INTRODUCTION
In today’s world we are more concerned with reducing the cost of failure
and maintenance in any industry. For this reason maintenance schedules of each
machine are more and more determined by the exact running condition of each
machine. To understand the running condition of the machine detailed condition
monitoring activities are to be undertaken. If done properly, condition
monitoring also helps in predicting the residual time before a particular machine
needs overhaul. Vibration monitoring and analysis is one such tool that can be
used for determining the condition of a rotating machine and its analysis gives a
clear picture of the any fault that there may be. The main objective of the
project is to utilize the vibration analysis utilize the vibration analysis
techniques to identify the faults of certain rotating machines of CESC Ltd. with
a special focus on manifestation of vibration due to improper oil clearance in
case of a journal bearing of turbine.
OBJECTIVE OF THE PROJECT
Condition monitoring is a continuous process which has enabled has to
monitor real-time condition of the machine. The aim of vibration monitoring is
to detect any change in the vibration as well as the technical condition of the
machine under investigation during operation. A noticeable change in vibration
may indicate fault.
In this context my project has the objective of learning the process of
vibration analysis and its use in fault detection extensively and then using this
process to identify the fault of three critical rotating equipments of the power
plant namely fans, pumps and turbo-generator sets.
Project work also focuses on identifying the high clearance problem of a
journal bearing of a turbine. The effect of pressure and temperature of the
lubricating oil on a high clearance journal bearing has also been critically
analysed. This will help to increase the damping of the shaft vibration by
influencing the oil wedge property.
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ORGANISATION OF THE PROJECT
The project is cascaded into parts
Part-1: Explanation of the basics of vibration technology
Part-2: Description of three types of critical rotating machines of the power
plant i.e. pump, turbine, compressor, fan etc
Part-3: Description of various faults of rotating machinery that causes high
vibration
Part-4: Detailed description of vibration analysis techniques and their
application in the fault detection.
Part-5:Vibration measurement of 14 different machines at Southern
Generating Station and Budge Budge Generating Station of CESC
Ltd. and analysis of the same for fault detection
Part-6: Reduction of vibration by balancing Seal Air Fan 2A at Southern
Generating Station
Part-7: Identification of high clearance problem of bearing 2 of turbine unit
3 at Budge Budge Generating Station.
Part-8: Effect of pressure and temperature of the lubrication oil on the
vibration originating due to high clearance
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PART-1
THE BASICS OF VIBRATION
Page 3
1.1 BASICS OF VIBRATION
The physical repeatable oscillatory motion of a rotating machine about a
neutral position is normally referred to as vibration. Although in theory the
vibration is defined as the periodic motion of any particle from its equilibrium
position. If the vibration level is too low, the frequency and the amplitude of the
vibration cannot be truely measured by sight or touch. But even in such low
levels (by the measure of absolute values) vibration can cause havoc in rotary
machineries. So the mechanical vibration is converted into electronic signal that
can be transmitted and amplified. The transducer converts mechanical vibration
into electronic signal, in which the frequency identifies how fast it is moving
and the amplitude says about how much it is deviating from its normal state.
The frequency tells us the cause of the vibration and the amplitude is the
measurement of severity of the problem.
1.2 CAUSES OF VIBRATION (GENERAL)
Forces generated within or outside the machine cause vibration. The force
can be generated because of the following reasons in general:
Change in direction with time, such as the force generated by a rotating
unbalance.
Change in amplitude or intensity with time, such as the unbalanced
magnetic forces generated in an induction motor due to un equal air gap
between the motor armature and stator (field).
Result in friction between rotating and stationary machine
components in much the same way that friction from a rosined bow causes a
violin string to vibrate.
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Cause impacts, such as gear tooth contacts or the impacts generated by
the rolling elements of a bearing over flaws in the bearing raceways.
There may be looseness or resonance due to coinciding natural and
generated frequencies of the machine.
1.3 WHY VIBRATION
The most important question faced while doing this project was “why are
we so excited about a machine’s vibration?” The answer is when a machine is
installed it is not expected to run smoothly for infinite time. Every machine has
its life time after that some of its component fail and those has to be replaced.
The most vulnerable component in respect to failure is the bearing of the
machine. In general the bearing life is inversely proportional to the cube of the
load, both static and dynamic. The amplitude of vibration indicates the amount
of unbalanced dynamic force on the bearing, which in turn signifies the
intensity of the problem. Sometimes there may be problem in the shaft or the
vanes, and then also there will be high vibration at the corresponding bearing.
So lesser the vibration lesser will be the unbalanced dynamic force and higher
will be the machine life. The basic aim is to reduce vibration by eliminating the
source problems and increase machine life and decrease cost of replacement of
the bearing and other parts of a machine.
1.4 CHARACTERISICS OF VIBRATION
Whenever vibration occurs, there are actually four properties involved that
determine the characteristics of the vibration. These forces are:
• The exciting force, such as unbalance or misalignment.
• The mass of the vibrating system, denoted by M.
• The stiffness of the vibrating system, denoted by the symbol K.
• The damping characteristics of the vibrating system, denoted by the
symbol C.
The exiting force is trying to cause vibration, where as the stiffness, mass
and damping forces are trying to oppose the exiting force and control or
minimize the vibration.
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The characteristics needed to define the vibration of a machine include:
1.
2.
3.
4.
5.
Frequency
Displacement
Velocity
Acceleration
Phase
1.4.1 Vibration Frequency
The amount of time required to complete one full cycle of the
vibration is called the period of the vibration. If, for example, the machine
completes one full cycle of vibration in 1/60th of a second, the period of
vibration is said to be 1/60th of a second. Although the period of the
vibration is a simple and meaningful characteristic, a characteristic
of
equal simplicity but more meaningful is the vibration frequency.
Vibration frequency is simply a measure of the number of complete
cycles that occur in a specified period of time such as "cycles-per-second"
(CPS) or "cycles-per-minute" (CPM). Frequency is related to the period of
vibration by this simple formula:
Frequency = 1/Period
In other words, the frequency of a vibration is simply the "inverse" of
the period of the vibration. Thus, the period of time required to complete
once cycle is 1 / 60th of a second, then the frequency of the vibration would
be 60 cycles-per-second or 60 CPS.
1.4.1.1 Significance of Vibration Frequency
There are literally hundreds of specific mechanical and operational
problems that can cause a machine to exhibit excessive vibration. The
majority of the forces that cause vibration are usually generated through the
rotating motion of the machine’s parts. Because these forces change in
direction or amplitude according to the rotational speed (RPM) of the
machine components, it follows that most vibration problems will have
frequencies that are directly related to the rotational speeds.
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Knowing this simple fact has eliminated literally hundreds of other
possible causes of vibration.
a) Predominant Frequency: Predominant frequency is the frequency
of vibration having the highest amplitude or magnitude.
b) Synchronous Frequency: Synchronous frequency is the vibration
frequency that occurs at 1x RPM.
c) Sub-synchronous Frequency: Sub-synchronous frequency is
vibration occurring at a frequency below 1x RPM. A vibration that
occurs at 1/2x RPM would be called a Sub-synchronous frequency.
d) Fundamental Frequency:
Fundamental frequency is the
lowest or first frequency normally associated with a particular
problem or cause. For example, the product of the number of teeth
on a gear times the RPM of the gear would be the fundamental
gear-mesh frequency. On the other hand, coupling misalignment
can generate vibration at frequencies of 1 x, 2x and sometimes 3 x
RPM. In this case, 1 x RPM would be called the fundamental
frequency.
e) Harmonic Frequency: A harmonic is a frequency that is an exact,
whole number multiple of a fundamental frequency. For example,
a vibration that occurs at a frequency of two times the fundamental
gear mesh frequency would be called the second harmonic of gear
mesh frequency. A vibration at 2x RPM due to, say, misalignment,
would be referred to as the second harmonic of the running speed
frequency (1 x RPM).
f) Order Frequency: An order frequency is the same as a harmonic
frequency.
g) Sub-harmonic Frequency: A sub-harmonic frequency is an exact
submultiples (1/ 2, 1/3, 1/4, etc.) of a fundamental frequency. For
example, a vibration with a frequency of exactly 1/2 the
fundamental gear-mesh frequency would be called a sub harmonic
of the gear mesh frequency. Vibration at frequencies of exactly
1/2, 1/3 or 1/4 of the rotating speed (1 x RPM} frequency would
also be called sub-harmonic frequencies; and these can also be
called Sub-synchronous frequencies. However, not all Sub
synchronous frequencies are sub harmonics.
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1.4.1.2 Vibration Amplitude
As mentioned earlier, vibration frequency is a diagnostic tool, needed
to help identify or pinpoint specific mechanical or operational problems.
Whether or not a vibration frequency analysis is necessary, depends on
how "rough" the machine is shaking. The magnitude of vibration or how
rough or smooth the machine vibration is, is expressed by its vibration
amplitude. Vibration amplitude can be measured and expressed as:
Displacement, Velocity, Acceleration,
The amplitude of vibration can be measured in four different ways,
namely peak-to-peak, zero-to-peak, RMS, average.
a) PEAK-TO-PEAK: Peak-to-peak is the distance from the top of
the positive peak to the bottom of the negative peak. The peakto-peak measurement of the vibration level is shown in fig
below. This type of measurement is most often used when
referring to displacement amplitude.
b) ZERO-TO-PEAK: Zero-to-peak or peak is the measurement from
the zero line to the top of the positive peak or the bottom of the
negative peak. The zero-to-peak value of the vibration level is
shown in fig below. This type of measurement is used to describe
the vibration level from a velocity transducer or accelerometer.
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c) ROOT MEAN SQUARE: It is the true measurement of the
energy under the curve. It is calculated by the square root of
the sum of the squares of a given number of points under the
curve. In the figure below the RMS is cosine 45 degree times
the peak of the wave or curve (0.707 x peaks, only applies to
pure sinusoidal curves).
d) AVERAGE: Sometimes we use the average value. The average
value is calculated by the analog meters, and then it is
converted to peak by multiplying with the constant 1.57. The
average value is thus 0.637 times the peak in case of pure
sinusoidal curves.
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1.4.2 Vibration Displacement
The vibration displacement is simply the total distance travelled by the
vibrating part from one extreme limit of travel to the other extreme limit of
travel. This distance is also called the "peak-to-peak displacement".
Peak-to-peak vibration displacement is normally measured in units called
mils, where one mil equals one-thousandth of an inch (1 mil = 0.001 inch).
Measured vibration amplitude of 10 mils simply means that the machine is
vibrating a total distance of 0.010 inches peak-to-peak. In Metric units, the
peak-to-peak vibration displacement is expressed in micrometers, sometimes
called microns, where one micrometer equals one-thousandth of a millimetre
(1 micrometer = 0.001 millimetre).
1.4.3 Vibration Velocity
The vast majority of machine failures caused by vibration problems
are fatigue failures, & the time required to fatigue failure is determined by
both how far an object is deflected. (displacement) and the rate at which the
object is deflected (frequency), of course, displacement is simply a measure
of distance travelled and frequency is a measure of the number of times that
“trip” is taken in a given period of time such as a minute or second, if it is
known how far one must travel in a given period of time, it is a simple
matter to calculate the speed or velocity required.
Vibration velocity is measurement of the speed at which a
machine or machine component is moving as it undergoes oscillating
motion. Vibration velocity is expressed in inches-per-second peak (in/secpike) for English units. In metric units, vibration velocity is expressed in
millimeters-per-second.
