PERFORMANCE ANALYSIS OF A COMPACT HEAT EXCHANGER Department of Mechanical Engineering

PERFORMANCE ANALYSIS OF A COMPACT HEAT EXCHANGER  Department of Mechanical Engineering
PERFORMANCE ANALYSIS OF A COMPACT
HEAT EXCHANGER
A THESIS SUBMITTED IN PARTIAL FULFILLMENT OF THE
REQUIREMENTS FOR THE DEGREE OF
Master of Technology
in
Mechanical Engineering
By
Akash Pandey
Department of Mechanical Engineering
National Institute of Technology
Rourkela
2011
PERFORMANCE ANALYSIS OF A COMPACT
HEAT EXCHANGER
A THESIS SUBMITTED IN PARTIAL FULFILLMENT OF THE
REQUIREMENTS FOR THE DEGREE OF
Master of Technology
in
Mechanical Engineering
By
Akash Pandey
Under the Guidance of
Prof. Ranjit Kumar Sahoo
Department of Mechanical Engineering
National Institute of Technology
Rourkela
2011
National Institute of Technology
Rourkela
CERTIFICATE
This is to certify that the thesis entitled “PERFORMANCE ANALYSIS OF A COMPACT HEAT
EXCHANGER”submitted to the National Institute of Technology, Rourkela (Deemed University)
by Akash Pandey, Roll No. 209ME3227 for the award of the Degree of Master of Technology in
Mechanical Engineering with specialization in “Thermal Engineering” is a record of bonafide
research work carried out by him under my supervision and guidance. The results presented in this
thesis has not been, to the best of my knowledge, submitted to any other University or Institute for
the award of any degree or diploma.
The thesis, in my opinion, has reached the standards fulfilling the requirement for the
award of the degree of Master of technology in accordance with regulations of the Institute.
Place: Rourkela
Date:
Dr. R. K. Sahoo
Professor
Department of Mechanical Engineering
National Institute of Technology, Rourkela
i
ACKNOWLEDGEMENT
I am extremely fortunate to be involved in an exciting and challenging research project
like “Performance Analysis of a Compact Heat Exchanger”. It has enriched my life, giving
me an opportunity to work in a new environment of Cryogenic heat transfer. This project
increased my thinking and understanding capability and after the completion of this project, I
experience the feeling of achievement and satisfaction.
I would like to express my greatest gratitude and respect to my supervisor Prof. Ranjit
Kumar Sahoo, for his excellent guidance, valuable suggestions and endless support. He has not
only been a wonderful supervisor but also a genuine person. I consider myself extremely lucky to
be able to work under guidance of such a dynamic personality. Actually he is one of such
genuine person for whom my words will not be enough to express.
I also express my special thanks to our research scholar Mr. S. A. Alur, Mr. Balaji
Kumar Choudhury, and my best friend and classmate Jitendra Bhushan for their support
during my experimentation. It was impossible for me to complete my project without their help.
I would like to express my thanks to all my classmates, all staffs and faculty members of
mechanical engineering department for making my stay in N.I.T. Rourkela a pleasant and
memorable experience and also giving me absolute working environment where I unlashed ,my
potential .
Last but not the least; I want to convey my heartiest gratitude to my parents for their
immeasurable love, support and encouragement.
Date:
Akash Pandey
Roll. No. 209ME3227
M.tech. (Thermal Engg.)
ii
CONTENTS
CERTIFICATE
i
ACKNOWLEDGEMENT
ii
CONTENTS
iii
ABSTRACT
vi
LIST OF FIGURES
vii
LIST OF TABLES
ix
NOMENCLATURE
x
CHAPTER 1
1
1. INTRODUCTION
2
1.1 Plate Fin Heat Exchanger
3
1.1.1 Advantages and Disadvantages
4
1.1.2 Materials
5
1.1.3 Manufacture
5
1.1.4Applications
6
1.1.5 Flow arrangement
7
1.2 Plate Fin Heat transfer surfaces
10
1.2.1 Plain fins
11
1.2.2 Wavy fins
12
1.2.3 Offset Strip fins
12
1.2.4 Louvered fins
13
1.2.5 Perforated fins
14
1.2.6 Pin fins
14
1.3 Heat Transfer and Flow Friction Characteristics
15
1.4 Objectives of the Study
16
iii
1.5 Organization of the Thesis
17
CHAPTER 2
18
2. LITERATURE SURVEY
18
CHAPTER 3
27
3. EXPERIMENTAL SETUP AND PROCEDURE
27
3.1 Detailed description of various equipment’s and instruments used
27
3.1.1 Plate Fin Heat Exchanger
27
3.1.2 Twin Screw Compressor
30
3.1.3 Heating Element
31
3.1.4 Resistance Temperature Detector (RTDs)
31
3.1.5 Orifice Mass Flow Meter
35
3.1.6 Variac or Autotransformer
36
3.2 Test Rig
3.2.1 Procedure for Hot Testing
37
39
CHAPTER 4
40
4. RATING PROCEDURE
40
CHAPTER 5
47
5. PERFORMANCE ANALYSIS
47
5.1 Calculations
48
5.2 Variation of Effectiveness with Mass Flow Rate
50
5.3 Variation Overall Thermal Conductance with Mas Flow Rate
51
5.4 Variation of Hot and Cold Effectiveness with Mass Flow Rate
52
5.5 Variation of Pressure Drop with Mass Flow Rate
53
CHAPTER 6
54
6. CONCLUSION
54
iv
6.1 Scope for Future Work
7. REFERENCES
54
55
v
ABSTRACT
Compact heat exchangers are one of the most critical components of many cryogenic
components; they are characterized by a high heat transfer surface area per unit volume of the
exchanger. The heat exchangers having surface area density (β) greater than 700 m2/m3 in either
one or more sides of two-stream or multi stream heat exchanger is called as a compact heat
exchanger. Plate fin heat exchanger is a type of compact heat exchanger which is widely used in
automobiles, cryogenics, space applications and chemical industries. The plate fin heat
exchangers are mostly used for the nitrogen liquefiers, so they need to be highly efficient
because no liquid nitrogen is produced, if the effectiveness of heat exchanger is less than 87%.
So it becomes necessary to test the effectiveness of these heat exchangers before putting them in
to operation.
The available plate fin heat exchanger has rectangular offset strip geometry and is tested
in the laboratory using the heat exchanger test rig. The experiment is conducted under balanced
condition i.e. the mass flow rate for both sides of fluid stream is same, and the experiment is
carried out at different mass flow rates. The effectiveness of heat exchanger is found out for
different mass flow rates. Various correlations are available in the literature for estimation of
heat transfer and flow friction characteristics of the plate fin heat exchanger, so the various
performance parameters like effectiveness, heat transfer coefficient and pressure drop obtained
through experiments is compared with the values obtained from the different correlations. The
longitudinal heat conduction through walls decreases the heat exchanger effectiveness, especially
of cryogenic heat exchangers, so the effectiveness and overall heat transfer coefficient is found
out by considering the effect of longitudinal heat conduction using the Kroeger’s equation.
vi
LIST OF FIGURES
Figure No.
Title
Page no.
CHAPTER 1
1.1
Basic Heat Transfer Mechanism
2
1.2
Exploded View of a Plate Fin Heat Exchanger
4
1.3
Cross Flow arrangement
8
1.4
Counter Flow arrangement
8
1.5
Cross Counter Flow arrangement
9
1.6
Some of the common Fin Geometries
10
1.7
Details of Boundary layer and Flow across Offset Strip and Wavy Fin
11
2.1
Typical j and f characteristics
19
2.2
Laminar Flow on the Fins and in the Wakes
20
2.3
Laminar Flow on the Fins and Oscillating Flow on the Wakes
20
3.1
Manufacturing details of plate Fin Heat Exchanger
29
3.2
Working Mechanism of Plate Fin Heat Exchanger
30
3.3
RTDs Construction
32
3.4
RTD Calibration Graph
33
3.5
Orifice Plate
35
3.6
Schematic P&I diagram of the Experimental Test Rig
37
3.7
Photograph of Experimental test rig
38
Geometry of Typical Offset Strip Fin Surface
41
Variation of Effectiveness with Mass flow rate (hot inlet temp. = 96
50
CHAPTER 2
CHAPTER 3
CHAPTER 4
4.1
CHAPTER 5
5.1
vii
5.2
Variation of Effectiveness with Mass flow rat (hot inlet temp. = 66
50
5.3
Variation of Overall Thermal conductance with Mass flow rate (hot inlet 51
Temperature = 96 )
5.4
Variation of Overall Thermal conductance with Mass flow rate (hot inlet 51
Temperature = 66 )
5.5
Variation of Hot and Cold Effectiveness with Mass flow rate (hot inlet
52
Temperature = 96 )
5.6
Variation of Hot and Cold Effectiveness with Mass flow rate (hot inlet
52
Temperature = 66 )
5.7
Variation of Pressure Drop with Mass Flow Rate
53
viii
LIST OF TABLES
Table No.
Title
Page No.
CHAPTER 2
2.1
Chronological listings of the correlations of OSF channels
24
3.1(a)
Dimension of Procured Plate Fin Heat Exchanger
27
3.1(b)
Dimension of Procured Plate Fin Heat Exchanger
28
3.2
Procured Design data of Plate Fin Heat Exchanger
28
3.3
Calibration Chart
34
Core Data
41
5.1
Experimentally Observed Data
47
5.2
Performance of Heat Exchanger
49
CHAPTER 3
CHAPTER 4
4.1
CHAPTER 5
ix
NOMENCLATURE
NTU = no. of transfer units
Re = Reynolds number
T = Temperature
C = Specific heat
P = Pressure
U = Overall heat transfer coefficient
h = Heat transfer Coefficient
j = Colburn factor
f = Friction factor
= Effectiveness
= Efficiency
h = hot fluid
c = cold fluid
Ao = Total heat transfer surface area
1 = inlet
2 = outlet
R = Gas constant
De = Equivalent Diameter
x
CHAPTER 1
Introduction
INTRODUCTION
A heat exchanger is a device to transfer heat from a hot fluid to cold fluid across an impermeable
wall. Fundamental of heat exchanger principle is to facilitate an efficient heat flow from hot fluid
to cold fluid. This heat flow is a direct function of the temperature difference between the two
fluids, the area where heat is transferred, and the conductive/convective properties of the fluid
and the flow state. This relation was formulated by Newton and called Newton’s law of cooling,
which is given in Equation (1.1)
Q = h*A*∆T
……………… (1.1)
Where h is the heat transfer coefficient [W/m2K], where fluid’s conductive/convective
properties and the flow state comes in the picture, A is the heat transfer area [m2], and T is the
temperature difference [K].Figure. 1.1 shows the basic heat transfer mechanism
Fig. 1.1 Basic heat transfer mechanism
Heat exchangers are one of the vital components in diverse engineering plants and
systems. So the design and construction of heat exchangers is often vital for the proper
functioning of such systems. It has been shown in [Barron, 1985] that the low temperature plants
based on Linde – Hampson cycle cease to produce liquid if the effectiveness of the heat
2
exchanger is below 86.9%. On the other hand in aircrafts and automobiles, for a given heat duty,
the volume and weight of the heat exchangers should be as minimum as possible.
So the main requirement for any heat exchanger is that it should be able to transfer the
required amount of heat with a very high effectiveness. In order to increase the heat transfer in a
basic heat exchanger mechanism shown below in Figure 1.1, assuming that the heat transfer
coefficient cannot be changed, the area or the temperature differences have to be increased.
