study of forced convection in smooth and ribbed ducts
TRANSCRIPT
Study of forced convection in smooth and ribbed ducts.
A Project Report
Submitted to the Department of Mechanical Engineering In partial fulfillment of the
Requirements for the degree of Bachelor of Technology (B Tech)
Submitted by:
Tabind Maqbool (01/11) Hamid Qadir shah (54/11)
Under the guidance of
Prof. (Dr.) Adnan Qayoum
Department of Mechanical Engineering
NIT Srinagar, J&K, India-190006
Mechanical Engineering Department, NIT, Srinagar i
Mechanical Engineering Department, NIT, Srinagar ii
CANDIDATES’ DECLARATION
We hereby certify that the work which is being presented in this project report titled,
“Study of forced convection in smooth and ribbed ducts” in the partial fulfillment of the requirements for the degree of Bachelor of Technology (B.Tech.) submitted in the Department of Mechanical Engineering, NIT Srinagar, is an authentic record of our own bonafide work carried out during 8th semester under the guidance of Prof. Dr. Adnan Qayoum. We have followed all ethics and publishing standard while preparing this project. The matter presented in this project has not been submitted by us or anyone else in any other University/Institute for the award of any other degree.
This is to certify that the above statement made by the candidates is true to the best of my knowledge.
Dr. Adnan Qayoum
(Supervisor)
Professor
Mechanical Engineering Department, NIT, Srinagar iii
ACKNOWLEDGEMENT
We would like to express our sincere gratitude to Dr. Adnan Qayoum,
Professor Mechanical Engineering Department our project guide, for
allowing us to undertake this project and providing us all the resources
required to successfully learn & complete this project.
Special thanks goes to Dr. G. A Harmain, HOD, Mechanical Engineering
Department and professor Sheikh Ghulam Mohammad for their kind
support in pursuing the project. A special debt of gratitude is owed to the
authors whose works we have consulted and quoted in this report.
We would also like to express our gratitude to all members of
Mechanical Engineering Department who have helped & supported us
throughout the project.
Last but not the least we thank Mr. Javaid Ahmad of machine shop for
his support and patience in working on lathe to create different profiles in
ducts.
Mechanical Engineering Department, NIT, Srinagar iv
ABSTRACT
This work presents an experimental study on the friction factor and
thermal enhancement factor characteristics in a circular tube with
different types of internal profile of under constant heat flux conditions.
In the experiments, measured data are taken at Reynolds number in range
of 7600 with air as the test fluid. The experiments were conducted on
circular duct with 28 mm internal diameter using plain profile duct lining
and internal square thread profile duct lining of constant pitch. The heat
transfer and friction factor data obtained in case of square threaded
profile is compared with the data obtained from a plain circular profile
under similar atmosphere and flow conditions.
The variation of heat transfer, pressure loss and friction factor (ƒ) is respectively determined and depicted graphically. The heat transfer enhancement for a test duct with square threads is 17 percent more as compared to plain duct.
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Table of contents
Page no
Certification i
Acknowledgement ii
Abstract iii
List of figures vi
List of plots vii
List of tables viii
Chapter 1: Introduction 1-9
1.1 Historical background 1
1.2 Heat transfer Augmentation techniques 2
1.3 Important definitions 4
1.4 Problem statement 5
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1.5 Objective 6
1.6 Literature survey 6
Chapter 2: Test Facility 10-25
2.1 Experimental setup 10
2.2 Components 11
Chapter 3: Methodology 26-29
3.1 Governing Equations 27
Chapter 4: Results and discussion 30-36
4.1 Smooth duct 31
4.2 Square thread duct 32
4.3 Pressure drop variation 33
4.4 Nusselt No and Reynolds No variation 34
4.5 Friction factor and Reynolds No variation 35
4.6 Convection coefficient and Re variation 36
Chapter 5: Conclusions 37REFERENCES 38-39APPENDIX A1 40-48APPENDIX A2 49-57APPENDIX B 58-59
Mechanical Engineering Department, NIT, Srinagar vii
S no LIST OF FIGURES Page no
Fig 2.1 Solid Works design of setup 10
Fig 2.2 Solid Works design of blower and motor assembly 11
Fig 2.3 Photograph of blower and motor assembly 12
Fig 2.4 Cut section of single phase motor 12
Fig 2.5 Photograph of AC motor 13
Fig 2.6 Solid Works design of test section 15
Fig 2.7 Photograph of test section 15
Fig 2.8 Schematic of heater plate 16
Fig 2.9 Solid Works design of orifice plate 17
Fig 2.10 Solid Works design of ball valve 18
Fig 2.11 Solid Works design of square threaded duct (cut section) 18
Fig 2.12 Solid Works design of square threaded duct 19
Fig 2.13 Photograph of square threaded duct 20
Fig 2.14 Photograph of square threaded duct insertion (isometric) 20
Fig 2.15 Solid Works design of smooth duct 21
Fig 2.16 Photograph of data logger 22
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Fig 2.17 Photograph of digital manometer 24
Fig 2.18 Wiring diagram of voltmeter 25
Fig 2.19 Wiring diagram of ammeter 25
Fig 3.1 Flow chart of methodology 27
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Nomenclature
A Area
d diameter ,m
h heat transfer coefficient kW (m C)
L length ,
M mass flow rate
P pressure , kPa
Pr Prandtl number
Re Reynolds number
U overall heat transfer coefficient kW (m C)
Cp Specific heat kJ (kg C)
ƒ friction factor
k thermal conductivity , kW (m C)
Nu Nusselt number
Mechanical Engineering Department, NIT, Srinagar x
p helical rib pitch , m
Q heat transfer rate , kW
T temperature C
U velocity , m/s
Ρ density , kg/m3
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Chapter 1
INTRODUCTION
1.1 Historical BackgroundHeat exchangers are used in different processes ranging from conversion, utilization
& recovery of thermal energy in various industrial, commercial & domestic
applications. Some common examples include steam generation & condensation in
power & cogeneration plants; sensible heating & cooling in thermal processing of
chemical, pharmaceutical & agricultural products; fl id heating in man fact ring &
water heat recovery etc. Increase in heat exchanger’s performance can lead to their
economical design which in turn will help to make energy, material & cost savings
related to a heat exchange process. The need to increase the thermal performance of
heat exchangers, thereby effecting energy, material & cost savings have led to
development & use of many techniques termed as heat transfer augmentation. These
techniques are also referred as heat transfer enhancement. Augmentation techniques
increase heat transfer by reducing the thermal resistance in a heat exchanger. Use of
heat transfer enhancement techniques lead to an increase in heat transfer coefficient
at the cost of increase in pressure drop. So, while designing a heat exchanger
incorporating augmentation techniques, one has to find an optimum design keeping
in view increase in heat transfer rate and pressure drop. Apart from this, issues like
long-term performance & detailed economic analysis of heat exchanger has to be
studied. To achieve high heat transfer rate in an existing or new heat exchanger while
taking care of the increased pumping power, several techniques have been proposed
in recent years.
Heat transfer augmentation techniques (passive, active or a combination of passive
and active methods) are commonly used in areas such as process industries, heating
and cooling in evaporators, thermal power plants, air-conditioning equipment,
refrigerators, radiators for space vehicles, automobiles, etc. Passive techniques, where
inserts are used in the flow passage to augment the heat transfer rate, are
advantageous compared with active techniques, because the insert manufacturing
process is simple and these techniques can be easily employed in an existing heat
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exchanger. In design of compact heat exchangers, passive techniques of heat transfer
augmentation can play an important role if a proper passive insert configuration can
be selected according to the heat exchanger working condition (both flow and heat
transfer conditions). The major challenge in designing a heat exchanger is to make
the equipment compact and achieve a high heat transfer rate using minimum pumping
power. In recent years, the high cost of energy and material has resulted in an
increased effort aimed at producing more efficient heat exchange equipment.
Furthermore, sometimes there is a need for miniaturization of a heat exchanger in
specific applications, such as space application, through an augmentation of heat
transfer. For example, a heat exchanger for an ocean thermal energy conversion
(OTEC) plant requires a heat transfer surface area of the order of 10000 m2/MW.
Therefore, an increase in the efficiency of the heat exchanger through an
augmentation technique may result in a considerable saving in the material cost.
Furthermore, as a heat exchanger becomes older, the resistance to heat transfer
increases owing to fouling or scaling. These problems are more common for heat
exchangers used in marine applications and in chemical industries. In some specific
applications, such as heat exchangers dealing with fluids of low thermal conductivity
(gases and oils) and desalination plants, there is a need to increase the heat transfer
rate.
The heat transfer rate can be improved by introducing a disturbance in the fluid flow
(breaking the viscous and thermal boundary layers), but in the process pumping
power may increase significantly and ultimately the pumping cost becomes high.