1.4.4 Vibration Acceleration
It is another important characteristic of vibration that can be used to
express the amplitude or magnitude of vibration. Technically, acceleration is
simply the rate of change of velocity. For a pure sinusoidal wave the
acceleration of the weight is highest or at its peak value at the upper limit of
travel where the velocity is zero. As the velocity of the weight increases, the
rate of change of velocity or acceleration decreases. At the neutral position,
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the weight has reached its maximum or peak velocity and at this point, the
acceleration is zero. After the weight passes through the neutral position, it
must begin to slow down or "decelerate" as it approaches the lower limit of
travel. At the lower limit of travel the rate of change of velocity
(acceleration) is, again, at its peak value.
1.4.5 Vibration Phase
Phase, with regards to machinery vibration, is often defined as "the
position of a vibrating part at a given instant with reference to a fixed point
or another vibrating part. Another definition of phase is: "that part of a
vibration cycle through which one part or object has moved relative to
another part". However, from a practical standpoint, phase is simply a
convenient mean of determining the "relative motion" of two or parts of a
machine or vibrating system. The units of phase are degrees, where one
complete cycle of vibration equals 360 degrees.
1.5 SOURCES OF FREQUENCIES:
There are in general three sources of frequencies generated frequencies,
excited frequencies and frequencies caused by electronic phenomena.
a) Generated frequencies, sometimes called forcing frequencies,
are those frequencies actually generated by the machine. Some
examples are imbalance, vane pass frequency (number of vanes
times speed), gear-mesh frequency (number of teeth times speed),
various frequencies generated by antifriction bearings, ball pass
frequency of the outer race, ball pass frequency of the inner race,
ball spin frequency, and fundamental train frequency. Generated
frequencies are the easiest to identify because they can be
calculated if the internal geometry and speed of the machine are
known. Some of the calculated frequencies may be present in most
machines without indicating a vibration problem. Calculated
frequencies should not be modulated with any degree of
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significance by other frequencies. If any of these frequencies are
generated, a vibration problem exists.
b) Excited frequenciesare property of the system. Amplified
vibration, called resonance occurs when a generated frequency is
tuned to a natural frequency. Vibration is amplified in a band of
frequencies around the natural frequency. The amplitude depends
upon damping. The natural frequency is the centre frequency of
this band. Natural frequency just acts like an electronics amplifier
and amplifies the signal of the vibration. Although natural
frequency is good where it acts as a carrier but in almost all the
cases it is detrimental for the machine due to the very high
vibration. The term ‘critical speed’ means the rotating speed of the
unit is equal to the natural frequency of the machine or any
component. This frequency is considered unacceptable by some
experts.
The natural frequency is not calculated frequency but modulated
by generate frequency. When the source of vibration or the
generated frequency is removed, the natural frequency will not
be excited. For this the natural vibration should be more that the
generated frequency.
c) Frequencies generated by electronic phenomena: In
certain cases there can be false or misleading vibration due to
some electronic phenomena. For example when a sinusoidal
wave is clipped there will be a lot of spurious frequencies.
Sometimes if there is a strong electromagnetic field, the signal
gets distorted, so we get frequencies those neither are generated
nor are excited. It is recommended to use filter to stop electronic
noise to enter and modulate the original required frequency
spectrum.
1.6
RELATIONSHIP
BETWEEN
DISPLACEMENT, AND ACCELERATION
VELOCITY,
Velocity is the measurement of how fast an object is moving from zeroto-peak and is normally measured in tenths of one inch per second (IPS). The
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effective frequency range of most velocity transducers is from about 10 to 2,000
Hz. Velocity is the most accurate measurement because it is not frequency
related. For example, 0.15 IPS is the same at 10 Hz as it is at 2,000 Hz.
Displacement is the measurement of how far an object is moving from
peak-to-peak and is normally measured in thousandths of one inch (mils).
Displacement is frequency related. So we need to specify the RPM while
measuring or stating vibration in displacement.
Acceleration measures the rate of change of velocity from zero-to-peak
and is normally measured in units of gravitational force (g's). This means that
high frequencies generate high g levels, and acceleration is frequency related.
Low frequencies generated high amplitude of displacement and high
frequencies generate high amplitude of acceleration. So displacement transducer
measures vibration at low frequency with greater accuracy and the acceleration
transducer does the same for higher frequency. As given in the following figure,
the response of velocity transducer is flat within 10 and 2000 Hz. So in this
range velocity transducer is a better option.
The relation between different measured quantities is given below
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Where
1.7 TIME DOMAIN AND FREQUENCY DOMAIN
Vibration analysis is basically done by analysing two different types of
signals. Time domain means the variation of vibration amplitude with respect to
the time and the frequency domain gives the same variation with respect to
frequency.
We get the time domain signal from the machinery under inspection. But
we know any time wave form is comprises of many frequencies. To get the
frequency we use Fast Fourier Transformation. It breaks the time domain signal
into a frequency domain signal.
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Fig: relation between time domain and frequency domain
In the above figure we can see the bottom wave form contains the
fundamental, the middle one is the second harmonic and the third one is the
third harmonic. The left hand axis denotes the frequency domain, where we get
one peak at each frequency. The right hand axis is the time domain axis. We get
the full wave. The top waveform contains all the three harmonics.
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PART-2
ROTATING MACHINES AND
COMPONENTS
Page 16
3 types of rotating equipments were kept under the purview of this project.
Short description of the equipments and their components are mentioned below.
2.1 TURBINE
Turbine is an equipment which converts heat energy into mechanical
energy and then into electrical energy. For an industrial turbine high pressure
and high temperature steam is forced into the turbine through valves where it
expands, thus rotating the rotor of the turbine at rated speed. The turbine rotor is
coupled to an AC/DC generator which produces electricity at rated voltage and
frequency (for AC generators). The expanded steam then moves through a
condenser for condensation. Large turbines use journal bearings for carrying the
load of the rotor.
2.2 PUMP
Pump is mechanical equipment which uses energy from a prime mover and
converts it into flow energy either in terms of pressure or flow rate or both. The
prime mover may be an electrical motor or high energized steam or oil. There
are basically two types of pumps, one being positive displacement pump where
the fluid particles are physically pushed towards the exit and another being
centrifugal pump where high pressure at exit of the pump is produced by
centrifugal action of the vanes of the pump as it rotates.
2.3 FAN
Fan is a mechanical component which deals with gaseous medium, and
pressurizes it by using energy from a prime mover.
2.4 BEARING
The term "bearing" comes ultimately from the verb "to bear", and a bearing
is thus a machine element that allows one part to bear another, usually allowing
(and controlling) relative motion between them.
This is typically to allow and promote free rotation around a fixed axis or
free linear movement; it may also be to prevent any motion, such as by
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controlling the vectors of normal forces. Bearings may be classified broadly
according to the motions they allow and according to their principle of
operation, as well as by the directions of applied loads they can handle.
2.4.1 TYPES OF BEARING
Type
Description
Friction Speed
Plain
bearing
Rubbing
surfaces, usually
with lubricant;
some bearings
use
pumped
lubrication and
behave similarly
to
fluid
bearings.
Depends
on
Low
materials
very
and
high
construct
ion,
Notes
Low to very
high
to
Widely
depends upon
relatively
application
friction,.
and
lubrication
used,
high
Moderat
e to high
(often
requires
cooling)
Moderate to
high (depends
on
lubrication,
often requires
maintenance)
Used for higher
loads than plain
bearings
with
lower friction
Very
high
(usually
Zero
Fluid is forced
limited
friction
between
two
to a few
at zero
faces and held
hundred
speed,
in by edge seal
feet per
low
second
at/by
seal)
Virtually
infinite
in
some
applications,
may wear at
startup/shutdown
in
some
cases. Often
negligible
maintenance.
Can fail quickly
due to grit or dust
or
other
contaminants.
Maintenance free
in continuous use.
Can handle very
large loads with
low friction.
Ball or rollers
Rolling are used to
element prevent
or
bearing minimise
rubbing
Fluid
film
bearing
Life
Low
friction
due less
are
of
contact
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2.4.1.1Thrust bearing
A thrust bearing is a particular type of bearing. Like other bearings they
permit rotation between parts, but they are designed to support a high axial load
while doing this (parallel to the shaft). Higher speed applications require oil
lubrication. Generally, they are composed of two washers (raceways) in which
those for ball bearings may be grooved, and the rolling elements. These
bearings can be of two types
a) Ball Thrust Bearings - Composed of ball bearings supported in a
ring, can be used in lower thrust applications where there is little radial
load. Washers may be flat or grooved. Miniature thrust bearings starting
with bores as small as 2 mm but as large as 340 mm.
b) Roller Thrust Bearings- Composed of cylindrical rolling elements,
Roller thrust bearings have higher load carrying capacities then equally
sized ball thrust bearings. Washers are typically flat (not grooved). Small
roller thrust bearings start at 45 mm bores.
Fig: Spherical roller thrust bearing.
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c) Tilting pad thrust bearing-Composed of number of pads supported in a
carrier ring. The pads can be tilted about a pivot to create a self sustaining
fluid film. It carries the load axially. They are used for high speed shafts
and thus require forced lubrication. It acts as an axial locator also.
Fig: Tilting pad thrust bearing
2.4.1.2 Journal bearing
Among all the bearing used here, the journal bearing is of greatest interest as it
carries highest load and requires forced lubrication by using equipmentslike Oil
Pump, Centrifuge, Strainer etc. This type of bearings are used in turbine, Boiler
feed pump etc. In general journal bearing consists of two halves or lobes; the
fluid at high pressure is pumped into the bearing. As the shaft starts rotating the
fluid pressure gradually builds up and makes the shaft crouching. Then the
pressure pushes the shaft to the other side of the bearing and creates a high
pressure zone there. The shaft is expected to be stable at that position. During
the start up and coasting down, due to low speed there is not enough pressure in
the fluid inside the bearing for holding the load. To stop shaft from rubbing,
another pump, i.e. jacking oil pump is used. This pump supplies high pressure
oil to just hold the shaft until the pressure is built up in the main oil film.
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Fig: Journal bearing
2.5 COUPLING
A coupling is a device used to connect two shafts together at their ends for the
purpose of transmitting power. Couplings do not normally allow disconnection
of shafts during operation, however there are torque limiting couplings which
can slip or disconnect when some torque limit is exceeded.
2.5.1 TYPES OF COUPLING
2.5.1.1 Rigid Coupling
A rigid coupling is a unit of hardware used to join two shafts within a
motor or mechanical system. It may be used to connect two separate systems,
such as a motor and a generator
When joining shafts within a machine, mechanics can choose between
flexible and rigid couplings. While flexible units offer some movement and give
between the shafts, rigid couplings are the most effective choice for precise
alignment and secure hold. By precisely aligning the two shafts and holding
them firmly in place, rigid couplings help to maximize performance and
increase the expected life of the machine.
Flanged rigid couplings are designed for heavy loads or industrial equipment.
They consist of short sleeves surrounded by a perpendicular flange. One
coupling is placed on each shaft so the two flanges line up face to face. A series
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of screws or bolts can then be installed in the flanges to hold them together.