Usually, the best solution is that the heat transfer surface area is extended although increasing the
temperature difference is logical, too. In reality, it may not be much meaningful to increase the
temperature difference because either a hotter fluid should be supplied to the heat exchanger or
the heat should be transferred to a colder fluid where neither of them are usually available. For
both cases either to supply the hot fluid at high temperature or cold fluid at lower temperature
extra work has to be done. Furthermore increasing the temperature difference more than enough
will cause unwanted thermal stresses on the metal surfaces between two fluids. This usually
results in the deformation and also decreases the life span of those materials. As a result of these
facts, increasing the heat transfer surface area generally is the best engineering approach.
The above requirements have been the motivation for the development of a separate class
of heat exchangers known as Compact heat exchangers. These heat exchangers have a very high
heat transfer surface area with respect to their volume and are associated with high heat transfer
coefficients. Typically, the heat exchanger is called compact if the surface area density (β ) i.e.
heat transfer surface area per unit volume is greater than 700 m2/m3 in either one or more sides
of two-stream or multi stream heat exchanger [R.K Shah, Heat Exchangers, Thermal Hydraulic
1980]. The compact heat exchangers are lightweight and also have much smaller footprint, so
they are highly desirable in many applications.
1.1 Plate fin heat exchanger
Plate fin exchanger is a type of compact heat exchanger where the heat transfer surface
area is enhanced by providing the extended metal surface interface between the two fluids and is
called as the fins. Out of the various compact heat exchangers, plate-fin heat exchangers are
unique due to their construction and performance. They are characterized by high effectiveness,
compactness, low weight and moderate cost. As the name suggests, a plate fin heat exchanger
(PFHE) is a type of compact exchanger that consists of a stack of alternate flat plates called
3
parting sheets and corrugated fins brazed together as a block. Streams exchange heat by flowing
along the passages made by the fins between the parting sheets. Separating plate acts as the
primary heat transfer surface and the appendages known as fins act as the secondary heat transfer
surfaces intimately connected to the primary surface. Fins not only form the extended heat
transfer surfaces, but also work as strength supporting member against the internal pressure. The
side bars prevent the fluid to spill over and mix with the second fluid. The fins and side bars are
brazed with the parting sheet to ensure good thermal link and to provide the mechanical stability.
Figure. 1.2 shows the exploded view of two layers of a plate fin heat exchanger. Such layers are
arranged together in a monolithic block to form a heat exchanger.
Fig. 1.2 Exploded view of a plate fin heat exchanger
1.1.1 Advantages and Disadvantages
Plate fin heat exchangers offer several advantages over the other heat exchangers:
1. Compactness: Large heat transfer surface area per unit volume (Typically 1000 m2/m3),
is usually provided by the plate fin heat exchanger. This in turn produces a high overall
heat transfer coefficient due to the heat transfer associated with the narrow passages and
corrugated surfaces.
2. Effectiveness: very high thermal effectiveness more than 95% can be obtained.
3. Temperature control: The plate heat exchanger can operate with relatively small
temperature differences. A close temperature approach (Temperature approach as low as
3K between single phase fluid streams and 1K between boiling and condensing fluids is
4
fairly common.),This is an advantage when high temperatures must be avoided. Local
overheating and possibility of stagnant zones can also be reduced by the form of the flow
passage.
4. Flexibility: Changes can be made to heat exchanger performance by utilizing a wide
range of fluids and conditions that can be modified to adapt to the various design
specifications. Multi steam operation is possible upto 10 streams.
5. True counter-flow operation (Unlike the shell and tube heat exchanger, where the shell
side flow is usually a mixture of cross and counter flow.).
The main disadvantages of a plate fin heat exchanger are:
1. Limited range of temperature and pressure.
2. Difficulty in cleaning of passages, which limits its application to clean and relatively noncorrosive fluids, and
3. Difficulty of repair in case of failure or leakage between passages.
1.1.2 Materials
Plate fin heat exchangers are generally, made from an alloy of aluminum or stainless
steel. However, the process temperature and pressure dictates the choice of the material.
Aluminum alloys are particularly suitable for low temperature applications because of their low
weight and excellent ductility and increasing strength under such conditions. In general, the fins
or secondary surfaces and the side bars are usually joined to the separating plate by using dip
brazing technology or more recently vacuum brazing technique. The brazing material in case of
aluminum exchangers is an aluminum alloy of lower melting point, while that used in stainless
steel exchangers is a nickel based alloy with appropriate melting and welding characteristics.
1.1.3 Manufacture
The basic principles of plate fin heat exchanger manufacture are the same for all sizes
and all materials. The heat exchanger is assembled from a series of flat sheets and corrugated
fins in a sandwich construction. Separating plates (i.e. parting sheets) provide the primary heat
transfer surface. Separating plates are positioned alternatively with the layers of fins in the stack
to form the containment between individual layers. These elements i.e. the corrugations, sidebars, parting sheets and cap sheets are now held together in a jig under a predefined load, and
5
placed in a furnace and brazed to form the plate fin heat exchanger block. After this the header
tanks and nozzles are welded to the block, taking care that the brazed joints remain intact during
the welding process. Differences arise in the manner in which the brazing process is carried out.
The methods in common use are salt bath brazing and vacuum brazing. In the salt bath process,
the stacked assembly is preheated in a furnace to about 5500 C, and then dipped into a bath of
fused salt composed mainly of fluorides or chlorides of alkali metals. The molten salt works as
both flux and heating agent, maintaining the furnace at a uniform temperature. In case of heat
exchangers made up of aluminum, the molten salt removes grease and the tenacious layer of
aluminum oxide, which would otherwise weaken the joints. Brazing takes place in the bath when
the temperature is raised above the melting point of the brazing alloy. The brazed block is
cleansed of the residual solidified salt by dissolving in water, and is then thoroughly dried.
In the vacuum brazing process, no flux or separate pre-heating furnace is required. The
assembled block is heated to brazing temperature by radiation from electric heaters and by
conduction from the exposed surfaces into the interior of the block. The absence of oxygen in the
brazing environment is ensured by application of high vacuum (Pressure ≈ 10-6 Mbar). The
composition of the residual gas is further improved (lower oxygen content) by alternate
evacuation and filling with an inert gas as many times as experience dictates. No washing or
drying of the brazed block is required. Many metals, such as aluminum, stainless steel, copper
and nickel alloys can be brazed satisfactorily in a vacuum furnace.
1.1.4 Applications
The plate-fin heat exchanger is suitable for use over a wide range of temperatures and
pressures for gas-gas, gas-liquid and multi-phase duties. They are used in a variety of
applications. They are mainly employed in the field of cryogenics for cryogenic separation and
liquefaction of air, natural gas processing and liquefaction, production of petrochemicals and
large refrigeration systems. The exchangers that are used for cryogenic air separation and LPG
fractionation are the largest and most complex units of the plate fin type and a single unit could
be of several meters in length. Brazed aluminum plate fin exchangers are widely used in the
aerospace industries because of their low weight to volume ratio and compactness. They are
being used mainly in environment control system of the aircraft, avionics and hydraulic oil
cooling and fuel heating. Making heat exchangers as compact as possible has been an everlasting
6
demand in automobile and air conditioning industries as both are space conscious. In the
automobile sector they are used for making the radiators. The other miscellaneous applications
are:
1. Fuel cells
2. Process heat exchangers.
3. Heat recovery plants.
4. Pollution control systems
5. Fuel processing and conditioning plants.
6. Ethylene and propylene production plants.
1.1.5 Flow arrangement
A plate fin heat exchanger can have two or more than two streams, which may flow in
directions parallel or perpendicular to one another. When the flow directions are parallel, the
streams may flow in the same or in opposite sense. So there are three primary flow arrangements
for a plate fin heat exchanger – (i) parallel flow, (ii) counter-flow and (iii) cross flow.
Thermodynamically, the counter-flow arrangement provides the highest heat (or cold) recovery,
while the parallel flow geometry gives the lowest. While the cross flow arrangement, gives an
intermediate thermodynamic performance, by offering superior heat transfer properties and
easier mechanical layout. Under some circumstances, a hybrid cross – counter-flow geometry
provides greater heat (or cold) recovery with superior heat transfer performance. Thus in general
engineering practice, there are three main configurations for the plate fin heat exchangers: (a)
cross flow, (b) counter-flow and (c) cross-counter flow.
(a) Cross flow:
In this type of heat exchangers as shown in the Fig. (1.3) the fluids flow in directions normal
to each other. Thermodynamically the effectiveness for cross flow heat exchangers falls in between
that for the counter flow and parallel flow arrangements. The largest structural temperature
difference exists at the corner of the entering hot and cold fluids. Only two streams are handled, in a
cross flow type of a heat exchanger which eliminates the need for distributors. For this type of heat
exchangers the header tanks are located on all four sides of the heat exchanger core, making this
arrangement simple and cheap. If high effectiveness is not necessary, and if the two fluid streams
have widely differing volume flow rates, or if either one or both streams are nearly isothermal (as in
7
single component condensing or boiling), then the cross flow arrangement should be preferred.
Typical applications include automobile radiators and some aircraft heat exchangers. (Fig.1.3 shows
a cross flow arrangement).
Fig.1.3 Cross flow arrangement
(b) Counter flow:
In a counter flow heat exchanger the two fluids flow parallel to each other but in opposite
directions. The counter-flow heat exchanger provides the most thermally effective arrangement for
recovery of heat or cold from process streams. A counter flow arrangement is thermodynamically
superior to any other flow arrangement. It is the most efficient flow arrangement, producing the
highest temperature change in each fluid compared to any other two-fluid arrangement for a given
overall thermal conductance (UA), fluid flow rates and fluid inlet temperatures. Cryogenic
refrigeration and liquefaction equipment use this geometry almost exclusively. But this type of heat
exchangers demands proper design because of the complex geometry of headers. (Fig. 1.4 shows a
counter flow arrangement for heat exchanger)
Fig.1.4 Counter flow arrangement
8
(c) Cross-Counter flow:
The cross-counter flow geometry is a hybrid of counter-flow and cross flow
arrangements, delivering the thermal effectiveness of counter-flow heat exchanger with the
superior heat transfer characteristics of the cross flow configuration. In this arrangement, one of
the streams flows in a straight path, whereas the second stream follows a zigzag path normal to
that of the first stream. While moving along the zigzag path, the second fluid stream covers the
length of the heat exchanger in a direction opposite to that of the direct stream. Thus the flow
pattern can be assumed to be globally counter-flow while remaining locally cross flow. Crosscounter flow PFHEs are used in applications similar to those of simple cross flow exchangers,
but they allow more flexibility in design and fabrication. They are particularly suited for the
applications where the two streams have considerably different volume flow rates, or permit
significantly different pressure drops. The fluid with the larger volume flow rate or that with the
smaller value of allowable pressure drop is made to flow through the straight channel, while the
other stream follows the zigzag path. For example, in a liquid-to-gas heat exchanger, the gas
stream with a large volume flow rate and low allowable pressure drop is assigned the straight
path, while the liquid stream with a high allowable pressure drop flows normal to it over a zigzag
path. This arrangement optimizes the overall geometry. (Fig.1.5 shows a cross-counter flow
arrangement for heat exchanger)
Fig.1.5 Cross-counter flow arrangement
9
1.2 Plate Fin heat transfer surfaces
The plate fin exchangers are mainly employed for liquid-to-gas and gas-to-gas
applications. Due to the low heat transfer coefficients in gas flows, extended surfaces are
commonly employed in plate-fin heat exchangers. By using specially configured extended
surfaces, heat transfer coefficients can also be enhanced. While such special surface geometries
provide much higher heat transfer coefficients than plain extended surfaces, but at the same time,
the pressure drop penalties are also high, though they may not be severe enough to negate the
thermal benefits. A variety of extended surfaces like the plain trapezoidal, plain rectangular
shown in Fig. 1.6 can perform such function, and we have included the offset strip fin geometry
in our present work.
