Study of forced convection in smooth and ribbed ducts

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.

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

Mechanical Engineering Department, NIT, Srinagar vi

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

Mechanical Engineering Department, NIT, Srinagar viii

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 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

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

Mechanical Engineering Department, NIT, Srinagar xi

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

Mechanical Engineering Department, NIT, Srinagar 1

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 definitionsIn this section a few important terms commonly used in heat transfer augmentation

work are defined.


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


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.


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 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


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


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 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


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 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


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.


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


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 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,


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


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: 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.


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


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 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}


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 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


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


6 Nusselt number Nu= hd/k Nu= 19.85

Table 4.2: Performance parameters (half open)S.No Parameters studied Experimental values


2 Mean velocity, vm vm =3.576 m/s3 Reynolds Number Re Re =58114 Heat transfer coefficient h h =14.3



Table 4.3: Performance parameters (¼ open valve)



2 Mean velocity, vm vm= 2.44 m/s3 Reynolds Number Re Re =3966.6254 Heat transfer coefficient h h =9.65


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/s3 Reynolds Number Re Re =8271.254 Heat transfer coefficient h h =29

5 Friction factor f 18



Table 4.5: Performance parameters (half open valve)



2 Mean velocity, vm vm =3.59 m/s3 Reynolds Number Re Re =58454 Heat transfer coefficient h h =21



Table 4.6: Performance parameters (¼ open valve)



2 Mean velocity, vm vm =2.46 m/s3 Reynolds Number Re Re =39984 Heat transfer coefficient h h =16.08



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


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.

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


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


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.


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


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.


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


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.


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


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


∆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


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



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,


∆ H air=1000× 0.031.15

=26.0869 m of air .


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


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=


(r )2πlK

---------------------------------------------(A8)


h (.049) (98.08-33.42) = (2.1826) (1.005) (41.7-25) + (5.24)

h =14.3





∆Hf = fL

2dgv2

∆ H f =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

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.




Area of cross section Ac = π4

×d2

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 = 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


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 =


(r )2πlK

---------------------------------------------

(A12)


h(0.0490) (120.3-35.83) = (1.491)(1.005)(46.67-25)+(7.47)

h =9.65




∆Hf = fL

2 dgv2

∆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


obtained


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.





Ac = π4

×d2

Ac =5.3066×10−4 m2

Surface area

As =2πrL

As=0.0490 m2



air,


∆Hair = 1000× 0.014

1.15 = 40 m of air.


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,


Vm = Q

A c

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 =


(r )2 πlK

----------------------------------(A16)


h(0.0490) (64-32.68) = (2.7025)(1.005)(40.36-25)+(3.33)

h =29




∆Hf = fL

2dgv2

∆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 =T 1+T 2+T 3+T 4+T 5+T 6

6 =98 is the mean wall temperature

Toutput =49.06


2 =37.03 is the bulk mean temperature


Ac = π4

×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 = 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


Re =5845


qo = q1 +q2 -------------------------------------------(A17)

where qo is the convective heat transfer from the walls of duct into the fluid, q1 is the



qo =h As (Tw - Tb) ----------------------------------------(A18)

q1 =m CP (Toutlet - Tinlet) -----------------------------------(A19)

q2 =


(r )2 πlK

----------------------------------------

(A20)


h(0.0490) (98.08-37.03) = (2.2)(1.005)(49.06-25)+(5.24)

h = 21




∆Hf = fL

2dgv2

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 =T 1+T 2+T 3+T 4+T 5+T 6

6 =120.3 is the mean wall temperature

Toutput =64.3


2 =44.64 is the bulk mean temperature


Ac = π4

×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 = 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


qo = q1 +q2 ------------------------------------------(A21)


Where qo is the convective heat transfer from the walls of duct into the fluid, q1 is the



qo =h As (Tw - Tb) ---------------------------------------(A22)

q1 =m CP (Toutlet - Tinlet) ---------------------------------(A23)

q2 =


(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

2dgv2

f = 35×10−5

Verifying the data

Using empirical relation to find Nusselt number by Dittus-Boelter equation





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


Study of forced convection in smooth and ribbed ducts

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