Because of their size and durability, flanged units can be used to bring shafts
into alignment before they are joined together. Rigid couplings are used when
precise shaft alignment is required; shaft misalignment will affect the coupling's
performance as well as its life. Examples:
Sleeve or muff coupling
Clamp or split-muff or compression coupling
2.5.1.2 Flexible Coupling
Flexible couplings are used to transmit torque from one shaft to another when
the two shafts are slightly misaligned. Flexible couplings can accommodate
varying degrees of misalignment up to 3° and some parallel misalignment. In
addition, they can also be used for vibration damping or noise reduction.
2.5.1.3 Gear Coupling
A gear coupling is a mechanical device for transmitting torque between two
shafts that are not collinear. It consists of a flexible joint fixed to each shaft. The
two joints are connected by a third shaft, called the spindle.
Each joint consists of a 1:1 gear ratio internal/external gear pair. The tooth
flanks and outer diameter of the external gear are crowned to allow for angular
displacement between the two gears. Gear couplings are generally limited to
angular misalignments, i.e., the angle of the spindle relative to the axes of the
connected shafts, of 4-50.
Single joint gear couplings are also used to connect two nominally coaxial
shafts. In this application the device is called a gear-type flexible or flexible
coupling.
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2.5.1.4 Lovejoy Coupling
Lovejoy coupling is a special type of flexible coupling which allows about
2degree angular misalignment and 0.04” parallel misalignment. It consists of a
spider and two hubs attached to the two shafts. The spider transmits the torque.
The Lovejoy coupling is used in Centrifuge.
Fig: Lovejoy coupling
2.5.1.5 MetaflexCoupling
It is another type of flexible coupling. It also allows some amount of axial
misalignment. It is used here in Boiler Feed Pump.
Fig: Metaflex Coupling
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2.5.1.6 Fluid Coupling
A fluid coupling is a hydrodynamic device used to transmit rotating mechanical
power. It has been used in automobile transmissions as an alternative to a
mechanical clutch. It has widespread application in industrial machine drives,
where variable speed operation and/or controlled start-up without shock loading
of the power transmission system is essential. Like in Boiler Feed Pump we use
VOITH Coupling, which is a modified fluid coupling.
Fluid coupling consists of three components, plus the hydraulic fluid:
1. The housing, also known as the shell (which must have an oil tight
seal around the drive shafts), contains the fluid and turbines.
2. Two turbines (fan like components):
•
One connected to the input shaft; known as the pump or
impellor primary wheel input turbine
•
The other connected to the output shaft, known as the
turbine, output turbine, secondary wheel or runner.
The driving turbine is rotated by the prime mover, which is typically an
electric motor. The fluid imparts force on the impeller of the driven part, which
in turn rotates the driven shaft. The amount of fluid or the angle of incidence
determines the relative velocity between the driving and the driven shaft.
2.6 FOUNDATION
One of the most important components of any rotating equipment is the
foundation. It holds the bearing and subsequently the machinery. The design
and construction of the foundation depends upon the total load to be carried,
amplitude of unbalance force etc. The turbine foundation requires specific
materials and construction due to the high load they have to carry.
The bearing casings are attached to the foundation by different means.
Sometimes they are bolted to the base plate. Depending upon requirement the
foundations sometimes have bolts grouted inside.
Foundations are made of concrete generally. Sometimes there are some
reinforcement to increase the strength.
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PART-3
COMMON FAULTS OF ROATING
MACHINES
Page 25
There are a lot of problems that can be diagnosed by good analysis of timewaveform and FFT of vibration taken at different position of the rotating
machine in different direction. The most common problems of a rotating
machine i.e. unbalance, bent shaft etc and their cause are written below.
3.1 UNBALANCE
The most common problem of a rotating machine is unbalance. It is a
condition of a rotating part where the centre of mass does not lie on the centre
of rotation. Unbalance of a rotor causes a centripetal force at the frequency of
the rotation rate to be applied to the bearings. If the unbalance is too high it can
shorten the life of the bearing due to periodic high stress induced fatigue. The
unbalance of a motor causes high vibration and the amplitude of the unbalanced
force as well as the vibration is proportional to the square of the rpm of the
machine. So a higher speed machine has to be balanced to a higher standard.
Unbalance of a rotating machine can be static, dynamic or couple
unbalance or any combination of these. In static unbalance the principal inertia
axis of a rotor is offset from and parallel to the axis of rotation.
A rotor with static unbalance will seek a position with the heavy spot at
the bottom if placed on level knife-edges. Static unbalance can theoretically be
corrected by the addition of a single correction mass.
Couple unbalance is the condition where the principal inertia axis
intersects the rotation axis of the rotor at the centre of gravity. A rotor with
couple unbalance will be stable in any position on knife edges, but will produce
out-of-phase unbalance forces on the bearings when rotated. Correction of
couple unbalance requires the addition of two correction masses.
Dynamic unbalance is a combination of these two types, and is the most
common type found in practice. In dynamic unbalance, the principal inertia axis
neither intersects nor is parallel to the axis of rotation. Correction of dynamic
unbalance requires at least two correction masses.
3.2 MISALIGNMENT
Misalignment of rotating machinery is one of the most important and
most common problems. It is responsible for most of the faults in rotating
machinery. A misaligned rotor generates bearing forces and excessive
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vibrations making diagnostic process more difficult. A perfect alignment can
never be achieved practically and misalignment is always present. Also, many
factors such as thermal growth, uneven applied loads, inappropriate
foundations, etc., can disturb the alignment.
Alignment can be of two types- radial and axial.
Radial: In this case the shaft centre-lines are not concentric; they are
offset from one another. The centerlines are parallel which makes the shafts go
up and down in an orbit around the coupling, which creates the force as well as
the vibration in the machinery. This can happen in both vertical and horizontal
plane.
Axial: In this case the centerlines are not parallel, they may intersect at
one of the faces, but chances of that are very slim. In most of the cases they
intersect somewhere between the two coupling faces. As the couplings are
joined by bolts, in angular or axial misalignment the shafts rotate in an arbitrary
way and causes vibration.
Page 27
In alignment the amplitude of vibration often is not very high, that does
not mean it is in good condition as most of the time we are unaware of the
misalignment until the couplings, shaft fits, or the bearings are worn out.
3.3 BENT SHAFT
Fig: Bent shaft
Another problem that is also quite common is bent shaft. In a perfectly
straight shaft, the centers of each shaft cross-section from end-to-end of the
shaft lie on a straight line. A shaft is bent if that is not the case. In a bent shaft,
the axis of the shaft is different than its axis of rotation. A bent shaft can be
caused by one or more of the following factors...
Mechanical overload:
Damage during rigging or improper handling
Impact during operation
Machine misalignment
Internal stress relief:
Unequal machining operations
Vibration during shipment
Improper material handling during heat-treating, rolling,
forging, thermal stress relieving
• Elevated temperature during operation
• Assembly stack-up stresses
•
•
•
•
•
•
•
•
A shaft that is initially straight can bend due to stresses caused by heavy
shrink fits with mating components such as a turbine wheel, for example. Bent
shafts caused by assembly stack-up stresses are usually not correctable by
simply straightening the shaft itself. The assembly needs to be analyzed as a
complete unit, and corrected accordingly. Sometimes, a change in the fits
Page 28
between the shaft and its mating components is required to accomplish the
straightness needed.
• If bent shaft is not corrected anyone or more from the
following can happen
• Equipment vibration due to unbalance
• Shaft misalignment
• Damage to bearings, seals, and couplings
• Contact and possible seizure with close-clearance surfaces
• Material fatigue and failure
3.4 SOFT FOOT
Fig: Problem due to soft foot
The common term for machine frame distortion, soft foot is caused when
one or more feet of a machine are shorter, longer or angled some way different
than the rest of the feet. Soft foot can be caused by the following:
1. Machinery foundations which are warped or twisted and have base
plates;
2. Damaged machine feet;
3. Inadequate number of shims in the machine feet;
4. Foreign materials under the feet of the machine or any dirt or trash
beneath it;
5. Flaws or dents in the base or feet of the machine; and
Page 29
6. The presence of jack bolts that warped the feet of the machine causing
it too much tension.
When machine frame distortion exists, the bearing housings are
misaligned with respect to one another. This offset, as well as angular
misalignment creates a load on the rotating shaft that results in shaft deflection.
When the shaft turns, these result in vibration since the shaft must deflect by
double the amount of the deflection at rest, at twice the speed of rotation. As the
bearing provide the force required, soft foot increases load on the bearing and
power consumption as well as provoke fatigue induced failure. Aside from this,
it has also become the cause of problems for reduction of the operational life of
the electric motors, and the disadvantage of internal clearance in pumps and
also gearboxes.
3.5 OIL WHIRL & OIL WHIP
Oil whirl is probably the most common cause of sub-synchronous
instability in hydrodynamic journal bearings. Typically, the oil film itself flows
around the journal to lubricate and cool the bearing. This develops an average
speed slightly less than 50 percent of the journal surface speed.
Normally, the shaft rides on the crest of an oil pressure gradient, rising
slightly up the side of the bearing somewhat off vertical at a given, stable
attitude angle and eccentricity. The amount of rise depends on the rotor speed,
rotor weight and oil pressure. With the shaft operating eccentrically relative to
the bearing centre, it draws the oil into a wedge to produce this pressurized
load-carrying film. If the shaft receives a disturbing force such as a sudden
surge or external shock, it can momentarily increase the eccentricity from its
equilibrium position. When this occurs, additional oil is immediately pumped
into the space vacated by the shaft. This results in an increased pressure of the
load-carrying film, creating additional force between the oil film and shaft. In
this case, the oil film actually drives the shaft upward, as soon as the clearance
is increased over a limiting value, the pressure again drops. The shaft drops to
the bottom. The shaft continues this motion of a whirling path around the
equilibrium position within the bearing clearance. If there is sufficient damping
within the system, the shaft can be returned to its normal position and stability.
Page 30
Otherwise, the shaft will continue in its whirling motion. When the whirl speed
i.e. less than half the speed of rotation, becomes equal to the system natural
frequency, the amplitude of instability due to oil whirl increases many times.
This phenomena is known as oil whip
The major causes of oil whirl and oil whip are ------• Low load corresponding to the oil pressure
• Unequal clearance
• Misalignment which provides the shaft to position itself so
as instability is created.
Oil whip and whirl result in severe wear of the bearing and the shaft
which in turn increases clearance and increases the vibration. When whip occurs
the amplitude of vibration or unbalanced force is too high which can cause
failure of the bearing as well as the whole machine.
3.6 LOOSENESS
There can be various types of looseness in a machine. But in general
looseness lessens the rigidity of machine, which in turn allows more vibration
or more unbalanced detrimental force on the machine.
Anti-friction bearing looseness can be of several forms---- sometimes the
outer-race gets loose from the Plummer block or the casing, sometimes the
inner race is not rigidly attached to the shaft. Sometimes the balls are loose in
the cage. In all the cases there is slip between to surfaces instead of pure rolling,
so the actual velocities of the inner, outer race or the balls are less than normal.
This results in significant amount of sub-harmonic vibration for inner race and
out race.
Bolts can be loose in parting plane or in the foot of the machine. Loose
bolt in the foot of the machine is completely different from soft foot. Sometimes
loose bolts in the foundation block allow more axial vibration than limit which
is detrimental to the machines which do not have axial locators or axial support.
Sometimes in journal bearing the inner ring may be loose or worn, in
those cases the clearance increases, which is more like looseness of a journal
bearing.
