Tribology- Design and Development of Experimental Setup for Determination of Temperature and Pressure Profile of Hydrodynamic Journal Bearings

1
Major Project Report
on
Design and Development of Experimental Setup for Determination of
Temperature and Pressure Profile of Hydrodynamic Journal Bearing
Submitted by
Deshpande Aditya N. (S-42 Gr No.142130)
Hande Gaurav P. (S-44, Gr No.142101)
Patil Tanaji B. (T-41 Gr No. 142154)
Deshmukh Sourabh H. (S-48 Gr No. 142134)
(Final year B. Tech)
Under the Guidance of
Prof. DR. S. P. CHIPPA
Department of Mechanical Engineering
Vishwakarma Institute of Technology, Pune
Pune-411037
May 2017

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Bansilal Ramnath Agarwal Charitable Trust’s
VISHWAKARMA INSTITUTE OF TECHNOLOGY, PUNE. 37
(An Autonomous Institute Affiliated to Savitribai Phule Pune University)
Department of Mechanical Engineering
CERTIFICATE
This is to certify that the Major Stage-III entitled “Design and Development of
Experimental Setup for Determination of Temperature and Pressure Profile of
Hydrodynamic Journal Bearing”, has been satisfactorily completed in the academic year
2016-17, by following group of students in the partial fulfilment of Bachelor’s Degree in Mechanical
Engineering.
DR. S. P. CHIPPA
(Guide)
DR. S. R. BAHULIKAR
Professor and Head,
Mechanical Engineering Department
(External Examiner)

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ACKNOWLEDGEMENTS
We would like to gratefully acknowledge the enthusiastic supervision of our project
guide, Prof. (Dr.) S. P. Chippa for his continuous, valuable guidance; patience, constant care,
and kind encouragement throughout the project work that made us to present this project report
of our Major Project-Stage III in an efficient manner.
We wish to express our gratitude to Prof. R. M. Jalnekar Director, VIT, Pune for
providing the facilities of the Institute and for his encouragement during the course of this
work. We also express our deep gratitude to Prof. (Dr.) S. R. Bahulikar, the Head of the
Department of Mechanical Engineering, VIT, Pune for his guidance and support. We would
also like to express our hearty thanks to Prof. L. D. Mangate for his valuable suggestions and
advice. We would like to express our thanks to Mr. V. R. Alawane for his support.
We would like to thank to all our colleagues, friends, teaching and non-teaching staff
members for creating a pleasant atmosphere during our work at VIT, Pune.
Finally, we wish to thank our Family members & our friends who have always been
very supportive and encouraging.

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ABSTRACT
The hydrodynamic journal bearing is special type of bearing where the load is
completely taken by the oil film between the journal (shaft) and bearing (sleeve). Pressure
development and consequently temperature development across the circumference and axis is
major concern to load carrying capacity. The performance of bearing is significantly affected
by change in speed. The excessive rise of pressure and temperatures in the lubricant film of
hydrodynamic journal bearings occurred as they are used for supporting high speed rotating
machinery. Main objectives of the project are to design and develop model for journal bearing
and experimentation on setup to measure the temperature distribution along the circumference
of bearing. The setup consist of facility to vary the speed of journal by VFD, varying load by
loading arrangement and also varying the lubricating oil. Here, in this a circular journal bearing
with journal diameter=60 mm, L/D ratio=1.0, eccentricity ratio=1.0 and clearance=60
micrometre has been designed and tested to access the pressure and temperature rise of the oil
film at the central plane of the bearing. The pressure and temperature profiles have been taken
at varying load of maximum 500N and at variable journal speeds up to 2800 rpm. The setup is
devised with the pressure transmitter and LM 35 with DAQ and Arduino for real time data
management system. The various results are discussed and their justification is made in the
report.

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INDEX
CHAPTER
NO.
TITLE
PAGE
NO.
Acknowledgement i
Abstract ii
Index iii
List of figures ,drawings, tables and graphs v-vii
1 Introduction
1.1 Introduction to hydrodynamic bearing
1.2 Historical background
1.3 Concepts, working of hydrodynamic bearing, Mechanism of
pressure development
1.4 Type of hydrodynamic bearings, applications
1.5 Aim of project work
1-7
1
2
2
3
8
2 Literature Review
2.1 Theories of bearing analysis
2.2 work done by others in concerned area
2.4 Selection criteria for journal bearing
2.4.1 L/D ratio
2.4.2 material
2.4.3 oil
2.4.4 radial clearance
2.4.5 Practical considerations
9-22
9
18
21
3 Setup Design
3.1 Determination of pressure profile using MATLAB
3.2 Theoretical analysis of proposed bearing using infinitely long
bearing theory
3.3 Possible configuration of setup and deciding best suited one
for project
23-35
23
28
32
4 Construction details
4.1 Shaft
4.1.1 Material
36-47
37
38

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4.1.2 Drawing
4.2 Bearing
4.2.1 Material
4.2.2 Successive iterations in bearing design
4.3 Test bearing Support
4.3.1 Material
4.3.2 Successive iterations in support design
4.4 Ball bearing support
4.4.1 Material
4.4.2 Drawing
4.5 Motor support plate
4.5.1 Material
4.5.2 Drawing
4.6 Loading arrangement assembly
4.6.1 Horizontal bar
4.6.2 Vertical support
4.6.3 Loading shaft
4.6.4 Loading disc
4.7 Base plate
4.7.1 Material
4.7.2 Drawing
4.8 Ball bearing – Size and dimensions
4.9 Driving motor -Specifications
4.10 Variable frequency drive ( VFD)
4.11.1 Specifications
4.11.2 Connection diagram
4.12 Base support selection
4.12 Eye bolt for fitting
39
39
40
40
41
42
42
43
43
45
5 Measurements and data acquisition
5.1 Data acquisition system (DAQ)
5.2 Arduino
48-56
48
49
50
53

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5.2.2 Connection diagram
5.2.3 Arduino programme
5.2.4 Result viewing procedure
5.3 Temperature sensor
5.3.1 Working and specifications
5.3.2 Connections with Arduino
5.4 Pressure sensor
5.4.1 Working and specifications
53
54
55
55
56
6 Discussion and analysis of results
6.1 Temperature vs time curve
6.2 Temperature vs sensor position/angle at various loads
57-58
57
58
8 Total Cost 59-62
8 Conclusion 63
9 Future Work 64
References 65-66
Appendix A (Process Data Sheet)
A1 Shaft
A2 Bearing
A3 Support Bearing Block
A4 Base Plate
A5 Test Bearing Support
A6 Motor Base Plate
67-73
67
68
69
70
71
73
Appendix B ( Drawings of all Parts)
B1 Shaft
B2 Bearing
B3 Test Bearing support
B4 Support Bearing Block
B5 Motor base plate
B6 Base plate
B7 Loading arrangement
B8 Overall assembly
74