Fig.1.6 some of the common fin geometries
In order to improve the gas side coefficients, surface features are
needed to provided on the gas side coefficients. These features may be divided into two
categories: the first, in which the surface remains continuous (wavy and herring-bone fins) and
the second in which it is cut (offset, louvered). In a continuous type fin, the corrugations causes
the gas to make sudden direction changes so that locally, the velocity and temperature gradients
10
are increased (Figure 1.7). This results in local enhancement of heat transfer coefficient. But an
undesirable consequence of such enhancement in heat transfer coefficient is an increase in the
friction factor and pressure drop
Fig 1.7 Details of boundary layer and flow across offset strip and wavy fin.
Whereas in a discontinuous type of fin geometry, boundary layers are interrupted which would
form on a continuous plate. Adjacent to the leading edge of the fin, both heat transfer coefficients
and friction factors are very much high due to generation of fresh boundary layers. But in
addition to this friction drag, form drag is also formed due to the finite thickness of the fin.
Although, the friction drag is associated with high heat transfer coefficient form drag has no
counterpart and represents the form of wasted energy. The form drag could be substantial
depending on the quality of the cutting edge. However, machined-formed fins are generally free
from this problem. Brief descriptions of application and associated mechanism of the extended
surfaces or fins depicted in Figure. 1.6 are given below.
1.2.1 Plain Fins
Plain fins are by far the most common of all compact cores or surfaces used in compact
heat exchangers. The plain- fin surfaces are characterized by long uninterrupted flow passages,
with performance similar to that obtained inside long circular tubes (Kays and London, 1984).
Although passages of triangular and rectangular cross section are more common, any desired
shape can be given to the fins, considering only manufacturing constraints. Straight fins in
triangular arrangement can be manufactured at high speeds and hence are less expensive than
rectangular fins. But generally they are structurally weaker than rectangular fins for the same
passage size and fin thickness. They also have lower heat transfer performance compared to
rectangular fins, particularly in laminar flow.
11
Plain fins are used in those applications where core pressure drop is critical. Their
application range from aerospace air conditioning duties to oil refining (Hesselgreaves 2001),
among many others. An heat exchanger with plain fins requires a smaller flow frontal area than
that with interrupted fins for specified pressure drop, heat transfer and mass flow rate. Of course,
the required passage length is higher leading to a larger overall volume. The heat transfer
enhancement achieved with plain fins results mainly from increased area density, rather than any
substantial rise in the heat transfer coefficient (Brockmeier, et al. 1993).
1.2.2 Wavy Fins
Wavy fins are uninterrupted fin surfaces with cross-sectional shapes similar to those of
plain fins, but with cyclic lateral shifts perpendicular to the flow direction. The resulting wave
form provides effective interruptions which causes the flow direction to change periodically and
induces a complex flow field. Consequently, the boundary layer separates and reattaches
periodically around the trough regions to promote enhanced heat transfer; increased pressure
drop penalty is also accompanied. Actually the Heat transfer is enhanced due to creation of
Goertler vortices. These counter-rotating vortices form while the fluid passes over the concave
wave surfaces, and produce a corkscrew-like flow pattern.
The heat transfer and pressure drop characteristics of a wavy fin surface lie between those of
plain and offset strip fins. The friction factor continues to fall with increasing Reynolds number.
Wavy fins are common in the hydrocarbon industry where exchangers are designed with high
mass velocities and moderate thermal duties. Unlike offset strip fins, the thickness of wavy fins
is not limited at high fin densities. Therefore, wavy fins are often used for streams at high
pressure, particularly those which can tolerate somewhat poor heat transfer coefficient.
1.2.3 Offset Strip fins
This is the most widely used fin geometry in high performance plate fin heat exchangers.
It consists of a type of interrupted surface, which may be visualized as a set of plain fins cut
normal to the flow direction at regular intervals, each segment being offset laterally by half the
fin spacing. Typical strip lengths are 3-6mm, and the Reynolds number based on strip length is
very small, which makes the flow to be always in laminar regime. The laminar boundary layer
develops on the short strip length, and then dissipates in the wake region between successive
12
offset strips. Surface interruption enhances heat transfer by two independent mechanisms. First,
it prevents the continuous growth of thermal boundary layer by periodically interrupting it. The
thinner boundary layer offers lower thermal resistance compared to continuous fin types. Above
a critical Reynolds number, interrupted surfaces offer an additional mechanism of heat transfer
enhancement. Oscillations in the flow field in the form of vortices shed from the trailing edges of
the interrupted fins enhance local heat transfer by continuously bringing in fresh fluid towards
the heat transfer surfaces but this enhancement is accompanied by an increase in pressure drop.
Considerable heat transfer enhancement is achieved compared to that of the plain fin. The heat
transfer performance of an offset strip fin is often as much as 5 times that of a plain fin surface of
comparable geometry, but at the expense of higher pressure drop. For specified heat transfer and
pressure drop requirements, the offset strip fin surface demands a somewhat higher frontal area
compared to those with plain fin, but results in a shorter flow length and lower overall volume. It
is believed that with the shorter strip lengths, the better heat transfer performance is achieved
(Manglik and Bergles, 1995).
An undesirable characteristic of this type of fin is that at high Reynolds numbers the friction
factor remains nearly constant (because of the higher contribution of form drag), while the heat
transfer performance goes down. Therefore, offset strip fins are used less frequently in very high
Reynolds number applications. On the other hand, they are extensively used in air separation and
other cryogenic applications where mass velocities are low and high thermal effectiveness is
essential.
1.2.4 Louvered Fin
The louvered fin geometry bears a similarity to the offset strip fin. Instead of shifting the
slit strips laterally, small segments of the fin are slit and rotated 20 to 45 degrees relative to the
flow direction. Actually the fin surfaces are cut and bent out into the flow stream at frequent
intervals in a louver or lanced-like fashion. The purpose of this fin surface louvering like a
“vinetia n blind” is to break up the boundary layers so as to yield high heat transfer coefficients
when compared to those in plain fins under same flow conditions (Kays and London, 1984). It
has been contended that the performance of the louvered fins is similar to or better than offsetstrip fins. The flow structure in the louvered fin flow passage is dependent on the flow rate Re at
very low flow rate, the main flow stream does not pass through the louvers, whereas at high flow
13
rate the flow becomes nearly parallel to the louvers (Webb, 1994, and Davenport 1983). The
base surface of the louvered fin geometry can be of triangular or rectangular shape, and louvers
can be cut in many different forms.
The multi-louvered fin has the highest heat transfer enhancement relative to pressure drop
in comparison with most other fin types. Flow over louvered fin surfaces is similar in nature to
that through the offset strip fin geometry, with boundary layer interruption and vortex shedding
playing major roles. An important aspect of louvered fin performance is the degree to which the
flow follows the louver. Louvered fins are widely used in automotive heaters and radiators,
where the latter is configured as a tube- fin exchanger.
1.2.5 Perforated Fins
This surface geometry is made by punching a pattern of spaced holes in the fin material
before it is formed into flow channels. The channels may be triangular or rectangular in shape
with either round or rectangular perforations.
If the porosity of the resulting surface is
sufficiently high, enhancement can occur due to boundary layer dissipation in the wake region
formed by the holes (Webb, 1994). The performance of the perforated fin is less than that of a
good offset strip fin, and thus the perforated fin is rarely used today. Perforated fins are now used
only in limited number of applications such as turbulators in oil coolers. Furthermore, the
perforated fin represents a wasteful way of making an enhanced surface, since the material
removed in making the perforated hole is wasted.
1.2.6 Pin Fin
In a pin fin exchanger, a large number of small pins are sandwiched between plates in
either an inline or staggered arrangement. Pins may have a round, an elliptical, or a rectangular
cross section. Due to their low compactness and high cost per unit surface area compared to
multi-louvered or offset strip fins these types of finned surfaces are not widely used these days.
Due to vortex shedding behind the pins, noise and flow-induced vibration are produced, which
are generally not acceptable in most heat exchanger applications. The potential application of pin
fin surfaces is at low flow velocities (Re < 500), where pressure drop is negligible. Pin fins are
used as electronic cooling devices with free-convection flow on the pin fin side.
14
1.3 Heat transfer and Flow Friction Characteristics
Accurate and reliable dimensionless heat transfer and pressure drop characteristics are a
key input for designing or analyzing a plate fin heat exchanger. For single-phase flow, the heat
transfer coefficient is generally expressed in terms of the Colburn correlation [Kern and Kraus,
1972]
h = j Cp G (Pr)-2/3
(1.2)
Where j called as colburn factor separates the effects of the fluid properties on the heat
transfer coefficient and permits correlations as a function of the Reynolds number (Re). While
the j data are expressed as functions of Prandtl number (Pr) and Re, temperature does not appear
directly in the expression. Temperature has the only role in determining the thermo-physical
properties such as density, viscosity, specific heat and thermal conductivity. Therefore, it is
generally recognized that j data determined at one temperature / pressure level and expressed in
dimensionless form are directly usable at another temperature / pressure level.
Since the plate fin heat exchangers are mainly used for gas to gas heat transfer applications and
most of the gases are low density gases, so the pumping power requirement in a gas-to-gas heat
exchanger is high as compared to that in a liquid-to-liquid heat exchanger. This fact makes it
mandatory to have an accurate estimation of friction characteristics of the heat exchanger
surfaces in gas application. The friction factor is defined on the basis of an equivalent shear force
in the flow direction per unit friction area. This shear force can be either viscous shear (skin
friction) or pressure force (form drag) or a combination of both. So without making an attempt to
differentiate between them, it is possible to express them by Fanning friction factor (f) given by
/ (1.3)
While equation (1.3) is the basic definition of friction factor, the pressure drop (∆P) for
internal flow through the ducts can be calculated from equation (1.4)
∆ (1.4)
It can be seen that temperature does not appear directly in the expression of friction factor
also. Therefore, the f data determined at one temperature / pressure level are directly usable at
15
other temperature / pressure level. But it is seen that j and f are strong functions of fin geometries
like fin height, fin spacing, fin thickness etc. Because fins are available in varied shapes, it
becomes necessary to test each configuration individually to determine the heat transfer and flow
friction characteristics for specific surface. For a given fin geometry, in general, increase in heat
transfer performance is associated with increase in flow friction and vice versa. Customarily the
ratio of j/f is taken as a measure of the goodness of the fin surface. Though the preferred fin
geometry would have high heat transfer coefficient without correspondingly increased pressure
penalty, the selection of particular fin geometry mainly depends on the process requirement; one
can sacrifice either of heat transfer or pressure loss at the cost of other.
The monograph Compact Heat Exchangers by Kays and London [1] remains one of the
earliest and the most authoritative sources of experimental j and f data on plate fin surfaces .
Although nearly two decades have passed after the latest edition, there has not been any
significant addition to this database in open literature. After that several attempts have been
made towards the numerical prediction of heat transfer coefficient and friction factor; but they
have generally been unable to match experimental data. Several empirical correlations, which
have been generated from the data of Kays and London, have found extensive application in
industry, particularly in less-critical designs. For critical applications, direct experimental
determination of j and f factors for each fin geometry remains the only choice.
In a plate fin heat exchanger the common range of Reynolds number is 500 to 3000 for
most of the applications. The Reynolds number is kept low because the hydraulic diameter of the
flow passages is generally small due to closely spaced fins and in such conditions operation with
low density gases leads to excessive pressure drop unless the gas velocity in the flow passage is
kept low.