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3.7 RESONANCE
Resonance is a very common cause of excessive vibration on machines
because
1. Machines consist of many individual elements or components such as
suction and discharge piping, bearing pedestals, bases, and accessory items such
as exciters and lube oil pumps, etc... Of course, each component has its own
mass and stiffness characteristics and, hence, its own unique natural frequency.
2. The stiffness of each machine component is different in different
directions. As a result, each machine component will likely have several
different natural frequencies. For example, consider a fan bearing. Most
likely, the stiffness of the bearing will be different in the horizontal,
vertical and axial directions. As a result, the natural frequencies of this
particular machine component will also be different in the horizontal, vertical
and axial directions.
3.8 RUBS
When there is low clearance between stator and rotor, the rotor may touch
once or more the stator part, which is called rubbing.
Rubbing at such a high speed of the rotor causes vibration. Interference
becomes too much, there is impact. Depending upon the cause of problem
rubbing or impact of this type can happen once in a revolution or more.
The only way to reduce this vibration is too increase the clearance, if the
only cause is interference. Sometimes broken off pieces can create rubs.
3.9 HIGH POINT IN GEAR AND ANTI-FRICTION
BEARING
There is another common cause of vibration i.e. high point on the gear
tooth or the bearing balls or the races. If there is a high point the number of
impact per revolution would be equal to the number of balls or number of teeth.
If there is more than one high point then the generated frequency as well as the
amplitude will be very high.
Improper machining, high impact load are the main causes. The only way
to reduce it is to overhaul the bearing or the gear.
Page 32
PARTPART-4
VIBRATION ANALYSIS
TECHNIQUES
Page 33
The fault detection of rotating machines is done by analyzing the
vibration signature from the machines. The vibration signature contains the
FFT, time waveform, phase, orbit etc. The FFT gives us the vibration frequency
and the amplitude at any specific frequency, i.e.1x, 2x etc. The basic analysis
starts from studying the FFTs. Sometimes the FFT alone is not sufficient; we
use time waveform or phase then.
4.1 TRANSDUCERS
We use various transducers to measure vibration. As explained before, we
use accelerometers where the frequency is high and displacement probe where
frequency of vibration is low. However, using the acceleration data and
software we can get the value of velocity and displacement, but not exact value.
The basic of the transducers are given below.
4.1.1 DISPLACEMENT TRANSDUCER
The two basic types are non-contacting and contacting. Both types
require firm mounting. Non-contacting displacement transducers require a
power supply, a modulator/demodulator, normally called a driver, and a volt
meter. The transducer has a small wound wire-coil at the end. The driver
generates a high frequency depending upon the distance between the tip and the
point to be measured. When the frequency passes through the coil, a magnetic
field is generated. The motion between the mount and target intersects the
magnetic field. This means the relative motion between turbine housing and
shaft across the coupling can be measured. But in case of contact type
transducers, the probe must touch the part whose vibration is to be measured.
4.1.2 VELOCITY TRANSDUCER
These transducers are voltage generators and do not require an external
power source. Most velocity transducers employ a permanent magnet mounted
on a stud. A coil of wire surrounds the magnet and the coil is supported by leaf
springs. When mounted on a machine, the motion causes the coil to move
through or intersect the magnetic field. So a voltage is produced. The output is
proportional to the change of position or the velocity. The sensitivity and the
resolution of the transducer depend upon the number of coils, the size of the
wire etc.
Page 34
Fig: velocity transducer
4.1.3 ACCELEROMETER
Most accelerometers use piezoelectric crystal principle, in which the
crystal generates voltage in response to a mechanical force. As the signal is too
weak to be transmitted, an electronic amplifier is added to near to the
transducer. These transducers give response to high frequency only. But, in
general, the larger the crystal, the lower is the frequency. Accelerometers
should be screwed or glued to the position where the measurement is being
taken. Otherwise there is a huge chance that the vibration can be overstated or
understated by about 12 dB or 4 times the actual.
Page 35
Fig: Accelerometer probe
Fig: CMXA 70
There are different accelerometers, eddy-current probes for measuring
shaft vibration or pedestal vibration continuously. There is a continuous
monitoring system in place for this. The online monitoring system shows only
the overall vibration for instantaneous work, but there is no provision for
analysis of the online data. The online vibration monitoring equipments or
probes are generally screwed to the bearing casings. Two eddy-current probes at
900 are placed to locate the shaft centreline perfectly.
Other than the usual online measuring instruments, there are several
offline vibration measuring instruments. In CTM department, CESC Ltd., I have
experienced firsthandsome of the latest vibration monitors and analysers have
used SKF make CMXA 50, CMXA 70 and CMXA 75 for this project.
4.2 ANALYSIS
There are literally hundreds of specific mechanical and operational
problems that can result in excessive vibration. However, since each type of
problem generates a unique vibration signature, a thorough analysis of vibration
characteristics can go a long way in reducing the number of possibilities and
hopefully to a single cause. A modern fish-bone analysis has been proven to be
successful in pinpointing the cause of vibration in day to day machinery
problems.
The major steps in the analysis are as below:
4.2.1 DEFINING THE PROBLEM
The following lists some of the reasons for performing a vibration
analysis:
1.
Establish "baseline data" for future analysis needs. At the
beginning of a predictive maintenance program, even machines in good
Page 36
operating condition should be thoroughly analyzed to establish their
normal vibration characteristics. Later, when problems do develop, this
baseline information can be - extremely useful in performing a follow-up
analysis to show precisely the vibration characteristics that have changed.
2.
Identify the cause of excessive vibration. Referring to the
vibration severity guidelines machines in service that have vibration levels in
the “alarm" regions or greater should be thoroughly. We analyze to identify
existing problems for immediate correction. Once corrections have been made,
a follow-up analysis should be performed to insure that problems have been
solved and the machine returned to satisfactory condition. If all significant
problems have been solved, the follow-up analysis data will serve as the
baseline data for future analysis as outlined in (1) above.
3.
Identify the cause of a significant vibration increase. Once a
developing problem has been detected by routine, periodic checks, the
obvious next step is to perform a detailed vibration analysis to identify the
problem for correction. Here also, a follow-up analysis will verify that the
problems have been corrected and provide a baseline for future comparison
4. Identify the cause of frequent component failures such as bearings,
couplings, seals, shafts, etc, if any.
4.2.1 COMPARISON WITH STANDARD
To identify if the measured vibration indeed represents a problem in the
machinery, there are some standards. Other than the manufacturer’s guideline,
in general, in industry we follow ISO 10816 and ISO 7919.
The following figure gives the details of applicability of the ISO with respect to
specific machines
Page 37
I have used the ISO 10816 only in this project. The detail alarm values and
advice trip values are given below
Page 38
4.2.3DETERMINE MACHINE DETAILS
Some of the important detailed features of the machine that need to be
known for accurate analyses include:
1. The rotating speed (RPM) of each machine component: Of course,
direct-coupled machines have only one rotating speed (RPM) that
needs to be known.
2. Types of bearings: Of course worn or defective sleeve or plain
bearings will have different vibration characteristics than defective rollingelement bearings. Therefore, it is most important to know whether the machine
has plain or rolling element bearings. If the machine has rolling-element
bearings, it is also beneficial to know the number of rolling elements and other
details of bearing geometry; with this information, the vibration analyst
can actually calculate the frequencies of vibration caused by specific bearing
defects such as flaws on the outer and inner raceways, rolling elements, etc.
3. Number of fan blades and pump impeller: Knowing the machine
RPM and number of blades on a fan will enable the analyst to easily calculate
the "blade-passing" frequency. This is simply the product of the number of fan
blades times fan RPM. Similar to fans and blowers, knowing the number of
vanes on a pump impeller allows the analyst to calculate the vane-passing
frequency, also called the "hydraulic-pulsation" frequency.
4. Number of gear teeth: The rotating speed and number of teeth on each
gear must be known in order to determine the possible "gear-mesh" frequencies.
5. Type of coupling: Gear and other lubricated types of couplings can
generate some unique vibration characteristics whenever their lubrication breaks
down or if lubrication is inadequate. Whereas a rigid coupling cannot have any
misalignment but flexible coupling can allow some amount of misalignment.
6. Machine critical speeds: Some machines such as high speed, multistage centrifugal pumps, compressors and turbines are designed to operate at
speeds above the natural or "resonant" frequency of the shaft. The resonant
frequency of the shaft or rotor is called its "critical" speed, and operating a near
this speed can result in extremely high vibration amplitudes. Therefore,
knowing the rotor critical speed relative to machine RPM and other potential
exciting force frequencies are very important.
Page 39
7. Background vibration sources: Many times the vibration being
measured on a machine is actually coming from another machine in the
immediate area. This is particularly true for machines mounted on the same
foundation or that are interconnected by piping or other structural means.
Therefore, it is important to be aware of potential "background" contributions.
.
4.2.4VISUAL INSPECTION
Before collecting data, the vibration analyst should first make a visual
check of the machine to determine if there are any obvious faults or defects that
could contribute to the machines condition. Some obvious things to look for
include;
1. Loose or missing mounting bolts
2. Cracks in the base, foundation or structural welds
3. Leaking seals
4. Worn or broken parts
5. Wear, corrosion or build-up of deposits on rotating elements such as
fans.
4.2.5HORIZONTAL, VERTICAL AND AXIAL SPECTRUMS
In many cases, the analysis steps carried out thus far may be sufficient to
pinpoint the specific problem causing excessive vibration. If not, the next step is
to obtain a complete set of amplitude-versus-frequency spectrums or FFTs at
each bearing of the machine train. For a proper analysis, the machine should
be operating under normal conditions of load, speed, temperature, etc.
In order to insure that the analysis data taken includes all the problemrelated vibration characteristics and, yet, is easy to evaluate and interpret, the
following way is taken, in general;
Page 40
4.3 INTERPRETING THE DATA
Once horizontal, vertical and axial FFTs have been obtained for each
bearing of the machine train the first thing we look for is the spectrum which
helps us
1. Identifying the machine component (motor, pump, gear box, etc.) of
the machine train that has the problem,
2. Reducing the number of possible problems from several hundred to
only a limited few.
4.3.1 Identifying the Problem Component Based On Frequency
Most problems generate vibration with frequencies that are exactly
related to the rotating speed of trip in trouble. These frequencies may be
exactly 1 x RPM or multiples (harmonics) of 1 x RPM such as 2x, 3x, 4x, etc.
In addition, some problem's may cause vibration frequencies that are exact sub
harmonics of 1 x RPM such as 1/2x, l/3x or 1/4 x RPM. In any event, the FFT
analysis data can identify the machine component with the problem based on
the direct relationship between the measured vibration frequency and the
rotating speed of the various machine elements.
4.3.2 Identifying the Problem Component Based On Amplitude
When we have directly coupled machine, both the motor and the pump or
fan has same frequency, we identify depending upon the amplitude of the
vibration. The machine part which shows maximum vibration is taken to be the
source, in general. However, there are exceptions to this rule such as
misalignment of direct coupled machines. In the case of coupling misalignment,
the vibratory force (action) is generated at the coupling between the driver a
driven components. As a result, the "reaction" forces on the driver and
driven unit; will be essentially equal, resulting in reasonably
comparable vibration amplitudes. The only reason one component may have a
slightly higher or lower amplitude than the other is because of
differences
in
the mass
and stiffness characteristics of
the two
Page 41
components.
