
for thermo fluidic transport. Thus self-sustained thermally driven oscillations are
obtained.
1.2 OBJECTIVES
The main objectives of this experiment are
To understand the different operational regimes (evaporator, adiabatic and
condenser section) of closed loop pulsating heat pipe.
To compare the thermal resistance of heat pipe for different orientation
of heat pipe
To compare the thermal resistance at different filling ratio
To study the evaporator and condenser temperature of heat pipe at different
orientation
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CHAPTER 2
LITERATURE REVIEW
2.1 Historical Development
Historically, the first application of gravity heat pipes was in boilers and bakeries, the
so-called Perkins tube was widely used in the 19th century. This is a bare, thick-
walled carbon steel tube filled with a certain amount of water, about 1/3 of the total
tube volume, and hermetically sealed. The lower tube end was heated by flue gases,
the upper end extended into the boiler where it was used to generate steam.
In 1938 a patent was granted in the USA which describes a tube incorporating
capillary grooves to aid liquid distribution and hence vaporization in boilers. The
first patent of a heat pipe employing a capillary wick for pumping liquid against
gravity was applied by Gaugler in 1944 as a two-phase heat transport device for
refrigerators. It was supposed to allow movement of the working fluid without
pumps and without natural convection, by utilization of the capillary force generated
by a capillary wick [16].
The heat pipe concept was first put forward by R.S.Gaugler of the general Motors
Corporation, Ohio, and USA. In a patent application dated December 21st, 1942, and
published as US Patent No. 2350348/ on June 6th, 1944, the heat pipe is described as
applied to a refrigeration system.
According to Gaugler, the object of the invention was to "cause absorption of heat, or
in other words, the evaporation of the liquid to a point above the place where the
condensation or the giving off of heat takes place without expending upon the liquid
any addition work to lift the liquid to an elevation above the point at which
condensation takes place". A capillary structure was proposed as the means for
returning the liquid from the condenser to the evaporator, and Gaugler suggested that
one form this structure might take would be a sintered iron wick. It is interesting to
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note the comparative small portion of the tube cross section allocated to vapor flow
in his designs.
The first heat pipe that Grover built used water as the working fluid and was
followed shortly by a liquid sodium heat pipe for operation at 1100 K. Both the high
temperature and ambient temperature regime soon explored by many workers in the
field. It was until 1966 that the first cryogenic heat pipe was developed by Haskin of
the Air Force Flight Dynamic Laboratory at Wright- Patterson Air Force Base.
Between 1964 and 1966, RCA (Radio Corporation of America) was the first
corporation to undertake research and development of heat pipes for the commercial
application. The concept of Variable conductance or Temperature Controlled Heat
pipe was first described by Hall of RCA in a patent application dated October 1964.
However, although the effect of a non-conducting gas was shown in Grover’s
original publication, its significance for achieving variable conductance was
immediate recognized. In subsequent years the theory and technology of variable
conductance Heat Pipes was greatly advanced, notably by Bienert and Brennan at
Dynatherm and Marcus at TRW. On April 5, 1967, the first “zero g” demonstration
of a heat pipe was conducted by a group of engineers of the Los Alamos scientific
Laboratory. This first successful flight experiment overcame the initial hesitation that
many spacecraft had for using this new technology to solve ever- present temperature
control problems on spacecraft. Subsequently, more and more spacecrafts have relied
on heat pipes either to control the temperature of individual components or of the
entire structure. Some examples of this trend were the ARS- E, OAO, ATS Fand G
spacecrafts and the sky labs.
The development of terrestrial applications of heat pipes progresses at much slower
pace. In 1968, RCA developed a heat pipe heat sink for transistors used in aircraft
transmitters. This probably represented the first commercial application of heat
pipes.
Publications in 1967 and 1968 by Feldman, Eastman, and Katzoff first discussed
applications of heat pipes to areas outside of government concern and that did not
fall under the high temperature classification such as: air conditioning, engine
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cooling and electronics cooling. These papers also made the first mentions of
flexible, arterial, and flat plate heat pipes. 1969publications introduced the concepts
of rotational heat pipe with its application to turbine blade cooling and the first
discussions of heat pipe applications to cryogenic processes.
2.2 The Revolution
Starting in the 1980s Sony began incorporating heat pipes into cooling schemes for
some of its commercial electronic products in place of both forced convection and
passive finned heat sinks. Initially they were used in tuners and amplifiers, soon
spreading to other heat flux electronics applications. During the late 1990s
increasingly hot microcomputer CPUs spurred a threefold increasing in the number
of U.S. heat pipe patent applications. As heat pipes transferred from specialized
industrial heat transfer component to a consumer commodity most development and
production moved from the U.S. to Asia. Modern CPU heat pipes are typically made
from copper and use water as the working fluid.
A wickless network heat pipe for high heat flux spreading applications was
developed by Cao, Y. and Gao, M. In this study the concept of the network heat pipe
employing the boiling heat-transfer mechanism in narrow space has been described.
Two flat-plate wickless network heat pipes (or thermal spreaders) were designed
fabricated and tested based on this concept by the authors. The fabricated thermal
spreaders, which were made of Copper or Aluminum, were wickless, cross-grooved
heat transfer devices that spread a concentrated heat source to a much larger surface
area. As a result, the high heat flux generated in the concentrated heat source could
be dissipated through a finned surface by air cooling. The network heat pipes were
tested under different working conditions and orientations relative to the gravity
vector, with water and methanol as the working fluids. The maximum heat flux is
achieved was about 40 W/cm2 for methanol and 110W/cm2 for water with a total heat
input of 393W.
A heat transfer analysis of an inclined two-phase closed thermosyphon was
developed by Zuo, Z. J. The inclination–induced circumferential flow was
unfavorable with respect to dry out because the thin top-side liquid film was easier to
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boil off, but contrastingly was favorable with respect to flooding because the thick
underside film corresponded to a large gravity force. Minimum working fluid
inventory remained almost constant for a large range of inclination angles (0-70 deg)
and then significantly increased for further increase of inclination angle. At a certain
inclination angle, the mean heat transfer co-efficient of the thermosyphon reached a
maximum value, which was related to the heat transfer behavior in both condenser
and evaporator. The highest flooding limit was at inclination angle ranging from 30
to 45 deg, which corresponded to the best balance of the two opposing effects:
secondary circumferential flow and gravity reduction.
Zhang, J. [3] and Wong, H. studied heat transfer and fluid flow in an idealized micro
heat pipe with the support of NASA and LaSPACE. They made an analysis for four
different values of length to width ratio of an idealized micro heat pipe, viz. 20, 50,
100 and 200. From the liquid temperature distribution along the length of the micro
heat pipe, they found that the temperature profile is relatively flat except the region
near the evaporator, and for a micro heat pipe with larger length to width ratio, the
length of the evaporator is shorter. From the vapor pressure distribution, they found
that the pressure goes approximately linearly and is not strongly affected by the
length to width ration. On evaluating the effective thermal conductivity of a micro
heat pipe increases with increase in the evaporation area at the evaporator, and length
or width of the micro heat pipe. They also added that a fluid with larger latent heart
would produce larger effective thermal conductivity.
In a study of micro and miniature heat pipes, developed by A.R. Anand, attempts
have been made to develop a one dimensional numerical model of micro heat pipes,
taking into account the effect of liquid vapor interfacial shear stress. Also governing
equations for conservation of mass, momentum and energy have been developed,
based on control volume to study the performance characteristics and validate the
experimental results. To identify and understand better the phenomena, which
governs the performance limitations and operating characteristics of micro heat
pipes, Babin et al. conducted a combined experimental and analytical investigation
on two micro heat pipes, one made of Copper and one of Silver, of length 57 mm and
cross section 1mm2 approximately with water as the working fluid. The steady-state
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experimental results obtained were compared with an analytical model and were
found to predict accurately the experimentally determined maximum heat transport
capacity for an operating temperature range of 400C to 600C. The results indicated
that the steady-state model could be used to predict accurately the level of
performance. In 1991, Wu and Peterson developed a transient numerical model
capable of predicting the thermal behavior of micro heat pipes during start up or
vibration in the thermal load of evaporator. The numerical model was used to
identify, evaluate and better understand the phenomena, which governs the transient
behavior of micro heat pipes as function of physical shape, the properties of the
working fluid, and the principal dimensions. The results were compared with the
steady state results obtained by Babin et al. In 1990 and were shown to accurately
predict the steady state dry out limit also. The wetting angle was also found to be one
of the most important factors contributing to the heat transport capacity. But no
experimental data were obtained on the transient operational characteristics.