1.4 Objectives of the study
The main objective of the present work is to evaluate the performance parameters of a
counter flow plate fin heat exchanger through hot testing, which includes1. Design and fabrication of the test rig for plate fin heat exchanger.
2. To determine the thermal performance parameters like overall heat transfer coefficient,
effectiveness and pressure drop of plate fin heat exchanger through hot testing under
balanced flow condition.
16
3. To compare the experimentally obtained values of effectiveness, overall heat transfer
coefficient with the values that are obtained from various correlations.
1.5 Organization of the Thesis
Thesis contains six chapters including references
Chapter 1 deals with the general introduction of the compact plate fin heat exchanger and the
scope of the future work.
In Chapter 2, a brief literature review on the topics related to the present work has been reported,
where emphasis has been laid on the literatures related to the prediction of j and f factors and the
thermal performance testing of heat exchangers.
A detailed description of the experimental setup has been given in Chapter 3. The design and
fabrication of experimental test rig and instrumentation are discussed in this chapter.
The basic steps involved in Rating, which is used for the calculation of performance parameters
of plate fin heat exchanger with offset strip fin geometry, like effectiveness, overall heat transfer
coefficient of an already designed heat exchanger is discussed in detail in Chapter 4.
In Chapter 5, the experimental results and detailed procedure for computing the performance
parameters of the heat exchanger in balanced condition are presented.
Finally the Chapter 6 is devoted for the concluding remarks and scope for future work.
17
CHAPTER 2
Literature Survey
LITERATURE SURVEY
Heat exchangers constitute the most important components of many industrial processes
and equipment’s covering a wide range of engineering applications. Increasing awareness for the
effective utilization of energy resources, minimizing operating cost and maintenance free
operation have led to the development of efficient heat exchangers like compact heat exchangers.
R.K Shah[15] in his elaborate discussion over the classification of heat exchangers has defined
the “compact heat exchangers” as one having a surface area density of more than 700 m2/m3.
Such compactness is achieved by providing the extended surfaces i.e. fin on the flow passages
which work as the secondary heat transfer area.
The main purpose of a recuperative heat exchanger is to facilitate the effective exchange
of thermal energy between the two fluids flowing on the either side of a solid portioning wall,
during which both the streams experience some viscous resistance and led to pressure drop. So in
any heat exchanger the information regarding the quantity oh heat transfer and pressure drop are
of utmost importance. Heat transfer and pressure drop characteristics of heat exchanger are
mainly expressed in terms of j and f factor respectively.
A large amount of study has been conducted to analyze the heat transfer and pressure drop
characteristics of compact heat exchangers in the past few decades. But this study mainly focuses
on the OSFs type of plate fin heat exchanger. And therefore the emphasis has been given on the
literatures related to the prediction of j and f factors and the thermal performance testing of heat
exchangers.
Patankar and Prakash [1] presented a two dimensional analysis for the flow and heat
transfer in an interrupted plate passage which is an idealization of the OSFs heat exchanger. The
main aim of the study is investigating the effect of plate thickness in a non-dimensional form t/H
on heat transfer and pressure drop in OSF channels because the impingement region resulting
from thick plate on the leading edge and recirculating region behind the trailing edge are absent
if the plate thickness is neglected. Their calculation method was based on the periodically fully
developed flow through one periodic module since the flow in OSF channels attains a periodic
fully developed behaviour after a short entrance region, which may extend to about 5 (at the
most 10) ranks of plates (Sparrow, et al. 1977). Steady and laminar flow was assumed by them
between Reynolds numbers 100 to 2000. They found the flow to be mainly laminar in this range,
18
although in some cases just before the Reynolds no. 2000 there was a transition from laminar to
turbulence. Especially for the higher values of t/H. They used the constant heat flow boundary
condition with each row of fins at fixed temperature. They made there analysis for different fin
thickness ratios t/H= 0, 0.1, 0.2, 0.3 for the same fin length L/H = 1, and they fixed the Prandtl
number of fluid = 0.7. For proper validation they compared there numerical results with the
experimental results of [ London and Shah] for offset strip fin heat exchangers. The result
indicate reasonable agreement for the f factors, but the predicted j factor are twice as large as the
experimental data. They concluded that the thick [plate situation leads to significantly higher
pressure drop while the heat transfer does not sufficiently improve despite the increased surface
area and increased mean velocity.
Joshi and Webb [2] developed an analytical model to predict the heat transfer coefficient
and the friction factor of the offset strip fin surface geometry. To study the transition from
laminar to turbulent flow they conducted the flow visualization experiments and an equation
based on the conditions in wake was developed.
Fig. 2.1 typical j and f characteristics
They also modified the correlations of Weiting [17]. There was some difference between there
correlation.
19
Fig 2.2 Laminar flow on the fins and in the wakes
(Source Joshi and Webb 1987)
Fig. 2.3 Laminar flow on the fins and oscillating flow on the wakes
(Source Joshi and Webb 1987)
Four different flow regimes (Figure. 2.2 and 2.3) were identified by Joshi and Webb [2]
from there experiment. The flow was found to be laminar and steady in the first regime. In the
second regime the oscillating flow structures were found in the transverse direction. The flow
oscillated in the wake region between two successive fins in the third regime. And in the fourth
regime the effect of vortex shedding came into picture. The laminar flow correlation of Joshi and
Webb started to under predict the j and f factors at the second regime. So they assumed the
Reynolds number at that point as the critical Reynolds number to identify the transition from
laminar to turbulent.
20
Suzuki et al [3] in order to study the thermal performance of a staggered array of vertical
flat plates at low Reynolds number has taken a different numerical approach by solving the
elliptic differential equations governing the flow of momentum and energy. The validation of
their numerical model has been done by carrying out experiments on a two dimensional system,
followed by those on a practical offset strip fin heat exchanger. The experimental result was in
good agreement with the performance study for the practical offset-strip-fin type heat exchanger
in the range of Reynolds number of Re<800.
Tinaut et al [4] developed two correlations for heat transfer and flow friction coefficients
for OSFs and plane parallel plates. The working fluid for OSF was engine oil and water was
taken for analyzing the parallel plate channels. By using the correlations of Dittus and Boelter
and some expressions of Kays and Crawford they obtained there correlations. For the validation
of their results they compared there correlations with correlations of Weitng [17]. Although there
were some differences between the results but there correlations have been found acceptable
upon comparing their results to the data obtained from other correlations.
Manglik and Bergles[5] carried an experimental research on OSFs. They investigated the
effects of fin geometries as non dimensional forms on heat transfer and pressure drop, for their
study they used 18 different OSFs. After their analysis they arrived upon two correlations, one
for heat transfer and another one for pressure drop. The correlations were developed for all the
three regions. They compared there results from the data obtained by other researchers in the
deep laminar and fully turbulent regions. There correlations can be acceptable when comparing
the results of the expressions to the experimental data obtained by Kays and London [16].
Hu and Herold [6] presented two papers to show the effect of Prandtl no. on heat transfer
and pressure drop in OSF array. Experimental study was carried out in the first paper to study the
effect for which they used the seven OSFs having different geometries and three working fluids
with different Prandtl number. At the same time the effect of changing the Prandtl number of
fluid with temperature was also investigated. The study was carried out in the range of Reynolds
number varying from 10 to 2000 in both the papers. The results of the two studies showed that
the Prandtl number has a significant effect on heat transfer in OSF channel. Although there is no
effect on the pressure drop.
Zhang et al [7] investigated the mechanisms for heat transfer enhancement in parallel
plate fin heat exchangers including the inline and staggered arrays of OSFs. They have also taken
21
into account the effect of fin thickness and the time dependent flow behavior due to the vortex
shedding by solving the unsteady momentum and energy equation. The effect of vortices which
are generated at the leading edge of the fins and travel downstream along the fin surface was also
studied. From there study they found that only the surface interruptions increase the heat transfer
because they cause the boundary layers to start periodically on fin surfaces and reduce the
thermal resistance to transfer heat between the fin surfaces and fluid. However after a critical
Reynolds number the flow becomes unsteady and in this regime the vortices play a major role to
increase the heat transfer by bringing the fresh fluids continuously from the main stream towards
the fin surface.
Dejong et al [8] carried out an experimental and numerical study for understanding the
flow and heat transfer in OSFs. In the study the pressure drop, local Nusselt number, average
heat transfer and skin friction coefficient on fin surface, instantaneous flow structures and local
time averaged velocity profiles in OSF channel were investigated. They compared there results
with the experimental results obtained by Dejong and Jacobi [1997] and unsteady numerical
simulation of Zhang et al [1997]. There results indicate that the boundary layer development,
flow separation and reattachment, wake formation and vortex shedding play an important role in
the OSF geometry.
H. Bhowmik and Kwan-Soo Lee [9] studied the heat transfer and pressure drop
characteristics of an offset strip fin heat exchanger. For their study they used a steady state three
dimensional numerical model. They have taken water as the heat transfer medium, and the
Reynolds number (Re)in the range of 10 to 3500. Variations in the Fanning friction factor f and
the Colburn heat transfer j relative to Reynolds number were observed. General correlations for
the f and j factors were derived by them which could be used to analyze fluid flow and heat
transfer Characteristics of offset strip fins in the laminar, transition, and turbulent regions of the
flow.
Saidi and Sudden [10] carried out a numerical analysis of the instantaneous flow and heat
transfer for OSF geometries in self-sustained time-dependent oscillatory flow. The effect of
vortices over the fin surfaces on heat transfer was studied at intermediate Reynolds numbers
where the flow remains laminar, but unsteadiness and vortex shedding tends to dominate. They
compared there numerical results with previous numerical and experimental data done by
Dejong, et al. (1998).
22
From the studies of few researchers like Patankar and Prakash [1], Kays and London [16]
it is easy to get information regarding the effects of OSFs on heat transfer and pressure drop. But
most of the researchers have not taken into account the effect of manufacturing irregularities
such as burred edges, bonding imperfections, separating plate roughness which also affect the
heat transfer and flow friction characteristics oh the heat exchanger. Dong et al [11] made
experiments and analysis considering the above factors to get better thermal and hydraulic
performance from the OSFs. Sixteen types of OSFs and flat tube heat exchangers were used to
make the experimental studies on heat transfer and pressure drop characteristics. A number of
tests were made by changing the various fin parameters and all the tests were carried out in
specific region of air side Reynolds number (500- 7500), at a constant water flow crate. The
thermal performance data were analyzed using the effectiveness-NTU method in order to obtain
the heat transfer coefficient. They also derived the j factor and f factor by using regression
analysis. Results showed that the heat transfer coefficient and pressure drop reduce with
enlarging the fin space, fin height and fin length.
Michna et al [12] investigated the effect of increasing Reynolds number on the
performance of OSFs. He conducted the experiment at Reynolds number between 5000 to
120000 and found that both heat transfer and pressure drop increased with increasing Reynolds
number, because the effect of vortex shedding and eddy formation at turbulent regime. Operation
of OSF heat exchangers under these Reynolds number may be useful in systems where
minimizing the heat exchanger size or maximising the heat transfer coefficient is more important
than minimising the pressure drop.
Various experiments are carried out in order to find out the j and f factors of the various
heat exchangers and are called as the thermal performance testing. These testing are needed for
heat exchangers, which do not have reported j and f data. Therefore, this test is conducted for
any new development or modification of the finned surfaces. T. Lestina & K. Bell, Advances in
Heat Transfer, told for heat exchangers already existing in the plants this test is done for the
following reasons:
a) Comparison of the measured performance with specification or manufacturing design rating
data.
b) Evaluation of the cause of degradation or malfunctioning.
c) Assessment of process improvements such as those due to enhancement or heat exchanger
23
replacement.