But,
in most cases with the coupling misalignment, the
vibration is fairly uniformly "shared" by the driver and driven units.
4.3.3 Reducing the List of Possible Problems Based On Frequency
In addition to identifying the problem machine component based on
frequency and/or amplitude characteristics, the second purpose of FFT analysis
data is to limit or reduce the list of possible problems based on the measured
vibration frequencies. As stated earlier, each mechanical and operational
problem generates its own unique vibration frequency characteristics.
Therefore, by knowing the vibration frequency, a list of the problems that
cause or generate that particular frequency can be made, which greatly reduces
the long list of possibilities.
The chart lists the most common vibration frequencies is they
relate to machine rotating speed (RPM), along with the common causes for each
frequency. The most common case is the high vibration at the frequency of
rotation or at 1x. There can be 10causes for vibration at 1x, they are:
1. Unbalance
2. Eccentric shaft
3. Misalignment
4. Bent shaft
5. Looseness
6. Distortion—from soft foot or piping strain conditions
7. Bad belts—if belt RPM
8. Resonance
9. High clearance in journal bearing
10. Electrical problems
Using this simple chart, along with the fact that the vibration frequency is
1x RPM of the fan has reduced the number of possible causes from literally
hundreds to only ten (10) likely causes, A little common sense can reduce this
Page 42
list even further. For an example if the 1x is different from the AC frequency,
then it is not an electrical problem.
Table: VIBRATION FREQUENCIES AND THE LIKELY CAUSES
Frequency in
Terms Of RPM
Most
Likely causes
1x RPM
Unbalance
2 x RPM
3 x RPM
Less than 1x RPM
Other possible causes &
Remarks
1. Eccentric shaft
2. Misalignment
3. Bent shaft
4. Looseness
5. Distortion—from soft foot or
piping strain conditions
6. Bad belts—if belt RPM
7. Resonance
8.High clearance in journal
bearing
9. Electrical problems
10.Reciprocating problems
Mechanical
looseness
1)Misalignment if high axial
vibration
2) Reciprocating forces
3) Resonance
4) bad belt if 2 x RPM of belt
Misalignment
Usually a combination of
misalignment
andexcessive
axial clearance (looseness).
Oil Whirl (Less 1) Bad drive belts
Page 43
than1/2 x RPM) 2) Background vibration
or Oil Whip
3)Sub-harmonic resonance
4) "Seat" Vibration
Synchronous(A.C
linefrequency)
Electrical
Problems
Common electrical problems
include broken rotor bars,
eccentric
rotor,
and
unbalanced phases in polyphase systems, unequal air gap.
2xSynch.
Torque Pulses
Rare as a problem
resonance is excited
Frequency Many Times
RPM
(HarmonicallyRelated
Freq.)
Bad Gears
Aerodynamic
Forces
Hydraulic forces
Mechanical
Looseness
Reciprocating
Forces
Gear teeth times RPM of bad
gear
Number of fan blade times RPM
Number of impeller vane times
RPM
May occur at 2, 3, 4 and
sometimes higher harmonics if
severe looseness
unless
High Frequency
Bad
Anti- 1) Bearing vibration may be
amplitude
and
(Not
Harmonically Frictionbearing unsteady
Related)
frequency
2)Cavitations, recirculation and
flow turbulence causes random
high frequency vibration
3)Improper
lubrication of
journal bearings
4)rubbing
4.3.5 Comparing Tri-Axial (Horizontal, Vertical and Axial) Data
Page 44
Not only can specific vibration problems be recognized by their
specific frequency characteristics, but in many cases by the direction in which
the vibration occurs. This is why it is necessary to take analysis data in the
horizontal, vertical and axial directions to further the process of elimination.
From the data we compare the vibration amplitude at any specific
frequency between the vertical and the horizontal, then between radial (vertical
and horizontal) and axial.
4.3.6 Comparing Horizontal and Vertical Readings
When comparing the horizontal and vertical data, it is important to
take note of how and where the machine is mounted and also, how the
bearings are mounted to the machine.
Basically, the vibration analyst
needs to develop a “feel" for the relative stiffness between the horizontal
and vertical directions in order to see whether the comparative
horizontal and vertical readings indicate a normal or abnormal situation.
Machines mounted on a solid or rigid base may be evaluated differently than
machines mounted on elevated structures or resilient vibration isolators such as
rubber pads or springs.
For example, if a fan is mounted on a rigid, solid concrete base
which, in turn, is mounted on a solid foundation located at ground level.
This would be regarded as a "rigid" installation and under normal conditions
the vertical stiffness would be greater than the horizontal stiffness. If
such is the case one would expect that normal problems, such as unbalance,
would cause higher amplitude of vibration in the horizontal direction than the
vertical direction, if a rigidly mounted machine has higher vibration in the
vertical direction than the horizontal direction, this would generally be
considered as 'abnormal', and may indicate a looseness or weakness condition.
On the other hand, if this same machine were mounted on springs or rubber
pads, a higher amplitude in. the vertical direction may not have been considered
unusual or an indication of structural problems.
Page 45
Another factor that needs to be considered is the "ratio" between
the horizontal and vertical Amplitudes. As explained, it is not unusual for
rigidly mounted machines to have higher amplitudes of vibration in the
horizontal direction, compared to the vertical direction, because of the
higher rigidity provided by the base.
4.3.7 Comparing Radial (Horizontal & Vertical) Data to Axial
Data
The second important comparison that needs to be made to tri-axial
analysis data is how the radial (horizontal and vertical) readings compare to the
axial readings. Relatively high amplitudes of axial vibration are normally the
result of:
1. Misalignment of couplings
2. Misalignment of bearings
3. Misalignment of pulleys or sheaves on belt drives
4. Bent shafts
5. Unbalance of "overhung" rotors such as the fans
A general rule, any time the amplitude of axial vibration
exceeds 50% of the highest radial (horizontal or vertical) amplitude, the
possibility of a misalignment or bent shaft condition should be considered.
4.3.8 Multiple Harmonics
The presence of multiple or "harmonically" related vibration frequencies is not
uncommon, and their presence in the FFT data can be easily explained by
examining the frequency characteristics of various vibration waveforms.
Figure below illustrates four (4) different types of vibration waveforms — a
sinusoidal is a sine wave, a square wave, triangular or "saw-tooth" wave and
a
spike
pulse. These waveforms can be readily generated by various
machinery problems, depending on the nature of the problem and the extent of
the exciting forces.
Page 46
A sinusoidal or "sine” wave could be the result, of a simple
unbalance
or misalignment problem. If a frequency analysis (FFT) is
performed on a true sinusoidal waveform, the result will be a single frequency
of vibration with certain amplitude and NO multiple frequencies.
By comparison, a frequency analysts (FFT) of sine& square waveform
will not only display the fundamental frequency (1x), but the odd multiple or
harmonic frequencies as well (i.e. 3x, 5x, 7x, etc.). A possibility is a mild
rubbing condition that might "flatten" the unbalance sine wave whenever the
rub occurs. The fundamental (1x) frequency accompanied by the odd multiple
or harmonic frequencies, similar to a square wave. However, the amplitudes of
the odd harmonics of a triangular waveform decrease more quickly at higher
frequency than do those of a square waveform as
shown
in Figure.
Triangular or saw tooth waveforms can also be generated by conditions
such as looseness or excessive bearing clearance that result in "distortion" of an
unbalance sine wave. Some problems such as a cracked or broken tooth on a
Page 47
gear, or a flaw on a bearing raceway or rolling element, will generate
vibration in the form of impact or spike-pulses. A frequency analysis or
FFT of a spike-pulse signal will reveal the fundamental impact frequency,
followed by the entire multiple or harmonic frequencies (i.e. 2x, 3x, 4x, 5x, 6x,
etc.) .
Distortion of a sinusoidal vibration waveform, resulting in
multiple vibration frequencies, may not only be the result of
mechanical problems such as looseness, bearing defects, rubbing or gear
defects as described above. (Waveform distortion can also result from the
setup and operation of the vibration analysis equipment.
For
example, if a magnetic holder is being used to mount the vibration
accelerometer to the machine, any looseness or rocking of the magnet on the
surface of the machine can result in the appearance of multiple frequencies
in the analysis data. In addition, if the amplitude of machine vibration
exceeds the full-scale amplitude range selected on theanalyzer instrument, the
true vibration signal may be "chopped",
resulting in multiple frequency
components in the FFT data that do not physically exist.
4.3.9 Side-Band Frequencies
"Side-band" frequencies are an additional vibration frequency that often
appears in the FFT data which may have less amplitude but are equally
important. Side band vibration frequencies are the result of a variation in the
amplitude of given vibration frequency signal as a function of time. This
variation in amplitude with time is also called "amplitude modulation".
For example, consider a Rolling element bearing with a significant
flaw or defect on the rotating inner raceway. As the inner raceway rotates, spike
pulses will be generated each time a rolling element impacts the flaw.
However, the amplitude or intensity of the pulses generated will vary as the
defect rotates into and out of the load zone of the bearing. This is shown in
figure below.
Page 48
Fig: Spike
pike pulses due to flaw on the inner race of the bearing
Impacts that occur when the defect is within the load zone will obviously
be more intense than those that occur out of the load zone. The
result
is a modulation the fundamental bearing defect frequency. The
fundamental bearing defect frequency in this case is the frequency at which
rolling
ing elements impact the inner raceway flaw and is called the "ball
passing frequency of the inner raceway" or simply BPFI. When
discussing side-band
band frequencies, the fundamental bearing frequency in this
case would be called the "carrier" frequency. The frequency at which the
amplitude of the carrier frequency varies is called the "modulating" frequency.
The modulating frequency in the case of a defect on the inner raceway will be 1
x RPM, since the defect is rotating into an
and
d out of the bearing load zone at the
rotating speed of the shaft.
4.3.10 Directional Nature
ature of Vibration
In addition to a comparison of tri
tri-axial
axial (horizontal, vertical and axial
data) other analysis techniques such as simple probing studies has been
discussed
cussed to show how the list of possible problems can be reduced. A
vibration frequency of 1 x RPM is probably the most common “predominant"
vibration encountered during analysis because so many different yet common
day-to-day
day problems can cause it. These problems include
1. Unbalance
Page 49
2. Bent shafts
3. Misalignment— of couplings, bearings and pulleys.
4. Looseness
5. Resonance
6. Distortion—from soft foot or piping strain conditions
7. Eccentricity---of pulleys and gears
8. Reciprocating forces
Of all the problems listed above, the only ones that generate uniform
radial forces and resultant vibration are unbalance and bent shafts. All of the
remaining problems typically generate forces and resultant vibration
which is very highly directional in nature. Therefore, determining
whether or not the radial vibration directional or non-directional can be an
extremely valuable analysis tool in reducing the list of possible problems.
To explain the difference between directional and uniform or non-directional
vibration, consider the response of a machine to a simple unbalance
problem. An unbalance condition generates a certain amount of radial force
which is governed by the amount of unbalance weight (ounces, grams, etc.), the
radius of the weight or its distance from the shaft centreline and the rotating
speed (RPM) of the machine. In any case, an unbalance generates a fixed
amount of force that is simply changing in direction with shaft rotation. If the
stiffness of the machine was the same in the horizontal and vertical directions,
the machine would literally move in a circular path, and the radial vibration
amplitudes would be the same in all radial directions. Of course, the horizontal
and vertical stiffness will probably not be exactly the same, so the radial motion
will probably be somewhat elliptical, resulting in slightly different amplitudes
measured in various radial directions. In any case, a simple unbalance,
uncomplicated by other problems, generates a fairly uniform, non-directional
radial vibration. In terms of radial vibration, a bent shaft reacts in much the
same way as simple unbalance.