In 1994, Faghri et al developed their mathematical model to examine the heat and
mass transfer processes in a micro heat pipe, taking into account the variation of the
curvature of the free liquid surface and the interfacial shear stress due to liquid-vapor
interaction. The model described the distribution of the liquid chart in micro heat
pipe and its thermal characteristics depending on the charge of the working fluid and
the heat load. It was observed from the modeling that for the same heat pipe, the
charge required when interfacial shear stress is considered, is greater than the charge
required if no shear is considered. Further for the same operating temperature the
maximum heat transfer, when interfacial shear stress is considered, is less than the
maximum heat transfer if no shear is considered.
2.3 Recent Researches
Pulsating heat pipes (PHP) are passive two phase thermal control devices first
introduced by Akachi et al. [1]. Mainly, PHPs consist of a capillary tube bent in
several curves to form parallel passages. In this application, reduced diameter
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channels are used, which are directly influenced by the selected working fluid. The
vapor plugs generated by the evaporation of the working fluid push the liquid slugs
toward the condensation section and this motion causes flow oscillations that guide
the device operation [2]. There are several applications for PHPs from electronics
and structural thermal control as well as microgravity thermal control. Due to simple
construction, light weight and low cost, PHPs have gained attention in a lot of
aspects to give isothermal characteristic for his component.
There are two possible configurations for PHPs, open loop and closed loop. In the
open loop configuration one end of the tube is pinched off or welded and the other
end may present a service valve for evacuation and charging. The closed loop
configuration has both ends connected and allowed the flow to be circulated.
Considering the sections of a PHP, it presents evaporation and a condensation section
and may also present an adiabatic section. The tube does not have a wick structure
and the construction is very simple. As any other two phase passive thermal control
device, heat is acquired from the source through the evaporation section transferring
to the working fluid and where the slug/plug pumping action will be generated. The
fluid then flows by the adiabatic section towards the condensation section. On a
closed loop configuration, the fluid is allowed to circulate and after being condensed,
the fluid returns to the evaporation section to complete the loop. On the open loop
configuration the liquid circulation is not possible.
Previous investigations have already addressed the operation and thermal behavior of
PHPs. Delil et al. [4] presented a survey on pulsating/oscillating devices suitable to
be used in microgravity and super gravity environments. Important contributions
related to the PHPs on closed loop configuration were given by Charoensawan et al.
[5], Khandekar et al. [6, 7], Rettidech et al. [15] and Tong et al. [9]. The critical
issues and an approximate operation behavior of PHPs have been already determined
[11]. As a relatively new field, most of the theory involved on PHPs design and
operation were derived from the classic two phase flow theory, which could be used
as a first approach in analyzing such a device.
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The PHP operation presents some unique characteristics and a very interesting
thermal behavior. One particularly of PHP operation is that it presents
thermodynamics instabilities associated with the plug/slug dynamics, even though
such a dynamics is in mechanical equilibrium. The vapor plugs formation and
collapse presents a chaotic behavior that is difficult to model. During its operation,
metastable conditions of the working fluid are present, which are directly related to
the thermo hydrodynamics particularities of this device [5-7]. Quasistationary
modeling has showed great potential in predicting PHPs operation, which was in
accordance to experimental results [12]. The slug flow dynamics is dependent on the
applied power to the evaporation section, tilt angle and amount of non condensable
gases [13]. A mathematical model has addressed the operation of PHPs, where the
chaotic behavior can be reflected which were in accordance to experimental results
[14]. Other mathematical models have been formulated to describe the PHP
operation, considering the geometric parameters and the effect of working fluid [15]
as well as the heat transfer effects on its operation [2].
The pulsating action of heat pipe is directly influenced by the inner diameter of tube.
The parameters influencing the plug/slug formation in reduced diameters must be
observed for this application, such as the correct working fluid selection, surface
tension and shear stress effects, etc. Without the pumping action the heat pipe will
operate as a solid bar conducting heat from one end to another end.
Another factor that influences the PHP performance is the number of turns. The
increase of number of turns improves the performance [5] and this higher heat fluxes
could be dissipated. For the proper working fluid the Clausius – Claperon relation
(dP/dT) Tsat = i lv/ (Tsat V iv)
In here higher magnitude of (dP/dT)Tsat is desirable. A comparison of this parameter
related to several working fluids was presented by Khandekar et al. [7]. This
represents a small change in saturation temperature will result in a large influence in
the saturation pressure which will affect the pumping force of PHP. Other parameters
are latent heat of vaporization and surface tension.
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CHAPTER 3EXPERIMENTAL METHOD
In order to study the heat transfer characteristics of pulsating heat pipe, an
experimental facility has been designed, fabricated and installed. The detailed
description of experimental apparatus and procedure are presented in the subsequent
sections of this chapter.
3.1 Experimental Apparatus
1. Pulsating Heat Pipe
2. Working Fluid (Water)
3. Test Stand
4. Heating Apparatus
a) Power Supply Unit
b) Ni-Cr Thermic Wire
c) Variac
5. Cooling Apparatus
a) Battery Fan
b) Adapter Circuit
6. Measuring Apparatus
a) Thermocouple K Type
b) Selector Switch
c) Digital Thermometer (Y Type)
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3.2 Pulsating Heat Pipe
A closed loop pulsating heat pipe or oscillating heat pipe consists of a
metallic tube of capillary dimensions wound in a serpentine manner and
joined end to end. It consists of 3 sections. They are:
Evaporator Section
Adiabatic Section
Condenser Section
Evaporator Section:
It is the section of the heat pipe where the refrigerant (water) absorbs heat and
evaporates. It is located on the bottom section of the heat pipe .The heat is
supplied on the heat pipe by Nicrome wire which is connected to the variac. As
the copper tube is good conductor of electricity so it is not directly connected
with Nicrome wire because it can make short circuit. So, the Nicrome wire is
surrounded to a Mica sheet and kept in a distance of Copper tube. Glass wool is
kept between the mica sheet and the Copper tube. So, the heat is transferred to
the Copper tube through the glass wool.
Condenser Section:
It is the section of heat pipe where heat is rejected from the working fluid on this
section the working fluid condenses and rejects the same amount of heat which
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it absorbed from the evaporator section. On this experiment it is located on upper
section of the heat pipe.
Adiabatic Section:
It is located between the evaporator section and condenser section. In here the
liquid and vapor phases of the fluid flow in opposite directions and no significant
heat transfer occurs between the fluid and surrounding medium.
Table 3.1: Experimental Parameter and Their Ranges
Parameters Condition
Inner diameter 2.2 mm
Outer diameter 2.3 mm
Total length 155 mm
Length of evaporator section 30mm
Length of adiabatic section 75mm
Length of condenser section 50mm
Material Copper
Air flow rate 3.5 m/s
3.3 Working Fluid
For choosing the right working fluid some properties have to be considered. They
are:
High value of (dP/dT) at saturation temperature
Low dynamic viscosity
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Low latent heat
High specific heat
Low surface tension
In this experiment water (H₂O) IS used as working fluid. The thermophysical
properties are
Table 3.2: Thermophysical Properties of Water
Freezing Temperature (t) 0⁰
boiling temperature 100⁰
Absolute pressure(p) 3.2 MPa
Density (ρ) 997 (kg/m3)
Specific volume (v) 1.00 *10-3 (m3/kg)
Specific Heat (cp) 4.181 (kJ/kgK )
Specific entropy (e) 0.367 (kJ/kgK)
Dynamic viscosity (μ) 0.890 centipoise
Kinematic viscosity (ν) 1.004
Expansion coefficient 0.257
Specific enthalpy 104.8
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3.4 Test Stand
The test stand is a wooden structure which holds the heat pipe .It has a base where an
aluminum box is located. It contains the evaporator section of the heat pipe .the
evaporator section is connected to Ni –Cr themic wire which is connected with the
variac . The test stand can be rotated and can be kept on any different orientation
between horizontal and vertical position. The whole structure is supported by two
columns which are situated in a large wooden base.
3.5. Heating Apparatus
Variac
Table 3.3: Variac specification
Type: 3ф
Rated capacity: 300 volt
Rated frequency: 60 Hz
Input voltage: 220 volt
Power Supply Unit
Type: AC
Voltage: up to 220 volt
Frequency: 50 Hz
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Ni-Cr Thermic wire
Diameter: 0.051 inch
Power: 120 W
3.6 Cooling Apparatus
Fan
For cooling the working fluid, forced convection is used. For forced convection a
D.C fan is used. It is located on test stand in front of the heat pipe.