Therefore, to achieve a desired heat transfer rate in an existing heat exchanger at an
economic pumping power, several techniques have been proposed in recent years and
are discussed in the following sections.
1.2 Heat transfer augmentation techniques
Generally, heat transfer augmentation techniques are classified in three broad categories:
a) Active method
b) Passive method
c) Compound method
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The active and passive methods are described with examples in the following
subsections. A compound method is a hybrid method in which both active and
passive methods are used in combination. The compound method involves complex
design and hence has limited applications.
1.2.1 Active method
Active method is a type of heat transfer augmentation technique in which some external power input is used for the enhancement of heat transfer. This technique has not shown much potential owing to complexity in design. Furthermore, external power is not easy to provide in several applications.Some examples of active methods are induced pulsation by cams and reciprocating
plungers, the use of a magnetic field to disturb the seeded light particles in a flowing
stream, etc.
1.2.2 Passive method
Passive method is a type of heat transfer augmentation technique in which no external
power input is used for the enhancement of heat transfer. This method enhances the
heat transfer by using the available power in the system, which ultimately leads to a
fluid pressure drop. The heat exchanger industry has been striving for improved
thermal contact (enhanced heat transfer coefficient) and reduced pumping power in
order to improve the thermo-hydraulic efficiency of heat exchangers. A good heat
exchanger design should have an efficient thermodynamic performance, i.e. minimum
generation of entropy or minimum destruction of available work (energy) in a system
incorporating a heat exchanger. It is almost impossible to stop energy loss
completely, but it can be minimized through an efficient design.
Although there are so many passive methods employ to enhance the heat transfer rate, following are the most commonly used methods are discussed here;
a) Treated Surfaces: They are heat transfer surfaces that have a fine scale
alteration to their finish or coating the alteration could be continuous or
discontinuous, where the roughness is much smaller than what affects single-
phase heat transfer, and they are used primarily for boiling and condensing
duties.
b) Rough surfaces: They are generally surface modifications that promote
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turbulence in the flow field, primarily in single phase flows, and do not
increase the heat transfer surface area. Their geometric features range from
random and grain roughness to discrete three dimensional surface
protuberances.
c) Extended surfaces: They provide effective heat transfer enlargement. The newer developments have led to modified fin surfaces that also tend to improve the heat transfer coefficients by disturbing the flow field in addition to increasing the surface area.
d) Displaced enhancement devices: These are the insert techniques that are used primarily in confined feed devices to improve the energy transfer directly at the heat exchange surface by displacing the fluid from the duct pipe with bulk fluid to the core flow.
e) Swirl flow devices: They produce and superimpose swirl flow or secondary
recirculation on the axial flow in a channel. These devices include helical
strip or cored screw type tube inserts, twisted tapes. They can be used for
single phase or two-phase flows heat exchanger.
f) Coiled tubes: These techniques are suitable for relatively more compact heat exchangers. Coiled tube produce secondary flows and vortices which promote higher heat transfer coefficient in single phase flow as well as in most boiling regions.
g) Surface tension devices: These consist of wicking or grooved surfaces, which directly improve the boiling and condensing surface. These devices are most used for heat exchanger occurring phase transformation.
h) Additives for liquids: These include the addition of solid particles, soluble trace additives and gas bubbles into single phase flows and trace additives which usually depress the surface tension of the liquid for boiling systems.
i) Additives for gases: These include liquid droplets or solid particles, which are introduced in single phase gas flows either as dilute phase (gas–solid suspensions) or as dense phase (fluidized beds).
1.3 Important definitionsIn this section a few important terms commonly used in heat transfer augmentation
work are defined.
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1.3.1 Thermo-hydraulic performanceFor a particular Reynolds No., the thermo-hydraulic performance of an insert is said
to be good if the heat transfer coefficient increases significantly with a minimum
increase in friction factor. Thermo-hydraulic performance estimation is generally
used to compare the performance of different inserts such as twisted tape, wire coil,
etc., under a particular fluid flow condition.
1.3.2 Overall enhancement ratioThe overall enhancement ratio is defined as the ratio of the heat transfer enhancement
ratio to the friction factor ratio. This parameter is also used to compare different
passive techniques and enables a comparison of two different methods for the same
pressure drop. The overall enhancement ratio is defined as where Nu, ƒ, Nu0
and ƒ0 are the Nusselt numbers and friction factors for a duct configuration with and
without inserts respectively. The friction factor is a measure of head loss or pumping
power.
1.3.3 Nusselt number, NuThe Nusselt number is a measure of the conductive resistance to the convective
resistance occurring at the surface and is defined as hd/k, where h is the convective
heat transfer coefficient, d is the diameter of the tube and k is the thermal
conductivity.
1.3.4 Prandtl number, PrThe Prandtl number is defined as the ratio of the molecular diffusivity of momentum
to the molecular diffusivity of heat.
1.3.5 PitchPitch is defined as the distance between two points that are on the same plane,
measured parallel to the axis of a twisted tape.
1.4 Problem Statement The aim of the project is to fabricate a forced convection apparatus and study the
effect on the heat transfer coefficient and the other related parameters in the smooth
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duct and the ribbed (square threaded) duct by using empirical relations and
comparing them with the experimental results.
The threads of a ribbed duct (introduced in the test section) acts as a disturbance in
the flow and hence enhances the heat transfer rate but at the cost of increase in
pressure drop. So while designing a heat exchanger, convective coefficient and
pressure drop has to be analyzed.
1.5 ObjectivesThe present study has been carried out to the performance analysis of heat transfer in
smooth and ribbed ducts. The analysis has been done for the following objectives.
1 To determine the variation of pressure drop with Reynolds number
2 To determine the variation of Nusselt number with Reynolds number
3 To determine the variation of heat transfer coefficient with Reynolds number
4 To determine the variation of friction factor with Reynolds number
5 To compare heat transfer enhancement in square threaded and smooth ducts
6 To compare the heat transfer coefficient at different mass flow rates
An experimental setup has been fabricated to compute the above mentioned
objectives.
1.6 Literature surveyThere are numerous techniques to embellish the heat transfer, such as fins, dimples,
additives, etc. A great deal of research effort has been devoted to developing
apparatus and performing experiments to define the conditions under which an
enhancement technique will improve heat transfer. Heat transfer enhancement
technology has been widely applied to heat exchanger applications in refrigeration,
automobile, process industries etc. The goal of enhanced heat transfer is to encourage
or accommodate high heat fluxes. Thus result of reduction in heat exchanger size,
generally leads to less capital cost. Another advantage is the reduction of temperature
driving force, which reduces the entropy generation and increases the second law
efficiency. The need to increase the thermal performance of heat exchangers, thereby
effecting energy, material & cost savings have led to development & use of many
techniques termed as Heat transfer Augmentation. These techniques are also referred
as heat transfer enhancement or Intensification.
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Augmentation techniques increase convective heat transfer by reducing the thermal
resistance in a heat exchanger. Use of Heat transfer enhancement techniques lead to
increase in heat transfer coefficient but at the cost of increase in pressure drop. So,
while designing a heat exchanger using any of these techniques, analysis of heat
transfer rate & pressure drop has to be done. Apart from this, issues like long-term
performance & detailed economic analysis of heat exchanger has to be studied. To
achieve high heat transfer rate in an existing or new heat exchanger while taking care
of the increased pumping power, several techniques have been proposed in recent
years.
Generally, heat transfer augmentation techniques are classified in three broad
categories: active methods, passive method and compound method. A compound
method is a hybrid method in which both active and passive methods are used in
combination. The compound method involves complex design and hence has limited
applications.
Kumar et al. experimentally showed the investigation to augment the heat transfer
rate by enhancing the heat transfer coefficient during the condensation of pure steam
and R-134a over horizontal finned tubes. Spines were found to be more effective in
the bottom side of the circular integral tube. Suresh Kumar et al. numerically studied
the thermo hydraulic performance of twisted tape inserts in a large hydraulic
diameter annulus. Authors found that the thermo-hydraulic performance in laminar
flow with a twisted tape is better than the wire coil for the same helix angle and
thickness ratio.
Sozen and Kuzay study showed that the enhanced heat transfer in round tubes filled
with rolled copper mesh at Reynolds number range of 5000-19000. With water as the
energy transport fluid and the tube being subjected to uniform heat flux, they
reported up to ten fold increase in heat transfer coefficient with brazed porous inserts
relative to plain tube at the expense of highly increased pressure drop.
Golriz and Grace experimentally found that the addition of an angled deflector to the
fin region of circular membrane water–wall heat exchanger surfaces in circulating
fluidized beds can lead to a significant increase in the local heat transfer. Wang and
Sunden reported correlations for ethyl glycol and polybutene (Pr. No.10000-70000),
They also concluded by considering the overall enhancement ratio, twisted tape is
effective for small Prandtl number fluids and wire coil is effective for high Prandtl
number fluids.