Compared to unbalance and bent shafts, the other listed
causes of 1 X RPM Vibration does not generate uniform radial vibration.
Instead, they create radial vibration which is very highly directional. For
example, consider the radial vibration generated by coupling misalignment.
Page 50
When a coupling is misaligned, obviously it
is misaligned in a certain
direction. As a result, the radial forces and, hence, the radial vibration will
be most pronounced in the direction of misalignment.
Similar to coupling misalignment, a distortion
problem from a soft foot or piping strain problem creates misalignment.
There are basically three ways to determine whether the
vibration of a machine reasonably uniform or highly directional in nature.
These include:
1. A comparison or horizontal, vertical and axial FFT data
2. Comparing the horizontal and vertical phase measurements
3. Multiple radial amplitude measurements.
4.3.11 Relative Phase
If the vibration is directional then for analyzing the problem further, we
measure the relative phase difference between various points.
900 phase difference between horizontal and vertical occurs in
general for unbalance.
For an eccentric shaft the phase difference between horizontal and
vertical is either 0 (zero) or 1800
For a bent shaft the axial phase difference between two ends of the
shaft is in general 1800
Misalignment always shows 1800 phase difference across the
coupling.
For soft foot 900 to 1800 phase difference between vertical
measurements on bolts and base or base-plate is common
Rotor rub causes significant, instantaneous change of phase
Mechanical looseness usually causes phase to be unsteady.
Sometimes for journal bearing we take the orbit of the shaft and decide
the cause of the problem. The orbit analysis and subsequent fault detection
along with case study is presented in the last chapter.
Page 51
By analysing in the above way we try to pinpoint a major cause of the
vibration. Sometimes, it is easy to point one cause. But often there are more
than one interconnected problems. So we try to come to conclusion from
analysis, previous case history and experience.
Page 52
PART-5
VIBRATION ANALYSIS OF
DIFFERENT ROTATING
MACHINES
Page 53
For analyzing vibration signature and learning how to detect
fault using vibration analysis we chose 14 rotating machine with
different construction and different speed. The machines comprised of
turbo-alternator sets, pumps, fans, compressors. We have measured
the vibration using vibration measuring instrument SKF CMXA70. I
have used an accelerometer for this offline measurement. Ianalyzed
the readings using Machine Analyst software. The FFT and Timewaveform of each machine and their analysis follows.
Page 54
5.1 BOILER FEE
FEED PUMP
5.1.1 BOILER FEED PUMP OF BBGS UNIT 3A
SPECIFICATION:
RATING: 10.6 MW
PUMP RPM: 4950
MOTOR RPM: 1493
NATURE OF COUPLING
COUPLING: VOITH
RATED VOLTAGE & VARIATION: 11KV, ± 10%
BEARING: JOURNAL BEARING
TYPE OF LUBRICATION: FORCED
TYPE OF LUBRICANT:: ISO VG 32
As the pump is horizontal with rated power with more than 300
30 KW but less
than 50 MW with flexible drive, it falls under Part
Part-3
3 of ISO 10816. As per ISO
10816 Part 3 for a machine having continuous running condition the alarm
setting is 7.1 mm/s and th
the advice trip setting is 11 mm/s.
Page 55
BOILER FEED PUMP 3A
MACHINE POSITION DIRECTION
PUMP
AXIAL
VALUE(OVERALL)(mm/s)
(RMS)
1.188
HORIZONTAL
1.715
VERTICAL
AXIAL
4.886
1.274
HORIZONTAL
2.972
VERTICAL
2.777
DRIVE
END
AXIAL
HORIZONTAL
VERTICAL
0.7472
0.9204
0.487
NON
DRIVE
END
AXIAL
HORIZONTAL
2.077
1.466
VERTICAL
0.8652
DRIVE
END
AXIAL
HORIZONTAL
VERTICAL
0.4534
1.011
0.87
NON
DRIVE
END
AXIAL
HORIZONTAL
0.5111
1.327
VERTICAL
2.298
NON
DRIVE
END
DRIVE
END
MOTOR
BOOSTER
PUMP
From the chart above, for BF pump 3A, the pump shows high vibration at non
drive end vertical direction.
Page 56
Fig: FFT of Pump Non drive end vertical
Fig: Time-wave of Pump Non drive end vertical
From the FFT, we can see high amplitude of 4.46 mm/s at 1x of the rotating
speed. The peak is only at 1x, there is no other significant peak.
Page 57
Fig: Pump drive end axial velocity
Fig: Pump drive end horizontal velocity
Fig: Pump drive end vertical velocity
Page 58
From the above FFTs of the drive end of the pump also show significant
amplitude at 1x. And the peak for axial direction is well under the alarm
settings. So we are concerned with only vertical and horizontal direction.
Analysis:
The FFT shows high peaks only at 1x.
To pin point the cause of the vibration the vibration reading at different
positions of the non-drive end were taken.
Also if there was a bent shaft the vibration amplitude would have been
equal in both horizontal and vertical direction. So the shaft is not bent.
We can say that there may be a slight misalignment in the vertical
direction but is well below alarm condition.
CAUSE
1. Unbalance
2. Eccentric shaft
3. Misalignment
4. Bent shaft
5. Looseness
POSSIBILITY REMARKS
NO
Huge difference between horizontal
and vertical vibration amplitude
NO
Phase difference between vertical
and horizontal is 900
YES
Only high vertical vibration at 1x,
no other significant peak,
NO
Huge difference between horizontal
and vertical vibration amplitude
NO
Little difference in values above
and below parting plane
5.1.2 BOILER FEED PUMP OF BBGS UNIT 3C
SPECIFICATION:
Rating: 10.6 MW
PUMP RPM: 4880
MOTOR RPM: 1493
NATURE OF COUPLING: VOITH
Page 59
RATED VOLTAGE & VARIATION: 11KV, ± 10%
RATED FREQUENCY & VARIATION: 50 Hz, ± 10%
BEARING: JOURNAL BEARING
TYPE OF LUBRICATION: FORCED
TYPE OF LUBRICANT: ISO VG 32
As the pump is horizontal with rated power with more than 300 KW but less
than 50 MW with flexible drive, it falls under Part-3 of ISO 10816. As per ISO
10816 Part 3 for a machine having continuous running condition the alarm
setting is 7.1 mm/s and the advice trip setting is 11 mm/s.
BOILER FEED PUMP 3C
MACHINE
POSITION
DIRECTION
VALUE(OVERALL)(mm/s)
PUMP
NON DRIVE AXIAL
1.407
END
HORIZONTAL 4.854
DRIVE END
VERTICAL
AXIAL
3.606
1.465
HORIZONTAL 4.036
VERTICAL
MOTOR
DRIVE END
3.009
AXIAL
0.7531
HORIZONTAL 1.336
VERTICAL
0.8573
NON DRIVE AXIAL
4.16
END
HORIZONTAL 1.15
VERTICAL
BOOSTER
PUMP
DRIVE END
0.7796
AXIAL
0.679
HORIZONTAL 1.516
VERTICAL
1.411
NON DRIVE AXIAL
1.253
END
HORIZONTAL 1.04
VERTICAL
3.15
Page 60
Here we have high vibration at the pump at both drive and non drive end. Also
there is high vibration at non drive end of the motor in the axial direction. All
these amplitudes are below the alarm level as specified by ISO 10816part-3 but
high with respect tosimilar machines.
Here are the FFTs of the zones highlighted above.
Fig: Pump non drive end horizontal
Fig: Pump non drive end vertical
Page 61
Fig: Pump drive end horizontal
Fig: Pump drive end vertical
From the above FFTs it is clear that the there is high vibration at 1x
Fig: Motor non drive end axial
There is considerable vibration at the non drive end of the motor.
Page 62
The high vibration in the vertical and horizontal direction at both drive and
non drive end of the pump is predominantly in 1x. Applying same vibration
analysis techniques as above, we rule out eccentric shaft.
There may be slight unbalance in the rotor but well below the limit.
CAUSE
1. Unbalance
2. Eccentric shaft
3. Misalignment
4. Bent shaft
5. Looseness
POSSIBILITY REMARKS
YES
Comparable vibration amplitude in
vertical and horizontal direction
with about 900 phase difference
NO
Phase difference between vertical
and horizontal is 900
NO
Huge difference in values at the
bearings nearer to the coupling
NO
No high amplitude at 2x near the
coupling
NO
Little difference in values above
and below parting plane
The motor non-drive end shows high vibration in the axial direction. The
vibration at 2x here is about 90% of the overall value. From the analysis we can
say the coupling between the motor non-drive end and booster pump may have
minor misalignment in the axial direction.
Page 63
5.2 SEAL AIR FAN
Fig: Schematic
chematic drawing of seal air fan 1A, 1B, 2A, 2B
5.2.1 SEAL AIR FAN 1A
SPECIFICATION:
Rating: 83.5 KW
PUMP RPM: 3000
MOTOR RPM: 3000
NATURE OF COUPLING: FLEXIBLE
BEARING: TAPER ROLLER BEARING DIA 65 mm
TYPE OF LUBRICANT: ISO VG 32 (AS PER DRAWING)
As the fan is horizontal with rated power with more than 15 KW but less than
300 KW with flexible drive, it falls under Part
Part-3
3 of ISO 10816. As per ISO
10816 Part 3 for a machine having continuous running condition the alarm
setting is 4.5
.5 mm/s and the ad
advice trip setting is 7.1 mm/s.
Page 64
SEAL AIR FAN 1A
MACHINE POSITION
FAN
NON
END
DIRECTION
DRIVE AXIAL
VALUE(OVERALL)(mm/s)
(RMS)
3.133
HORIZONTAL 3.081
DRIVE END
VERTICAL
AXIAL
0.7036
2.201
HORIZONTAL 2.917
MOTOR
DRIVE END
NON
END
VERTICAL
AXIAL
HORIZONTAL
VERTICAL
3.764
0.8299
1.484
1.059
DRIVE AXIAL
0.6975
HORIZONTAL 1.194
VERTICAL
1.266
There is high vibration in the vertical direction of the drive end of the fan, in
axial and horizontal direction at the non-drive end and in the horizontal
direction at the drive end, but all the amplitudes are well below the
recommended alarm settings as mentioned earlier.
Fig: Fan drive end vertical velocity
Page 65
Fig: Fan drive end horizontal velocity
Fig: Fan drive end axial velocity
From the FFTS of the horizontal and axial direction at the drive end of the fan,
there are significant peaks at 2x. The vibration at 2x is about 60% of the overall
vibration amplitude.Analyzing the FFTs it can be said that the shaft may have
radial misalignment. There is a high amplitude in axial direction also, which
may be due to slight axial misalignment but that is within limit.
Fig: Fan non-drive end axial
Page 66
Fig: Fan non-drive end horizontal
ANALYSIS
From the above two FFTs at the non-drive end, there are peaks at 4x and 5x.
Taking 5x as the main affecting frequency and others as the modulation of 5x
and 1x. The vibration at 5xmay be its characteristics vibration as it is not
associated with ball pass frequency, misalignment or any other causes.The
amplitude of vibration is well below alarm value.