Adapter Circuit
As the fan is a DC fan and the power supply is AC. So, for running the fan a
converter circuit is required to convert the AC current to DC current. It consists of a
transformer, full wave rectifier circuit to convert the AC to DC current.
3.7 Measuring Apparatus
Thermocouple
A number of thermocouple is glued to the wall of pulsating heat pipe. They are
calibrated and connected to different sections of heat pipe for measuring temperature.
This thermocouple is Ty (Chromel/Alumel) Type K. It is ‘general purpose'
thermocouple. It is low cost and, owing to its popularity, it is available in a wide
variety of probes. Thermocouples are available in the -200°C to +1200°C range.
Sensitivity is approx 41uV/°C. Use type K unless you have a good reason not to.
15
Selector Switch
Selector switch is a type of rotating connector. That can be rotate to different
positions to make contact with the particular position of the heat pipe through the
thermocouple. For this experiment three selector switches are used. Each of them has
8 points and used to measure the temperature of different points in heat pipe.
Digital Thermometer
Table 3.4: Specification of digital thermometer
AC voltage 2V 20V 200V 750V
DC current 2 mA, 20 mA, 200 mA
Resistance 200 Ώ -200M Ώ
Frequency 20khz
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3.8 Experimental setup
For studying the thermal characteristics of heat pipe, an experimental setup [fig 3.6]
has been built. It is made by a circular copper tube. Its outer diameter is 2.3 mm and
inner diameter is 2.2 mm. The total length of this pipe is 155 cm. The tubes are bent
to U shape. Two of the extreme bents are connected by a T connector. This forms the
closed loop heat pipe whether the open loop heat pipe has no connection between the
two extreme ends. For pulsating action of this heat pipe distinct liquid slug and
bubble formation are essential. This liquid slug and bubble formation are related to
the Eotvos number or Bond number. Eotvos number is the ratio of buoyancy force to
surface tension force.
Eo=∆ ρg L2
σ
Eo = Eotvos number.
Δρ = Change of density.
σ = Surface tension.
L – Characteristics length.
Selecting the perfect diameter
There is a critical diameter above which the heat pipe will not function as pulsating
heat pipe. This critical diameter is related to Eotvos number. For pulsating action the
Eotvos number has to be less than or equal to 4.For selecting the inner diameter of
heat pipe, the working diameter should be less than the critical diameter. In this
experiment the diameter of the heat pipe has been taken 2.2 mm which makes the
Eotvos number less than 4.
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Number of turns
The number of turns increases the level of perturbations inside the device. If the
number of turns is less than a critical value, then there is a possibility of a stop-over
phenomenon to occur. As the thermal performance of heat pipe is a function of
number of turns, so this has to be selected properly. On this experiment the number
of turns of heat pipe is ten (10) which is a optimum number for this diameter and
corresponding working fluid.
Selection of working fluid
Working fluid is very important part of heat pipe performance. In this experiment
water is used as the working fluid which posses all the characteristics of a good
working fluid such as high latent heat, high thermal conductivity, high specific heat
and high surface tension.
Filling process
In this experiment, water is selected as the working fluid. The filling process is done
by using the filling valve which is known as T connector [fig: 3.15]. The liquid
filling can be done by different processes. In this experiment the filling process is
done by making the pressure difference at the two ends of the pipe. Due to some
technical problem t the filling ratio could not be controlled at desired level.
For 450 inclinations the filling ratio is 85.6% and for horizontal position (900
inclinations) the filling ratio is 85%.
Heating process
Heating process is done at the evaporator section. It is situated at the bottom of the
heat pipe. The heating process is done by passing the AC current through the
Nicrome wire. As the resistance of the wire is very high, heat is generated during the
current flow. This heat is passed to water and the water is evaporated when the
evaporator temperature is higher than the vaporization temperature of water.
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Cooling process
The cooling process is occurred at the condenser section. It is at the top of the heat
pipe where the water vapor is condensed. The cooling process is done by a DC
operated fan [fig: 3.12]. An adapter circuit [fig: 3.16] is used to convert the AC to
DC. The fan speed is 3.5 m/s.
Measuring system
For recording the temperature at different position of the heat pipe, K type
thermocouples are used. The data is recorded at a regular time interval. The time
interval is 10 minutes. A digital thermometer [fig: 3.14] is used for observing the
temperature. Selector switch [fig: 3.13] is used for selecting different thermocouples
to observe their corresponding temperature.
Orientation
The heat pipe is tilted by changing the angle of the rotating plate of the test stand.
Different thermal characteristics are observed on different orientation of heat pipe by
changing the tilt angle of this plate. The different orientation angles are:
Vertical orientation (00)
300, 450 and 600 inclinations
Horizontal orientation (900)
Fig 3.1 to 3.5 show the different orientation of heat pipe and fig 3.7 to 3.16 shows the various components and essential apparatus to conduct the experiment.
19
20
Fig 3.6: Experimental Setup of Pulsating Heat Pipe (PHP)
21
22
Fig 3.9: Copper tube used for constructing Heat Pipe
Fig 3.7: Beaker used for measurement of water
Fig 3.10: Glass wool
Fig 3.8: Araldite used for sealing and connecting
materials
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Fig 3.15: Filling valve for filling water
Fig 3.13: Selector switch for selecting different thermocouple
Fig 3.16: Adapter circuit
Fig 3.11: Variac used for variable cooling
Fig 3.14: Digital Thermometer
Fig 3.12: Fan used for Forced Power supply
3.9 Mathematical Equation
Eotvos Number or Bond Number:
Dcrit=2[ σg( ρliq−ρ vap)
]12
Eo=(B¿¿ o)2¿
Thermal Resistance:
(Te-Tc) / Q ……….. (1)
Where, Te= Evaporator temperature (0C)
Tc= Condenser temperature (0C)
Q= Heat input (W)
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CHAPTER 4
RESULTS AND DISCUSSION
To evaluate and understand the heat transfer characteristics of closed loop pulsating
heat pipe, the wall temperature at different points of the CLPHP is measured. The
temperature profiles are plotted against the heat input and time. By using equation
(1), thermal resistance is determined and then thermal resistances are plotted against
heat input.
Vertical Mode
In vertical mode, the vapor bubbles take up heat in the evaporator and grow in size.
Their own buoyancy helps them to rise up in the tube section. Simultaneously other
bubbles, which are above in the tube, are also helped by their respective buoyant
forces. These rising bubbles in the tube also carry the liquid slugs trapped in between
them. In this mode of operation there is a natural tendency for the liquid slugs to
travel toward the evaporator helped by gravity force. Simultaneously the vapor
bubbles have the natural tendency to travel towards the condenser helped by buoyant
force.
Fig 4.1 shows that the thermal resistance of the PHP decreases with the increase of
the heat input. At 100% filling ratio, PHP acts as single phase buoyancy driven
thermosyphon. In this case, there is no bubble formation and the liquid starts
circulating inside the device due to density difference associated with the
temperature gradient. So decreasing trend is higher till the heat input of 35 W and
then with the increase of heat input the temperature difference does not change much
between evaporator and condenser due to higher buoyancy force is required to
overcome the liquid phase. So the trend remains nearly steady.
Fig 4.2, with 100% filling ratio, shows that the average evaporator and condenser
temperature increase with the heat input. At 100% filling ratio the maximum
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temperature 101.78 C is attained at 56.67 W heat input. The maximum condenser⁰
temperature 46.24 C is attained at same heat input and time. In this figure the slope⁰
of evaporator temperature is higher than the slope of condenser temperature.
With 82.5% filling ratio, small amount of bubble formation occurs and natural
circulation of flow is hindered. Evaporator temperature increases which results high
temperature difference between evaporator and condenser. Fig 4.3 shows that the
slope of decreasing trend of thermal resistance is higher till the heat input of 12 W.
Bubble formation increases with heat input leading to higher pumping pressure and
thermal instabilities. Then the pulsating mode starts and thermal resistance decreases
smoothly.
Fig 4.4, with 82.5% filling ratio, shows that the average evaporator and condenser
temperature increase with the heat input. At 82.5% filling ratio the maximum
temperature 101.39 C is attained at 62.31 W heat input. The maximum condenser⁰
temperature 56.53 C is attained at same heat input. ⁰
Fig 4.5, with 63% filing ratio, shows that the thermal resistance of the PHP decreases
with the increase of the heat input. At this filling ratio the partial dry out of
evaporator section occurs and the true pulsating mode of PHP starts. The
performance of the CLPHP improves. The pulsating action starts properly and the
thermal resistance is expected to decrease smoothly. But the curve does not show the
smooth trend due to the entrapped air in the heat pipe which enters during the filling
process.