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LIST OF FIGURES:
Figure
No.
Figure name Page No.
1.1. Journal Bearing Principle 1
1.2. Working Of Journal Bearing 3
1.3. Hydrodynamic Journal Bearing 4
1.4. Hydrostatic Journal Bearing 5
1.5. Radial Bearing 6
1.6. Axial Bearing 6
1.7 Radial Thrust Bearing 7
1.8 Thick Film Lubrication 7
1.9 Thin Film Lubrication 7
1.10 Boundary Lubrication 8
2.1.
Circumferential Pressure Distribution For Full Sommerfeld
Boundary Conditions
11
2.2 Gumbel Boundary Conditions 12
2.3
Circumferential Pressure Distribution For Reynolds Boundary
Condition.
13
2.4
Circumferential Pressure Distribution For Reynolds Boundary
Condition
14
2.5 Swift-Stieber Boundary Conditions 15
2.6
Friction Variable With Sommerfeld Number For An Idealized
Full Journal Bearing
16
2.7
Load Carrying Capacity Along Radial And Axial Direction.
17
2.8 Temperature Distribution By Amit Singla,Et All 19
2.9 Temperature Distribution By Pirto,Et All 19
2.10 Variation Of Temperature With Time By Shinde ,Et All 20
3.1
RPM 1500, Mesh Plot Of Non Dimensional Pressure Vs Rolling
Direction
26
3.2
Rpm 1500, Mesh Plot Of Non Dimensional Pressure Vs Axial
Direction
26
3.3 Pressure Distribution Curve 1500 RPM 27
3.4 Pressure Distribution Curve 2500 RPM 27
3.5 Simply Supported Test Setup 33
3.6 Overhanged Supported Test Setup 34
3.7 Proposed Simply Supported Test Setup 35

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4.1 Final Setup 36
4.2 Shaft 38
4.3 Sleeve Configuration 38
4.4 Test Bearing Support 39
4.5 Support Bearing Support 40
4.6 Motor Support Plate 40
4.7 Drawing Of Lever Arm 41
4.8 Analysis Stress Results Of Lever Arm Of Loading Arrangement 41
4.9 Base Plate 42
4.10 Connection Diagram Of VFD 43
4.11 Channel Section Table Of Dimensions As Per IS 800-2007 44
4.12 Analysis Results Of Channel Of Support Structure 45
4.13 Table Of Eye Bolt Dimensions 46
4.14 Analysis Results Of Eye Bolt Of Loading Arrangement 47
4.15 Final Assembly 47
5.1 Data Acquisition System NI Made 48
5.2 Arduino Circuit Board Diagram 49
5.3 Lm 35 54
5.3 LM 35 Connection Diagram 55
5.4 Pressure Transmitter 55
5.5 Actual Project Demonstration 56

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LIST OF TABLES:
Table
No.
Table name Page No.
1 Journal Bearing Specification 39
2 Ball Bearing Specifications 42
3 Driving Motor Specifications 42
4 Variable frequency Drive Specifications 43
5 Data Acquisition System (DAQ) Specifications 48
6 Arduino Specifications 49
7 LM-35 Specifications 54
8 Pressure Transmitter Specifications 56
9 Cost of Project 59
LIST OF GRAPHS:
Graph
No.
Graph names Page No.
1 Temperature vs time (varying load at 1500 RPM) 57
2 Temperature vs sensor position/angle at various loads 58

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Chapter 1
INTRODUCTION
1.1 Introduction to hydrodynamic bearing:
Hydrodynamic journal bearings are typical critical power transmission components that
carry high loads in different machines. In machine design, therefore, it is essential to know the
true or expected operating conditions of the bearings. These operating conditions can be studied
both by experimental and mathematical means, for example in test rig experiments, in field or
laboratory tests with engines and by calculation or simulation.
Numerous studies of the operating conditions of hydrodynamic journal bearings have
been made during the last decades. Still, the case is far from closed. For example, there are a
limited number of studies that carry out an in-depth examination of the true operating
conditions of bearings in true-scale experiments. There is also a need for experimental studies
to verify the theoretical ones.
Fluid friction i.e. viscosity which exists in the lubricant being used is studied alongside
the pressure effect which is being generated in the bearing, thus the effect of lubricants with
different viscosities are considered.
Figure No 1.1: Journal Bearing

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A simple journal bearing consists of two rigid cylinders. The outer cylinder (bearing)
wraps the inner rotating journal (shaft). A lubricant fills the small annular gap or clearance
between the journal and the bearing. The amount of eccentricity of the journal is related to the
pressure that will be generated in the bearing to balance the radial load. The lubricant is
supplied through a hole or a groove and may or may not extend all around the journal. The
pressure around the journal is measured on various manometers by means of pressure
pipe/tubes. This is done at various speeds to get the relationship between speed and the
pressure. In this project stage II, aim is to design prototype of the setup. In this stage we need
to decide dimensions and size of parts to be manufactured, i.e. bearing, journal, shaft, loading
system, support structure etc.
1.2 Historical Background
In 1883 Beauchamp Tower carried an experimental investigation to determine suitable
methods of lubricating railway axle bearings. Tower performed tests on partial journal bearings
in which each bearing was lubricated from an oil bath. He made an unexpected discovery that
bearings under loaded condition developed peak pressure which was several times higher than
mean pressure calculated on the basis of projected area. Putting a number of pressure gauges
at the mid plane in the circumferential direction the pressure in the bearing clearance was
measured. The pressure distribution shows a peak value. This experiment probably has led to
the development of hydrodynamic theory of lubrication.
Shortly after the work of Tower, Reynolds in1886 could make a review of the detailed
experimental results of Tower and deduced that the lubrication of bearing was dependent on
the hydrodynamic action. Reynolds also suggested that oil was dragged into the clearance space
which converged in the direction of motion.
1.3 Working of hydrodynamic bearing
Here the general working of hydrodynamic bearing is given.

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(a) (b) (c)
Figure No: 1.2 working of journal bearing
A journal bearing designed to support a radial load, is the most familiar of all bearings. The
sleeve of the bearing system is wrapped partially or completely around a rotating shaft or
journal.
In Fig. 1.2 Journal is loaded and there is no rotation and the clearance space is filled with
lubricant. If the journal is given motion in the direction shown in fig.(b), the journal will climb
on bearing surface due to solid friction between journal and bearing surfaces. As the journal
will have some rotation and the clearance space is filled with lubricant, that film in the left half
will be convergent. We have seen that convergent film will developed a positive pressure
around the journal. The hydrodynamic pressure thus developed will shift the journal and
operate stably as given in fig.(c).the amount of rotation of the line of centres of journal and
bearing from the load line is dependent on the magnitude of applied radial load, the journal
speed and the viscosity of lubricant. The angle subtended by the line of centres between the
journal and the bearing with the load line is a measure of attitude angle.
1.4 Types of journal bearing and applications
Based on modes of lubrication sliding contact bearings are classified into two types
1. Hydrodynamic Bearing
2. Hydrostatic Bearing
1. Hydrodynamic Bearing
In hydrodynamic bearing, the load supporting high pressure fluid film is crested due to the
shape of zone between the two contacting surfaces and the relative motion between the two
surfaces.
The moving surface pulls the lubricant into a wedge shape zone, at a velocity sufficiently high
to create the high pressure film necessary to separate the two surfaces against the load.

