thesis 0510044,0510064,0510091 heat pipe 15.02.11

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:

1

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

2

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

3

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

4

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

5

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

6

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

7

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.

8

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.

9

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)

10

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

11

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

12

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

13

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

14

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

16

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.

17

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.

18

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

23

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)

24

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

25

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 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.


temperature increases with the heat input. At 63% filling ratio, the maximum



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

26


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.


temperature increase with the heat input. At 28% filling ratio the maximum



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


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)


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


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


Heat Input (W)

ThermalResistance

(°C/W)


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


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


Heat Input (W)

Thermal Resistance

(°C/W)


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)



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


Heat Input (W)

Thermal Resistance(0C/W)


0 10 20 30 40 50 60 700

20

40

60

80

100

120


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


Heat Input (w)

ThermalResistance

(°C/W)


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)


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)


0 10 20 30 40 50 600

20

40

60

80

100

120


Heat Input

Avg. Tevap (°C)

Avg. Tcond. (°C)

Heat Input (w)

Temp. (°C)


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)


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)


Inclination: 30°

37

FR=79%

0 10 20 30 40 50 60 700

0.5

1

1.5

2

2.5


Heat Input (W)

ThermalResistance

(°C/W)


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)


38


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


41

Change of heat input with inclination angle


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.

58

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.

59

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

61

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

62

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

63


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


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


FR=82.5%

ime(sec.)

Heat Input(w)

Avg. Tevap

(°C)Avg. Tcond.


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


FR=100%

65

Time(sec.)

Heat Input(w)

Avg. Tevap

(°C)Avg. Tcond.


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.


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%

66

Time(sec.)

Heat Input(w)

Avg. Tevap

(°C)Avg. Tcond.


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.


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%

67

Time(sec.)

Heat Input(w)

Avg. Tevap

(°C)Avg. Tcond.


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

68

Documents

thesis 0510044,0510064,0510091 heat pipe 15.02.11