Fig 4.6, with 63% filling ratio, shows that the average evaporator and condenser
temperature increases with the heat input. At 63% filling ratio, the maximum
temperature 102.67 C is attained at 62.81 W heat input. The maximum condenser⁰
temperature 61.56 C is attained at same heat input. ⁰
At 41.3% filling ratio, the dry out of evaporator section increases. So, the bubble
formation increases and the pumping action improve. So, the figure 4.7 shows that
the decreasing trend of thermal resistance is almost smooth throughout the heat input.
Fig 4.8, with 41.3% filling ratio, shows average evaporator and condenser
temperature increases with the heat input. At 41.3% filling ratio, the maximum
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temperature 102.43 C is attained at 64.27 W heat input. The maximum condenser⁰
temperature 62.83 C is attained at the same heat input. ⁰
At 28% filling ratio, the internal bubble size increases in the evaporator which
creates the dry out of evaporator section. The flow instabilities increase the level of
perturbation. So, the decrement of thermal resistance is high. Fig 4.9, with 28% filing
ratio, shows that the decreasing trend of thermal resistance higher till the heat input is
20 W and than the decreasing trend becomes slow. The maximum performance is
observed at this filling ratio.
Fig 4.10, with 28% filling ratio, shows that the average evaporator and condenser
temperature increase with the heat input. At 28% filling ratio the maximum
temperature 104.01 C is attained at 63.36 W heat input. The maximum condenser⁰
temperature 72.41 C is attained at same heat input. ⁰
Fig 4.11 shows that horizontal mode (90 inclinations) of operation with 82% filling⁰
ratio. It shows the thermal resistance of the PHP nearly steady with the heat input.
This suggests that gravity force is uniform throughout the PHP. So, the gravity force
is inactive and the only dominating force is surface tension. At horizontal inclination
(90°), all the temperatures of the adiabatic section rapidly become equal and there is
no movement of bubble plugs. Bubbles only oscillate about a mean position with
high frequency. The temperature difference between the evaporator and condenser
increases with the increment of heat input proportionally so the thermal resistance
remains nearly steady.
Fig 4.12, at horizontal inclination with filling ratio 82%, shows that average
evaporator and condenser temperature increase with the heat input. At 82% filling
ratio, the maximum temperature 101.87 C is attained at 40.92 W heat input. The⁰
maximum condenser temperature 48.42 C is attained at same heat input. ⁰
At 60 inclination with 79% filling ratio, the pressure force is created due to the⁰
temperature difference between the evaporator and condenser. The pressure force
acts with the surface tension force inside the tubes. At first the bubbles move slowly
to condenser, then the bubbles move faster with the increase of heat input, so the heat
27
transfer increases and Fig 4.13 shows that the thermal resistance of the PHP
decreases with increase of heat input.
Fig 4.14, at 60 inclination with filling ratio 79%, shows that the average evaporator⁰
and condenser temperature increase with the heat input. At 79% filling ratio, the
maximum temperature 104.57 C is attained at 56.78 W heat input. The maximum⁰
condenser temperature 57.44 C is attained at same heat input.⁰
Fig 4.15, 45⁰ inclination of with 85.6% filling ratio, shows that the thermal resistance
of the PHP decreases with increase of heat input. Before 30W heat input the decrease
trend of thermal high and after that the decrease trend becomes slow.
Fig 4.16, at 45 inclination with filling ratio 85.6%, shows that the average⁰
evaporator and condenser temperature increase with the heat input. At 85.6% filling
ratio, the maximum temperature 103.54 C is attained at 58.68 W heat input. The⁰
maximum condenser temperature 61.08 C is attained at same heat input. ⁰
Fig 4.17, 30 inclination with 79% filling ratio, shows that the decreasing trend of⁰
the thermal resistance of the PHP is uniform throughout the heat input. At this
position the pressure force due to temperature difference is higher than the gravity
force. The surface tension force is also active.
Fig 4.18, at 30 inclination with filling ratio 79%, shows that the average evaporator⁰
and condenser temperature increase with the heat input. At 79% filling ratio, the
maximum temperature 104.74 C is attained at 61.23 W heat input. The maximum⁰
condenser temperature 64.33 C is attained at same heat input. ⁰
Fig 4.19 shows that the comparison of thermal resistance vs. heat Input at different
filling ratios. A closer look at comparison curve, it has been found that in between
25% and 65% filling ratio, the PHP functions in a true pulsating mode. Thermal
performance improves at lower filling ratio with partial or total dry out in the
evaporator section. The maximum performance was observed at about 25% to 30%
filling ratio.
Fig 4.20 shows that the maximum heat input vs. filling ratio. The maximum heat
input was found at lower filling ratio and minimum heat input at 100% filling ratio.
28
At 100% filling ratio, the buoyancy induced natural liquid circulation. Liquid
circulation gets hampered due to surface tension which results insufficient
perturbations. So the performance of the device is hampered. But at lower filling
ratio, more bubbles are formed which produce more perturbations. The PHP operates
as true pulsating mode and give high throughput.
Fig 4.21 shows that comparison of thermal resistance vs. heat input at different
inclination nearly same filling ratio. A closer look at comparison curve, at 30⁰
inclination closed loop PHP performs better than other position but comparatively at
vertical mode of operation it gave maximum heat throughput.
Fig 4.22 shows that the maximum heat input vs. inclination angle nearly at same
filling ratio. The maximum heat input decreases with decreasing the inclination
angle. The maximum heat input was obtained at vertical inclination (0°) and
minimum heat input was obtained at horizontal position (90°). At vertical inclination,
the motion of the liquid slugs and vapor bubbles at one section of the PHP moves
toward the condenser. This works as driving force. On the other section the motion
of slugs and bubbles moves toward the condenser. This works as restoring force. The
inter-play between the driving force and restoring force leads to oscillation of the
vapor bubbles and liquid slugs in the axial direction. So heat input is high at vertical
position. At horizontal inclination (90°), all the temperatures of the adiabatic section
rapidly equalize and no movement of bubble plugs. Bubbles only oscillate about a
mean position with high frequency. The input heat has to be stopped for safety and
surface tension predominates in capillary dimensions of the tubes.
From all the curves, some deviations are found from the expected theoretical value.
During the filling of working fluid, theoretically the tube must be vacuumed
perfectly. But due to technical limitations, the tubes could not be maintained
vacuumed. Some air infiltrated in the heat pipe. So, during heating the entrapped air
is also heated with the working fluid and the resistance of the air cannot be
neglected. So, the actual thermal resistance is greater than expected theoretical value.