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Liao and Xin carried out experiments to study the heat transfer and friction
characteristics for water, ethylene glycol and ISOVG46 turbine oil flowing inside four
tubes with three dimensional internal extended surfaces and copper continuous or
segmented twisted tape inserts within Prandtl number ranging from 5.5 to 590 and
Reynolds numbers from 80 to 50,000. They found that for laminar flow of VG46
turbine oil, the average Stanton number could be enhanced up to 5.8 times with
friction factor increase of 6.5 fold compared to plain tube.
Chang et al. experimentally showed that, in a duct fitted with transverse ribs, the flow
cells behind the 908 ribs are no longer stagnant but periodically shed when the duct
reciprocates. The typical zigzag stream-wise heat transfer variation along the ribbed
wall in a stationary system yields a large wavy pattern in the reciprocating duct.
Angirasa proved through experimental study which shows augmentation of heat
transfer by using metallic fibrous materials with two different porosities; 97% and
93%. The experiments were carried out for different Reynolds numbers (17000-
29000) and power inputs (3.7 and 9.2 W). The improvement in the average Nusselt
number was about 3-6 times in comparison with the case when no porous material
was used.
Fu et al. experimentally demonstrated that a channel filled with high conductivity
porous material subjected to oscillating flow is a new and effective method of cooling
electronic devices.
Afanasyev et al. studied different surfaces shaped by a system of spherical cavities in
a turbulent flow and found that such shaping of the heating surface has no
appreciable effect on the hydrodynamics of flow but results in considerable (up to
30–40 per cent) heat transfer enhancement. The experimental investigations of Hsieh
and Liu reported that Nusselt numbers were between four and two times the bare
values at low Re and high Re respectively.
Bogdan et al. numerically investigated the effect of metallic porous materials,
inserted in a pipe, on the rate of heat transfer. The pipe was subjected to a constant
and uniform heat flux. The effects of porosity, porous material diameter and thermal
conductivity as well as Reynolds number on the heat transfer rate and pressure drop
were investigated. The results were compared with the clear flow case where no
porous material was used. The results obtained lead to the conclusion that higher heat
transfer rates can be achieved using porous inserts at the expense of a reasonable
pressure drop.
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Smith et. al. investigated the heat transfer enhancement and pressure loss by insertion
of single twisted tape, full length dual and regularly spaced dual twisted tapes as swirl
generators in round tube under axially uniform wall heat flux conditions.
Chinaruk Thianpong et. al. experimentally investigated the friction and compound
heat transfer behaviour in dimpled tube fitted with twisted tape swirl generator for a
fully developed flow for Reynolds number in the range of 12000 to 44000. Whitham
studied heat transfer enhancement by means of a twisted tape insert way back at the
end of the nineteenth century. Date and Singham numerically investigated heat
transfer enhancement in laminar, viscous liquid flows in a tube with a uniform heat
flux boundary condition. They idealized the flow conditions by assuming zero tape
thickness, but the twist and fin effects of the twisted tape were included in their
analysis.
Saha et al. have shown that, for a constant heat flux boundary condition, regularly
spaced twisted tape elements do not perform better than full-length twisted tape
because the swirl breaks down in-between the spacing of a regularly twisted tape.
Rao and Sastri while working with a rotating tube with a twisted tape insert, observed
that the enhancement of heat transfer offsets the rise in the friction factor owing to
rotation. Sivashanmugam et. al. and Agarwal et.al. studied the thermo-hydraulic
characteristics of tape-generated swirl flow.
Peterson et al. experimented with high-pressure (8–16 MPa) water as the test liquid
in turbulent flow with low heat fluxes and low wall–fluid temperature differences
typical of a liquid–liquid heat exchanger. Benzenine et al., Saim and Abboudi, Imine
[found that the heat transfer can be enhanced by the use of transversal waved baffles.
Wban and Pil found that the heat transfer can be enhanced in case of smooth ducts by
using rough surfaces and it depends upon properties and size of the fluid molecules.
Dutta and Hossain studied the effect of local heat transfer and friction factor in a
rectangular pipe with inclined and perforated baffles. The effect of baffle size,
position, and orientation were studied for heat transfer enhancement. Ko and Anand
studied the effect of local heat transfer in a rectangular pipe with porous baffles. The
conclusion of this study is that the heat transfer increases 2 to 4 times than the solid
baffle. Karwa and Maheshwari studied the heat transfer and friction in an
asymmetrical rectangular duct with some solid and perforated baffles with relative
roughness. The friction factor for the solid baffle was found between 9.6-11.1 times
than smooth duct which decreases in perforated baffle.
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Xinyi and Dongsheng studied the turbulent flow and heat transfer enhancement in
ducts or channels with rib, groove or rib-groove tabulators. The present experimental
study investigates the increase in the heat transfer rate between a tubes heated with a
constant uniform heat flux with air flowing inside it using internal threads.
As per the available literature, the enhancement of heat transfer using internal threads
in turbulent region is limited. So, the present work has been carried out with turbulent
flow (Re number range of 7000-14000) as most of the flow problems in industrial
heat exchangers involve turbulent flow region.
Chapter 2
EXPERIMENTAL SETUP
2.1 Experimental setupThe test facility consists of a centrifugal blower unit fitted with a circular tube, which
is connected to the test section located in horizontal orientation. Nichrome bend
heater encloses the test section to a length of 60cm. Input to heater is given using
220V AC. Thermocouples Tinlet, T2, T3, T4, T5 T6, and Twall at a distance of 150mm,
250mm, 350mm and 450mm from the origin of the heating zone are embedded on the
walls of the tube and two thermocouples are placed in the air stream, one at the
entrance (Tinlet) and the other at the exit (Toutlet) of the test section to measure the
temperature of flowing air.
A digital device is used to display the temperature measured by thermocouple at
various position. The temperature measured by instrument is in oC. The test tube of 4
mm thickness is used for experimentation.
A manometer measures the pressure drop across the test section for calculating the
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friction factor. It is also used to measure the mass flow rate using an orifice place.
The pipe system consists of a valve, which controls the airflow rate through it. The
diameter of the orifice is 14 mm and coefficient of discharge is 0.62.
Display unit consists of voltmeter, ammeter and temperature indicator. The circuit is
designed for a load voltage of 0-220 V with a maximum current of 55 A.
Fig 2.1: Solid works design of setup
2.2. Components
The experimental test set up is designed for determining the convection coefficient
and is essentially a set up with a circular duct & consists of the following
components.
2.2.1. BlowerThe common radial blower shown in Figure 2.2 is used. The inlet is an opening of
diameter 25 mm. The space between the blower casing and the inlet is made air tight
by the help of high temperature silicone rubber seal.
The outlet of the blower is a cylindrical pipe of diameter 40 mm which forms the
bottom part of the set up. Two 90o elbows and a reducer are used to get the desired
experimental set up. A ball valve is installed upstream between the two 90o elbows of
the pipe to regulate the volume flow rate of the air in the blower.
The reducer changes the diameter of the pipe from 40 mm to 32 mm to form the top
part of the set up. The portion (600mm between two flanges-forms the test section )
of the top part is heated by a heating element spread on 400 mm on this portion. The
orifice plate is installed after the test section, at the top for measuring the volume
flow rate. The blower motor has a single phase AC supply of 220 V supplied. The
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mean rpm of the blower motor is 1440 and the shaft diameter of the motor is 75mm.
Fig 2.2: Solid works design of blower and motor assembly
Fig 2.3: Photograph of blower and motor assembly
2.2.2 Single phase motor
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An induction or asynchronous motor is an AC electric motor in which the electric
current in the rotor needed to produce torque is induced by electromagnetic induction
from the magnetic field of the stator winding.
Fig 2.4: Cut section of single phase motor (source: Wikipedia)
An induction motor (see Figure 2.4 and 2.5) therefore does not require mechanical
commutation, separate-excitation or self-excitation for all or part of the energy
transferred from stator to rotor, as in universal, DC and large synchronous motors. An
induction motor's rotor can be either wound type or squirrel-cage type. Three-phase
squirrel-cage induction motors are widely used in industrial drives because they are
rugged, reliable and economical. Single - phase induction motors are used extensively
for smaller loads, such as fans, blowers
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Fig 2.5: Photograph of AC motor
Although traditionally used in fixed-speed service, induction motors are increasingly
being used with variable-frequency drives (VFDs) in variable-speed service. VFDs
offer especially important energy savings opportunities for existing and prospective
induction motors in variable-torque centrifugal fan, pump and compressor load
applications. Squirrel cage induction motors are very widely used in both fixed-speed
and VFD applications
Single phase induction motors require just one power phase for their operation. They are commonly used in low power rating applications, in domestic as well as industrial use.