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Figure No 1.3 Hydrodynamic journal bearing
The figure 1.3 demonstrates a hydrodynamic journal bearing and a journal rotating in a
clockwise direction. Journal rotation causes pumping of the lubricant (oil) flowing around the
bearing in the rotation direction. If there is no force applied to the journal its position will
remain concentric to the bearing position. However a loaded journal displaces from the
concentric position and forms a converging gap between the bearing and journal surfaces. The
pumping action of the journal forces the oil to squeeze through the wedge shaped gap
generating a pressure. The pressure falls to the cavitation pressure (close to the atmospheric
pressure) in the diverging gap zone where cavitation forms.
Thus in hydrodynamic bearings, it is not necessary to supply the lubricant under pressure. The
only requirement is to ensure sufficient and continues supply of the lubricant.
Applications of hydrodynamic bearings:
1. Engines
2. Large centrifugal pumps
3. Hydraulic turbine
4. Guide ways of machine tools etc.
2. Hydrostatic Bearing
In a hydrostatic bearing an external source of pressurized fluid forces lubricant between two
surfaces; thus enabling non-contacting operation and the ability to support a load. Hydrostatic

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bearings can support large loads without journal rotation and provide large (accurate and
controllable) direct stiffness as well as damping (energy dissipation) coefficients.
Figure No: 1.4 Hydrostatic journal bearing
In hydrostatic bearings as shown in figure 1.4, the fluid which is already present in a device
can act as a lubricant. For example, liquid oxygen in rocket engines, water in hydraulic
machinery, kerosene in aircraft engines.
Applications of hydrostatic bearing:
1. Vertical turbo generators
2. Rolling mills
3. Large telescopes
4. Gyroscope
5. Machine tools
6. Rocket engines
Bearing classification based on types of load carried
1. Radial bearing
2. Thrust bearing or axial bearing
3. Radial-thrust bearing
1. Radial bearing
These bearing can carry on radial loads. They are shown in figure 1.5.

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Figure No 1.5: Radial Bearing
2. Thrust bearing or axial bearing
These bearing can carry only axial loads as shown in figure 1.6.
Figure No 1.6: Axial bearing
3. Radial-thrust bearing
These bearing can carry both radial and thrust loads and shown in figure 1.7.

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Figure No 1.7 Radial-thrust bearing
Bearing classification based on type of lubrication
a) Thick film lubrication
Figure No 1.8: Thick film lubrication
b) Thin film lubrication
Figure No 1.9: Thin film lubrication

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c) Boundary lubrication
Figure No 1.10 Boundary lubrication
1.5 Aim of project work
The main objectives of project is to design and develop an experimental setup of circular
hydrodynamic journal bearing for study of following:
1. For studying temperature distribution across periphery
2. To study the effect of load, speed and viscosity on temperature and pressure profile.
It is important to know the exact temperature distribution along the circumference and how it
affects the load carrying capacity of bearing is also aim of project work.

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Chapter 2
LITERATURE REVIEW
2.1 Theories of bearing analysis
2.1.1 Assumptions in journal bearing design:
 Inertia & body terms are negligible in comparison of pressure and viscous terms.
 There is no variation of pressure across the film,
∂p
∂y
= 0
 There is no slip at fluid-solid boundaries.
 No external force acting on the film.
 The flow is viscous and laminar.
 Due to geometry of the fluid film,
∂u
∂y
&
∂w
∂y
are larger than other derivative of velocity
components.

h
l
≪ 13-3
2.1.1 Reynolds Equation
2.1.2 Infinitely Long Bearing Analysis
Assumption: L/D >3 or to 4

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Sommerfeld substitution:
2.1.2.1 Full Sommerfeld Boundary
Condition
i) P=0 at θ(0) =0
ii) P=0 at P=2π
The first boundary condition p= 0 at θ=0
yields the constant of integration A=0 in
equation
The second boundary condition
∂p
∂θ
=0 at
h=Ћ gives the value of thickness as
Ћ=
2C (1−ε2)
(1+ε2)

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Load Capacity:
Components of the load along the line of centre and its perpendicular to it, are given as:
Figure No 2.1 circumferential pressure
distribution for full Sommerfeld boundary
condition

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This can also be expressed in terms of dimensionless parameters known as Sommerfeld
number, S as given below,
2.1.2.2 Half Sommerfeld Boundary Condition
Alternatively, the bearing performance parameters were determined assuming that film
ruptures at θ=π where pressure becomes equal to zero or ambient. This boundary condition was
suggested by Gumbel and is more commonly known as half Sommerfeld boundary condition
or π- film boundary condition.
Boundary conditions:
i) P=0 at θ(0) =0
ii) P(0)=P(2π)
Capitation condition:
P=0 π<θ≤2π
Figure no 2.2: Gumbel boundary configuration

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Figure No 2.3: Mechanism of
pressure development
2.1.2.3 Reynold’s Boundary Condition
Half Sommerfeld boundary condition yields satisfactory and realistic results, it does not satisfy
the requirement of flow continuity in the cavitation boundary satisfying following boundary
conditions
Boundary conditions:
i) P=0 at θ(0) =0
ii) P=
∂u
∂y
=0 at θ= θc
Cavitation condition:
P=0 θ<θc≤2π

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Figure No 2.4 Circumferential pressure distribution for Reynolds boundary condition.
Using Full Sommerfeld BC:
Friction force

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And coefficient of friction on the journal surface is given as:
Figure No 2.5: Swift-stieber boundary conditions
2.1.3 Petroff Theory
Assumption: Bearing is lightly loaded
Radius of Journal = R

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Radial Clearance = c
Radius of Bearing =R+c
Figure No 2.6: Friction variable with Sommerfeld number for an idealized full journal
bearing
S≥0.15 Lightly loaded bearing condition
0.04 ≤ S ≤ 0.15 Moderately loaded bearing condition
S ≤ 0.04 heavily loaded bearing conditions
2.1.4 Infinitely Short Bearing Analysis
Dubois and Ocvirk (1953) Assumption: L/D < ¼
With above assumptions Reynolds equation reduces to,
Substituting x= Rθ above equation becomes
Integrating equation twice with respect to y pressure distribution in the lubricant film is
obtained as:

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P=0 at z ± L/2
Figure No 2.7 Load carrying capacity along
radial and axial direction.

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2.3 Work Done By Others In Concerned Area:
Here, we have considered what other people have done in the field of journal bearing
temperature measurement.
There are various paper published in various international journals many papers have
experimentally defined and analyse regimes and curve for temperature or pressure distribution.
A paper published by Amit singla, et all, [1] for 500 N load is concerned with elliptical journal
bearing with loading by pneumatic cylinder plots are shown below

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Figure 2.8 temperature distribution by Amit singla, et all
As per authors conclusion maximum temperature observed was 540
c at 1000 RPM .as peed
increase to 2000 RPM at 680
c
As per authors pirto,et all [2] maximum temperatures observed was 580
c at 2000 RPM for load
of 10 kN and about 520
c at same speed
Figure 2.9 temperature distribution by pirto, et all

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L/D ratio used 1 in both cases and shaft dia. 100 mm for both cases
In vertical downward direction Temperature observed on upper half of bearing is more than
lower half of bearing. In all cases temperature difference was observed to be 5 to 60
c in many
cases temperature was measured by K-type thermocouple PT-100 and RTD the resolutions of
search sensor used upto 0.10
c
Some authors like shinde, et all have evaluated variation of temperature in boundary lubrication
regime it was found that temperature goes on increasing with respect to time .Temperature
difference almost about 150 c in time period of 1.5 hrs.
Figure 2.10 variation of temperature with time by shinde, et all
As we have used journal bearing circular profile, the theoretical temperature difference
observed to be 2 to 30
c so there might be increase temperature difference due to slight change
in viscous forces and friction present in the system

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2.4 Selection criteria for journal bearing
Theoretical considerations
(i) load capacity,
(ii) Coefficient of friction and
(iii) Oil flow.
A bearing designed should be able to carry the applied thrust or radial load calculated from
other design considerations .The coefficient of fluid friction upon which depends the frictional
power loss should be as low as possible. A higher coefficient of friction means higher heat
generated in bearing .Finally, the flow rate of lubricant through the bearing is to be ascertained
correctly so that the bearing does not starve and consequently oil film which supports the load
does not break.
Theoretical load carrying capacity indicates that, it increases with increase in film thickness in
eccentricity ratio. The reduction in film thickness gives rise to enhanced load capacity.
Amongst various parameters,
L/D ratio, Radial clearance(C/R), and Unit load carrying capacity are important.
1. Length to Diameter Ratio (L/D):
Theoretical predication of good capacity shows that it can be increased with
increase in bearing length because of larger projected area. It has also found that rate of
flow of oil decreases with increase in bearing length. So larger bearing give higher load
and less flow rate whereas, reverse for short bearing.
Rigidity of shaft and bearing plays important role, to estimate correct L/D ratio,
if the rigidly supported and can be defected with shaft the journal may be damage at
ends due to metal to metal contact.
2. Radial clearance :
Under given operating conditions, the clearance in a journal bearing has significant
effect on load. Capacity and flow rate keeping constant, an increase in radial journal
clearance decreases load capacity and increases oil flow rate. As oil flow rate is directly
proportional to cube of radial clearance, a slight increase in clearance will increase oil
flow rate significantly, so reducing temp. of bearing.