Inclination: Vertical
FR=100%
29
0 10 20 30 40 50 600
0.5
1
1.5
2
2.5
Thermal Resistance vs Heat Input
Heat Input (W)
ThermalResistance
(°C/W)
Fig 4.1: Thermal Resistance vs. Heat Input
0 10 20 30 40 50 600
20
40
60
80
100
120
Avg. Evaporator & Condenser Temp.vs
Heat Input
Avg. Tevap (°C)
Avg. Tcond. (°C)
Heat Input (W)
Temp.(°C)
Fig 4.2: Avg. Evaporator and Condenser Temp vs. Heat Input
Inclination: Vertical
FR=82.5%
30
0 10 20 30 40 50 60 700
0.2
0.4
0.6
0.8
1
1.2
1.4
1.6
1.8
2
Thermal Resistance vs. Heat Input
Heat Input (W)
ThermalResistance
(°C/W)
Fig 4.3: Thermal Resistance vs. Heat Input
0 10 20 30 40 50 60 700
20
40
60
80
100
120
Avg. Evaporator and Condenser Temp. vs
Heat input
Avg. Tevap (°C)
Avg. Tcond. (°C)
Heat Input (W)
Temp.(°C)
Fig 4.4: Evaporator Temp vs. Heat Input
Inclination: Vertical
FR=63%
31
0 10 20 30 40 50 60 700
0.2
0.4
0.6
0.8
1
1.2
1.4
1.6
1.8
2
Thermal Resistance vs Heat Input
Heat Input (W)
ThermalResistance
(°C/W)
Fig 4.5: Thermal Resistance vs. Heat Input
0 10 20 30 40 50 60 700
20
40
60
80
100
120
Avg. Evaporator and Condenser Temp.vs
Heat Input
Avg. Tevap (°C)
Heat Input (W)
Temp (0C)
Fig 4.6: Evaporator and Condenser Temperature vs. Heat Input
Inclination: Vertical
FR=41.3%
32
0 10 20 30 40 50 60 700
0.2
0.4
0.6
0.8
1
1.2
1.4
1.6
1.8
Thermal Resistance vs Heat Input
Heat Input (W)
Thermal Resistance
(°C/W)
Fig 4.7: Thermal Resistance vs. Heat Input
0 10 20 30 40 50 60 700
20
40
60
80
100
120
Avg. Evaporator and Condenser temp.vs
Heat Input
Avg. Tevap (°C)
Avg. Tcond .(°C)
Heat Input (W)
Temp.(°C)
Fig 4.8: Avg. Evaporator and Condenser Temp vs. Heat Input
Inclination: Vertical
Filling ratio: 28%
33
0 10 20 30 40 50 60 700
0.2
0.4
0.6
0.8
1
1.2
1.4
1.6
1.8
2
Thermal Resistance vs Heat Input
Heat Input (W)
Thermal Resistance(0C/W)
Fig 4.9: Thermal Resistance vs. Heat Input
0 10 20 30 40 50 60 700
20
40
60
80
100
120
Avg. Evaporator and Condenser Temp.vs
Heat Input
Avg. Tevap (°C)
Avg. Tcond .(°C)
Heat Input (W)
Temp.(°C)
Fig 4.10: Evaporator and Condenser Temperature vs. Heat Input
Inclination: 90° (Horizontal)
34
FR=82%
0 5 10 15 20 25 30 35 40 450
0.2
0.4
0.6
0.8
1
1.2
1.4
1.6
1.8
2
Thermal Resistance vs Heat Input
Heat Input (w)
ThermalResistance
(°C/W)
Fig 4.11: Thermal Resistance vs. Heat Input
0 5 10 15 20 25 30 35 40 450
20
40
60
80
100
120
Avg. Evaporator and Condensor Temp.vs
Heat Input
Avg. Tevap (°C)
Avg. Tcond. (°C)
Heat Input (W)
Temp.(°C)
Fig 4.12: Avg. Evaporator and Condenser Temp vs. Heat Input
Inclination: 60°
FR=79%
35
0 10 20 30 40 50 600
0.5
1
1.5
2
2.5Thermal Resistance vs. Heat Input
Heat Input (W)
Thermal Resistance(°C/W)
Fig 4.13: Thermal Resistance vs. Heat Input
0 10 20 30 40 50 600
20
40
60
80
100
120
Avg. Evaporator and Condenser Temp.vs
Heat Input
Avg. Tevap (°C)
Avg. Tcond. (°C)
Heat Input (w)
Temp. (°C)
Fig 4.14: Avg. Evaporator and Condenser Temp vs. Heat Input
Inclination: 45°
36
FR=85.6%
0 10 20 30 40 50 60 700
0.5
1
1.5
2
2.5
Thermal Resistance Vs Heat Input
Heat Input (W)
ThermalResistance
(°C/W)
Fig 4.15: Thermal Resistance vs. Heat Input
0 10 20 30 40 50 60 700
20
40
60
80
100
120
Avg. Evaporator and Condenser Temp vs
Heat Input
Avg. Tevap (°C)
Avg. Tcond. (°C)
Heat Input (W)
Temp.(°C)
Fig 4.16: Avg. Evaporator and Condenser Temp vs. Heat Input
Inclination: 30°
37
FR=79%
0 10 20 30 40 50 60 700
0.5
1
1.5
2
2.5
Thermal Resistance vs Heat Input
Heat Input (W)
ThermalResistance
(°C/W)
Fig 4.17: Thermal Resistance vs. Heat Input
0 10 20 30 40 50 60 700
20
40
60
80
100
120
Avg. Evaporator and Condenser Temp vs
Heat Input
Avg. Tevap (°C)
Avg. Tcond. (°C)
Heat Input (W)
Temp.(°C)
Fig 4.18: Avg. Evaporator and Condenser Temp vs. Heat Input
38
Inclination: Vertical
0 10 20 30 40 50 600
0.5
1
1.5
2
2.5
Comparision of Thermal Resistance vs. Heat Input at dfferent Filling Ratio(At vertical position)
FR=28%(Vertical)
FR=41.3%(Vertical)
FR=63%(Vertical)
FR=82.5%(Vertical)
FR=100%(Vertical)
Heat Input (W)
ThermalResistance
(°C/W)
Fig 4.19: Comparison of Thermal Resistance vs. Heat Input at different Filling Ratio
(At vertical position)
39
20 30 40 50 60 70 80 90 100 11052
54
56
58
60
62
64
66
Maximum Heat Input vs Filling ratio
Filling Ratio (%)
MaximumHeatInput(W)
Fig 4.20: Maximum Heat Input vs. Filling Ratio
40
0 10 20 30 40 50 60 700
0.5
1
1.5
2
2.5
Comparision of Thermal Resistance vs. Heat Input at dfferent Inclina-tion
(Nearly same filling ratio)
Inclination 90°( FR=82%)Inclination 30°( FR=79%)Inclination 45°( FR=85.6%) Inclination 60°( FR=79%)
Heat Input(W)
Fig 4.21: Comparison of Thermal Resistance vs. Heat Input at Different Inclination
(Nearly same filling ratio)
41
Change of heat input with inclination angle
(Nearly same filling ratio)
0 10 20 30 40 50 60 70 80 90 1000
10
20
30
40
50
60
70
Maximum Heat Input vs Inclination Angle
Inclination Angle (deg.)
Maximum Heat Input (W)
Fig 4.22: Maximum Heat Input vs. Inclination angle
42
CHAPTER 5
SUMMARY AND CONCLUSION
The following facts summarize the essential aspects of this study:
1. Valuable information related to the fundamental characteristics and operational
regimes of a PHP were generated. An operationally better performance and self-
sustained thermally driven pulsating action of the device was only observed in the
filling ratio range 25–65%. Above this range, the overall degree of freedom and the
pumping action of bubbles were insufficient for rendering good performance. Below
a certain range of filling ratio, partial dry out of the evaporator was detected.
2. The results also indicate that a 100% filled PHP (not working in the pulsating
mode but instead as a single-phase buoyancy-induced thermosyphon) is thermally
better performing than a partially filled pulsating mode device under certain
operating conditions.
3. The tested PHP did not operate in the vertical mode constantly for the working
fluid tested. The reasons are attributed to fixed number of turns and atmospheric
pressures existing at testing conditions.
4. Although the Eotvos number of water was much below the prescribed maximum
limit of Eo = 4, gravity forces were definitely seen to affect the performance. This
suggests that, in the vertical mode fluid transport is mainly by the bubble pumping
action thereby providing substantial heat transfer.
5. Closed loop pulsating heat pipes are complex heat transfer systems with a very
strong thermo- hydrodynamic coupling governing the thermal performance.
6. Different heat input to these devices give rise to different flow patterns inside the
tubes. This intern is responsible for various heat transfer characteristics. The study
indicates that design of these devices should aim at thermo-mechanical boundary
conditions which resulting convective flow boiling conditions in the evaporator
leading to higher local heat transfer co-efficients.
43
7. The inclination operating angle changes the internal flow patterns thereby resulting
in different performance levels. The best performance is obtained from vertical
direction and the worst performance is obtained from the horizontal orientation.
44
RECOMMENDATIONS
1. A comprehensive investigation can be carried out on the closed loop heat
pipe by changing the filling ratio, working fluid, cross sections, shape and
using stainless steel and aluminum.
2. Further investigation can be carried out on closed loop pulsating heat pipe by
introducing more angle of inclination with different filling ratios.
3. Air velocity should be varied to test the thermal performance and water
cooling can be incorporated for the comparison of heat transfer performance
between water cooling and air cooling.
4. For accuracy, temperature reading should be taken electronically via
interfacing system with real time data acquisition.
5. Room temperature should be controlled and atmospheric properties should be
uniform.
6. A Computational Fluid Dynamics (CFD) analysis can be done to investigate
the reasons of difference in thermal resistance.
45
REFERENCE
1. Akachi, H. polasek F., Stule P., “Pulsating Heat Pipes”, Proceedings of the 5 th
International Heat Pipe Symposium, 1996, pp.208-217, Melbourne, Australia.
2. Khandekar, S., Dollinger, N., Groll, M., “Understanding Operational
Regimes of Closed Loop Pulsating Heat Pipes: An Experimental Study”,
2003, Applied Thermal Engineering, vol. 23, PP. 707-719.
3. Zhang, Y.,Faghri, A., “Heat Transfer in a Pulsating Heat Pipe with an Open
End”, International journal of Heat and Mass Transfer, Vol. 45, 2002, PP.
755-764.