The main components of a single phase motor are the rotor and stator winding. The
rotor is the rotating part, the stator winding helps in rotating the rotor. The winding
has got 2 parts; One main winding and an auxiliary winding. The auxiliary winding is
placed perpendicular to the main winding. A capacitor is connected to the auxiliary
winding.
Speed-Torque curve
The Speed-Torque curve is shown in plot 2.1. The torque of the induction motor is
zero when the motor is driven slower than synchronous speed, and it becomes braking
torque at synchronous speed. The maximum torque can also be obtained at the rated
rpm. The rotation region that provides maximum or high torque varies depending on
the electric resistance, so changing the cage materials or geometry makes it possible
to study the motor characteristics that most closely meet the goals of the design.
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Plot 2.1 (source: Wikipedia)
2.2.3 Test SectionThe test section is made of mild steel of 4 mm thickness. The internal diameter of the
test section is 24 mm. It is 600 mm long having an effective length of 500 mm. The
inlet bulk temperature taping is situated 50 mm after the start of test section and outlet
bulk temperature taping is situated 50 mm before the end of test section of duct. The
tapings are inserted via two through bolts, each of 6 mm diameter. In order to make
the removing of the taping time and again easier the thermocouples are glued to these
M6 bolts. The thermocouple wire is inserted through the longitudinal hole of the
through bolt. This ensures the uniform depth of thermocouple inside the duct each
time we re-insert the bolt. The duct has a heating element fitted circumferentially to it
over a length of 400 mm. The test section is shown in Fig 2.7
Fig 2.6: Solid works design of test section
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Fig 2.7: Photograph of test section
In addition to the two bulk temperature tapings there are seven wall temperature
tapings, each placed on opposite sides of diameter at 150 mm, 250 mm, 350 mm and
450 mm downstream the start of test section. These are used to calculate the average
temperature of the duct wall.
2.2.4. Heater Plate
The heating element is present between the sole plate and pressure plate. It is pressed
hard between the two plates. The heating element consists of nichrome wire wound
around a sheet of mica. The two ends of the nichrome wire are connected to the
contact strips. The contact strips are connected to the terminals of the iron. There are
two reasons for which mica is chosen in the heating material. Mica is a very good
insulating material. Besides that mica can also withstand very high temperatures. The
entire assembly of mica sheet, nichrome wire and contact strips are riveted together
resulting in a mechanically sound and robust construction. There is an asbestos sheet,
which separates and thermally insulates the top plate from the heating element.
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Fig 2.8: Schematic of heater plate
2.2.5 Orifice PlateThe orifice plate is located near the exit of the blower and is used for measuring the
volume flow rate of the blower. It is connected to the digital manometer (HTC-6205)
for measuring the pressure drop across the orifice. The value of Cd = 0.62.
Orifice Plate. Position
Fig 2.9: Solid works design of orifice plate
2.2.6. Thermocouple
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Thermocouples are temperature measuring devices consisting of two dissimilar
conductors that are in contact with each other at one or more spots, when the two
metals are subjected to temperature it produces a voltage differential. Thermocouples
are a widely used type of temperature sensor for measurement and control and can
also convert a temperature gradient into electricity. Commercial thermocouples are
inexpensive, interchangeable, are supplied with standard connectors, and can measure
a wide range of temperatures. In contrast to most other methods of temperature
measurement, thermocouples are self powered and require no external form of
excitation. The k type thermocouples are used for measuring the flow temperature and
the temperature of the wall. These temperature values are used to calculate the value
of convection coefficient.
The k type thermocouples are made of Alumel and Chromel having a range of −200
°C to +1350 °C. They are connected to the PC via the Ajankiya IM 2000 series, eight
channel USB data acquisition card for higher sensitivity and accuracy.
2.2.7 Ball ValveIn our experimental setup, the ball valve has been used to control the mass flow rate
of air. The valve opened or closed by turning the lever (0o-90o). The valve can thus
be fully opened or partially opened depending upon the needs or requirements.
The diameter of the valve is 40mm which is installed upstream between the two 90o
elbows of the pipe to regulate the volume flow rate of the air in the blower.
Mechanical Engineering Department, NIT, Srinagar 18
Fig 2.10: Solid works design of ball valve
2.2.8. Duct lining profiles
Threaded profile
To enhance the heat transfer rate in a circular duct, turbulence needs to be created in a
duct. One of the best methods is to use internal threads and tapered profile. In our
experimentation we are using square threads to create turbulence in air.
Fig 2.11: Solid works design of square duct (cut section)
Fig 2.12: Solid works design of square threaded duct
Fabrication of square threads
Mechanical Engineering Department, NIT, Srinagar 19
A rod of mild steel of 900 mm is cut down into four equal parts having 216 mm length and diameter of 32mm. Different operations were performed on them as follows:
Lathe setting: gear combination of HJN is set to create 2 threads per inch.
1. Plain turning is done on outside of two pieces to make them of diameter of 28mm exact in order to fit them into the duct fitted with thermocouples, heating coil and insulation.
2. Finishing is done at speed of 250 rpm
3. Facing is done on each piece to have uniform cross-section and to set revolving centre right in centre to support the job while plain turning.
4. Drilling is done at an rpm of 88 with a 17 mm drill bit to make a bore in a piece so as to make internal threads.
5. Again boring is done with a drill bit of 22 mm to finalize the internal diameter.
6. In order to make internal threads a square thread cutting tool is designed and fabricated on grinding machine.
7. Now this cutting tool is fixed in tool holder of .5inches diameter and internal threading is done.
Fig 2.13: Photograph of square threaded duct
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Fig 2.14: Photograph of square threaded duct insertion (isometric)
A square threaded helical profile of 12.7 mm pitch is created inside the circular duct
with a depth of 1.5mm. Now for enhanced the heat transfer P/e i.e. ratio of pitch to
depth of thread, is calculated and comes out to be 8.46 which is close to optimum
value (optimum value is in between 7 & 10)
Plain profile:
Plain profile is created by simply boring two pieces with a drill bit of first 17mm and
then 25mm which is the finalized inside diameter of the plain profile. Outside
diameter is kept same as before equal to 28mm.
Mechanical Engineering Department, NIT, Srinagar 21
Fig 2.15: Solid works design of smooth duct
2.2.9 Data acquisitionData acquisition is the process of sampling signals that measures real world physical
conditions and converts the resulting samples into digital numeric values that can be
manipulated by a computer. DAQs typically convert analog waveforms into digital
values for processing. The components of data acquisition system include:
i. Sensors that convert physical parameters to electrical signals
ii. Signal conditioning circuitry to convert sensor signals into a form that can be converted to digital values
iii. Analog-to-digital converters, which convert conditioned sensor signals to digital values
Data acquisition applications are controlled by software programs developed using
various general purpose programming languages. Data acquisition begins with the
physical phenomenon or physical property to be measured. Examples of this include
temperature, light intensity, gas pressure, fluid flow, and force. Regardless of the type
of physical property to be measured, the physical state that is to be measured must
first be transformed into a unified form that can be sampled by a data acquisition
system. The task of performing such transformations falls on devices called sensors.
A data acquisition system is a collection of software and hardware that lets you
measure or control physical characteristics of something in the real world. A
complete data acquisition system consists of DAQ hardware, sensors and actuators,
signal conditioning hardware, and a computer running DAQ software.
A sensor, which is a type of transducer, is a device that converts a physical property
into a corresponding electrical signal (e.g., strain gauge, thermistor). An acquisition
system to measure different properties depends on the sensors that are suited to detect
those properties. Signal conditioning may be necessary if the signal from the
transducer is not suitable for the DAQ hardware being used. The signal may need to
be filtered or amplified in most cases. Various other examples of signal conditioning
might be bridge completion, providing current or voltage excitation to the sensor,
Mechanical Engineering Department, NIT, Srinagar 22
isolation, linearization. For transmission purposes, single ended analog signals, which
are more susceptible to noise can be converted to differential signals. Once digitized,
the signal can be encoded to reduce and correct transmission errors. DAQ hardware
is what usually interfaces between the signal and a PC. DAQ device drivers are
needed in order for the DAQ hardware to work with a PC
We have used Ajinkya IM 2000 series 8 channel DAQ. It has RS 483 connection
protocol. It connects directly the computer via a USB.
Fig 2.16: Photograph of data logger
2.2.10 ManometerA manometer is a pressure measuring instrument, or pressure gauge, often limited to
measuring pressures near to atmospheric pressure. Manometers come in two types
Analog manometers and Digital manometers. A digital manometer is used. Digital
manometers are microprocessor based instruments that can be stationary or mobile.