32
It has been recommended in use a slightly higher radial clearance for small
bearing.
3. Bearing Unit Load (P=W/(2RL)):
Unit load is defined as load per unit projected area of bearing. Unit load is
function of Sommerfeld number. If the bearing C/R is constant and Sommerfeld no. is
inversely proportional to unit load, so for successful hydrodynamic lubrication, bearing
unit load should be such that Sommerfeld no. does not become critical.
i.e. S > 0,04.
Practical Considerations:
Also space and size requirement, alignment of all components and overall cost also plays
important role in practical considerations.

33
CHAPTER 3
SETUP DESIGN
Here, we have designed the parameters using MATLAB software and also by meaning of
theoretical calculations, and it is discussed below.
3.1 Determination of pressure profile using MATLAB
Here, we have used a MATLAB program for finding attitude angle and angle at which the
maximum pressure occurs. Also we can find the pressure distribution graph along length and
circumference.
Here, we have considered some values and by putting these values in program so that after
running the program we get the required values.
The program is as follows:
%%%%%%%%%%%%%%%%%This is a programme to Calculate Coefficients of Bearing
%Bearing Data %% All data in non-dimensional form
clear all
%load stability_1 freq_signal
format short
Load=500; % N
clrnc=60e-6; % m--------------;
eta=0.01; % pa s;
% freq_signal=2:0.5:100;
% freq_signal=1500/60;
% for fr=1:length(freq_signal)
speed=500; %rpm
% radius=0.0127; % mm;
radius=0.03; % mm;
omega1=(2*pi*speed*radius)/60; % Velocity (m/s)
% length1=0.0127*2; % m;
length1=radius*2; % m;
rh1=8e-2;
rh2=1;
WW=Load*clrnc^2/(eta*omega1*radius^2*length1); % Non-dimensional load
n_comp=0;
pre_W=WW;
L_D=1.0;
D_L=1/L_D;
e_c=0.2; % initially assumed
%%%Data for mesh generation
x_ele=100;
z_ele=50;
pi=22/7;
th=2*pi;
del_theta=th/x_ele;

34
theta_help=zeros(1,x_ele);
theta_help(1)=0;
for ii=2:x_ele
theta_help(ii)=theta_help(ii-1)+del_theta;
end
it = 0;
del_z=1/z_ele;
z_help=zeros(1,z_ele);
z_help(1)=1-del_z/2;
for jj=2:z_ele
z_help(jj)=z_help(jj-1)-del_z;
end
for jj=1:z_ele
theta(jj,:)=theta_help;
end
for ii=1:x_ele
Z(:,ii)=z_help';
end
[z_ele,x_ele]=size(theta);
P_0=zeros(z_ele,x_ele);
PP_0=P_0;
st=0;
while st~=1
h_0=1+e_c*cos(theta);
P_0(end,:)=P_0(end-1,:);
for i=2:z_ele-1
for j=1:x_ele
sa=1;
sb=1;
if j==1
sb=-(x_ele-1);
elseif j==x_ele
sa=-(x_ele-1);
end
A1=3*(h_0(i,j+sa*1)-h_0(i,j-sb*1))*(P_0(i,j+sa*1)-P_0(i,j-sb*1))/(4*h_0(i,j)*del_theta^2);
A2=-6*(h_0(i,j+sa*1)-h_0(i,j-sb*1))/(2*del_theta*(h_0(i,j))^3);
B=(P_0(i,j+sa*1)+P_0(i,j-sb*1))/(del_theta^2)+(D_L^2)*(P_0(i+1,j)+P_0(i-1,j))/(del_z^2);
C=2/del_theta^2+2*D_L^2/del_z^2;
P_0(i,j)=(A1+A2+B)/C;
end
end
for ii=1:z_ele
for jj=1:x_ele
if P_0(ii,jj)<0
P_0(ii,jj)=0;
end
end
end
Wx=-del_theta*del_z*sum(sum(P_0.*cos(theta)));
Wy=del_theta*del_z*sum(sum(P_0.*sin(theta)));
phi1=atan(Wy/Wx);
W1=sqrt(Wx^2+Wy^2);
W=W1;
chkk=abs(sum(sum(abs(P_0)-abs(PP_0)))/sum(sum(abs(P_0))));
if chkk<=1e-5
if abs((W1-WW)/WW)<=1e-3
st=1;
end

35
e_c=e_c*(1-rh2*rh1*((W1-WW)/WW))
end
if n_comp==0
if W1>WW
if ((pre_W-W1)/W1)>1e-4
rh2=0;
else
rh2=1;
end
end
end
if W1<WW
%%%%%%%%%%%%%%%%%%%
if 1==1
if ((W1-pre_W)/WW)<1e-3
rh2=1;
else
rh2=0;
end
end
%%%%%%%%%%%%%%%%%%%
if 1==0
if conp<1e-3
rh2=1;
else
rh2=1;
end
end
%%%%%%%%%%%%%%%%%%%%%%%
end
% if it>5
% e_c=e_c*(1-rh2*rh1*((W1-WW)/WW))
% end
pre_W=W1;
PP_0=P_0;
it = it+1;
end
% W=W1/6;
S=1/(pi*W1);
disp(' ')
disp('*****************************************************************************')
disp(['Eccentricity = ',num2str(e_c),' , attitude angle = ',num2str(round(phi1*180/pi)),' , Load (non-dims) =
',num2str(W)])
disp([' Sommerfeld number = ',num2str(S)])
disp('*****************************************************************************')
%mesh(theta,Z,P_0)
% perturb_pressure
% N=freq_signal(fr)*60;
% omega=2*pi*N/60;
% st_damp(fr).K_b=((eta*omega*radius^3*length1)/clrnc^3)* Stiffness;
% st_damp(fr).C_b=((eta*radius^3*length1)/clrnc^3)*Damping;
% end
% save stiff_damp_jrnbrg_btecproj st_damp freq_signal

36
For 1500 RPM, we obtained following results;
Figure No 3.1 Rpm 1500, mesh plot of non-dimensional pressure vs rolling direction
Figure No 3.2 Rpm 1500, mesh plot of non-dimensional pressure vs axial direction

37
Figure No 3.3 pressure distribution curve 1500 RPM
Attitude Angle = 36.37770
For RPM 2500, we obtained following results,
Load (non-dims) = 0.42482, Sommerfeld number = 0.74898
Eccentricity = 0.17578, Load (non-dims) = 0.42482
Sommerfeld number = 0.74898
Figure No 3.4 pressure distribution curve 2500 RPM
>> phi1*180/(2*pi)
ans = 39.5039

