Study of forced convection in smooth and ribbed ducts

Mechanical Engineering Department, NIT, Srinagar i
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 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.

Mechanical Engineering Department, NIT, Srinagar v
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
1.5 Objective 6
1.6 Literature survey 6
Chapter 2: Test Facility 10-25

Mechanical Engineering Department, NIT, Srinagar vi
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 37
REFERENCES 38-39
APPENDIX A1 40-48
APPENDIX A2 49-57
APPENDIX 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
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

Mechanical Engineering Department, NIT, Srinagar viii

Mechanical Engineering Department, NIT, Srinagar ix
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
p helical rib pitch , m
Q heat transfer rate , kW
T temperature C

Mechanical Engineering Department, NIT, Srinagar x
U velocity , m/s
Ρ density , kg/m3

Mechanical Engineering Department, NIT, Srinagar 1
Chapter 1
INTRODUCTION
1.1 Historical Background
Heat 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 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
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
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 definitions
In this section a few important terms commonly used in heat transfer augmentation work
are defined.
1.3.1 Thermo-hydraulic performance
For 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 ratio
The 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, Nu
The 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, Pr
The Prandtl number is defined as the ratio of the molecular diffusivity of momentum to
the molecular diffusivity of heat.
1.3.5 Pitch
Pitch 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 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 Objectives
The 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 survey
There 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.
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.
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.
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.
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 setup
The 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
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. Blower
The 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 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
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
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.
Plot 2.1 (source: Wikipedia)
2.2.3 Test Section
The 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
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.
Fig 2.8: Schematic of heater plate
2.2.5 Orifice Plate
The 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
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 Valve
In 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.
Fig 2.10: Solid works design of ball valve
2.2.8. Duct lining profiles
Threadedprofile
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
Fabricationof square threads
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
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.
Fig 2.15: Solid works design of smooth duct
2.2.9 Data acquisition
Data 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, 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 Manometer
A 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
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

Fig 2.17: Photographof digital manometer
2.2.11 Voltmeter
The 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.

Fig 2.18: Wiring diagram of voltmeter
2.2.12 Ammeter
An 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 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)
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 equations
The 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}
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)
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 internal thread
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
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 meanvelocity, 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 increasedthe 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 increasedfrom 20 to 29 for full openvalve, 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 inbetter 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/s
3 Reynolds Number Re Re = 7277.725
4 Heat transfer coefficient h h =20.03
5 Friction factor f 4.234
6 Nusselt number Nu= hd/k Nu= 19.85
Table 4.2: Performance parameters (half open)
3 Reynolds Number Re Re =5811
Table 4.3: Performance parameters (¼ open valve)
2 Mean velocity, vm vm= 2.44 m/s
3 Reynolds Number Re Re =3966.625
6 Nusselt number Nu= hd/k Nu =14

4.2 Square threaded duct
Table 4.4: Performance parameters (Full open valve)
1 Mass flow m m = 2.7205 kg/s
2 Mean velocity, vm vm= 5.09 m/s
3 Reynolds Number Re Re =8271.25
4 Heat transfer coefficient h h =29
5 Friction factor f 18
Table 4.5: Performance parameters (half open valve)
4 Heat transfer coefficient h h =21
Table 4.6: Performance parameters (¼ open valve)

4.3 Pressure drop variation
As 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
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 the convective heat
transfer. Mean Nusselt numbers for test tubes with internal square threads is better than
that for the plain tube.
Plot 5.2 Nusselt Number Vs Reynolds Number
0
5
10
15
20
25
30
0 2000 4000 6000 8000
NusseltNumber
Reynolds Number
Plain Lining
Threaded Lining

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 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
ΔP =
fL(ρv2
)
2d
Plot 5.3 Friction factor Vs Reynolds No
0
5
10
15
20
25
30
35
40
0 2000 4000 6000 8000
FrictionFactorx10-5
Reynolds Number
Smooth Lining
Threaded Lining

4.5 Heat transfer coefficient and Reynolds number variation
Heat 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
Plot 5.4 Heat Transfer Coefficient Vs Reynolds Number
0
5
10
15
20
25
30
35
0 2000 4000 6000 8000
HeattransferCoefficient
Reynolds Number
Smooth lining
Threaded Lining

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.
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.