They have output capabilities that can be used for process control or transferring the
measurement data. Digital manometers are excellent for in-the-field measurement
and process control tasks because they can be networked. The manometer we are
using for our project has following specifications:
Table 2.1 Specification of manometer
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Model HTC 6205
Accuracy ±0.3% FSD
Reliability ±.2% FSO
Pressure range ± 5 psi
psi .001
Units and mBar .1
resolution kPa .01
cmH2O .001
Mechanical Engineering Department, NIT, Srinagar 24
Fig 2.17: Photograph of digital manometer
2.2.11 VoltmeterThe voltmeter is used to measure the voltage drop across the heating element. An
EBRIT V11 single phase digital voltmeter with range 0-750 V AC is used in this
setup. It is connected in parallel to the heating element. The resolution of the device
is 0.1volts. The wiring diagram is shown below in the figure 2.18.
Mechanical Engineering Department, NIT, Srinagar 25
Fig 2.18: Wiring diagram of voltmeter
2.2.12 AmmeterAn ammeter is used to measure the current flowing through the heating element. An
EBRIT A11 single phase digital ammeter with range 0-99 A AC is used in this setup.
It is connected in series to the heating element. The resolution of the device is
0.01Amps. The wiring diagram is shown below in the figure 2.19.
Fig 2.19: Wiring diagram of ammeter
Chapter 3
METHODOLOGY
The test facility and experimental setup used has already been discussed in previous
chapters. The ducts (smooth and ribbed) inserted separately in the test section are
under study. At various positions of the ball valve (full open, half open and ¼ open),
the temperature readings are recorded using the data logger interfaced with the
computer to calculate mean wall temperature, bulk mean temperature, inlet and outlet
temperature. The heat transfer coefficient is then determined using Newton’s law of
Mechanical Engineering Department, NIT, Srinagar 26
cooling for each case separately.
The pressure drop across the ducts is measured using the manometer. This reading is
then used to calculate friction factor using the suitable equation.
The pressure drop across an orifice plate is recorded using the manometer. This
pressure drop is used to calculate mass flow rate across the orifice plate from which
mean velocity of air is determined using equation of continuity. This now allows us to
calculate Reynolds number which will tell us whether the flow is laminar or turbulent.
To validate our results, theoretical Nusselt number is calculated using Dittus-Boelter
equation and compared with the experimentally calculated Nusselt number.
The experiment is performed on two different types of duct lining.
i. Smooth lining
ii. Square threaded lining
The performance parameters that need to be investigated in both the cases are:
i. Mass flow rate, ṁ
ii. Mean velocity of air, Vm
iii. Reynolds number, Re
iv. Heat transfer coefficient, h
v. Friction factor, ƒ
vi. Nusselt number, Nu
Ducts
Smooth Ribbed (Square threaded)
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Full open half open ¼ open Full open half open ¼ open
ṁ ṁ ṁ ṁ ṁ ṁ
Vm Vm Vm Vm Vm Vm
Re Re Re Re Re Re
h h h h h h ƒ ƒ ƒ ƒ ƒ ƒ
Nu Nu Nu Nu Nu Nu
Fig 3.1 Flow chart of methodology
3.1 Governing equationsThe heat input to the system is given by
Q = VI
Where V is the voltage measured and I is the current measured.
Also,
Q1 = hA∆T
Where Q1 is the convective heat transfer to the flowing air
And,
∆T = Tw - Tb
Where,
Tw= {(T2 + T3 + T4 + T5 + T6 + T7)/6}
And,
Tb= (Tinlet + Toutlet)/2
Also heat loss,
Q2 = (Tw-T8)/Rth
Where,
Rth = {ln (r2/r1)/2 LK}
Mechanical Engineering Department, NIT, Srinagar 28
Where L is the length of duct
And, K is the thermal conductivity of the insulating material.
From the equilibrium equation
Heat input = convective heat transfer + heat losses through insulation
Q= Q1 + Q2
The experimentation is divided into two stages with mass flow rate as the varying
parameter.
a) Experimentation is carried out without internal threads.
b) Experimentation is carried out with internal threads throughout duct
(p=12.7mm).
The data reduction of the measured results is summarized in the following
procedures:
We have
Tw = {(T2 + T3 + T4 + T5 + T6 + T7)/6}
And
T b = (Tinlet + Toutlet) /2
Volumetric flow rate of air,
Q= Cd√2gh
Where, Cd= Coefficient of discharge
Velocity of air flow,
V = (Q/A)
Where, A = area of circular duct, πd2/4
Reynolds Number (Re)
Re = ( ρd/µ)
Nusselt number (Nu)
Nu=hd/k
Where, K is the thermal conductivity of air.
Friction factor (ƒ)
Mechanical Engineering Department, NIT, Srinagar 29
ƒ =2∆PdρL 2
Where ∆p is the pressure drop in the duct, measured by digital manometer.
Thermal enhancement factor (η):
The enhancement efficiency (η) is defined as the heat transfer coefficient for the tube
with internal to that for the plain tube without internal threads at constant Reynolds
number as follows
η = h with internalthread
h without internal threads
Where h is the convective heat transfer coefficient.
Chapter 4
RESULTS AND DISCUSSION
Controlling the mass flow rate with the help of ball valve, we can control the air
flown in the duct. There can be lot of possibilities to control the valve position
and hence mass flow rate, but, we have taken only three positions.
1. Fully open valve
Mechanical Engineering Department, NIT, Srinagar 30
2. Half open valve
3. ¼ open valve
The change in mass flow rate has its effect on the mean velocity, Reynolds no,
heat transfer coefficient, friction factor and Nusselt no. The decrease in mass flow
rate tends to decrease mean velocity, Reynolds no, heat transfer coefficient,
friction factor and Nusselt no. The summary of variations is given in tables, 4.1 to
4.6, for the three different valve positions and two surface profile cases.
The duct surface profile also has an effect on the heat transfer. The introduction of
threads increased the Reynolds no from 7277.72 to 8271.25 for full open valve,
from 5811 to 5845 for half open valve and from 3966 to 3998 for quarterly open
valve. The, heat transfer coefficient increased from 20 to 29 for full open valve,
from 14.3 to 21 for half open valve and from 9.65 to 16.08 for quarterly open
valve. This is because of the turbulence caused in the fluid by the threaded profile.
Turbulence results in better mixing of the fluid layers which in turn results in
better heat transfer. The friction factor and the Nusselt no also increased
significantly. An increasing variation was seen across all three valve positions by
introduction of threads.
The results have been tabulated overleaf.
4.1 Smooth Duct
Table 4.1: Performance parameters (full open)
S.No Parameters studied Experimental values
1 Mass flow m m =2.7324 kg/s
2 Mean velocity, vm vm =4.4786 m/s3 Reynolds Number Re Re = 7277.7254 Heat transfer coefficient h h =20.03
5 Friction factor f 4.234
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6 Nusselt number Nu= hd/k Nu= 19.85
Table 4.2: Performance parameters (half open)S.No Parameters studied Experimental values
1 Mass flow m m =2.186 kg/s
2 Mean velocity, vm vm =3.576 m/s3 Reynolds Number Re Re =58114 Heat transfer coefficient h h =14.3
5 Friction factor f 6.64
6 Nusselt number Nu= hd/k Nu= 16.03
Table 4.3: Performance parameters (¼ open valve)
S.No Parameters studied Experimental values
1 Mass flow m m =1.4918 kg/s
2 Mean velocity, vm vm= 2.44 m/s3 Reynolds Number Re Re =3966.6254 Heat transfer coefficient h h =9.65
5 Friction factor f 14.254
6 Nusselt number Nu= hd/k Nu =14
4.2 Square threaded duct
Table 4.4: Performance parameters (Full open valve)
S.No Parameters studied Experimental values
1 Mass flow m m = 2.7205 kg/s
2 Mean velocity, vm vm= 5.09 m/s3 Reynolds Number Re Re =8271.254 Heat transfer coefficient h h =29
5 Friction factor f 18
6 Nusselt number Nu= hd/k Nu= 26.8
Mechanical Engineering Department, NIT, Srinagar 32
Table 4.5: Performance parameters (half open valve)
S.No Parameters studied Experimental values
1 Mass flow m m =2.2 kg/s
2 Mean velocity, vm vm =3.59 m/s3 Reynolds Number Re Re =58454 Heat transfer coefficient h h =21
5 Friction factor f 25
6 Nusselt number Nu= hd/k Nu= 20.68
Table 4.6: Performance parameters (¼ open valve)
S.No Parameters studied Experimental values
1 Mass flow m m =1.515 kg/s
2 Mean velocity, vm vm =2.46 m/s3 Reynolds Number Re Re =39984 Heat transfer coefficient h h =16.08
5 Friction factor f 35
6 Nusselt number Nu= hd/k Nu= 15.83
4.3 Pressure drop variationAs air flows through a duct its total pressure drops in the direction of flow. The
pressure drop is due to:
1. Fluid friction
2. Momentum change due to change of direction and/or velocity
The pressure drop due to friction is known as frictional pressure drop or friction loss,
Δpf. The pressure drop due to momentum change is known as momentum pressure
drop or dynamic loss, Δpd. Thus the total pressure drop Δpt is given by
Δpt = Δpf+ Δpd
Mechanical Engineering Department, NIT, Srinagar 33
Pressure drop increases with increase in Reynolds number. Pressure drop is observed
to be more in a threaded duct compared to that of plain test duct. The large increase in
the pressure drop can be attributed to the large value of friction factor in threaded
ducts and the increased velocity associated more intense swirl flow in case of more
depth.