38
3.2 Theoretical analysis of proposed bearing using infinitely long bearing theory
Theoretical Considerations
The three important parameters concerning bearing design are
(i) load capacity, (ii) coefficient of friction and (iii) oil flow.
Practical Considerations
1. Length to Diameter Ratio (L/D): Rigidity, Influence on Load capacity & oil flow
rate, Space constrained.
2. Radial Clearance: Influence on Load capacity & oil flow rate
3. Bearing Unit Load (P=W/(2RL)) : Somerfield number S > 0,04, Fatigue life

40
Bearing design based on empirical relations to find Sommerfeld number
Here, radial load= 50 kg=500N
Speed= 1400 rpm, Diameter= 60 mm
L/D=1
So length is 60 mm
Now select material as Cu-Pb alloy so, C/R= 0.001
So C= 3 x 10^-5 mm but we consider C= 6x 10^-5 mm
Now
Minimum film thickness is half of C
So Hmin= 3x 10^-5mm
Now eccentricity ratio
E=(1-3x 10^-5 /6x 10^-5 )=0.5
So for L/D=1 and E=05
So mu x (R/C)= 4.31
So, mu= 8.62 x 10^-3
Frictional force
F= mu x W
F=4.31 N

41
Here,
Q= (U x L x C x Qc)/2
= 4.39 * 0.06* 6 x 10^-5 * 4.445/2
= 3.51x 10^-5 m^3/sec
Heat generated= heat dissipated
(F x U)/J= rho x S x Q x temp diff
4.31*4.39= 880*1760*3.51x 10^-5*temp diff
Temp diff=2.340
c

42
Initial temp=300
c
So eq temp=32.340
c
Viscosity=0.018 Pa.s
So, S=0.7
From table S=0.17
So design is safe.
3.3 Possible configuration of setup and deciding best suited one for project
Here, as per the mounting feasibility and size required, there has been two possible
setup configurations possible. The factors considered for setup selection are
1. Space requirement
2. Ease of mounting sensors
3. Proper alignment requirement
So, the setups are of two types:
1. Simply supported test bearing
2. Overhanging test bearing
The configurations are discussed in following section.

43
3.3.1 Simply supported test bearing
Figure No 3.5: Simply supported test setup
Figure 3.5shows simply supported test setup in which the test bearing is simply
supported over the two support ball bearing. Here, the advantage is that the mounting
is convenient and the load is equally distributed among the bearings. The setup consists
of two support bearing equally spaced from test bearing mounted on the same shaft.
The shaft is couple to motor via jaw coupling. The problem of positioning and alignment
is reduced here than other type of support.
3.6.2 Overhanging test bearing

44
Figure 3.6 shows simply supported test setup in which the test bearing is overhanged
supported over the two support ball bearing. The bearing under test is at the end of shaft
on one side. The setup shows ease of mounting of sensors and loading. The setup is
prone to the improper misalignment and mounting.
Figure No 3.6: Overhanged supported test setup

45
3.3.1 Proposed Setup:
Figure No 3.7: Proposed simply supported test setup
Here, after considering all parameters of the setup selection, the simply supported setup is
selected as final setup for the project. Figure 3.7 shows the proposed setup. Test bearing is
supported between two support bearings and is loaded with loading arrangement. The motor
and shaft are connected via coupling and motor is connected to variable frequency drive i.e.
VFD.

46
CHAPTER 4
CONSTRUCTIONAL PARTS
In this chapter, the various parts, their construction, design considerations are
discussed. Setup has main following parts:
1. Shaft
2. Sleeve
3. Test Bearing
4. Support Bearing
5. Loading Mechanism
6. Lubrication system
7. Pressure and Temperature measurement
8. Power Drive
9. Data Acquisition
The constructional details are given in the subsequent sections.
Figure No. 4.1 Final setup

47
4.1 Journal (Shaft):
Material selected: Mild steel(30C8)
Considering material with Syt= 200 Mpa
FOS=4
First find Torque on shaft
P= 2 pi x N xT/60
Here P= 1HP =746 watt N=1440RPM
So, T=5000 N.mm
T= pi /16 x Ss x d^3 ............ D=10mm
By considering Torsion stiffness
θ = (584 x T x l) / (G x d^4 )
So taking θ = 0.250
d = 21 mm
Taking higher values
1.d = 25 mm D= 35 mm Df= 45 mm
2. d= 30 mm D= 40 mm Df= 50 mm
This dia is calculated on basis of only torsion It will used as referenceFurther
dead weights will be addedSo it will be higher.
PROBLEM
In many literature for 500N bearing dia.is 65 mm

48
But, the main objective of project is to find temperature distribution along the
circumference so Diameter 25 is to small. So we modified the diameter to large value
i.e 60mm to get appropriate temperature distribution. The material used here is EN 8.
Figure No 4.2 shaft
4.2 Sleeve Configuration:
Without anchor with anchor
Fig No: 4.3 Sleeve Configuration
Here, the sleeve use to mount the sensors and to fit bearing, there are 2 possible
configurations available. First is use of anchored sleeve and other without anchor. Here,
due to cost considerations selection is made of sleeve without anchor.
Diameter of journal d 60.00 mm
Length of the bearing L 60.00 mm

49
Radial clearance c 60 micros
Aspect ratio L/D 1.00
Clearance ratio C/R 0.001
Table No:1 Journal Bearing Specification
4.3 Test bearing support:
Here, we have selected the profile of support to fix the bearing as well as to give loading
arrangement facility. It is made of MS and its drawing is as per shown in fig 4.4.
Fig No: 4.4 Test Bearing Support
4.4 Ball bearing support:
It is 2 in no and used to house ball bearings at both ends. It is made of MS and its drawing is
shown.

50
Fig No: 4.5 Support Bearing Support
4.5 Motor support plate:
It is used to increase the centre distance of shaft from base and is of 20 mm thick of MS. Its
drawing is shown.
Fig No: 4.6 Motor Support Plate
4.6 Loading arrangement assembly:
It is used to load the bearing and we can apply max 500 N force with leverage ratio of 5.

51
It has vertical bar, a horizontal lever and a hook to hold the eye bolt. One side connected to
loading discs with weights of known quantity. Its drawing is as shown.
Here, the maximum applied load is 500N and so there is need to use leverage principle. Here,
the constraint is of space required. Here, minimum distance is of radius of sleeve with
considering clearances.
Here, distance of pivot from bearing centre is taken as 85 mm.
Here, leverage ratio of 5 is used as it gives 10 kg at the lever end.
It means to have length of arm 5 times more than load arm.
Here, selection is made of aluminium HE 30 solid section of size as shown below
Figure No 4.7 drawing of lever arm
Here, structural analysis of the loading arrangement is performed and the results are given
below.
Figure No 4.8 analysis stress results of lever arm of loading arrangement
Here, the maximum stress coming as 103 MPa so it is having factor of safety 2.5. So
design is safe.
4.7 Base plate:
It houses all the parts and made of MS. Its drawing is shown.