4. Delil, A., A.M, “pulsating and oscillating Heat Transfer Devices in
Acceleration Enviorments from Microgravity to Supergravity”, SAE paper
#2001-01-2240.
5. Charoenswan, P., Khandekar, S., Groll, M., Terdtoon, P., “Closed Loop
Pulsating Heat Pipes Part A: Parametric Experimental Investigations”,
Applied Thermal Engineering, 2003, Vol. 23, pp. 2009-2020.
6. Khandekar, S. and Groll, M., “Pulsating Heat Pipes : A Challenge and Still
Unsolved Problem in Heat Pipe Science”, Proceedings of the 3rd International
Conference on Transport Phenomena in Multiphase systems , 2002,35-44
(ISBN 83-88906-03-8) pp. 2002 Vol. 3.
7. Khandekar, S, Schneider, M ,Schafer, P., Kulenovic R., Groll, M.,
“Thermofluid Dynamic Study of Flat Plate Closed Loop Pulsating Heat
Pipes”, Microscale Thermo physical Engineering, Taylor and Francis, 2002,
(ISSN 1089-3954) pp. 303-318, Vol. 6/4.
46
8. Rettidech, S. and Roger R. Riehl, “Characteristics of an Open Loop pulsating
Heat Pipe”, SAE paper #2004-01-2509.
9. Tong, B. Y.,Wong, T. N., Ooi, K.T., “Closed Loop Pulsating Heat Pipe” ,
2001, Applied Thermal Engineering, Vol. 21, pp. 1845-1862.
10. Khandekar, S., Groll, M., Charonsawan, P., Terdtoon, P., “Pulsating Heat
Pipes: Thermo-Fluidic Charecterstics and Comparative Study with Single
Phase Thermosyphon”, Proceedings of 12th International Heat Transfer
Conference , Vol 4, pp.459-464, Grenoble, France,2002.
11. Chowdhury, F., et. al. “Study on Heat Transfer Characteristics of Looped
Parallel Thermosyphon,” Proceedings of 4th European Thermal Science
Conference., s10-HPI-2004.
12. Swanepoel, G., Thermal Management of Hybrid Electrical Vehicles Using
Heat Pipes, M. Sc. Thesis, University of Stellenbosch. 2001.
13. Zhang, Y, Faghri, A “Heat Transfer in a pulsating with an open End “,
International journal of Heat and Mass Transfer, vol 45, 2002, pp .755-764.
14. Roger R. Riehl, “Characteristics of an Open Loop pulsating Heat Pipe”, SAE
paper #2004-01-2509.
15. Rittidech, S., Terdtoon, P., Murakami, M., Kamonpet, P., Jompakdee, W.,
“Correlation to predict Heat Transfer Characteristics of a Closed End
Oscillating Heat Pipe at Normal Temperature Condition”, 2003, Applied
Thermal Engineering, Vol. 23, pp. 497-510.
16. Asselman, G.A. and Groll, D.B., "Heat Pipes," Philips Tech. Rev., No. 4, PP 104 - 113, 1973.
47
APPENDIX A
A.1 TYPICAL HEAT PIPE
A heat pipe is a heat transfer mechanism that combines the principles of both thermal
conductivity and phase transition to efficiently manage the transfer of heat between
two solid interfaces.
At the hot interface within a heat pipe, which is typically at a very low pressure, a
liquid in contact with a thermally conductive solid surface turns into a vapor by
absorbing the heat of that surface. The vapor condenses back into a liquid at the cold
interface, releasing the latent heat. The liquid then returns to the hot interface through
either capillary action or gravity action where it evaporates once more and repeats
the cycle. In addition, the internal pressure of the heat pipe can be set or adjusted to
facilitate the phase change depending on the demands of the working conditions of
the thermally managed system.
A typical heat pipe consists of a sealed pipe or tube made of a material with high
thermal conductivity such as copper or aluminum at both hot and cold ends. A
vacuum pump is used to remove all air from the empty heat pipe, and then the pipe is
filled with a fraction of a percent by volume of working fluid (or coolant) chosen to
match the operating temperature. Examples of such fluids include water, ethanol,
acetone, sodium, or mercury. Due to the partial vacuum that is near or below the
vapor pressure of the fluid, some of the fluid will be in the liquid phase and some
will be in the gas phase. The use of a vacuum eliminates the need for the working gas
to diffuse through any other gas and so the bulk transfer of the vapor to the cold end
of the heat pipe is at the speed of the moving molecules. In this sense, the only
48
practical limit to the rate of heat transfer is the speed with which the gas can be
condensed to a liquid at the cold end.
Inside the pipe's walls, an optional wick structure exerts a capillary pressure on the
liquid phase of the working fluid. This is typically a sintered metal powder or a series
of grooves parallel to the pipe axis, but it may be any material capable of exerting
capillary pressure on the condensed liquid to wick it back to the heated end. The heat
pipe may not need a wick structure if gravity or some other source of acceleration is
sufficient to overcome surface tension and cause the condensed liquid to flow back to
the heated end.
The materials chosen depend on the temperature conditions in which the heat pipe
must operate, with coolants ranging from liquid helium for extremely low
temperature applications (2–4 K) to mercury (523–923 K) and sodium (873–1473 K)
and even indium (2000–3000 K) for extremely high temperatures. The vast majority
of heat pipes for low temperature applications use some combination of ammonia
(213–373 K), alcohol (methanol (283–403 K) or ethanol (273–403 K)) or water
(303–473 K) as working fluid. Since the heat pipe contains a vacuum, the working
fluid will boil and hence take up latent heat at well below its boiling point at
atmospheric pressure. Water, for instance, will boil at just above 273 K (0 degrees
Celsius) and so can start to effectively transfer latent heat at this low temperature.
The advantage of heat pipes over many other heat-dissipation mechanisms is their
great efficiency in transferring heat. They are a fundamentally better heat conductor
than an equivalent cross-section of solid copper Some heat pipes have demonstrated
a heat flux of more than 230 MW/m², nearly four times the heat flux at the surface of
the sun.
49
Fig A.1: Operation of Heat Pipe
A.2 TYPES OF HEAT PIPE:
Heat pipes can be divided into different categories depending on their structure and
shape.
On the basis of structure and operation:
Flexible heat pipe
Looped parallel heat pipe
Rotating heat pipe
Pulsating heat pipe
On the basis of shape
Flat type
Tubular and cylindrical type
On the basis of cooling system in the condenser
Water cooled heat pipe Air cooled heat pipe
50
A.3 PULSATING HEAT PIPE
The Pulsating Heat Pipe is an innovating technology that has gained attention in the
last 5 years. This is a special type of heat pipe and the driving force is the slug/plug
motion of the working fluid in the tube, generated by the evaporation. PHPs consist
of a meandering tube bent to form several parallel channels. It can be configured as
an:
open loop pulsating heat pipe
closed loop pulsating heat pipe
In the first one, one end of the PHP is pinched-off and welded, while the other end
presents a service valve for vacuum and charge the closed loop PHP is an endless
tube as both ends are welded together. Each PHP configuration presents particular
operation modes, which are mainly guided by the chaotic slug/plug motion. Either
PHP configuration presents a high dependence on their thermal behavior related to
the gravity vector during operation, which must be carefully considered. Higher
operation temperatures are achieved when the PHP operates at the vertical
orientation, while at horizontal orientation, the operation temperatures are lower.
Fig A.2: Closed Loop Pulsating Heat Pipe
51
Fig A.3: Closed Loop Pulsating Heat Pipe
A.4 APPLICATIONS OF HEAT PIPE
Heat pipes are very efficient heat transport elements. They can be described as light
weight devices with high thermal conductance. Heat pipes allow the transportation of
high fluxes with small temperature difference with no change in the operating
temperature. In addition, there are no moving mechanical parts in heat pipes, and
special sets of them can be used for temperature control, as thermal diodes and
thermal switches. Also, they can be built in difference geometries and sizes.
Most suitable where:
Low humidity level necessary
Humidity control required
Air reheated after cooling in traditional HVAC system
Large quantities of ventilation air needed
Electronic component production, assembly and storage
Film drying, processing and storage
Drug, chemical and paper manufacturing and storage
Candy, chocolate processing and storage
52
Swimming pool enclosures
Hospital operating rooms
Grocery stores
Telephone exchanges, relay stations, clean rooms
Underground silos
Other Heat Pipe Applications
Heat pipes have been used for many applications:
a. Remote heat rejection from a concentrated source (e.g. computer chip)
b. Obtain uniform temperature
c. Efficient heat exchangers.
d. Space technology
e. Note book and Desktop application
f. Laptop heat solution
g. Solar thermal
h. Pipeline over permafrost.