4.4 Nusselt number and Reynolds number variation
The variation of Nusselt number with Reynolds number in the plain tube and square
threaded test tube with threads of constant pitch is shown in the graph. It is observed
that Nusselt number increases with increasing Reynolds number. It is observed that
for tube with internal threads the heat transfer rates are higher than those from the
plain tube. This is due to the fact that the threads increase the turbulent intensity of air
across the range of Reynolds numbers which results in better intermixing of the air in
the test duct. Due to this the average bulk temperature of the air is increased and so
Mechanical Engineering Department, NIT, Srinagar 34
the convective heat transfer. Mean Nusselt numbers for test tubes with internal square
threads is better than that for the plain tube.
3500 4000 4500 5000 5500 6000 6500 7000 75000
5
10
15
20
25
30
Plain LiningThreaded Lining
Reynolds Number
Nus
selt
Num
ber
Plot 5.2 Nusselt Number Vs Reynolds Number
4.5 Friction factor and Reynolds number variation
The variation of friction factor v/s Reynolds number for the plain tube and square
threaded one of constant pitch is shown in figure. The friction factor for the test tube
using internal threads is more than that for plain test tube. Also friction factor
decreases with increase in Reynolds number for the square threaded. This shows that
the turbulence formation advanced due to artificial turbulence exerted by internal
threads. Due to increase in swirl of flow and formation of eddies flow there is a
significant increase in the head loss or pressure energy loss in threaded duct ,however
Mechanical Engineering Department, NIT, Srinagar 35
as Reynolds number increases the flow there is marked decline in the friction factor .
This can be attributed to the inverse relation of friction factor with velocity from
Darcy Wiesbach equation
Δ P=
fL( ρ v2)2 d
3500 4000 4500 5000 5500 6000 6500 7000 75000
5
10
15
20
25
30
35
40
Smooth LiningThreaded Lining
Reynolds Number
Fricti
on F
acto
r x 1
0-5
Plot 5.3 Friction factor Vs Reynolds No
4.5 Heat transfer coefficient and Reynolds number variationHeat transfer coefficient (h) generally increases with increase in Reynolds number
(Re). However there is a steep rise in the case of internal threaded ducts.
The variation of heat transfer coefficient with Reynolds number in the plain tube and
square threaded test tube with threads of constant pitch is shown in the graph. It is
observed that h increases with increasing Reynolds number as is seen in Nusselt
number
Mechanical Engineering Department, NIT, Srinagar 36
3500 4000 4500 5000 5500 6000 6500 7000 75000
5
10
15
20
25
30
35
Smooth liningThreaded Lining
Reynolds Number
Heat
tran
sfer
Coe
fficie
nt
Plot 5.4 Heat Transfer Coefficient Vs Reynolds Number
Chapter 6
CONCLUSIONS
Experimental investigations of heat transfer, friction factor and thermal enhancement
factor of a plain circular tube and a circular tube with internal square threads of
constant pitch were studied. The following conclusions are drawn.
Mechanical Engineering Department, NIT, Srinagar 37
1. The heat transfer for a duct with square threads increases by 13.9%. This is due to the fact that the threads hinder the free movement of air particles in the test section, which increases the turbulence of air. Due to increase in turbulence, better intermixing of air particles takes place which result in average increase of bulk mean temperature of air.
2. The friction factor increases for a duct with square threads as compared to a plain duct due to swirl flow caused by wake formation in the square threads. The increase in friction factor is about 200 percent.
3. The enhancement of Nusselt number is much higher than enhancement in friction factor for the square type internal threads that justifies the usage of internal threads in horizontal tube.
4. The performance of circular tube can be improved by the use of internal threads. The cost involved for making internal threads is minimal compared to energy efficiency improvement provided by this technique.
Future scope:The heat transfer enhancement for different types threads viz acme, buttress, knuckle,
etc, different grooves viz square, triangular etc and rib profiles can be calculated to
optimise the heat transfer enhancement.
Further varying p/e ratio of the different profiles can be done to optimise the p/e ratio
for swirl flow formation. The optimum values are between 7 and 10.
REFERENCESAgarwal, S.K. and Raja Rao, M. (1996), “Heat transfer augmentation for flow of
viscous liquid in circular tubes using twisted tape inserts”, International Journal of
Heat Mass Transfer, Vol.99, pp.3547-3557.
Angirasi, D. (2001), “Experimental investigation of forced convection heat transfer
augmentation with metallic porous materials”, International Journal of Heat Mass
Mechanical Engineering Department, NIT, Srinagar 38
Transfer, pp. 919-922.
Date, A.W. and Singham, J.R.(1972), “Numerical prediction of friction and heat
transfer characteristics of fully developed laminar flow in tubes containing twisted
tapes”, Trans. ASME, Journal of Heat Transfer, Vol. 17, pp.72
Eiamsa-ard, S., Thianpong, C., Eiamsa-ard, P. and Promvonge P.(2009), “Convective
heat transfer in a circular tube with short-length twisted tape insert”, International
communications in heat and mass transfer (2009).
Fu, H.L., Leong, K.C., Huang X.Y. and Liu C.Y. (2001), “An experimental study of
heat transfer of a porous channel subjected to oscillating flow”, ASME Journal of
Heat Transfer, Vol. 123, pp. 162-170.
Hsieh, S.S.,Liu, M.H. and Tsai, H.H. (2003). “Turbulent heat transfer and flow
characteristic in a horizontal circular tube with strip-type inserts part-II (heat
transfer)”, International Journal of Heat and Mass Transfer, Vol.46, pp.837-849.
Liao,Q., and Xin, M.D. (2000),”Augmentation of convective heat transfer inside
tubes with three-dimensional internal extended surfaces and twisted-tape inserts”,
Chemical Engineering Journal,Vol.78, pp.95-105.
Pavel, B.L., and Mohamad, A.A. (2004), “An experimental and numerical study on
heat transfer enhancement for gas heat exchangers fitted with porous media”,
International Journal of Heat and Mass Transfer, Vol.47, pp.4939-4952.
Peterson, S.C., France, D.M. and Carlson, R.D. (1989), “Experiments in high
pressure turbulent swirl flow”, Trans. ASME, Journal of Heat Transfer, Vol.108,
pp.215-218.
Rao, M.M. and Sastri, V.M.K. (1995), “Experimental investigation for fluid and heat
transfer in a rotating tube twisted tape inserts”, International Journal of Heat and
Mass Transfer, Vol.16, pp.19-28.
Mechanical Engineering Department, NIT, Srinagar 39
Saha, S.K., Gaitonde, U.N. and Date, A.W. (1989), “Heat transfer and pressure drop
characteristics of laminar flow in a circular tube fitted with regularly spaced twisted-
tape elements”, Journal of Exp. Thermal Fluid Sci., Vol.2, pp. 310-322.
Sivashanmugam, P. and Suresh, S. (2007), “Experimental studies on heat transfer and
friction factor characteristics of turbulent flow through a circular tube fitted with
regularly spaced helical screw tape inserts”, Experimental Thermal and Fluid
Science, Vol.31, pp. 301-308.
Sozen, M. and Kuzay, T.M.(1996), “Enhanced heat transfer in round tubes with
porous inserts”, International Journal Heat and Fluid Flow, Vol.!7, pp.124-129.
Thianpong, C., Eiamsaard, P., Wongcharee, K. and Eiamsaard, S. (2009),
“Compound heat transfer enhancement of a dimpled tube with a twisted tape swirl
generator”, International Communications in Heat and Mass Heat and Mass transfer,
Vol.36, pp.698-704.
Whitham, J.M. (1896), “the effects of retarders in fire tubes of steam boilers”, Street
Railway, Vol.12(6), pp.374.
APPENDIX-A1
CALCULATIONS FOR SMOOTH DUCT
A1.1 Full open valve
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In this case the experiment where carried out and the following readings were
obtained
S No. T1 T2 T3 T4 T5 T6 Tsurface Tinput Toutput Twall Tbulk
1 69 70 78 69 70 66 45 25 39 70.33 32
2 73 70 82 80 71 70 45 25 39 74.33 32
3 74 71 80 85 71 70 45 25 40 75.16 32.5
4 79 73 82 88 72 74 45 25 40 78 32.5
5 78 74 85 90 72 74 46 25 40 78.83 32.5
6 79 74 85 93 74 76 46 25 40 80.16 32.5
7 86 80 100 94 80 83 47 25 41 87.16 33
8 89 81 104 94 81 85 48 25 42 89 33.5
9 90 82 105 95 83 87 48 25 42 90.33 33.5
Average 46.1 25 40.3
3
80.366 32.6
Twall =T 1+T 2+T 3+T 4+T 5+T 6
6 = 80.366 is the mean wall temperature.