Another reason for developing these correlations is that, generally in most heat exchanger
problems the working fluid, the heat flow rate and mass flow rate are usually known, so if certain
correlations between geometry and fin performance is also known, then the problem can be
greatly simplified. For that purpose developing the correlations for fanning friction factor f and
Colbourn factor j are important for heat exchanger. Some of the correlations and their
investigators are given in Table 2.1. Generally the correlations include three distinct nondimensional ratios depending on OSF geometry. These are the ratio of free flow area ( = s/h),
the ratio of heat transfer area (β= t/l), and the ratio of fin density ( = t/s).
Table 2.1 Chronological listing of the correlations for OSF channels
Investigators
Correlation
j
Manson
0.61⁄D" $.% Re$.% ( , 1⁄D" * 3.5
$.% (
0.321Re
1⁄D" . 3.5
,
/
For Re<3500:
11.81⁄D" Re$.23 ( , 1⁄D" * 3.5
/
f
$.23
3.371Re ( ,
1⁄D" . 3.5
For Re>3500:
f
11.81⁄D" Re$.23 ( , 1⁄D" * 3.5
$.23 (
3.371Re
,
1⁄D" . 3.5
/
Where D" 2sh⁄s 7 h
j 0.665Re9$.%
8
Kays
f 0.44t⁄I 7 1.328Re9$.%
8
Re<1000:
j 0.4831⁄D" 9$.2 α9$.>? Re9$.%@2
24
f 7.6611⁄D" [email protected]>? α9$.$A Re9$.3
Wieting
Re>2000:
j 0.2421⁄D" [email protected] t⁄D" 9$.$>A [email protected]>
f 1.1361⁄D" 9$.3> t⁄D" 9$.%@? Re9$.A>
Where D" 2sh⁄s 7 h
Re< Re*
j 0.53Re9$.% 1⁄D" 9$.% α9$.?
f 8.12Re9$.3? 1⁄D" 9$.? α9$.$
Re> Re*+1000
Joshi and Webb:
j 0.21Re9$.? 1⁄D" 9$.? t⁄D" 9$.$
f [email protected] 1⁄D" 9$.3> t⁄D" 9$.3
Where DB 2s C thHDs 7 h 7 I G
E"
I [email protected]
ReI 257 J L
K
E $.%>
JL
I
9
RO 9$.%
D" Mt 7 1.328 JID L
Q
R
Re<2000:
j 1.37I⁄D" 9$.% α9$.>? Re9$.23
f 5.55I⁄D" [email protected] α9$..$A Re9$.23
Mochizuki et al.
Re>2000:
j 1.17I⁄D" 7 3.759E t⁄D" $.$>A [email protected]
f 0.83I⁄D" 7 0.339$.% t⁄D" $.%@? Re9$.$
Where
D" 2sh⁄s 7 h
j 0.6522Re9$.%[email protected] α9$.%? δ9$.?AA γ.9$.$23>
X1 7 5.269X109% [email protected]?$ α$.%$? δ$.?%2 γ9.$%% ($.
Manglik and Bergles
f 9.6243Re9$.3? α9$.>%2 δ[email protected]$%@ γ.9$.2%A
X1 7 7.669X109> Re?.?A α$.A$ δ@.3232 γ[email protected] ($.
25
Where
D" 4shl⁄2sl 7 hl 7 th 7 ts(
The plate fin heat exchangers find a variety of applications in the field of cryogenics,
where high heat transfer performance and high effectiveness are the foremost requirement. But
there are many factors which affect their performance, like flow maldistribution, heat in leak
from the atmosphere and wall longitudinal heat conduction. Prabhat Gupta, M.D. Atrey [13],
have evaluated the performance of a counter-flow heat exchangers for low temperature
applications by considering the effect of heat in leak and longitudinal conduction. They
developed a numerical model considering the effect of heat in leak and the longitudinal wall heat
conduction and made predictions which were compared with the experimental results to
understand the quantitative effect of heat in leak and axial conduction parameters on degradation
of heat exchanger performance. From there study it was found that in addition to operating and
design parameters, the thermal performance of these heat exchangers is strongly governed by
various losses such as longitudinal conduction through wall, heat in leak from surrounding, and
flow maldistribution, etc.
Randall F Barron, Cryogenic heat exchanger [14], has showed the effect of longitudional
wall heat conduction on the performance of cryogenic heat exchangers. Cryogenic heat
exchangers operate at low temperatures where the longitudinal wall heat conduction results in
serious performance deterioration these is because they have small distances ( on the order of
100 to 200 mm or 4 to 8 in) between the warm and cold ends i.e. they have short conduction
lengths. Because of the inherent requirement of high effectiveness for cryogenic heat exchangers,
the NTU values are usually large (as high as 500 to 1000), so the effect of longitudinal
conduction is most pronounced for heat exchangers having short conduction lengths and large
NTU. The wall longitudinal heat conduction reduces the local temperature difference between
the two streams, thereby reducing the heat exchanger effectiveness and the heat transfer rate.
26
CHAPTER 3
Experimental Setup and
Procedure
EXPERIMENTAL SETUP AND PROCEDURE
3.1 Detailed Description of various Equipment’s and Instruments used
3.1.1 Plate fin heat exchanger
The test section consists of a counter flow plate fin heat exchanger with offset strip fin
geometry. This Plate Fin Heat Exchanger was sent here for its performance analysis and to
establish the correlation for j and f factors, which is manufactured by APOLLO heat exchangers
Mumbai, for BARC Mumbai. Figure 3.3 shows the plate fin heat exchanger with all its
dimensions and arrangements of Inlet and Outlet ports. This plate fin heat exchanger consists of
offset strip fins. And table 3.1 and 3.2 provides the details of core dimensions and thermal data
respectively.
This Project is basically an experimental set-up, which is build up for the thermal
performance testing of the plate fin heat exchanger for studying its performance. The procured
heat exchanger is an Aluminum Plate Fin Heat Exchanger and which was manufactured at
Apollo Heat exchangers For BARC Mumbai. As per the information gathered from the BARC
Mumbai this heat exchanger is designed for operating at high pressure and is to be used for low
temperature applications. The properties such as effectiveness, NTU, overall heat transfer
coefficient, colburn factor j and skin friction co-efficient f etc are calculated in order to measure
its performance.
Table 3.1(a) Dimensions of procured plate fin heat exchanger.
CORE DATA
INTERNAL
EXTERNAL
(HOT SIDE)
(COLD SIDE)
FIN
OSF
OSF
NO. OF PASSAGE
4
5
NO. OF PASS
1
1
FLOW RATE
COUNTER FLOW
COUNTER FLOW
27
Table 3.1(b) Dimensions of procured plate fin heat exchanger
CORE SIZE
FLOW LENGTH/EFFECTIVE FLOW LENGTH
1000mm/900mm
TOTAL HEIGHT
105mm
TOTAL WIDTH/EFFECTIVE WIDTH
85mm/73mm
Table 3.2 Procured design data of plate fin heat exchanger
THERMAL DATA
HEAT LOAD
5.5 KW
HOT SIDE
COLD SIDE
FLUID
HELIUM (HP)
HELIUM (LP)
FLOW RATE
5 g/s
4.8 g/s
INLET TEMP.
36.65 0C
-189.45 0C
OUTLET TEMP.
-177.25 0C
24.39 0C
PRESSURE DROP
0.003 kg/cm2
0.02 kg/cm2
OPERATING PRESSURE
7.35 kg/cm2
7.05 kg/cm2
28
Fig. 3.1 Manufacturing details of Plate fin heat exchanger
29
3.1.2 Twin – Screw Compressor
Our air supply system consists of a Twin screw rotary compressor which is a positive
displacement machine that uses two helical screws known as rotors to compress the gas. The
rotors comprise of helical lobes affixed to a front and rear shaft. One rotor is called the male
rotor and it will typically have three bulbous lobes. The other rotor is the female rotor and this
has valleys machined into it that matches the curvature of the male lobes. Typically the female
rotor has five valleys. In a dry running rotary screw compressor, timing gears ensure that the
male and female rotors maintain precise alignment. In oil flooded rotary compressor lubricating
oils bridges the space between the rotors, both providing a hydraulic seal and transferring
mechanical energy between the driving and driven rotors. Gas enters at the suction side and
moves through the threads as the screws rotate. The meshing rotors force the gas through the
compressor and the gas exits at the end of the screw. A 3-5 rotor combination is provided in the
compressor so that, the male rotor turns 1.66 times to every one time of the female rotor. Screw
compressors have relatively high rotational speed compared to other types of positive
displacement machines which make them compact. They have the ability to maintain high
efficiencies over a wide range of operating pressure and flow rates and have long service life and
high reliability. All these things make them widely acceptable by various industries of the world.
(a)
Suction
(b)
Entrapment
(c)
Compression
(d)
Discharge
Fig. 3.2 working mechanism of Twin Screw Compressor
The twin-screw compressor has been purchased from KAESER KOMPRESSOREN GmBh. The
Compressor specification is given below:
30
Make:
Kaeser (Germany)
Model:
BSD 72
Profile of screw:
Sigma
Free air delivery:
336 m3 /hr
Suction pressure:
Atmospheric
Maximum Pressure:
11 bar
Operating temperature:
75-1000c
Motor:
37kw, 74amps, 3Φ, 50Hz, 415V±10%, 3000rpm
Oil capacity:
24 L
Cooling:
Air
3.1.3 Heating Element
This heating element was fully designed and developed in our cryogenics lab the location
of the heater is as shown in the P&I diagram. It is basically having a shell and tube type of
configuration in which incoming cold side fluid i.e. air enters the equipment and leaves the
heater through the series of baffles. Our heater contains seventeen number of aluminium baffles
through which a five set of heating tube passes. Load of heater is 1575 W, power source 220/230
V, single phase 50Hz AC Supply.
3.1.4 Resistance Temperature Detectors (RTDs)
Resistance thermometers also called as resistance temperature detectors (RTD’s) or
resistive thermal devices are temperature sensors that exploit, the predictable change in electrical
resistance of some materials with changing temperature. It is a positive coefficient device, which
means that the resistance increases with temperature. So the material whose resistance increases
with temperature is used for making the RTD’s. Typical elements used for RTD’s include nickel
(Ni), copper (Cu), but platinum (Pt) is by far the most common, because it has the best accuracy
and stability in comparison to other RTD materials. For it the resistance versus temperature
curve is fairly linear and the temperature range is widest and has a very high resistivity, which
means that only a small of platinum is required to fabricate a sensor and making platinum costs
competitive with other RTD materials. The RTD’s are slowly replacing the use of thermocouples
in many industrial applications below 600 0C, due to higher accuracy and reliability.
31
Fig. 3.3 RTD’s construction
Figure 3.3 shows the construction of a RTD. RTDs are constructed by one of two
different manufacturing configurations. First one is the wire wound RTD which are constructed
by winding a thin wire into a coil. Thin fin element is a more common configuration, which
consists of a very thin layer of metal laid out on a plastic or ceramic substrate. Thin-film
elements are cheaper and more widely available because they can achieve higher nominal
resistances with less platinum. In order to protect the RTD, a metal sheath encloses the RTD
element and lead wires are connected to it. RTD’s are available with three different lead wire
configurations. The selection of a particular configuration depends on the desired accuracy and
instruments to be used for the measurement.
(a) Two wire configuration
(b) Three wire configuration and
(c) Four wire configuration.
RTDs are popular because of their excellent stability and exhibit the most linear signal
with respective to temperature when compared to any other electronic temperature sensor. They
are generally more expensive than alternatives, however, because of the careful construction and
use of platinum. RTDs are also characterized by a slow response time and low sensitivity; and
because they require current excitation they can be prone to self heating. And there main
limitation is that they cannot be used for measurement of temperature above 660 0C and below 270 0C. Also they are less sensitive to small temperature changes.