53
Fig A.4: Heat Pipe in Miniature Form
Fig A.5: Application of Heat Pipe on Computer Technology
54
A.5 LIMITATIONS OF HEAT PIPE
When heated above a certain temperature, all of the working fluid in the heat
pipe will vaporize and the condensation process will cease to occur: in such
cases, the heat pipes thermal conductivity is effectively reduced to heat
conduction properties of its solid metal casing alone. As most heat pipes
constructed of copper (a metal with high heat conductivity), an overheated
heat pipe will generally continue to conduct heat at around 1/80 of the
original conductivity.
If the heat source temperature drops below a certain minimum value,
depending on the specific fluid and gas combination in the heat pipe, a
complete shutoff can occur. So the control feature is particularly useful for
fast worm up application in addition to its value as a temperature leveler for
variable load conditions.
The rate of heat transfer through the heat pipe is solely dependent on the rate
of evaporation and condensation. If the temperature difference is not high
enough, the heat transfer rate at condenser section would decrease. Natural
convection by air is not high enough to support high rate of cooling.
If non condensable gases are present in the gas mixture, then the heat transfer
will be affected. To ensure effective heat transfer, a mechanism has to be
introduced in the heat pipe system.
Most manufacturers cannot make a traditional heat pipe smaller than 2 mm
due to material limitation. Experiments have been conducted with micro heat
pipes, which use piping with sharp edges, such as triangular or rhombus like
tubing. In these cases, the sharp edges transfer the fluid trough capillary
action, and no wick is necessary.
Heat pipes are excellent heat transfer devices but their sphere of application is
mainly confined to transferring relatively small heat loads over relatively
55
short distances when the evaporator and condenser are at same horizontal
level. This limitation on the part of the heat pipes is mainly related to the
major pressure losses associated with the liquid flow through the porous
structure, present along the entire length of the heat pipe and viscous
interaction between the vapor and liquid phases, also called entrainment
losses. For the applications involving transfer of large heat loads over long
distances, the thermal performance of the heat pipes is badly affected by
increase in these losses. For the same reason conventional heat pipes are very
sensitive to the change in orientation in gravitational field. For the
unfavorable slopes in evaporator-above-condenser configuration, the pressure
losses due to the mass forces in gravity field adds to the total pressure losses
and further affect the efficiency of the heat transfer process.
As a result of these limitations, different solutions involving structural
modifications to the conventional heat pipe have been proposed. Some of
these modifications incorporate arterial tubes with considerably low hydraulic
resistance for liquid return to the heat source (arterial heat pipes), while
others provide spatial separation of the vapor and liquid phases of the
working fluid at the transportation section (separated line heat pipes).
Though these new forms of heat pipes are able to transfer significant heat
flows and can increase heat transport length, they remain very sensitive to
spatial orientation relative to gravity. To extend functional possibilities of
two-phase systems towards applications involving otherwise inoperable
slopes in gravity, the advantages provided by the spatial separation of the
transportation line and the usage of non-capillary arteries are combined in the
loop scheme. This scheme allows heat pipes to be created with higher heat
transfer characteristics while maintaining normal operation in any directional
orientation. The loop scheme forms the basis of the physical concept of Two-
Phase Loops (TPLs).
56
APPENDIX B
DESIGN OF PULSATING HEAT PIPE
B.1 WHY HEAT PIPE
Limited space budget
No electrical consumption
Zero noise or noise reduction
Low maintenance and high reliability
Stagnation region
Low weight
B.2 COMPONENTS OF HEAT PIPE
The three basic components of a heat pipe are:
1. The container
2. The working fluid
3. The wick or capillary structure
Container
The function of the container is to isolate the working fluid from the outside
environment. It has to therefore be leak-proof, maintain the pressure differential
across its walls, and enable transfer of heat to take place from and into the working
fluid.
Selection of the container material depends on many factors. These are as follows:
Compatibility (both with working fluid and external environment)
Strength to weight ratio
Thermal conductivity
57
Ease of fabrication, including welding, machineability and ductility
Porosity
Wettability
Most of the above are self-explanatory. A high strength to weight ratio is more
important in spacecraft applications. The material should be non-porous to prevent
the diffusion of vapor. A high thermal conductivity ensures minimum temperature
drop between the heat source and the wick.
Working Fluid
A first consideration in the identification of a suitable working fluid is the operating
vapor temperature range. Within the approximate temperature band, several possible
working fluids may exist, and a variety of characteristics must be examined in order
to determine the most acceptable of these fluids for the application considered. The
prime requirements are:
Compatibility with wick and wall materials
Good thermal stability
Wettability of wick and wall materials
Vapor pressure not too high or low over the operating temperature range
High latent heat
High thermal conductivity
Low liquid and vapor viscosities
High surface tension
Acceptable freezing or pour point
The selection of the working fluid must also be based on thermodynamic
considerations which are concerned with the various limitations to heat flow
occurring within the heat pipe like viscous, sonic, capillary, entrainment and nucleate
boiling levels.
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In heat pipe design, a high value of surface tension is desirable in order to enable the
heat pipe to operate against gravity and to generate a high capillary driving force. In
addition to high surface tension, it is necessary for the working fluid to wet the wick
and the container material i.e. contact angle should be zero or very small. The vapor
pressure over the operating temperature range must be sufficiently great to avoid
high vapor velocities, which tend to setup large temperature gradient and cause flow
instabilities.
A high latent heat of vaporization is desirable in order to transfer large amounts of
heat with minimum fluid flow, and hence to maintain low pressure drops within the
heat pipe. The thermal conductivity of the working fluid should preferably be high in
order to minimize the radial temperature gradient and to reduce the possibility of
nucleate boiling at the wick or wall surface. The resistance to fluid flow will be
minimized by choosing fluids with low values of vapor and liquid viscosities.
Tabulated below are a few mediums with their useful ranges of temperature.
B.3 PHP DESIGN
The cooling device performance depends on it’s structure, shape, material and
length. Thermal performance of any device vastly depends on a parameter known as
thermal resistance.
Thermal resistance is
Rth = (ΔT/Q)
Where. ΔT= temperature drop along the device
Q= heat load
The overall thermal resistance of a pulsating heat pipe composed of several
components from evaporator to condenser.
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Two conductive thermal resistance in the wall Rwall
Thermal resistance due to evaporation at evaporator R evap
Thermal resistance due to condensation at condenser Rcond
Thermal resistance along the heat pipe length R l-v
Two contact resistance due to surface roughness Rcont
The total heat transfer capacity of PHP, Q
Q = (ΔT)/ (2 Rwall + R evap +R cond + R l-v + 2 Rcont)
Where,
ΔT = temperature drop along the device
Q = heat load
Wall Resistance, R wall
The conductive thermal resistance of the wall is negligible as the wall material has
high thermal conductivity. The copper is the most common wall material and 1 mm
Copper material introduces 2 × 10-6 °C/W
Evaporation Resistance, R evap
Resistance in the evaporator of heat pipe can be estimated to be between (.001A) 0C/W and (1.180* 10 -4 A) 0C/W. water has been widely approved to have the best
transport capabilities. Best evaporation resistance is achieved due to the best heat
transfer in the case of square channel s due to the liquid film evaporation
enhancement in the channel angles and best bubble rise in that case.
60
Condensation Resistance, R cond
As the matter of cause a similar range for the heat transfer coefficient in the
condensation region can be applied.
Liquid Vapor Thermal Resistance, R l-v
Liquid vapor thermal resistance along the PHP , R l-v , is the most important part
of the thermal chain and is a function of the pressure /temperature state conditions
from the evaporator to the condenser .this resistance determines the PHP heat
transfer rate .It can be summarized altogether with R evap , R cond , R l-v depends on
following effects
Effect of number of turns
Effect of filling ratio
Effect of evaporator/condenser section size area
Effect of inclination angle.
Contact Resistance, R cont
Generally PHP should introduce small contact resistances. Usually in power
electronics, contact thermal resistances appear between the power module and the
cooling device, heat sink or heat exchanger due to the surface roughness.
B.4 INFLUENCING DESIGN PARAMETERS
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Looking into the available literature, it can be seen that six major thermo mechanical
parameters bottom heat mode of operation is possible. have emerged as the primary
design parameters have emerged as the primary design parameters affecting the
CLPHP system dynamic. These include:
Internal diameter of the CLPHP
Input heat flux to the device
Volumetric filling ratio of the working fluid
Total number of turns
Device orientation with respect to gravity
Working fluid thermo physical properties
Other conditions which influence the operation are:
Use of flow direction control check valves
Tube cross sectional shape
Tube material and fluid combination
Rigidity of the tube material.