Tbulk = T input+T output2 = 32.68 is the bulk mean temperature.
Area of cross section
Ac = π4
×d2
Ac =5.3066×10−4 m2
Surface area
As =2πrL
As=0.0490 m2
Head measured at orifice plate, ∆H =4.7 cm of water.
Mechanical Engineering Department, NIT, Srinagar 41
As we know that air flow through orifice plate is calculated in terms of
the head of the air,
ρwater ∆Hwater =ρair ∆Hair
∆Hair = 1000× 0.0471.15 = 40.86 m of air.
As we know that volumetric flow rate for the orifice plate is given by
Q = C d A c A√ A c2−A2
√ 2× g ×∆Hair
Therefore,
Q = 0.613× 5.306 ×1.3266 ×28.35.138
×10−4
Hence,
Q =2.376×10−3 m3/s
Mass flow
ṁ =ρair Q
ṁ =2.7324×10−3 kg/s
Mean velocity,
vm = QA c
vm =4.4786 m/s
Reynolds Number
Re = vm×d/β
Mechanical Engineering Department, NIT, Srinagar 42
Re =7277.725
From the equilibrium equation we have,
qo = q1 + q2 ---------------------------------------------(A1)
Where qo is the convective heat transfer from the walls of duct into the
fluid and q1 is the heat absorbed by the air while passing through heated
duct, q2 is the heat loss by conduction through insulation at temperature
Tsurface
qo =h As (Tw - Tb) ----------------------------------
(A2)
q1 =m CP (Toutlet - Tinlet) -------------------------------
(A3)
q2 =
T w−T surfaceln (r 2)
(r )2πlK
-----------------------------------------
(A4)
Using equations (A2), (A3) and (A4) in the equation (A1) we get:
h (0.0490) (80.366-32.68) = (2.7324)(1.005)(40.36-
24.7)+(3.8)
h =20.03
Coefficient of friction f
Friction factor f = 4f
Losses due to friction will create differential pressure head that is given by ∆Hf
∆Hf = fL
2dgv2
Mechanical Engineering Department, NIT, Srinagar 43
∆Hf = 0.1 cm of water
f = 4.234×10−5
Verifying the data
Using empirical relation to find Nusselt number by Dittus-Boelter equation.
Theoretically, Nusselt number is
Nu theoretical = 0.0243(Re)0.8(Pr)0.4
Nu theoretical = 24
Experimentally, Nusselt number is
Nu= hd/k
Nu= 19.85
A1.2 Half open valve
In this case the experiment where carried out and the following readings were
obtained
S No. T1 T2 T3 T4 T5 T6 Tsurface Tinput Toutput Twall
1 93 94 108 95 90 89 50 25 40 94.83
2 98 99 108 100 100 94 50 25 42 99.83
3 97 103 111 98 95 94 52 25 43 99.6
Average 50.67 25 41.7 98.08
Twall =T 1+T 2+T 3+T 4+T 5+T 6
6 =98 is the mean wall temperature.
Tbulk = T input+T output
2 = 33.4 is the bulk mean temperature
Area of cross section
Mechanical Engineering Department, NIT, Srinagar 44
Ac = π4
×d2
A c=5.3066 ×10−4 m2
Surface area
As =2πrL
A s=0.0490 m2
Head measured at orifice plate, ∆ H=3 cm of water.
As we know that air flow through orifice plate is calculated in terms of the head of the
air,
ρwater ∆Hwater =ρair ∆Hair
∆ H air=1000× 0.031.15
=26.0869 m of air .
As we know that volumetric flow rate for the orifice plate is given by
Q = C d A c A√ A c2−A2
√ 2× g ×∆Hair
Therefore,
Q = 0.613× 5.306 ×1.3266 ×22.61
5.138× 10−4
Hence,
Q=1.8979 ×10−3 m3/ s
Mass flow
ṁ =ρair Q
ṁ = 2.1826×10−3 kg/s
Mechanical Engineering Department, NIT, Srinagar 45
Mean velocity,
Vm = QA c
Vm =3.576 m/s
Reynolds Number
Re = vm×d/β
R e=5811
From the equilibrium equation we have
qo = q1 +q2 -------------------------------------------(A5)
where qo is the convective heat transfer from the walls of duct into the fluid and q1 is
the heat absorbed by the air while passing through heated duct, q2 is the heat loss by
conduction through insulation at temperature Tsurface
qo =h As (Tw - Tb) ------------------------------------------------(A6)
q1 =m CP (Toutlet - Tinlet) ---------------------------------------(A7)
q2=
T w−T surfaceln (r 2)
(r )2πlK
---------------------------------------------(A8)
Using equations (A6), (A7) and (A8) in the equation (A5) we get:
h (.049) (98.08-33.42) = (2.1826) (1.005) (41.7-25) + (5.24)
h =14.3
Coefficient of friction f
Friction factor f = 4f
Mechanical Engineering Department, NIT, Srinagar 46
Losses due to friction will create differential pressure head that is given by ∆Hf
∆Hf = fL
2dgv2
∆ H f =0.1 cm of water
f = 6.64×10−5
Verifying the data
Using empirical relation to find Nusselt number by Dittus-Boelter equation.
Theoretically, Nusselt number is
Nutheoritical = 0.0243(Re)0.8(Pr)0.4
Nutheoretical =20.78
Experimentally, Nusselt number is
Nu= hd/k
Nu= 16.03
A1.3 ¼ Open valve
Again mass flow rate is varied through ball valve and following readings were
obtained.
S No. T1 T2 T3 T4 T5 T6 Tsurface Tinput Toutput Twall
1 107 100 118 116 109 104 52 25 45 109
2 135 120 145 130 131 130 52 25 47 131.83
3 126 115 138 104 118 120 53 25 48 120.16
Averag
e52.33 25 46.67 120.3
Twall =T 1+T 2+T 3+T 4+T 5+T 6
6 = 120.3 is the mean wall temperature.
Mechanical Engineering Department, NIT, Srinagar 47
Tbulk = T input+T output
2 = 35.83 is the bulk mean temperature
Area of cross section Ac = π4
×d2
Ac =5.3066×10−4 m2
Surface area As =2πrL
As=0.0490 m2
Head measured at orifice plate, ∆H =1.4 cm of water.
As we know, air flow through orifice plate is calculated in terms of head of air,
ρwater ∆Hwater =ρair ∆Hair
∆Hair = 1000× 0.014
1.15 = 12.174 m of air.
As we know that volumetric flow rate for the orifice plate is given by
Q = C d A c A
√A c2−A 2√ 2× g ×∆Hair
Therefore,
Q = 0.613× 5.306 ×1.3266 ×15.447
5.138×10−4
Hence, Q =1.297×10−3 m3/s
Mass flow ṁ =ρair Q
ṁ =1.4918×10−3 kg/s
Mean velocity,
Vm = QA c
Vm =2.44 m/s
Reynolds Number
Mechanical Engineering Department, NIT, Srinagar 48
Re = Vm×d/β
Re =3966.625
From the equilibrium equation we have
qo = q1 +q2 ----------------------------------------------------(A9)
where qo is the convective heat transfer from the walls of duct into the fluid and q1 is
the heat absorbed by the air while passing through heated duct, q2 is the heat loss by
conduction through insulation at temperature Tsurface
qo =h As (Tw - Tb) ---------------------------------------------(A10)
q1 =m CP (Toutlet - Tinlet) ---------------------------------------(A11)
q2 =
T w−T surfaceln (r 2)
(r )2πlK
---------------------------------------------
(A12)
Using equations (A10), (A11) and (A12) in the equation (A9) we get:
h(0.0490) (120.3-35.83) = (1.491)(1.005)(46.67-25)+(7.47)
h =9.65
Coefficient of friction f
Friction factor f = 4f
Losses due to friction will create differential pressure head that is given by ∆Hf
∆Hf = fL
2 dgv2
∆Hf = 0.1 cm of water
f = 14.254×10−5
Verifying the data
Using empirical relation to find Nusselt number by Dittus-Boelter equation.