RTD’s are commonly categorized by their nominal resistance at 0 0C. By far the most
common RTD’s used in the industry have a nominal resistance of 100 ohms at 0 0C are called as
the PT-100 sensors. The relationship between resistance and temperature is very nearly linear
and follows the equation.
For < 0 0C RT = R0 [1+ aT+bT2+cT3 (T-100)]
For > 0 0C RT = R0 [1+ aT+bT2]
32
Where,
RT = resistance at temperature T
R0 = resistance at nominal temperature
a, b, and c are the constants used to scale the RTD.
We have used four RTD’s for the measurement of temperature at the inlet and outlet ports of
both hot and cold fluid. And we have done the calibration of RTD’s and the calibration chart is
and calibration graph are given below.
140
120
100
80
RTD2
60
RTD3
40
RTD4
20
0
0
20
40
60
80
100
120
140
Fig.3.4 RTD Calibration graph
33
Table 3.3 Calibration Chart
THERMOMETER(0C)
RTD1(0C)
RTD2(0C)
RTD3(0C)
RTD4(0C)
30.5
31.85
31.8
31.98
37.04
32.8
33.67
33.64
33.82
32.84
34
34.87
34.84
35.01
34.03
39
40.31
40.31
40.46
39.48
43.5
44.98
44.98
45.13
44.11
47
48.9
48.89
45.13
44.11
50
51.72
51.69
51.84
50.76
54
55.76
55.74
55.92
54.85
57
58.66
58.7
58.82
59.71
60
61.55
61.6
61.69
60.51
62.8
63.93
63.98
64.07
62.87
65
66.67
66.71
66.78
65.54
68
70.04
70.06
70.24
69.1
71
72.45
72.43
72.63
71.44
74.5
76.1
76.12
76.34
75.12
85
86.69
86.81
86.87
85.57
90
91.64
91.79
91.89
90.64
94.5
96.21
96.34
96.45
95.04
100
101.8
101.91
101.99
100.51
105.5
107.55
107.73
107.77
106.38
107.5
107.9
108.01
108.08
106.55
110.5
113.5
113.36
113.42
111.95
112.5
114.71
114.85
114.89
113.33
115
117.04
117.18
117.22
115.65
117
120
120.07
120.12
118.54
34
3.1.5 Orifice mass flow meter
A flow meter or flow sensor is an instrument used in almost all mechanical and electrical
instrumentation process to measure the flow rate of liquid or gas. An orifice meter is a device
used for measuring the rate of fluid flow. Its working is based on the Bernoulli’s principle which
says that there is a relationship between the pressure of the fluid and the velocity of the fluid.
When the velocity increases, the pressure decreases and vice versa. An orifice plate is basically a
thin plate with a hole in the middle. It is usually placed in a pipe in which fluid flows. When the
fluid flows through the pipe, it has a certain velocity and a certain pressure. When the fluid
reaches the orifice plate, with the hole in the middle, the fluid is forced to go through the small
Fig.3.5 Orifice plate
Hole of varying cross section, the point of maximum convergence actually occurs shortly
downstream of the physical orifice, and is called as the point of vena contracta(see drawing to
the right). Both the velocity and pressure of the fluid changes while it passes through the orifice
plate . Beyond the vena contracta, the fluid expands and the velocity and pressure change once
again. The volumetric and mass flow rates are obtained from the Bernoulli’s equation by
measuring the difference in fluid pressure between the normal pipe section and at the vena
contracta, shown in fig 3.4. The pressure recovery is limited for an orifice plate and the
permanent pressure loss depends primarily on the area ratio. For an area ratio of 0.5, the head
loss is about 70 - 75% of the orifice differential.
The mass flow rate can be calculated from the formula given below:
35
Q = CCd√2YZ
Where,
Q = mass flow rate
C = is area constant =
[
]
\9]^
Cd = coefficient of discharge
H = head of air
3.1.6 Variac or Autotransformer
Variac also called as an Autotransformer is an electrical transformer with only one
winding. The auto prefix refers to the single coil acting on itself rather than any automatic
mechanism. In an autotransformer portions of the same winding act as both the primary and
secondary. The winding has at least three taps where electrical connections are made.
Autotransformers are often used to step up or step own between voltages in the 110-117-120 volt
range and voltages in the 220-230-240-volt range. We have used two variacs
An autotransformer has a single winding with two end terminals, and one or more
terminals at intermediate tap points. The primary voltage is applied across two of the terminals,
and the secondary voltage taken from two terminals, almost always having one terminal in
common with the voltage source and electrical load. The primary and secondary circuits
therefore have a number of windings turns in common. Since the volts-per-turn is the same in
both windings, each develops a voltage in proportion to its number of turns. The other end of the
source and load are connected to taps along the winding. Different taps on the winding
correspond to different voltages, measured from the common end. In a step-down transformer
the source is usually connected across the entire winding while the load is connected by a tap
across only a portion of the winding. In a step-up transformer, conversely, the load is attached
across the full winding while the source is connected to a tap across a portion of the winding.
36
3.2 Test Rig
1: Compressor
4,8: U- Tube manometer
2: Control Valve
3,7: Pressure Taps
5: Heater
6:Test section
T1, T2, T3, T4 are RTD’s
Fig. 3.6 Schematic P&I diagram of the Experimental Test Rig
37
Fig.3.7 Photograph of the Experimental Setup
38
3.2.1 Procedure for Hot Testing
Air is used as the as working fluid in this experiment. The apparatus was connected to a
compressor system which is capable of continuously delivering dry air .The compressed air from
the compressor enters the laboratory through a control valve which is used to regulate the flow
rate through the heat exchanger and then routed to the testing heat exchanger. This is the cold
side fluid which is made to enter the heat exchanger from the bottom side and when it comes out
it is made to pass through the heater, where it gets heated up and which is then again fed into the
heat exchanger from top end and which finally results in hot and cold fluid streams. The heat
supplied to the heater is controlled with the help two variacs. The pressure taps are located on the
upstream and downstream of heat exchanger to measure the pressure drop across the heat
exchanger. These pressure taps was connected with tubing and which is connected to a U-tube
manometer to give an average reading of the pressure drop. The air inlet and outlet temperatures
at both ends of heat exchanger core were measured using four RTD’s. The air flow rate was
measured using the Rotameter and the mass flow rate of both the fluids can be measured using
orifice meter. The orifice meter is used only when the test is carried out in unbalanced condition
i.e. when the mass flow rates on both sides is different. The pressure drop across the orifice plate
can be measured by using U-tube manometers.
It was ensured that there is no mass leak from the system. And the test section was
carefully insulated, by using glass wool sheets and asbestos tapes to eliminate heat losses from
the system to the surrounding. The air flow rate through the test section was set using the control
valve, and the temperatures, core pressure drop across the heat exchanger and the room pressures
were recorded for flow in the required range. The system was then allowed to run until the steady
state is achieved. The system was considered to be at steady state when all the temperature
readings steadily decrease and steadily increase for at least one minute. Once the steady state was
achieved for a particular mass flow rate the air flow rate and the temperature and pressure
differentials of the air stream across the core are accurately measured for estimating the rate of
heat transfer, pressure drop and various performance parameters like effectiveness, NTU, and
heat transfer coefficient. In experimental calculation in order to take into account the effect of
wall longitudinal heat conduction the KROGERES formula was used. For theoretical calculation
the usual procedure for rating was followed and the calculations were done in the excel sheet.
39
CHAPTER 4
Rating Procedure
RATING PROCEDURE
Design of heat exchanger involves two types of problem – (a) Sizing and (b) Rating.
Sizing involves the determination or we can say selection of type of heat exchanger, flow
arrangement, material of heat exchanger and physical dimensions of the heat exchanger to meet
the specified heat transfer and pressure drop requirements. Whereas, Rating of the heat
exchanger consists of finding the thermal performance parameters like, effectiveness, heat
transfer coefficient and pressure drop of an already designed heat exchanger whose dimensions
are known to us. We are working on the rating problem. Since the outlet temperatures are not
known for the rating problem, the average fluids mean temperatures have to be projected first.
The heat transfer coefficient and the effectiveness of the plate fin heat exchanger are found based
on different correlations existing in literature. The outlet temperatures and the average fluid
temperatures are calculated from the effectiveness value and then compared with the values
assumed earlier. The above procedure is carried out until the calculated values of the mean fluid
temperatures matches with the assumed values. Following steps show the detailed rating
procedure:
1. The first step in rating procedure is to calculate the various surface geometrical properties of
the heat exchanger. We are using a plate fin heat exchanger with offset strip fin geometry, and
geometry of the offset strip fin surface is described by the following parameters:
i) Fin spacing (s), excluding the fin thickness,
ii) Fin height (h), excluding the fin thickness,
iii) Fin thickness(t) and,
iv) Fin strip length(l or Lf)
The lateral fin offset is generally the same and equal to half the fin spacing (including fin
thickness). Figure 4.1 shows a schematic view of the rectangular offset strip fin surface and
defines the basic geometric parameters. But the present heat exchanger has different fin
geometries for the hot and cold side fluids. Table 4.1 shows the fin specifications for hot and
cold side of the heat exchanger.
40
Figure 4.1 Geometry of Typical Offset Strip Fin Surface
Table 4.1 Core Data (fin specifications)
Hot Fluid
Cold Fluid
No. Of passes (N)
4
5
Fin thickness (t)
0.2mm
0.2mm
Fin frequency (f)
588
714
Fin length (l)
5mm
3mm
Fin height (h)
9.5mm
9.5mm
Plate thickness (a)
0.8mm
0.8mm
There are some secondary geometrical parameters which are derived from the above basic fin
geometries, which are calculated as follows. The calculation is done for the hot side fluid and by
following the same steps the results can be obtained for cold side.
41
i) Fin spacing, s =
(1 − f * t )
= 0.001501 m
(f)
ii) Free flow area to Frontal area,
_
`
H`a =
( s − t )h
(0.001501 − 0.0002) × 0.0093
=
= .00001615
( h + t )( s + t ) (0.0093 + 0.0002)( 0.001501 + 0.0002)
iii) Heat transfer area per fin, as
as = 2hl + 2ht + 2sl = 2 * 0.0093* 0.005+ 2 * 0.001501* 0.005+ 2 * 0.0093* 0.0002= 0.00011173
iv) Ratio of fin area to heat transfer area of fin,
2h(l + t )
2 × 0.0093(0.005 + 0.0002)
=
= 0.86565
2( hl + sl + ht ) 2(0.0093 × 0.005 + 0.005 × 0.001501 + 0.0093 × 0.0002)
v) Equivalent diameter, De =
=
( 4 * Freeflowar ea * length)
heattransferarea
2( s − t ) hl
2(0.001501 − 0.0002) × 0.0093 × 0.005
=
= 0.002165810m
hl + sl + ht (0.0093 × 0.005 + 0.001501 × 0.005 + 0.0093 × 0.0002)
vi) Distance between plates, b = h + t = 0.0093 + .0002 = .0095m
2. Heat transfer area, A
The various heat transfer area for hot side are calculated as follows
i) Total area between plates, A frh = b × N h × W = 0.0095 × 4 × 0.073 = 0.0028
ii) Total free flow area, A ffh = σ × A frh = 0.748742226*0.0028= 0.0021 m2
iii) Wall conduction area, a wh = A frh − A ffh = 0.0028 − 0.0021 = 0.001213 m2
iv) Total heat transfer area, Ah =
4 × A ffh × L
De
=
4 × 0.0021 × 0.9
= 3.4524 m2
0.002165810
v) Total wall conduction area = awh + a wc = 0.0007+0.001=0.0017 m2
3. Heat exchanger input data
Temperature of hot gas at inlet = 368.81 K
Temperature of cold gas at inlet = 311.93 K
Pressure at inlet of cold gas = 1.21 bar
Pressure of gas at hot inlet = 1.17 bar
42
Mass flow rate of cold gas = 0.0095 kg/s
Mass flow rate of hot gas = 0.0095 kg/s
4. Estimation of average temperature
Since the fluid outlet temperatures are not known for the rating problem, the average
fluids mean temperatures have to be predicted first. The fluid properties at the predicted mean
temperatures of 342.02 K and 338.71 K for hot and cold fluid are obtained from property
package, Gaspak.