52
Fig No: 4.9 Base Plate
4.8 Ball bearing – Size and dimensions:
Ball bearing used are total 2 and its specifications as follows;
Table 2 Ball Bearing Specifications
4.9 Driving motor:
Its specifications are as follows
Table 3 Driving Motor Specifications
SR
NO
PARAMETER VALUE
1 phase 3 (Crompton made)
2 Input voltage 220V/ 415V
3 RPM 2800 RPM
4 Current 3.84 A
5 Connections Star/ Delta
6 Frequency 50 Hz
7 Power 1 HP/ 746 W
8 Efficiency 77%
9 Frame type 80 B 3
SR
NO
PARAMETER VALUE
1 Bearing number 6008-2Z( SKF make)
2 Internal diameter 40 mm
3 Outer diameter 68 mm
4 Width 15 mm
5
Load Rating capacity
Dynamic
Static
12.9 kN
9.9 kN
6 Speed limit with grease 10000 RPM

53
4.10 Variable frequency drive (VFD):
Its specifications are as follows. It is used to change the speed of shaft as per desired with
resolution of 1 RPM.
Table 4 Variable frequency Drive Specifications
SR
NO
PARAMETER VALUE
1 Input 1 PH/ 3 PH
2 Input current 5.1 A
3 Input voltage 200-220 V
4 Frequency input 50 Hz
5 Output 3 PH
6 Output current 4.2 A
7 Output voltage 0-240 V
8 Power 1 HP
9 Frequency range 0-400 Hz
:
Figure No 4.10 Connection diagram of VFD
4.12 Support Selection:
Here, selection is made of channel section amongst other choices
Possible following choices:
1. I section
2. C section

54
3. L section
But considering the weight per unit length and mounting ease, we have selection is
made of following C section.
Figure No 4.11 Channel Section Table of Dimensions as per IS 800-2007
This is as per Indian Standards of IS 800-2007 Steel table.

55
Figure No 4.12 analysis results of channel of support structure
Here, FEM analysis has been performed considering the unit loads and locations of
fixed support. Width taken is 350 mm between two channels. This is as per motor size.
As shown, maximum stress comes to around 60MPa which is far less than yield strength
of material used i.e. Mild steel. So we get factor of safety of 4.
4.13 Eye Bolt for Lifting
Here, for loading the bearing, options available are:
1. Providing the anchor in the sleeve itself,
2. Use of eye bolt
So selection is made of anchor arrangement firstly but due to fact that the manufacturing
cost went too high, the other option is selected. Here, eye bolt is threaded to the sleeve
with upper hole as dedicated to eye bolt.

56
Lifting Eye Bolt Design
Max. Load-500N
Material- 30C8 Syt=400N/mm2
Factor of Safety=6
σt=400/6=66.67N/mm2
σt=
𝑃
𝜋
4
𝑑𝑐2
dc=3.09mm
d=dc/0.8=3.87mm
For safety and load fluctuations, selection is made of M6 bolt for lifting.
Figure No4.13 table of eye bolt dimensions

57
Figure No 4.14 analysis results of eye bolt of loading arrangement
4.14 Final Assembly Drawing:
It is as shown in figure.
Figure No 4.15 Final Assembly

58
CHAPTER 5
MEASUREMENTS AND DATA ACQUISITION
In this chapter, the measuring instruments and their specifications are given. Also the data
acquisition and their specifications are given.
5.1 Data acquisition system (DAQ)
It is used to generate and store the outputs in numerical form in real time scenario. Here, to
generate lifetime data, the DAQ is need to used. Selection is made of National Instruments
DAQ USB 6001.
Figure No 5.1: Data Acquisition System NI made
Table 5 Data Acquisition System (DAQ) Specifications
SR
NO
PARAMETER VALUE
1
Inputs
analogue inputs
digital inputs
8
13
2
Outputs
Analogue
Digital
2
13
3
Number of channels
Single
differential
4
8
4 Resolution 14 bit
5 Maximum sample rate 20Ks/s
6 Input voltage 10-12 V
7 Software used NI Max, Lab-view
8 Analogue input CMRR 56 dB
9 Analogue output slew rate 3 V/μs

59
Here, the results are obtained through the NI MAX software provided by the NI Company
itself. We need to select the task to be performed and by setting all parameters and clicking on
RUN we can get real-time data. It is stored and can be used in MS EXCEL for plots.
5.2 Arduino:
Figure No 5.2: Arduino Circuit Board Diagram
Here, use is made of the Arduino for the purpose of collecting information of pressures and
temperatures at various points located in bearing. They are given to programmed Arduino.
Table 6 Arduino Specifications
SR
NO
PARAMETER VALUE
1 Type UNO
2 Analogue inputs 6
3 Input power 12 V
4 Connections Through computer
5 Program used Arduino
6 Output Through computer

60
5.2.3 Arduino Program
Its program is given below. Here, tempC is variable and assigned to each input AO, A1,
A2,…A4. We have used 2 Arduino as 5 sensors each and both can work simultaneously.
Delay is in microseconds and can be adjusted.
TempC is variable, and Pin are inputs 0,1,2,3,4 respectively. Here, TempC is floating that is
variable. By default values are in o
F and by using factor we have converted in o
C. We can
adjust the delay as 1000Millisec or any value as per wish. It is nothing but delay or time
between two readings.
Program:
float tempC;
int reading;
int tempPin = 0;
float tempC1;
int reading1;
int tempPin1 = 1;
float tempC2;
int reading2;
int tempPin2 = 2;
float tempC3;
int reading3;
int tempPin3 = 3;
float tempC4;
int reading4;
int tempPin4 = 4;
void setup()
{
analogReference(INTERNAL);

61
Serial.begin(9600);
}
void loop()
{
{
reading = analogRead(tempPin);
tempC = reading / 9.31;
//Serial.println("t");
//Serial.println(tempC,DEC);
////delay(1000);
}
{
reading1 = analogRead(tempPin1);
tempC1 = reading1 / 9.31;
//Serial.println(tempC1,DEC);
//delay(1000);
}
{
//delay(1000);

62
//delay(1000);
}
{
Serial.println(tempC);
Serial.println(tempC1);
Serial.println("t");
//Serial.print("%f %f %f %f %f", tempC,tempC1,tempC2,tempC3,tempC4);
delay(1000);
}

63
//Serial.println(tempC,DEC);
//
//
//
//
//
}
Here we have used directly readings which are saved as text file and can be imported in MS
EXCEL for plotting the results.
5.3 Temperature sensor
LM35 Thermocouple
The LM35 series are precision integrated-circuit temperature devices with an output
voltage linearly-proportional to the Centigrade temperature. The LM35 device has an
advantage over linear temperature sensors calibrated in Kelvin, as the user is not required to
subtract a large constant voltage from the output to obtain convenient Centigrade scaling. The

64
LM35 device does not require any external calibration or trimming to provide typical
accuracies of ±¼°C at room temperature and ±¾°C over a full −55°C to 150°C temperature
range. Lower cost is assured by trimming and calibration at the water level. The low-output
impedance, linear output, and precise inherent calibration of the LM35 device makes
interfacing to readout or control circuitry especially easy.
Table 7 LM-35 Specifications
SR
NO
PARAMETER VALUE
1 scale factor 10.0 mV/°C
2 Temperature range −55° to +150°C
3 Operating voltage 5 V
4 calibration Directly in Celsius
5 Type IC sensor
Figure No 5.3 LM 35
The 3 ported sensor has one pin for power, middle for sensing element, third for ground.

65
Figure No 5.4 LM 35 connection diagram
5.4 Pressure sensor
Baumer pressure transmitter
Figure No 5.5 Pressure transmitter
Series CT transmitters are designed for extremely diverse industrial applications:
ventilation, heating, pumps, pressure measurement on water systems (building complexes,
etc.), maintaining pressure set points on oil compressors, integration in production
machinery, agriculture machines, etc.