Mechanical Engineering Department, NIT, Srinagar 49
Theoretically, Nusselt number is
Nutheoritical = 0.0243(Re)0.8(Pr)0.4
Nutheoretical = 16
Experimentally, Nusselt number is
Nu= hd/k
Nu= 14
APPENDIX-A2
CALCULATIONS FOR THREADED DUCT
A2.1 Full open valve
In this case the experiment where carried out and the following readings were
obtained
S No. T1 T2 T3 T4 T5 T6 Tsurface Tinput Toutput Twall
1 58 60 74 80 54 58 33 25 40 64
2 80 74 76 50 54 80 36 25 41 69
3 50 70 75 70 50 40 34 25 40. 59
Averag
e40.3
Twall =T 1+T 2+T 3+T 4+T 5+T 6
6 = 64 is the mean wall temperature.
Tbulk = T input+T output
2 = 32.68 is the bulk mean temperature
Area of cross section
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Ac = π4
×d2
Ac =5.3066×10−4 m2
Surface area
As =2πrL
As=0.0490 m2
Head measured at orifice plate, ∆H =4.6 cm of water.
As we know that air flow through orifice plate is calculated in terms of the head of the
air,
ρwater ∆Hwater =ρair ∆Hair
∆Hair = 1000× 0.014
1.15 = 40 m of air.
As we know that volumetric flow rate for the orifice plate is given by
Q = C d A c A√ A c2−A 2
√ 2 × g×∆Hair
Therefore,
Q=0 .613 ×5 . 306× 1. 3266 × 405 . 138
× 10−4
Hence,
Q =2.35×10−3 m3/s
Mass flow
ṁ=ρair Q
ṁ =2.7025×10−3 kg/s
Mean velocity,
Mechanical Engineering Department, NIT, Srinagar 51
Vm = Q
A c
V =5.09 m/s
Reynolds Number
Re = vm×d/β
Re =8271.25
From the equilibrium equation we have
qo = q1 +q2 ----------------------------------------(A13)
where qo is the convective heat transfer from the walls of duct into the fluid, q1 is the
heat absorbed by the air while passing through heated duct and q2 is the heat loss by
conduction through insulation at temperature Tsurface
qo =h As (Tw - Tb) ---------------------------------(A14)
q1 =m CP (Toutlet - Tinlet) -----------------------------(A15)
q2 =
T w−T surfaceln (r 2)
(r )2 πlK
----------------------------------(A16)
Using equations (A14), (A15) and (A16) in the equation (A13) we get:
h(0.0490) (64-32.68) = (2.7025)(1.005)(40.36-25)+(3.33)
h =29
Coefficient of friction f
Friction factor f = 4f
Losses due to friction will create differential pressure head that is given by ∆Hf
∆Hf = fL
2dgv2
∆Hf = 0.55cm of water
Mechanical Engineering Department, NIT, Srinagar 52
f = 18.03×10−5
Verifying the data
Using empirical relation to find Nusselt number by Dittus-Boelter equation.
Theoretically, Nusselt number is
Nutheoritical = 0.0243(Re)0.8(Pr)0.4
Nutheoretical =28.88
Experimentally, Nusselt number is
Nu= hd/k
Nu= 26.8
A2.2 Half open valve
In this case the experiment was carried out in a similar way, as in smooth lining and
the following observations were recorded
Twall =T 1+T 2+T 3+T 4+T 5+T 6
6 =98 is the mean wall temperature
Toutput =49.06
Tbulk = T input+T output
2 =37.03 is the bulk mean temperature
Area of cross section
Ac = π4
×d2
Ac =5.3066×10−4 m2
Surface area
Mechanical Engineering Department, NIT, Srinagar 53
As =2πrL
As=0.0490 m2
Head measured at orifice plate, ∆H = 3.2 cm of water.
As we know that air flow through orifice plate is calculated in terms of the head of the
air,
ρwater ∆Hwater =ρair ∆Hair
∆Hair = 1000× 0.032
1.15 = 28.346 m of air.
As we know that volumetric flow rate for the orifice plate is given by
Q = C d A c A
√ A c2−A 2√ 2× g ×∆Hair
Therefore,
Q = 0.613× 5.306 ×1.3266 ×22.61
5.138× 10−4
Hence,
Q =1.913×10−3 m3/s
Mass flow
ṁ=ρair Q
ṁ =2.2×10−3 kg/s
Mean velocity,
Vm = Q
A c
Vm =3.59 m/s
Reynolds Number
Re = vm×d/β
Mechanical Engineering Department, NIT, Srinagar 54
Re =5845
From the equilibrium equation we have
qo = q1 +q2 -------------------------------------------(A17)
where qo is the convective heat transfer from the walls of duct into the fluid, q1 is the
heat absorbed by the air while passing through heated duct and q2 is the heat loss by
conduction through insulation at temperature Tsurface
qo =h As (Tw - Tb) ----------------------------------------(A18)
q1 =m CP (Toutlet - Tinlet) -----------------------------------(A19)
q2 =
T w−T surfaceln (r 2)
(r )2 πlK
----------------------------------------
(A20)
Using equations (A18), (A19) and (A20) in the equation (A17) we get:
h(0.0490) (98.08-37.03) = (2.2)(1.005)(49.06-25)+(5.24)
h = 21
Coefficient of friction f
Friction factor f = 4f
Losses due to friction will create differential pressure head that is given by ∆Hf
∆Hf = fL
2dgv2
f = 25×10−5
Verifying the data
Using empirical relation to find Nusselt number by Dittus-Boelter equation.
Theoretically, Nusselt number is
Mechanical Engineering Department, NIT, Srinagar 55
Nutheoritical = 0.0243(Re)0.8(Pr)0.4
Nutheoretical =21.7
Experimentally, Nusselt number is
Nu= hd/k
Nu= 20.68
A2.3 ¼ Open valve
In this case the experiment was carried out in a similar way, as in smooth lining and
the following observations were recorded
Twall =T 1+T 2+T 3+T 4+T 5+T 6
6 =120.3 is the mean wall temperature
Toutput =64.3
Tbulk = T input+T output
2 =44.64 is the bulk mean temperature
Area of cross section
Ac = π4
×d2
Ac =5.3066×10−4 m2
Surface area
As =2πrL
As=0.0490 m2
Head measured at orifice plate, ∆H =1.53 cm of water.
As we know that air flow through orifice plate is calculated in terms of the head of the
air,
ρwater ∆Hwater =ρair ∆Hair
Mechanical Engineering Department, NIT, Srinagar 56
∆Hair = 1000× 0.0153
1.15 = 13.29 m of air.
As we know that volumetric flow rate for the orifice plate is given by
Q = C d A c A
√ A c2−A 2√ 2× g ×∆Hair
Therefore,
Q = 0.613× 5.306 ×1.3266 ×15.447
5.138×10−4
Hence,
Q =1.321×10−3 m3/s
Mass flow
ṁ =ρair Q
ṁ =1.51×10−3 kg/s
Mean velocity,
Vm = Q
A c
Vm =2.46 m/s
Reynolds Number
Re = Vm×d/β
Re =3998
From the equilibrium equation we have
qo = q1 +q2 ------------------------------------------(A21)
Mechanical Engineering Department, NIT, Srinagar 57
Where qo is the convective heat transfer from the walls of duct into the fluid, q1 is the
heat absorbed by the air while passing through heated duct and q2 is the heat loss by
conduction through insulation at temperature Tsurface
qo =h As (Tw - Tb) ---------------------------------------(A22)
q1 =m CP (Toutlet - Tinlet) ---------------------------------(A23)
q2 =
T w−T surfaceln (r 2)
(r )2πlK
------------------------------------(A24)
Using equations (22), (23) and (24) in the equation(21) we get:
h (0.0490) (120.3-44.64) = (1.515)(1.005)(64.3-25)+(7.47)
h =16.08
Coefficient of friction f
Friction factor f = 4f
Losses due to friction will create differential pressure head that is given by ∆Hf
∆Hf = fL
2dgv2
f = 35×10−5
Verifying the data
Using empirical relation to find Nusselt number by Dittus-Boelter equation
Theoretically, Nusselt number is
Nutheoritical = 0.0243(Re)0.8(Pr)0.4
Nutheoretical =16.02
Experimentally, Nusselt number is
Nu= hd/k
Mechanical Engineering Department, NIT, Srinagar 58
Nu= 15.83
APPENDIX-B
Cost Analysis
Mechanical Engineering Department, NIT, Srinagar 59
Table B1: Fabrication cost
S. No. Items Cost (Rs.)
1. Temperature indicators 2,400
2. Data acquisition system 20,000
3. Digital Voltmeter ac/dc 900
4. Digital Ammeter ac/dc 900
5. Heating element 1,000
6. Digital manometer 6,500
7. Miscellaneous 3,300
Total build up cost 35,000
Table B2: Market price
S.No. Items Cost (Rs.)
Mechanical Engineering Department, NIT, Srinagar 60
1. Forced convection apparatus 35,000
2. Data acquisition system 25,000
Total market cost 60,000
Savings:
Total market – Actual fabrication cost = Rs. (60,000-35,000)
= Rs.25,000
Mechanical Engineering Department, NIT, Srinagar 61