The properties of hot gas at the mean temperature are,
Specific heat, C p = 1.04086 K J / kgK
Viscosity, µ = 0.0000198 Pa − s
Prandtl number, Pr = 0.7169882
Density, ρ = 1.14249kg / m 3
The properties of cold gas at the mean temperature are,
Specific heat, C p = 1040.67 K J / kgK
Viscosity, µ = 0.0000192 Pa − s
Prandtl number, Pr = 0.7167
Density, ρ = 1.321kg / m 3
5. Heat Transfer coefficients and surface effectiveness of fins
The calculations for the heat transfer coefficients for the hot and cold gas are analogous.
So the calculations are presented for the hot fluid and can be obtained for cold gas by following
the same procedure
i) Core mass velocity, G =
m
0.0095
=
= 4.5739 kg / s − m 2
A ffh 0.0021
4.5379 × 0.002165810
= 500.31
µ
0.0000198
iii) The critical Reynolds number Re* proposed by Joshi and Webb is
ii) The Reynolds no., Re =
GDe
=
9
I [email protected] t $.%>
Re 9$.%
e
Re 257 b c
bc
D" dt 7 1.328 b c
s
I
ID"
I
$.$$%
[email protected]
= 257 f J$.$$%$ L
$.$$$ $.%>
f J $.$$% L
%[email protected]$
f 0.002165 M0.0002 7 1.328 J$.$$%f.$$2% L
9$.% 9
R
= 956.1409069
iv) The Colburn factor j for (Re*>Re) is given by correlation propsed by Joshi and Webb is
43
j 0.53Re9$.% 1⁄D" 9$.% α9$.?
0.53 f 500.319$.% f 2.308604639$.% f 0.1613978499$.?
0.02698
v) The convective heat transfer coefficient is given by
(0.02698 × 1040.86 × 4.5739)
= 160.34 W/m2K
0.667
(Pr)
(0.7169882)
vi) The fin parameter is given by
hh =
M =
( jh * C h * Gh )
0.667
=
(2 * hh )
( 2 × 160.3430)
=95.5787
=
(K f * t )
(165 × 0.0002)
vii) l h = Height of fins for hot side= b
And for cold side passages which are the outer layers l h = b/2.
viii) The fin effectiveness is given by
n f = tanh(Ml h ) /( Ml h ) = 0.93280
vi) Overall surface effectiveness is given by
η 0 h = 1 − (a f / a s ) * (1 − η f ) = 1 − (0.86565)(1 − 0.93280) = 0.94183
Similarly, convective heat transfer coefficient and overall surface effectiveness for cold fluid are
hc = 179.948 W/m2K, η 0c = 0.87593 and jc = 0.03536
6. Overall heat transfer coefficient and NTU
The overall heat transfer coefficient is given by
1
1
a
1
=
+
+
(U O AO ) h (η oh hh Ah ) K w A w (η oc hc Ac )
Where,
Aw = lateral conduction area = W × L × (2 N P + 2) = 0.073 × 0.886(2 × 4 + 2) = 0.657 m2
1
1
0.0008
1
=
+
+
(U O AO ) h (0.94183 × 160.34 × 304524) 165 × 0.657 (0.87593 × 179.9486 × 5.2152)
= 0.0031419
44
(U O AO ) h = 318.27751W/K
Overall heat transfer coefficient, Uoh =
And U OC =
(U O Ao) h 318.27751
=
= 92.19027 W/m2 K
Aoh
3.4524
(U O Ao) c 318.2775
=
= 61.0288
Aoc
5.452
Number of transfer units, N tu =
U O AO 318.27751
=
= 32.18956
C min
9.8876
7. Effectiveness of heat exchanger without considering the effect of longitudinal
conduction
ε =
1 − e − N tu (1−C r )
1 − e −32 .18956 (1− 0 .9999 )
=
= 0 .96989
1 − C r e − N tu (1− C r ) 1 − 0 .9999 e −32 ..189561 − 0 .9999 )
8. Effect of wall longitudinal heat conduction
The consequence of longitudinal heat conduction is to decrease the effectiveness of heat
exchanger. The decrease in the effectiveness of heat exchanger is found out by using the
Kroeger’s equation.
i) Wall conduction area, aw = 0.0017 m2
ii) Conductivity of fin, K w = 165 W/m2K
iii) Wall conduction parameter, λ =
K w a w 165 × 0.0017
=
= 0.03225
LC min 0.9 × 9.8876
iv) y = λN tu Cr = 0.03225 × 32.18956 × 0.9999 = 1.03795
v) γ =
(1 − Cr )
(1 − 0.9999)
=
= .0000141432
(1 − Cr )(1 + y ) (1 − 0.9999)(1 + 1.037955)
 (1 + γ ) y 
−5
vi) φ = γ ( y /(1 + y )1 / 2 
 = 1.04768 × 10
1 − γ (1 + γ ) y 
vii) ϕ =
(1 + φ ) (1 + 1.04768 × 10 −5 )
=
= 1.00002
(1 − φ ) (1 − 1.04768 × 10 −5 )
viii) r1 =
(1 − C r ) N tu
(1 − 0.9999) × 32.18956
=
= 0.000910499
1 + λN tu Cr 1 + 0.032247 × 32.18956 × 0.9999
ix) (1 − ε ) =
(1 − Cr )
(1 − 0.9999)
=
= 0.0582595
ϕ exp( r1 ) − Cr 1.0002 × exp(0.000910499) − 0.9999
45
x)
ε = [1 − (1 − ε )] = 1 − .058254 = 0.94174
This is the value of the actual effectiveness of heat exchanger subsequently considering
wall longitudinal heat conduction effect. Outlet temperatures of fluids based on this value of
effectiveness are found as follows.
The outlet temperature of hot fluid is given by
Tho = Thi −
εC min (Thi −T ci )
0.94174 × 9.8876(368.81 − 311.93)
= 368.81 −
= 315.246 K
Ch
9.88817
The outlet temperature of cold fluid is,
Tco = Tci +
εC min (Thi −T ci )
0.94174 × 9.8876(368.81 − 311.93)
= 311.93 +
= 365.4964 K
Cc
9.8876
Mean temperature of hot and cold fluid is given by,
Mean temperature of hot fluid is, Thm =
Thi +T ho 368.81 + 315.246
=
= 342.028 K
2
2
Mean temperature of cold fluid is, Tcm =
Tci + Tco 311.93 + 365.4964
=
= 338.713K
2
2
Since our calculated mean temperatures match with the assumed values of mean
temperature we can stop the iteration here, otherwise we would have taken the values of mean
temperature obtained in the above step and carried out the iteration once again, till the assumed
and found values are identical
9. Pressure drop
Since the pressure drop of cold fluid is more critical so the calculation is shown for cold
fluid
i) The friction factor (f) for Re*>Re is given by
f 8.12Re9$.3? 1⁄D" 9$.? α9$.$
= 0.05999
ii) The pressure drop, ∆p =
4 fLG 2 4 × 0.05999 × 0.9 × (3.9169) 2
=
= 836.0924 Pa
2 De ρ b
2 × 0.001674 × 1.18351
46
CHAPTER 5
Performance Analysis
PERFORMANCE ANALYSIS
The main aim of present work is to calculate the performance parameters like,
effectiveness, overall heat transfer coefficient of the plate fin heat exchanger. In order to find the
performance of present heat exchanger a number of experiments were carried out at different
mass flow rates and at different hot fluid inlet temperature under balanced flow. Table 5.1 shows
the experimentally observed data
Table 5.1 experimentally observed data.
Flow
Rate
P1
P2
(Kg/cm2) (Kg/cm2)
(litr/min)
∆gh
∆gB
(mm of
(mm of
Hg)
Hg)
T1
T2
T3
T4
()
()
300
0.08
0.06
9
6
42.24
87.34
96.2
47.15
400
0.14
0.12
15
12
38.35
87.02
95.12
43.01
500
0.2
0.17
25
22
38.93
88.49
96.12
43.11
550
0.24
0.20
30
26
39.82
88..83
96.66
43.48
588
0.28
0.24
31
27
40.41
88.45
96.20
43.99
650
0.32
0.26
40
35
41.16
87.86
95.95
44.17
300
0.08
0.06
8
6
40.92
62.06
66.48
43.06
400
0.135
0.10
16
14
42.77
62.90
66.43
44.56
500
0.2
0.16
24
22
39.57
66.02
41.69
600
0.28
0.23
31
30
39.94
62.44
65.98
41.73
650
0.34
0.28
37
34
42.72
62.77
66.34
44.06
62.52
47
5.1 CALCULATIONS:
The temperatures values which are obtained experimentally are firs of all corrected using
the calibration chart, and also the pressure values are converted in units of Pa or bar, and then
used for further calculations.
T1 = 38.93 , T2 = 87.23, T3 = 95.96, T4 = 48.10, P1 = 1.21 bar, P2 = 1.18 bar,
Flow rate=500litr/min
1.Mass flow rate = Volume flow ratef
%$$f$mn
2$
f
ij
klj
[email protected]%f$o
>[email protected]>
= 0.01 kg/s
2. Heat capacity of hot and cold fluids,
Cc = mc*Cpc = 0.010087 KW/K
for hot fluid, Ch = mh*Cph = 0.010093 KW/K
3. Capacity Rate ratio, Cr = Cmin/Cmax = . 0.9994
4. Effectiveness, B
h pqrstuvmquwrv
pxrs lrstuv 9lyrstuv
pyqyuwrvmqyrstuv
py lrstuv 9lyrstuv
= 91.134
= 86.920
5. Number of transfer units, NTU = 15.009
After considering the effect of longitudinal heat conduction. Same steps as described in Chapter
4 are followed, but here the NTU value is assumed in such a way that the effectiveness obtained
from Kroger’s equation matches with the experimental value of effectiveness.
6. Overall Heat transfer conductance,UA0
UA0 = NTUfCmin = 15.009f 0.010087
= 160.898 W/K
Here also the surface geometrical properties are calculated by following the procedure as
mentioned in Chapter 4. Table 5.2 shows the performance parameters of heat exchanger obtained
after calculation.