5.2.2 SEAL AIR FAN 1B
SPECIFICATION:
Rating: 83.5 KW
PUMP RPM: 3000
MOTOR RPM: 3000
NATURE OF COUPLING: FLEXIBLE
BEARING: TAPER ROLLER BEARING DIA 65 mm
TYPE OF LUBRICANT: ISO VG 32 (AS PER DRAWING)
As the fan is horizontal with rated power with more than 15 KW but less than
300 KW with flexible drive, it falls under Part-3 of ISO 10816. As per ISO
10816 Part 3 for a machine having continuous running condition the alarm
setting is 4.5 mm/s and the advice trip setting is 7.1 mm/s.
Page 67
SEAL AIR FAN 1B
MACHINE
POSITION
FAN
DIRECTION
VALUE(OVERALL)(mm/s)
(RMS)
1.569
NON DRIVE AXIAL
END
HORIZONTAL 2.222
DRIVE END
VERTICAL
AXIAL
2.341
1.905
HORIZONTAL 3.759
MOTOR
DRIVE END
VERTICAL
AXIAL
HORIZONTAL
VERTICAL
2.233
0.9471
1.57
2.584
NON DRIVE AXIAL
0.9568
END
HORIZONTAL 2.128
VERTICAL
1.22
The vibration values are well within limits. But there is a little high vibration in
the horizontal direction at the drive end of the fan. The FFT for the same is
given below
Fig: Fan drive end horizontal velocity
From the FFT it is clear that there is peaks at all harmonics and the maximum
amplitude is a 15000 cpm or at 5x. The individual peaks are not that high and
they contribute equally for a high overall value.
Page 68
This 5x may be considered characteristics as characteristics frequency of these
fans. All the other values are significantly low. The fan is in good condition
overall.
5.2.3 SEAL AIR FAN 2A
SPECIFICATION
Rating: 83.5 KW
PUMP RPM: 3000
MOTOR RPM: 3000
NATURE OF COUPLING: FLEXIBLE
BEARING: TAPER ROLLER BEARING DIA 65 mm
TYPE OF LUBRICANT: ISO VG 32 (AS PER DRAWING)
As the fan is horizontal with rated power with more than 15 KW but less than
300 KW with flexible drive, it falls under Part-3 of ISO 10816. As per ISO
10816 Part 3 for a machine having continuous running condition the alarm
setting is 4.5 mm/s and the advice trip setting is 7.1 mm/s.
SEAL AIR FAN 2A
MACHINE
POSITION
FAN
NON
END
DRIVE AXIAL
DRIVE END
MOTOR
DRIVE END
NON
END
DIRECTION
VALUE(OVERALL)(mm/s)(RMS)
0.8077
HORIZONTAL
1.114
VERTICAL
2.142
AXIAL
1.028
HORIZONTAL
1.316
VERTICAL
AXIAL
1.004
0.7524
HORIZONTAL
1.212
VERTICAL
1.899
DRIVE AXIAL
0.7293
HORIZONTAL
1.205
VERTICAL
0.9555
Page 69
Thefan shows low vibration amplitudes in all the directions at all the positions.
The overall values are well within the limits specified by ISO 10816. The fan is
running in good condition without any obvious fault.
5.2.4 SEAL AIR FAN 2B
SPECIFICATION:
Rating: 83.5 KW
PUMP RPM: 3000
MOTOR RPM: 3000
NATURE OF COUPLING: FLEXIBLE
BEARING: TAPER ROLLER BEARING DIA 65 mm
TYPE OF LUBRICANT: GREASE
As the fan is horizontal with rated power with more than 15 KW but less than
300 KW with flexible drive, it falls under Part-3 of ISO 10816. As per ISO
10816 Part 3 for a machine having continuous running condition the alarm
setting is 4.5 mm/s and the advice trip setting is 7.1 mm/s.
Page 70
SEAL AIR FAN 2B
MACHINE
POSITION
FAN
DIRECTION
VALUE(OVERALL)(mm/s)
(RMS)
2.069
NON DRIVE AXIAL
END
HORIZONTAL 2.118
DRIVE END
VERTICAL
AXIAL
1.06
3.264
HORIZONTAL 2.638
VERTICAL
MOTOR
DRIVE END
1.318
AXIAL
1.336
HORIZONTAL 2.297
VERTICAL
1.877
NON DRIVE AXIAL
1.591
END
HORIZONTAL 2.613
VERTICAL
1.804
The fan shows low vibration amplitudes in all the directions at all the positions.
The overall values are well within the limits specified by ISO 10816. The fan is
running in good condition.
Page 71
5.3 DMCW PUMP
Fig: DMCW PUMP
5.3.1 DMCW PUMP 3A
SPECIFICATION:
Rating: 125 KW
PUMP RPM: 1485
MOTOR RPM: 1485
NATURE OF COUPLING: FLEXIBLE
BEARING: BALL BEARING
TYPE OF LUBRICANT: GREASE
As the pump is horizontal with rated power with more than 15 KW but less than
300 KW with flexible drive, it falls under Part-3 of ISO 10816. As per ISO
10816 Part 3 for a machine having continuous running condition the alarm
setting is 4.5 mm/s and the advice trip setting is 7.1 mm/s.
Page 72
DMCW PUMP 3A
MACHINE POSITION
PUMP
NON
END
DIRECTION
DRIVE AXIAL
VALUE(OVERALL)(mm/s)
(RMS)
2.26
HORIZONTAL 2.241
DRIVE END
VERTICAL
AXIAL
3.717
2.032
HORIZONTAL 3.515
MOTOR
DRIVE END
NON
END
VERTICAL
AXIAL
HORIZONTAL
VERTICAL
5.182
1.275
1.198
1.776
DRIVE AXIAL
1.582
HORIZONTAL 0.5576
VERTICAL
0.4923
Page 73
Fig: Pump non-drive end vertical
Fig: Pump drive end vertical
There is vibration in the alarm level in the horizontal direction at the drive end.
The vertical vibration at the non drive end is high but below the alarm zone. All
the other vibrations are within limit.
From the FFTs above we can see that there is high vibration in 6x and apart
from that there is no other significant peak.
Page 74
5.3.2 DMCW PUMP 3B
SPECIFICATION:
Rating: 125 KW
PUMP RPM: 1485
MOTOR RPM: 1485
NATURE OF COUPLING: FLEXIBLE
BEARING: BALL BEARING
TYPE OF LUBRICANT:
As the pump is horizontal with rated power with more than 15 KW but less than
300 KW with flexible drive, it falls under Part-3 of ISO 10816. As per ISO
10816 Part 3 for a machine having continuous running condition the alarm
setting is 4.5 mm/s and the advice trip setting is 7.1 mm/s.
DMCW PUMP 3B
MACHIN POSITION
DIRECTION
E
PUMP
NON
DRIVE AXIAL
END
HORIZONTAL
DRIVE END
VERTICAL
AXIAL
VALUE(OVERALL)(mm/s)
(RMS)
1.568
1.557
1.542
1.385
HORIZONTAL 2.24
VERTICAL
MOTOR
DRIVE END
NON
END
2.637
AXIAL
1.255
HORIZONTAL 1.322
VERTICAL
0.4139
DRIVE AXIAL
1.271
HORIZONTAL 0.718
VERTICAL
0.5404
The pump shows low vibration amplitudes in all the directions at all the
positions. The overall values are well within the limits specified by ISO 10816.
The pump is running in good condition without any obvious fault.
Page 75
5.3.3 DMCW PUMP 3C
SPECIFICATION:
Rating: 125 KW
PUMP RPM: 1485
MOTOR RPM: 1485
NATURE OF COUPLING: FLEXIBLE
BEARING: BALL BEARING
TYPE OF LUBRICANT:
As the pump is horizontal with rated power with more than 15 KW but less than
300 KW with flexible drive, it falls under Part-3 of ISO 10816. As per ISO
10816 Part 3 for a machine having continuous running condition the alarm
setting is 4.5 mm/s and the advice trip setting is 7.1 mm/s.
DMCW PUMP 3C
MACHIN POSITION
DIRECTION
E
PUMP
NON
DRIVE AXIAL
END
HORIZONTAL
DRIVE END
VERTICAL
AXIAL
VALUE(OVERALL)(mm/s)
(RMS)
0.9582
1.32
1.465
0.7014
HORIZONTAL 1.498
VERTICAL
MOTOR
DRIVE END
NON
END
2.1497
AXIAL
0.6559
HORIZONTAL 0.4204
VERTICAL
0.143
DRIVE AXIAL
0.7465
HORIZONTAL 0.2396
VERTICAL
0.1658
The pump shows low vibration amplitudes in all the directions at all the
positions. The overall values are well within the limits specified by ISO 10816.
The pump is running in good condition without any obvious fault.
Page 76
5.4 TURBINE
5.4.1 BBGS TURBINE UNIT 1
SPECIFICATION:
Maker: PARSONS
Rating: 250 MW
Rated RPM:3000
NATURE OF COUPLING: RIGID
BEARING: JOURNAL BEARING
TYPE OF LUBRICANT: ISO VG 46
Though there are ISO standards for the alarm and advice trip settings for the
turbine, we generally follow the maker’s instruction. For this turbine the alarm
settings for bearing 1-4 is 5.5 mm/s and the advice trip is 8.5 mm/s at the
pedestal. The alarm for shaft vibration is at 150 micron, and advice trip is at 175
micron. For the bearings 4-8, the alarm is at 8.5 mm/s (pedestal) and 175
micron (shaft) and advice trip is at 11 mm/s (pedestal) and 200 micron (shaft).
Pedestal 1
Pedestal 2
ADSC
6.41
AVEL
0.53
HDSC
17.216
HVEL
0.939
VDSC
10.102
VVEL
0.832
ADSC
6.191
AVEL
0.594
HDSC
12.683
Page 77
Pedestal 3
Pedestal 4
Pedestal 5
Pedestal 6
HVEL
1.179
VDSC
5.266
VVEL
0.479
ADSC
4.854
AVEL
0.295
HDSC
12.062
HVEL
1.222
VDSC
4.119
VVEL
0.362
ADSC
29.046
AVEL
3.347
HDSC
6.636
HVEL
0.674
VDSC
13.562
VVEL
1.437
ADSC
27.509
AVEL
3.165
HDSC
9.056
HVEL
0.895
VDSC
35.736
VVEL
3.804
ADSC
15.651
AVEL
1.931
HDSC
20.164
HVEL
2.364
VDSC
27.027
Page 78
Pedestal 7
Pedestal 8
VVEL
3.152
ADSC
14.395
AVEL
2.129
HDSC
26.542
HVEL
3.42
VDSC
36.959
VVEL
4.382
ADSC
177.335
AVEL
19.63
HDSC
25.683
HVEL
4.555
VDSC
40.747
VVEL
4.7
All the values are within limit. But some of these have slightly higher values,
the FFTs of those are given below
Fig: Bearing 5 vertical velocity
Page 79
Fig: Bearing 6 vertical velocity
Fig: Bearing 7 horizontal velocity
Fig: Bearing 7 vertical velocity
Page 80
Fig: Bearing 8 horizontal velocity
Fig: Bearing 8 vertical velocity
Fig: Bearing 8 axial velocity
From the FFTs above there are clear peaks at 1x and 2x.