B.5 TUBE DIAMETER
The internal tube diameter is one of the parameters which essentially defines a PHP.
The physical behavior adheres to the ‘pulsating’ mode only under a certain range of
diameters. The critical Bond number (or Eötvös) criterion gives the tentative design
rule for the diameter
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This criterion ensures that individual liquid slugs and vapor bubbles are formed in
the device and they do not agglomerate leading to phase separation, if the device is
kept isothermally in a non-operating period. This is most crucial, especially if top
heating mode is employed. In bottom heat mode, though at EO > 4 the tendency of
slug flow diminishes as surface tension tends to reduce, a certain amount of liquid
transport is still possible by the bubble pumping action thereby providing substantial
heat transfer. For a given specified heat power, decreasing the diameter will increase
the dissipative losses and lead to poor performance. Increasing the diameter much
above the critical diameter will change the phenomenological operation of the
device. It will no more act as a pulsating heat pipe but will transform into an
interconnected array of two phase thermosyphons. In this case then, only
B.6 TOTAL NUBER OF TURNS
The number of turns increases the level of perturbations inside the device. If the
number of turns is less than a critical value, then there is a possibility of a stop-over
phenomenon to occur. In such a condition, all the evaporator U-sections has a vapor
bubble and the rest of the PHP has liquid. This condition essentially leads to a dry
out and small perturbations cannot amplify to make the system operate self-
sustained. If the total heat throughput is defined, increasing the number of turns leads
to a decrease in heat flux handled per turn. Thus, an optimum number of turns exits
for a given heat throughput.
APPENDIX C
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Inclination: Vertical
FR=28%
Time(sec.)
Heat Input(w)
Avg. Tevap
(°C)Avg. Tcond
.(°C)Thermal Resistance(°C/W)
1200 7.94 46.54 32.25 1.81800 16.01 57.78 36.82 1.312400 21.13 64.34 42.98 1.01 3000 25.57 70.41 46.12 0.953600 31.42 78.84 53.39 0.814500 36.72 84.37 56.46 0.765400 42.4 93.08 62.98 0.716000 47.51 97.71 67.2 0.646600 57.02 101.21 70.43 0.547200 63.36 104.01 69.71 0.547800 63.38 103.32 71.31 0.518400 63.38 103.72 72.41 0.51
Inclination: VerticalFR=41.3%
Time(sec.)
Heat Input(w)
Avg. Tevap
(°C)Avg. Tcond
.(°C)Thermal Resistance(°C/W)
600 4.73 35.78 28.31 1.581200 8.43 40.31 28.43 1.411800 15.29 48.54 28.78 1.302400 23.37 57.13 29.32 1.193000 29.91 64.12 33.61 1.023600 38.41 73.76 36.50 0.974800 43.73 79.49 44.07 0.815400 48.38 84.78 49.46 0.736000 55.45 94.27 57.12 0.677200 64.27 102.23 62.01 0.637800 64.27 102.43 62.83 0.628400 64.27 102.37 63.41 0.61
Inclination: Vertical
FR=63%
64
Time(sec.)
Heat Input(w)
Avg. Tevap
(°C)Avg. Tcond.
(°C)Thermal Resistance(°C/W)
600 4.12 35.79 28.39 1.81200 9,71 43.57 29.01 1.491800 15.81 52.17 30.43 1.372400 21.78 59.97 31.34 1.313000 28.48 66.73 36.826 1.053600 35.37 75.84 41.88 0.964200 41.82 83.64 49.35 0.824800 54.32 96.71 56.51 0.745400 62.73 101.73 59.70 0.676000 62.81 102.55 61.35 0.666600 62.81 102.42 61.34 0.6547200 62.81 102.67 61.56 0.6537800 62.81 102.57 61.74 0.65
Inclination: Vertical
FR=82.5%
ime(sec.)
Heat Input(w)
Avg. Tevap
(°C)Avg. Tcond.
(°C)Thermal Resistance(°C/W)
600 4.78 38.14 29.27 1.91200 8.64 45.49 31.39 1.621800 12.1 51.23 33.12 1.502400 21.49 63.78 33.73 1.433000 28.67 71.43 35.02 1.273600 33.23 76.08 37.21 1.174200 38.78 81.78 39.89 1.084800 46.52 89.93 42.95 1.015400 52.64 95.83 45.29 0.966000 62.31 101.23 49.51 0.836600 62.31 101.43 52.21 0.797200 62.31 101.39 56.53 0.727800 62.31 101.54 56.76 0.72
Inclination: Vertical
FR=100%
65
Time(sec.)
Heat Input(w)
Avg. Tevap
(°C)Avg. Tcond.
(°C)Thermal Resistance(°C/W)
600 4.73 39.29 28.37 2.311200 7.384 44.43 28.79 2.121800 8.25 45.57 30.80 1.792400 15.12 54.63 31.08 1.533600 17.73 55.71 31.12 1.504200 23.41 62.68 31.67 1.414800 27.58 66.45 31.78 1.266000 34.46 75.12 36.56 1.126600 37.35 79.23 39.133 1.0737200 43.63 87.31 40.24 1.077800 49.51 95.04 45.03 1.028400 56.67 101.78 46.24 0.989000 56.67 101.53 46.54 0.9710200 56.67 101.67 46.64 0.97
Inclination: Horizontal
FR=82%
Time(sec.)
Heat Input(w)
Avg. Tevap
(°C)Avg. Tcond.
(°C)Thermal Resistance(°C/W)
600 3.61 35.49 29.01 1.801200 7.42 43.13 30.37 1.721800 12.59 54.91 34.64 1.612400 19.12 68.37 38.16 1.583600 23.34 79.47 43.99 1.524200 28.16 87.78 45.56 1.514800 33.42 96.25 46.12 1.505400 39.41 101.21 48.16 1.506000 40.92 101.43 48.53 1.396600 40.92 101.87 48.42 1.377200 40.92 101.62 48.32 1.35
Inclination: 30°
FR=79%
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Time(sec.)
Heat Input(w)
Avg. Tevap
(°C)Avg. Tcond.
(°C)Thermal Resistance(°C/W)
600 4,97 38.35 28.34 2.011200 14.81 56.74 29.79 1.821800 19.28 61.14 30.87 1.572400 27.06 72.08 37.44 1.283600 32.87 79.27 44.10 1.074200 32.87 79.54 44.69 1.064800 39.25 87.45 49.77 0.965400 39.28 87.56 50.24 0.956600 42.56 94.72 57.27 0.887200 56.78 104.57 57.44 0.838400 56.78 104.47 57.91 0.829000 56.78 104.31 57.75 0.82
Inclination: 45°FR=85.6%
Time(sec.)
Heat Input(w)
Avg. Tevap
(°C)Avg. Tcond.
(°C)Thermal Resistance(°C/W)
600 5.07 38.8 28.12 1.931200 14.73 56.72 33.23 1.611800 19.47 62.83 34.14 1.482400 27.53 71.69 41.95 1.083600 33.57 79.13 47.91 0.935100 33.57 79.54 48.99 0.916000 39.43 86.23 51.93 0.876600 39.48 86.78 52.43 0.877200 44.32 93.37 57.47 0.818400 58.68 103.24 60.99 0.7259000 58.68 103.54 61.08 0.729600 58.68 103.43 61.84 0.7110800 58.68 103.78 62.12 0.71
Inclination: 60°FR=79%
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Time(sec.)
Heat Input(w)
Avg. Tevap
(°C)Avg. Tcond.
(°C)Thermal Resistance(°C/W)
600 5.12 38.92 28.14 1.911200 15.04 57.41 33.80 1.571800 19.56 63.28 37.07 1.342400 26.53 72.77 43.06 1.123600 33.68 80.45 47.78 0.974800 33.68 80.91 48.58 0.965400 39.48 85.34 53.36 0.816600 47.35 91.04 55.53 0.757200 53.47 97.82 60.93 0.697800 61.23 104.74 64.33 0.668400 61.24 104.57 64.76 0.659000 61.24 104.81 65.62 0.64
Heat Input vs. Filling Ratio (Vertical)
Filling Ratio (%) Maximum Heat Input(W)28 63.3841.3 64.2763 62.8182.5 62.31100 56.67
Heat Input vs. Inclination Angle (Nearly same filling ratio)
Inclination Angle(deg.) Filling Ratio (%) Maximum Heat Input(W)90 82 40.9260 79 56.7845 85.6 58.6830 79 61.240 82.5 62.31
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