48
Table 5.2 Performance of heat exchanger
Flow
Mass
Rate
flow
B
h
NTU
UA0
Reh
Rec
W/K
∆z{|l}~ ∆zp|}~
(lit/min) rate
(kg/s)
300
0.0057 89.902 83.749 13.64
400
0.0074 90.236
298.67 199.72
8.73
5.87
15.24 117.37 416.04 278.22
7.97
5.53
500
0.0103 91.134 86.920 15.00 160.89 542.44 362.75
7.44
5.05
550
0.0116
406.46
7.72
4.53
588
0.0127 92.001 86.247 15.70 207.64 668.83
445
7.64
4.45
650
0.0142 92.786 85.371 17.49 258.57 747.83 497.56
7.98
3.94
300
0.0057 88.108 83.064 11.76
280.24 186.48
4.3
3.02
400
0.0078 89.937 85.480 13.25 107.60 419.28 277.56
3.36
2.33
500
0.0101
88.48
102.89 491.54 399.77
3.48
3.04
600
0.011
90.004 86.597 11.28 135.41 702.63 465.13
3.46
2.58
650
0.014
90.572 85.114 12.00 174.87 789.66 525.48
3.49
2.21
92.08
85.90
80.95
86.365 16.24 196.09
86.814
9.79
69.74
610.9
In order to compare our experimental results with the values that are obtained from
theoretical correlations, some graphs are plotted for which the experiment is conducted at
49
different mass flow rates and at two different hot inlet temperatures of 66 and 96. Some of the
graphs are shown below:
5.2 Variation of Effectiveness with Mass Flow Rate
Effectiveness V/s Mass Flow Rate
95
Effectiveness
94
Muse
93
Experimental
92
Joshi
91
Mangalik
90
Maiti
89
88
0
0.005
0.01
0.015
Mass Flow Rate (kg/s)
Fig. 5.1 Variation of effectiveness with mass flow rate ( hot inlet temperature=96
Effectiveness
Effectiveness V/s Mass Flow Rate
95
94
93
92
91
90
89
88
87
experimental
Joshi
Mangallik
Maiti
Muse
0
0.005
0.01
Mass flow rate (kg/s)
0.015
Fig. 5.2 Variation of effectiveness with mass flow rate (hot inlet temperature=66
Figure 5.1 and 5.2 shows the variation of effectiveness obtained experimentally as well as
with theoretical correlations and that obtained with simulation software Aspen with mass flow
rate. It is seen that in both the cases effectiveness increases with mass flow rate. Experimental
hot effectiveness first increases, then becomes almost constant for certain mass flow rates and
50
then again increases. However from two figures it can be seen that the value of experimental
effectiveness is more when hot inlet temperature is 96 as compared to effectiveness value
when hot inlet temperature is 66. So it can be concluded that with increase in hot inlet
temperature effectiveness increases.
Overall Thermal Conductance
5.3 Variation of Overall thermal Conductance with Mass flow rate
500
Overall Thermal Conductance V/s Mass
Flow Rate
400
UA experimental
300
UA Joshi
UA Mangalik
200
UA Maiti
100
0
0
0.005
0.01
0.015
Mass Flow Rate (Kg/s)
Overall Thermal Conductance
Fig. 5.3 Variation of overall thermal conductance with mass flow rate ( hot inlet
temperature of 96)
Overall thermal conductance V/S Mass flow
Rate
500
400
UA experimental
300
UA Joshi
200
UA Mangalik
100
Ua maiti
0
0
0.005
0.01
0.015
Mass flow Rate( kg/s)
Fig. 5.4 Variation of overall thermal conductance with mass flow rate (hot inlet
temperature of 66
Figure 5.3 and 5.4 shows the variation of overall thermal conductance with mass flow
rate for hot inlet temperature of 96 and 66 respectively. It can be seen that the theoretical as
51
well as experimental overall heat transfer coefficient increases with increasing mass flow rate. It
is due to the fact that with increasing mass flow rate the Reynolds number increases and as a
result Colburn factor (j) also increases which is directly proportional to heat transfer coefficient,
so overall thermal conductance increases.
5.4 Variation of Hot and Cold Effectiveness with Mass Flow Rate
Hot Effectiveness V/s Cold Effectiveness
94
Effectiveness
92
90
Hot Effectineness
88
Cold effectiveness
86
84
82
0
0.003
0.006
0.009
0.012
0.015
Mass Flow Rate( kg/s)
Fig. 5.5 Variation of Hot and Cold effectiveness with mass flow rate ( hot inlet temperature of
96)
Effectiveness
Hot Effectiveness V/s Cold Effectiveness
91
90
89
88
87
86
85
84
83
82
hot effectiveness
cold effectiveness
0
0.005
0.01
0.015
Mass flow rate(kg/s)
Fig. 5.6 Variation of Hot and Cold effectiveness with mass flow rat (hot inlet temperature of
66)
Figure 5.5 and 5.6 show how the experimental hot and cold effectiveness varies with the
mass flow rate for hot inlet temperature of 96 and 66 respectively. It is seen that both hot and
52
cold effectiveness increases with increasing mass flow rate and try to approach other and there is
an optimum mass flow rate for each hot inlet temperature at which the gap between the two
effectiveness is minimum and then again increases. Also after the optimum point the cold
effectiveness again decreases, this is because a heat exchanger is designed for a particular mass
flow rate and inlet temperatures at which it gives maximum effectiveness, after which its
performance deteriorates. Also the imbalance increases because of heat loss to the environment
as we are not able to provide the complete insulation. It can also be seen from the graphs that at
lower hot inlet temperature the imbalance i.e. difference between the two effectiveness is less as
compared to the imbalance at high temperatures.
5.5 Variation of Pressure Drop with Mass Flow Rate
Pressure Drop V/s Mass Flow Rate
Pressure Drop (KPa)
7
6
cold side
5
4
Joshi
3
Mangalik
2
Maiti
1
0
0
0.005
0.01
0.015
Mass Flow Rate (Kg/s)
Fig. 5.7 Variation of pressure drop with mass flow rate
Figure 5.7 shows how the pressure drop in the heat exchanger varies with varying mass
flow rate and also the comparison between experimental and theoretical pressure drop. It can be
seen that the pressure drop increases with mass flow rate for each case. However the
experimental pressure drop is much more as compared to the theoretical pressure drop because in
theoretical calculations we have not taken in to account the pressure drop taking place in piping’s
and also the manufacturing irregularities and header losses.
53
CHAPTER 6
Conclusions
CONCLUSIONS
The hot test is conducted to determine the thermal performance parameters of the
available plate fin heat exchanger at different mass flow rates and two different hot inlet
temperatures of 96 and 66. An average effectiveness of 91% is obtained. It is found in both the
cases that the effectiveness and overall thermal conductance increases with increasing mass flow
rate It is also found that hot fluid effectiveness increases with flow rate of the fluid and agrees
within 4% with the effectiveness value calculated by different correlations and that obtained by
using the simulation software, Aspen. Also the pressure drop increases with increasing mass
flow rate and experimental values are more as compared to theoretical results because the losses
in pipes and manufacturing irregularities have not been taken in to account.
For a particular hot inlet temperature there is an optimum mass flow rate at which the
difference between the hot and cold effectiveness of the heat exchanger is minimum and at this
point the imbalance is also minimum. We found that the insulation which is provided in the heat
exchanger has a significant effect on its performance. It is expected that the imbalance i.e.
difference between the hot and cold end temperature can be brought to a minimum level if a
perfect insulation like vacuum is provided.
6.1 Scope for Future Work
Present tests are conducted at room temperatures and in future we can perform the
experiment at low temperatures in order to check the performance of the present heat exchanger
for Cryogenic applications. In cold testing air at about 100K will be used as the cold fluid. In
cold test in place of heater a cold box will be used.
54
References
REFERENCES
[1] Patankar S. V. and Prakash C. 1981 An Analysis of Plate Thickness on Laminar Flow and
Heat transfer in Interrupted Plate passages. International Journal of Heat and Mass Transfer 24:
1801-1810.
[2] Joshi H. M. and Webb R. L. 1987. Heat Transfer and Friction in Offset Strip Fin Heat
Exchanger, International Journal of Heat and Mass Transfer. 30(1): 69-80
[3] Suzuki, K., Hiral, E., Miyake, T., Numerical and Experimental studies on a two Dimensional
Model of an Offset-Strip-Fin type Compact Heat Exchanger used at low Reynolds Number.
International Journal of Heat and Mass Transfer 1985 28(4) 823-836.
[4] Tinaut F. V., Melgar A. and Rehman Ali A. A. 1992 Correlations for Heat Transfer and Flow
Friction Characteristics of Compact Plate Type Heat Exchangers. International Journal of Heat
and Mass Transfer. 35(7):1659:1665
[5] Manglik and Bergles A. E. 1995 Heat Transfer and Pressure drop Correlations for
Rectangular Offset Strip Finn Compact Heat Exchangers. Experimental Fluid Science 10:171180.
[6] Hu S and Herold K. E. 1995a Prandtl Number Effect on Offset Strip Fin Heat Exchanger
Performance: Predictive Model for Heat Transfer and Pressure Drop. International Journal of
Heat and Mass Transfer 38(6) 1043-1051
Hu S and Herold K. E. 1995b Prandtl number Effect on Offset Strip Fin Heat Exchanger
Performance: Experimental Results. International Journal of Heat and Mass Transfer 38(6)
1053-1061.
[7] Zhang L. W., Balachandar S., Tafti D. K. and Najjar F. M. 1997. Heat Transfer Enhancement
Mechanisms in Inline and Staggered Parallel Plate Fin Heat Exchanger. International Journal of
Heat and Mass Transfer 40(10):2307-2325
55
[8] Dejong N. C., Zhang L. W., Jacobi A. M., Balchandar S. and Tafti D. K. 1998. A
Complementary Experimental and Numerical Study of Flow and Heat Transfer in Offset Strip
Fin Heat Exchangers. Journal of Heat Transfer 12:690:702
[9] Bhowmik H., Kwan- Soo Lee 2009. Analysis of Heat Transfer and Pressure Drop
Characteristics in an Offset Strip Fin Heat Exchanger. International Journal of Heat and Mass
Transfer 259-263
[10] Saidi A. and Sudden B. 2001. A Numerical Investigation of Heat Transfer Enhancement in
Offset Strip Fin Heat Exchangers in Self Sustained Oscillatory Flow. International Journal of
Numerical Methods for Heat and Fluid Flow. 11(7): 699-716
[11] Dong J., Chen J., Chen Z. and Zhou Y. 2007. Air Side Thermal hydraulic Performance of
Offset Strip Fin Heat Exchangers Fin Alumunium Heat Exchangers. Applied Thermal
Engineering 27:306-313
[12] Michna J. G., Jacobi A. M. and Burton L. R. 2005. Air Side Thermal- Hydraulic
Performance of an Offset Strip Fin Array at Reynolds Number up to 12, 0000. Fifth
International Conference on Enhanced Compact and Ultra Compact Heat Exchangers. Science,
Engineering and Technology 8-14.
[13] Prabhat Gupta, Atrey M. D. Performance Evaluation Of Counter Flow Heat Exchangers
Considering the Heat In Leak and Longitudinal Conduction for Low Temperature applications.
Cryogenics Volume 40, issue 7, Pages 469-474.
[14] Barron R. F., Cryogenic Heat Transfer, Taylor and Francis (1999) 311-318.
[15] Shah R. K. and Sekulic D. P. Fundamentals of Heat Exchangers, John Willey & Sons Inc.,
pp 10-13
[16] Kays W. M. and London A. L. Compact Heat Exchangers. 2nd Edition, McGraw-Hill, New
York, 1964
[17] Weiting A. R.1975. Empirical Correlations for Heat Transfer and Flow Friction
Characteristics of Rectangular Offset strip Fin Plate Fin Heat Exchangers. Transactions of
ASME, Journal of Heat Transfer 97:488-497
56
[18] Manson S. V. Correlations of Heat Transfer Data and of Friction Data for Interrupted Plane
Fins Staggered in successive Rows NACA Technical Note 2237(1950)
[19] Maiti D.K. Heat Transfer and Flow Friction Characteristics of Plate Fin Heat Exchanger
Surfaces- A Numerical Study PhD Dissertation Indian Institute of Technology, Kharagpur
(2002)
[20] DURMAZ, Gurcan, Experimental and Numerical Analysis of Heat Transfer Performance of
Offset Strip Fins, Master of Science thesis, The Graduate School Of Engineering and Science of
IZMIR Institute of Technology:2009
57
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