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APPENDIX-A1
CALCULATIONS FOR SMOOTH DUCT
A1.1 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 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 40.33 80.366 32.6
Twall =
𝐓𝟏+𝐓𝟐+𝐓𝟑+𝐓𝟒+𝐓𝟓+𝐓𝟔
𝟔
= 80.366 is the mean wall temperature.
Tbulk =
𝐓𝐢𝐧𝐩𝐮𝐭+𝐓𝐨𝐮𝐭𝐩𝐮𝐭
𝟐
= 32.68 is the bulk mean temperature.
Area of cross section
Ac =
𝝅
𝟒
× 𝒅 𝟐
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.
As we know that air flow through orifice plate is calculated in terms of the
head of the air,
𝝆water ∆Hwater = 𝝆air ∆Hair
∆Hair =
𝟏𝟎𝟎𝟎×𝟎.𝟎𝟒𝟕
𝟏.𝟏𝟓
= 40.86 m of air.
As we know that volumetric flow rate for the orifice plate is given by
Q =
𝐂𝐝 𝐀𝐜 𝐀
√ 𝐀𝐜^𝟐−𝐀^𝟐
√𝟐 × 𝒈 ×∆Hair
Therefore,
Q =
𝟎.𝟔𝟏𝟑×𝟓.𝟑𝟎𝟔×𝟏.𝟑𝟐𝟔𝟔×𝟐𝟖.𝟑
𝟓.𝟏𝟑𝟖
× 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 =
𝐐
𝐀𝐜
vm =4.4786 m/s

Reynolds Number
Re = vm×d/β
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 =
𝐓𝐰−𝐓𝐬𝐮𝐫𝐟𝐚𝐜𝐞
𝐥𝐧(𝐫𝟐)
(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
𝟐𝒅𝒈
𝒗 𝟐
∆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
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 =
𝐓𝟏+𝐓𝟐+𝐓𝟑+𝐓𝟒+𝐓𝟓+𝐓𝟔
𝟔
=98 is the mean wall temperature.
Tbulk =
𝟐
= 33.4 is the bulk mean temperature
Ac =
𝝅
𝟒
× 𝒅 𝟐
𝐀𝐜 = 𝟓. 𝟑𝟎𝟔𝟔 × 𝟏𝟎−𝟒
𝐦𝟐

Surface area
As =2𝝅𝒓𝑳
𝐀𝐬 = 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,
∆𝐇𝐚𝐢𝐫 =
𝟏𝟎𝟎𝟎×𝟎.𝟎𝟑
𝟏.𝟏𝟓
= 𝟐𝟔. 𝟎𝟖𝟔𝟗 𝐦 𝐨𝐟 𝐚𝐢𝐫.
Q =
√ 𝐀𝐜^𝟐−𝐀^𝟐
√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
Mean velocity,
Vm =
𝐐
𝐀𝐜
Vm =3.576 m/s

Reynolds Number
Re = vm×d/β
𝐑𝐞 = 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=
𝐥𝐧(𝐫𝟐)
(𝐫)
𝟐𝛑𝐥𝐊
---------------------------------------------(A8)
h (.049) (98.08-33.42) = (2.1826) (1.005) (41.7-25) + (5.24)
h =14.3
∆Hf =
𝒇𝑳
𝟐𝒅𝒈
𝒗 𝟐
∆𝐇𝐟 = 𝟎. 𝟏 𝐜𝐦 𝐨𝐟 𝐰𝐚𝐭𝐞𝐫
f = 6.64× 𝟏𝟎−𝟓

Verifying the data
Nutheoritical = 0.0243(Re)0.8(Pr)0.4
Nutheoretical =20.78
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.
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
Average 52.33 25 46.67 120.3
Twall =
𝐓𝟏+𝐓𝟐+𝐓𝟑+𝐓𝟒+𝐓𝟓+𝐓𝟔
𝟔
= 120.3 is the mean wall temperature.
Tbulk =
𝟐
Area of cross section Ac =
𝝅
𝟒
× 𝒅 𝟐
Ac =5.3066× 𝟏𝟎−𝟒
m2
Surface area As =2𝝅𝒓𝑳
As=0.0490 m2

As we know, air flow through orifice plate is calculated in terms of head of air,
∆Hair =
𝟏𝟎𝟎𝟎×𝟎.𝟎𝟏𝟒
𝟏.𝟏𝟓
= 12.174 m of air.
Q =
√𝐀𝐜𝟐−𝐀𝟐
√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 =
𝐐
𝐀𝐜
Vm =2.44 m/s
Reynolds Number
Re = Vm×d/β
Re =3966.625
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
qo =h As (Tw - Tb) ---------------------------------------------(A10)

q1 =m CP (Toutlet - Tinlet) ---------------------------------------(A11)
q2 =
𝐥𝐧(𝐫𝟐)
(𝒓)
𝟐𝝅𝒍𝑲
---------------------------------------------(A12)
h(0.0490) (120.3-35.83) = (1.491)(1.005)(46.67-25)+(7.47)
h =9.65
∆Hf =
𝐟𝐋
𝟐𝐝𝐠
𝐯 𝟐
∆Hf = 0.1 cm of water
f = 14.254× 𝟏𝟎−𝟓
Verifying the data
Nutheoretical = 16
Nu= hd/k
Nu= 14

APPENDIX-A2
CALCULATIONS FOR THREADED DUCT
A2.1 Full open valve
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
Average 40.3
Twall =
𝐓𝟏+𝐓𝟐+𝐓𝟑+𝐓𝟒+𝐓𝟓+𝐓𝟔
𝟔
= 64 is the mean wall temperature.
Tbulk =
𝟐
Ac =
𝜋
4
× 𝑑2
Ac =5.3066× 𝟏𝟎−𝟒
m2
Surface area
As =2𝜋𝑟𝐿
As=0.0490 m2
air,
𝜌water ∆Hwater = 𝜌air ∆Hair
∆Hair =
𝟏𝟎𝟎𝟎×𝟎.𝟎𝟏𝟒
𝟏.𝟏𝟓
= 40 m of air.