66
Table 8 Pressure Transmitter Specifications
SR
NO
PARAMETER VALUE
1 Pressure range 0-200 bar
2 Output signal 4-20 mA
3 Supply voltage 15 V DC
4 Operating temperature -25+85°C
5 Construction Stainless steel
Figure No 5.6 Actual Project Demonstration

67
CHAPTER 6
DISCUSSION AND ANALYSIS OF RESULTS
The graph 1 shows the variation of temperature of all sensors with time in degrees Celsius.
T1 is 10 min, T2, is in 20 min, T3 in 30 min. It is as per literature review paper by Shinde et all.
Here, temperature for T1 is less than T2 which is less than T3. This is because of lower load at
T1 than T2 and T3. Blue curve is for 75N, green for 100 N and red for 150N all at same speed
of 1500 RPM and of same viscous oil.
Graph 1 Temperature vs time (varying load at 1500 RPM)
Following graph 2 shows temp variation along circumference at the various loads shown below.
It shows that as load increases max temperature increases.
Temperature is maximum at sensor numbers 4 to 6 which are located at bottom of bearing. It
follows same profile for all loads and it is having shifting to upward for higher loads.
30
30.5
31
31.5
32
32.5
33
33.5
34
34.5
35
35.5
0 2 4 6 8 10 12 14
TemperatureinoC
LM 35 number
T1
T2
T3

68
LM 35 numbers are as follows:
1 to 3 at top right of bearing while 4 to 6 at bottom and 7 to 10 at left of bearing at top.
Graph 2 Temperature vs sensor position/angle at various loads
40
45
50
55
60
65
70
75
80
85
0 2 4 6 8 10 12
temperatureoC
LM35 number
75 N 1500 RPM
100N 1500RPM
150N 1500RPM

69
CHAPTER 7
COST OF PROJECT
Below table shows the total cost of project with supplier name, address and various items
purchased. Table 9 Cost of Project
Sr.
No
Item
Name
Description Supplier Qty
Price Per
Quantity
Total
Price
Date
1
Driving
Motor
3 Phase 1425 Rpm,
Crompton
NK Mehta
Company,
Shukrawar Peth,
Pune
1 6290 6290
02-
03-
17
2 NI DAQ
Data Acquisition
For Pressure
Sensors NI USB
6001
NI India, Starcom
Information
Technology,
Banglore
1 18350 18350
31-
12-
16
3
Pressure
Sensors
Baumer CTX Type
Shree Instruments,
Bhosari, Pune
2 4000 8000
23-
12-
16
4
Arudino
Board
Arduino Uno Type
Akronsys Systems,
Dhankawadi, Pune
2 375 750
01-
03-
17
5 LM 35 Temperature Sensor
Akronsys Systems,
Dhankawadi, Pune
1 70 70
01-
03-
17
6
Cables
For LM
35
Connection With
Arduino
Akronsys Systems,
Dhankawadi, Pune
5 3 15
01-
03-
17
7
Ball
Bearing
Support Bearing ID
40 Mm OD 68 Mm,
SKF Italy
Pratimma
Enterprises,
Shukrawar Peth,
Pune
2 475 950
01-
03-
17
8 Eye Bolt Loading
Arrangement
National Trades,
Parvati, Pune
1 20 20
03-
04-
17

70
Connection To
Bearing, Dia 8mm
9 Coupling
Connect Shaft With
Motor
Pratimma
Enterprises,
Shukrawar
Peth,Pune
1 530 530
03-
04-
17
10
Laoding
Horizontal
Bar
Al HE 30
20x30x500
Nakoda Metals,
Narhegaon, Pune
1 317 317
11-
04-
17
11 Plug
Connecting Bearing
With Sensors, 1/8”
BSP
Technomet
Enterprises, Narhe,
Pune
1 35 35
17-
03-
17
12 Adopter
Reduction Of Size
Of Connection Port
Of Pressure Sensor
Technomet
Enterprises, Narhe,
Pune
2 80 160
17-
03-
17
13 LM 35 Temperature Sensor
Akronsys Systems,
Dhankawadi, Pune
10 52 520
11-
04-
17
14 Plug
Connecting Bearing
With Sensors
Technomet
Enterprises, Narhe,
Pune
12 30 350
04-
04-
17
15
Base
Channel
Section
Structure Support
Nakoda Metals,
Narhegaon, Pune
4 230 905
07-
04-
17
16
Base
Vertical
Legs
Structure Support
Nakoda Metals,
Narhegaon, Pune
4 100 400
07-
04-
17
17
Loading
Vertical
Plate
Loading
Nakoda Metals,
Narhegaon, Pune
1 293 293
07-
04-
17
18
Cables
For Lm 35
Connection With
Arduino
Akronsys Systems,
Dhankawadi, Pune
11 8 88
11-
04-
17

71
19
Cables
For Lm 35
Connection With
Arduino
Akronsys Systems,
Dhankawadi, Pune
1 10 10
11-
04-
17
20
Report
Printing
Forbs Marshall
Competition
Na 1 100 100
13-
04-
17
21
Test
Bearing
Support
Blank Material Ms
75x75x260 Mm
Nakoda Metals,
Narhegaon, Pune
1 782 782
15-
04-
17
22
Loading
Disc And
Bar
Ms Dia 5 X 320
Mm, Disc Dia 110x
10 Mm
Nakoda Metals,
Narhegaon, Pune
1 52 52
15-
04-
17
23
Aluminiu
m
Loading
Bar
550 Mm Long,
30x30 Solid
Nakoda Metals,
Narhegaon, Pune
1 350 350
17-
04-
17
24 Plug
Bared 1/8” BSP To
Close The Holes
Technomet
Enterprises, Narhe,
Pune
5 31 155
24-
04-
17
25
Oil Inlet
Plug
1/8” BSP Plug
Technomet
Enterprises, Narhe,
Pune
1 40 40
27-
04-
17
26 Oil Hose
2 M Hose Length
Blue Colour
Technomet
Enterprises, Narhe,
Pune
2 55 110
27-
04-
17
27
Travelling
Expenses
From Alawane To
College
NA - 200 200
27-
04-
17
28 Plug
Connecting Bearing
With Sensors
Technomet
Enterprises, Narhe,
Pune
2 35 70
28-
04-
17
29 Allen Bolt M10×40 Mm
Ganesh Hardware
And
Electricals,Narhe,
Pune
4 11 44
28-
04-
17

72
Ganesh Hardware
And
Electricals,Narhe,
Pune
2 15 30
28-
04-
17
Ganesh Hardware
And
Electricals,Narhe,
Pune
6 21 126
28-
04-
17
Ganesh Hardware
And
Electricals,Narhe,
Pune
2 22 44
28-
04-
17
33 Nut
M6 Loading And
Hook Of Eye Bolt
Ganesh Hardware
And
Electricals,Narhe,
Pune
6 - 10
28-
04-
17
34
Earthing
Wire
Motor And Vfd
1mm^2 Green
Colour 5 M Length
Pradeep Hardware
Near Vit College,
Pune
5 8 40
28-
04-
17
35
Lubricatin
g Oil
Castrol 20w-40-2t,
1 Litre
Narhe 1 335 335
21-
05-
17
37
Machinin
g And
Material
Cost
Alawane Engg
Works, Narhe
Narhe NA 19000 19000
29-
04-
17
Total Cost Rs 59611

73
CHAPTER 8
CONCLUSION
After completing the objectives of project, various conclusions can be made. The conclusions
are nothing but output of efforts or learning from the project. They are listed as follows:
1. The design of setup has been finalised as per all considerations with bearing.
2. Also the design is checked with Sommerfeld Number calculations based on analytical
solution, which found to be correct and safe.
3. The design is also checked for MATLAB program so that the location of pressure range
is found.
4. Loading arrangement and appropriate positioning of test rig sensors and various parts
has been done.
5. As per results, it is observed that temperature is about 60-800
c. This is because of metal
to metal contact but this also having the temperature difference of 2-50
c which is as per
the theoretical considerations.
6. The temperature is increased up to some time and then reaches to optimum level and
fixes there. This shows the steady state condition. Also this remains steady at given
load, speed, type of oil used.
7. The graph of temperature vs angle shows that max temperature occurs at bottom side
and minimum at upper side of bearing due to vertically upwards loading.

74
CHAPTER 9
FUTURE WORK
After completion of project objectives there can be some area in which further work can
be possible:
1. Use of different textured bearings to study their behaviour
Here, we can use the textured bearings in internal or external texture instead of
plane journal bearing. So that temperature distribution can be found out as
research field operation
2. To implement this set up for axial thrust bearing using an modified loading
arrangement
Here we have applied load radially, we can also apply the axial load by
modifying loading arrangement. This would be useful for testing hydrodynamic
thrust bearing.
3. Use of external pressurized oil as a fluid
Here we used oil at atmospheric pressure. Here occurs metal to metal contact as
there is no development of oil film at low speed. By using oil at pressure more
than atmospheric we can avoid metal to metal contact and can get better result at
low speed also.
4. Use of oil collecting tank for collecting oil
This can be done by providing a slot on base plate on both sides of test bearing
support.

75
REFERENCES
1. Introduction to tribology of bearings, text book by B.C. Mujumdar, S-Chand
Publications, 2008.
2. Design and Construction of Journal Bearing Demonstration Rig, Inyama Godwin
Okechukwu, Caritas University, Nigeria, 2008.
3. Design of machine elements third edition ,text book by V.B Bhandari
4. Experimental Evaluation Of Temperature Distribution In Journal Bearing
Operating At Boundary/Mixed Lubrication Regimes Pankaj Shinde Prof. P. N.
NagareIjiert
5. “Experimental Determination of Temperature and Pressure Profile of Oil Film
of Elliptical Journal Bearing”. In ISSN 2250-3234 volume 4 (pp. 469-474). Amit
Singla, Paramjit Singh and Amit Chauhan (2014)
6. Chaitanya Desai and Dilip Patel (2005) “Experimental Analysis of Pressure
Distribution of Hydrodynamic Journal Bearing: A Parametric Study”. in
ICME05-AM-30.
7. S. Kasolang, M. Ali Ahmad (2012) “Experimental Study of Temperature Profile in
a Journal Bearing”.
8. Performance Characteristics Of Twin Holes Hydrodynamic Journal Bearing,
Roopak Kumar,Krishan Kumar Gupta
9. Journal bearing design as related to maximum load speeds,and operating
temperatures, Mckee,NBS,1937.
10. Manual for journal bearing test rig DUCOM instruments ,VIT pune
11. Thermal behaviour and performamce characteristics of a twin axial groove
journal bearing as a function of applied and rotational speed ,Brito et all
,ICMMD, 2008.
12. Oil film pressure in hydrodynamic journal bearing,Doctoral Dissertation ,Antti
Valkonen,2009
13. EPO-PHEN data spreadsheet ,Sherwin Williams corporation, 2016
14. LM-35 data spreadsheet ,National semiconductor corporation ,2000
15. Operational manual NI USB-6001,National instruments,2016

76
16.CTX-CTL Data spreadsheet ,Baumer corporation,2015
17.Experimental study of hydrostatic thrust pad bearing with positive micro-textures
,Veerraju,2011
18.Arduino programming notebook , Brian W.Evans ,2013
19.Realisation and evaluation of start stop journal bearing test rig,Gradle,et all,2014
20.Operational manual VFD-L ,Delta,2010

77
APPENDIX A
PROCESS DATA SHEET
Part No 1 shaft
Date 13-03-2017 Part Name Journal ( Shaft)
Machine Centre LATHE Part No. MP-001
Machine No. P1L1S-2716 7 Cycle Time
Raw Material EN 8 Colour
Blank Weight 8 kg Blank Size 70 x 400 mm
Finished job weight 5 kg Heat Treatment No
Rough Turning Ø60.5 x 380.5
Spindle Speed 980 rpm Coolant On
Feed rate 0.15mm/rev Tool Single point cutting tool
Finish turning Ø60 x 380 mm
Grinding Ø60 -0.060
Machine Surface grinding machine Part No. MP-001
Machine No. S-2516_7 Cycle Time 0.5 hrs

78
Part No 2 bearing(sleeve)
Date 23-03-2017 Part Name Bearing ( Sleeve)
Machine HMC Part No. MP-002
Machine No. HMC-400XL Setting Time 2 Hrs.
Model No. M21- 324 Cycle Time 1.5 Hrs.
Raw Material Phosphorous bronze Colour
Blank Weight 4 kg Blank Size Ø110 x 65 mm
Finished job weight 2.475 kg Heat Treatment No
Turning Ø100 X 60mm
Drilling I/D Ø60 mm
Feed rate 650mm/min Tool Boring Tool
Drilling 18 holes Ø8.5 mm
Feed rate 650mm/min Tool Drilling Tool
Tapping M 1/8th
BSP
Feed rate 100mm/min Tool 1/8th
BSP Tapping Tool
Grinding Ø60 +0.060
Machine Centrifugal grinding
machine
Part No. MP-002
Machine No. CSG-2515_G5 Cycle Time 45 min.

79
Part No 3 support bearing block
Date 05-03-2017 Part Name Support bearing block
Machine Power Saw Part No. MP-003
Raw Material M. S Colour Grey
Blank Weight 7.5kg Blank Size 210 x 165 x 35 mm
Cutting
Feed rate 0.15mm/rev Tool Multi point saw blade
Drilling I/D Ø68 mm
Feed rate 650mm/min Tool Boring Tool
Tapping M12x1.5 mm
Feed rate 100mm/min Tool 1/8th
BSP Tapping Tool
Grinding Ø68 +0.060
Machine Centrifugal grinding
machine
Part No. MP-003
Machine No. CSG-2515_G5 Cycle Time 45 min.

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Part No 4 base plate
Date 08-03-2017 Part Name Base plate
Machine Gas Cutting Machine Part No. MP-002
Raw Material MS Colour Grey
Blank Weight 7.2 kg Blank Size 715 x 260 x 12 mm
Milling 701x251x11mm
Feed rate 0.15mm/rev Tool Multi point cutting tool
Surface grinding 750x250x10 mm
Feed rate 50mm/min Tool Grinding tool
Tapping M 10 x 1.5 mm
Feed rate 100mm/min Tool M10 Tapping Tool
Feed rate 100mm/min Tool M12Tapping Tool

81
Part No 5 test bearing support
Date 27-03-2017 Part Name Bearing support
Machine Centre Lathe,
horizontal milling
m/c
Part No. MP-005
Milling 701x251x11mm
Center Arc boring Ø 100 mm
Feed rate 650mm/min Tool Boring tool
Groove Milling
Feed rate 100mm/min Tool End mill Ø 16 mm
Mill Finishing
Feed rate 10mm/min Tool End mill Ø 16 mm

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Part No 6 motor base plate
Date 09-03-2017 Part Name motor base plate
Machine Gas Cutting Machine Part No. MP-006
Machine No. GCM-1L Setting Time 1 Hrs.
Model No. - Cycle Time 0.75 Hrs.
Cutting 201x201 x 21mm
Feed rate 100mm/min Tool M10 Tapping Tool

84
APPENDIX B
DRAWINGS OF ALL PARTS

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Tribology- Design and Development of Experimental Setup for Determination of Temperature and Pressure Profile of Hydrodynamic Journal Bearings