Q =
Cd Ac A
√Ac2−A2
√2 × 𝑔 ×∆Hair
Therefore,
𝐐 =
𝟎. 𝟔𝟏𝟑 × 𝟓. 𝟑𝟎𝟔 × 𝟏. 𝟑𝟐𝟔𝟔 × 𝟒𝟎
𝟓. 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,
Vm =
Q
Ac
V =5.09 m/s
Reynolds Number
Re = vm×d/β
Re =8271.25
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
qo =h As (Tw - Tb) ---------------------------------(A14)

q1 =m CP (Toutlet - Tinlet) -----------------------------(A15)
q2 =
Tw−Tsurface
ln(r2)
(𝑟)
2𝜋𝑙𝐾
----------------------------------(A16)
h(0.0490) (64-32.68) = (2.7025)(1.005)(40.36-25)+(3.33)
h =29
∆Hf =
𝑓𝐿
2𝑑𝑔
𝑣2
∆Hf = 0.55cm of water
f = 18.03× 𝟏𝟎−𝟓
Verifying the data
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 =
𝐓𝟏+𝐓𝟐+𝐓𝟑+𝐓𝟒+𝐓𝟓+𝐓𝟔
𝟔
=98 is the mean wall temperature
Toutput =49.06
Tbulk =
𝟐
=37.03 is the bulk mean temperature
Ac =
𝜋
4
× 𝑑2
Ac =5.3066× 𝟏𝟎−𝟒
m2
Surface area
As =2𝜋𝑟𝐿
As=0.0490 m2
Head measured at orifice plate, ∆H = 3.2 cm of water.
air,
∆Hair =
𝟏𝟎𝟎𝟎×𝟎.𝟎𝟑𝟐
𝟏.𝟏𝟓
= 28.346 m of air.
Q =
Cd Ac A
√Ac2−A2
√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
Ac
Vm =3.59 m/s
Reynolds Number
Re = vm×d/β
Re =5845
qo = q1 +q2 -------------------------------------------(A17)
where qo is the convective heat transfer from the walls of duct into the fluid, q1 is the heat
qo =h As (Tw - Tb) ----------------------------------------(A18)
q1 =m CP (Toutlet - Tinlet) -----------------------------------(A19)
q2 =
Tw−Tsurface
ln(r2)
(𝑟)
2𝜋𝑙𝐾
----------------------------------------(A20)
h(0.0490) (98.08-37.03) = (2.2)(1.005)(49.06-25)+(5.24)
h = 21

∆Hf =
𝑓𝐿
2𝑑𝑔
𝑣2
f = 25× 𝟏𝟎−𝟓
Verifying the data
Nutheoretical =21.7
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 =
T1+T2+T3+T4+T5+T6
6
=120.3 is the mean wall temperature
Toutput =64.3
Tbulk =
Tinput+Toutput
2
=44.64 is the bulk mean temperature
Ac =
𝜋
4
× 𝑑2

Ac =5.3066× 𝟏𝟎−𝟒
m2
Surface area
As =2𝜋𝑟𝐿
As=0.0490 m2
air,
∆Hair =
𝟏𝟎𝟎𝟎×𝟎.𝟎𝟏𝟓𝟑
𝟏.𝟏𝟓
= 13.29 m of air.
Q =
Cd Ac A
√Ac2−A2
√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
Ac
Vm =2.46 m/s

Reynolds Number
Re = Vm×d/β
Re =3998
qo = q1 +q2 ------------------------------------------(A21)
Where qo is the convective heat transfer from the walls of duct into the fluid, q1 is the heat
qo =h As (Tw - Tb) ---------------------------------------(A22)
q1 =m CP (Toutlet - Tinlet) ---------------------------------(A23)
q2 =
Tw−Tsurface
ln(r2)
(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
∆Hf =
𝑓𝐿
2𝑑𝑔
𝑣2
f = 35× 𝟏𝟎−𝟓
Verifying the data
Using empirical relation to find Nusselt number by Dittus-Boelter equation Theoretically,
Nusselt number is

Nu= hd/k
Nu= 15.83

APPENDIX-B
Cost Analysis
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.)
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

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