Page 81
Analysis:
By analyzing the FFTs, it can be said that slight misalignmentmay be
causing this vibration. Because the amplitude of vibration is low in the
horizontal direction, the unbalance is not considered.
Analyzing all the possibilities, I came to the conclusion that the high
vibration may be due to minor misalignment.
The horizontal and vertical vibration may be due to lack of rigidity.
But in case of the high axial vibration in bearing 8, there is high
amplitude only at 1x. After analyzing all the possible causes, I came to the
conclusion that the high vibration may be due to the fact that the pedestal is not
grouted here. It is removable bolted type, so we observe lesser damping by the
bearing and higher vibration than a similar type of machine with a grouted
pedestal.
5.4.2 BBGS TURBINE UNIT 2
SPECIFICATION:
Maker: PARSONS
Rating: 250 MW
Rated RPM:3000
NATURE OF COUPLING: RIGID
BEARING: JOURNAL BEARING
TYPE OF LUBRICANT: ISO VG 46
Page 82
Though there are ISO standards for the alarm and advice trip settings for the
turbine, we generally follow the maker’s instruction. For this turbine the alarm
settings for bearing 1-4 is 5.5 mm/s and the advice trip is 8.5 mm/s at the
pedestal. The alarm for shaft vibration is at 150 micron, and advice trip is at
175 micron. For the bearings 4-8, the alarm is at 8.5 mm/s (pedestal) and 175
micron (shaft) and advice trip is at 11 mm/s (pedestal) and 200 micron (shaft).
Pedestal 1
Pedestal 2
Pedestal 3
ADSC
29.054
AVEL
2.44
HDSC
16.102
HVEL
1.371
VDSC
4.832
VVEL
0.437
ADSC
4.966
AVEL
0.49
HDSC
9.749
HVEL
0.844
VDSC
6.059
VVEL
0.642
ADSC
4.357
Page 83
Pedestal 4
Pedestal 5
Pedestal 6
Pedestal 7
AVEL
0.388
HDSC
10.184
HVEL
1.087
VDSC
4.104
VVEL
0.464
ADSC
50.249
AVEL
5.537
HDSC
10.477
HVEL
1.037
VDSC
9.383
VVEL
1.049
ADSC
50.401
AVEL
5.559
HDSC
17.62
HVEL
2.162
VDSC
30.815
VVEL
3.507
ADSC
8.337
AVEL
0.971
HDSC
12.143
HVEL
1.622
VDSC
13.468
VVEL
1.676
ADSC
7.893
Page 84
Pedestal 8
AVEL
1.075
HDSC
12.571
HVEL
1.832
VDSC
17.26
VVEL
2.324
ADSC
97.613
AVEL
10.76
HDSC
20.75
HVEL
3.62
VDSC
25.719
VVEL
3.172
Apart from the high axial vibration at bearing 4 and 8, the vibration at all other
directions and positions are well within alarm limit. The FFTs are below
Fig: Bearing 4 axial
Page 85
Fig: Bearing 8 axial
Both have high peaks at 1x, and no other significant peak. All the others are in
good condition, and there is no obvious fault in the machinery.
5.4.3 BBGS TURBINE UNIT 3
SPECIFICATION:
Maker: BHEL
Rating: 250 MW
Rated RPM:3000
NATURE OF COUPLING: RIGID
BEARING: JOURNAL BEARING
TYPE OF LUBRICANT: ISO VG 46
Though there are ISO standards for the alarm and advice trip settings for the
turbine, we generally follow the maker’s instruction. For this turbine the alarm
settings is normal running condition plus 50 micron but not more than 200
micron and advice trip is 320 micron. At pedestal those are 84 and 106 micron
respectively.
Page 86
Pedestal 1
Shaft 1
Pedestal 2
Shaft 2
Pedestal 3
ADSC
4.929
AVEL
0.724
HDSC
3.973
HVEL
0.657
VDSC
8.977
VVEL
2.15
Y Overall
49.32
Y 1X
28.39
Phase
189
X Overall
76.94
X 1X
56.67
Phase
34
ADSC
4.877
AVEL
0.515
HDSC
5.564
HVEL
0.612
VDSC
2.887
VVEL
0.41
Y Overall
47.652
Y 1X
29.17
Phase
290
X Overall
96.99
X 1X
76.56
Phase
262
ADSC
9.043
AVEL
0.966
HDSC
4.308
HVEL
0.456
Page 87
Shaft 3
Pedestal 4
Shaft 4
Pedestal 5
Shaft 5
VDSC
34.57
VVEL
3.669
Y Overall
21.896
Y 1X
10
Phase
200
X Overall
47.92
X 1X
39.362
Phase
322
ADSC
61.96
AVEL
6.737
HDSC
16.619
HVEL
1.835
VDSC
13.609
VVEL
1.841
Y Overall
39.95
Y 1X
30.35
Phase
20
X Overall
31.045
X 1X
21.06
Phase
216
ADSC
19.87
AVEL
2.573
HDSC
3.842
HVEL
0.348
VDSC
27.285
VVEL
2.942
Y Overall
60.183
Y 1X
47.87
Phase
320
Page 88
Pedestal 6
Shaft 6
X Overall
53.098
X 1X
40.29
Phase
191
ADSC
9.029
AVEL
1.281
HDSC
6.475
HVEL
0.828
VDSC
32.02
VVEL
3.522
Y Overall
60.389
Y 1X
48.02
Phase
321
X Overall
54.815
X 1X
28.94
Phase
155
This turbine has high shaft vibration at bearing 2. That is analyzed later. Apart
from that there is high axial vibration at bearing 4
Fig: Bearing 4 axial velocity
Page 89
Analysis
From the FFT above, it is clear that there is high vibration at 1x in the axial
direction. After analyzing all the possible causes, I came to the conclusion that
it may be due to minor axial misalignment.
5.4.4 SGS TURBINE UNIT 1
SPECIFICATION:
Maker: BHEL
Rating: 67.5 MW
Rated RPM: 3000
NATURE OF COUPLING: RIGID
BEARING: JOURNAL BEARING
TYPE OF LUBRICANT: ISO VG 46
Though there are ISO standards for the alarm and advice trip settings for the
turbine, we generally follow the maker’s instruction. For this turbine the alarm
settings for bearing pedestal is 50 micron and the advice trip is 100 micron. The
alarm for shaft vibration is at 150 micron, and advice trip is at 175 micron.
Page 90
Pedestal 1
Pedestal 2
Pedestal 3
Pedestal 4
ADSC
5.596
AVEL
0.611
HDSC
28.195
HVEL
3.676
VDSC
10.186
VVEL
1.247
ADSC
17.525
AVEL
1.959
HDSC
10.274
HVEL
1.322
VDSC
8.076
VVEL
0.888
ADSC
24.327
AVEL
3.487
HDSC
23.724
HVEL
2.573
VDSC
30.656
VVEL
3.435
ADSC
43.855
AVEL
4.858
HDSC
10.853
HVEL
1.949
VDSC
12.101
VVEL
1.567
Page 91
There is high vibration in the vertical direction of bearing 3.
Fig: Bearing 3 vertical velocity
From the FFT above there is high vibration amplitude at 1x. After considering
all the possible causes and analyzing them, I came to the conclusion that this
may be due to minor unbalance, as the amplitude of vibration at 1x in the
vertical direction is near that in the horizontal direction.
5.4.5 SGS TURBINE UNIT 2
SPECIFICATION:
Maker: BHEL
Rating: 67.5 MW
Rated RPM: 3000
NATURE OF COUPLING: RIGID
BEARING: JOURNAL BEARING
TYPE OF LUBRICANT: ISO VG 46
Though there are ISO standards for the alarm and advice trip settings for the
turbine, we generally follow the maker’s instruction. For this turbine the alarm
settings for bearing pedestal is 50 micron and the advice trip is 100 micron. The
alarm for shaft vibration is at 150 micron, and advice trip is at 175 micron.
Page 92
Pedestal 1
Pedestal 2
Pedestal 3
Pedestal 4
ADSC
5.741
AVEL
0.649
HDSC
9.041
HVEL
1.276
VDSC
7.416
VVEL
0.764
ADSC
45.77
AVEL
5.076
HDSC
24.079
HVEL
2.717
VDSC
24.038
VVEL
2.718
ADSC
24.361
AVEL
3.172
HDSC
5.882
HVEL
1.223
VDSC
20.478
VVEL
2.3
ADSC
22
AVEL
2.507
HDSC
6.767
HVEL
0.898
VDSC
10.197
VVEL
1.191
Page 93
There is high vibration in the vertical direction of bearing 3.
Fig: Bearing 2 vertical velocity
Fig:Bearing 2 horizontal velocity
From the FFT above there is high vibration amplitude at 1x. After considering
all the possible causes and analyzing them, I came to the conclusion that this
may be due to minor unbalance, as the amplitude of vibration at 1x in the
vertical direction is near that in the horizontal direction.
Page 94
PART-6
BALANCING
Page 95
After learning about vibration analysis and analyzing rotating machineries, Seal
Air Fan 2A was analyzed. The report of the proceedings of the analysis, fault
detection and balancing is written below.
6.1 FAN specification:
SEAL AIR FAN 2A
Make: BHEL (Modified)
Type: NDFC 9.5
Weight of rotating parts: 350 Kgs (approximately, inclusive of shaft,
coupling flange and impeller)
Rated RPM: 3000
Balancing radius : 500mm (approx.)
Equipment used for taking reading: SKF make CMXA 70 Vibration
analyzer with LASER tachometer.
Schematic:
6.2 Prelude:
The machine was running with a vibration of 13.4 mm/s (rms, vertical direction)
at its non-drive end bearing. The machine was last balanced in 2007 where the
final balanced vibration in the vertical direction of the said bearing was 6.98
mm/s.
Page 96
During the visit of SKF personnel to the Southern Generating Station site for
collecting monthly readings of various equipments, they identified Seal Air Fan
2A for having high vibration.
6.3 Problem:
The first reading taken on the non-drive end bearing by the vibration analyzer
showed an overall vibration of 14.57 mm/s.
6.4 Details of the reading:
Time waveform
FFT
As evident from the FFT that the 1X component forms 93% of the overall
vibration reading, so the 1X vibration is said to be dominant.
It is also evident from the FFT that there are insignificant higher order
harmonics, so other causes for such high vibration such as misalignment, bent
shaft, looseness in the system or problems with the bearing have been ruled out.
Page 97
So it was decided that this problem could be rectified by balancing the rotating
impeller.
6.5 Balancing table:
(All vibration values are in rms and all angles are measured in clock wise
direction viewed from motor end.)
Initial Run
Vertical vibration (in rms at 1X):
Phase (with theoretical axis)
13.4 mm/s
117º
Trial weight (1) added
Angle (with real axis)
55 gm
207º
Trial Run 1
Vertical vibration (in rms at 1X):
Phase (with theoretical axis)
Vector diagram after Trial Run 1
16.5 mm/s
157°
Page 98
INDEX
U
WT
U+WT
Vibration
due to initial
unbalance
Vibration
due to only
trial weight
Vibration
due to both
trial weight
and
initial
unbalance
So as per
the above Vector diagram, the
trail weight was added at 211º from the theoretical axis of reference and the
vector (WT) was 10.4 mm/s.
But the actual trial weight was added at 208º from the mark on the shaft
(real axis of reference) against the direction of rotation.
So there was deviation of 3º between the Theoretical axis of rotation and
Real axis of rotation.
So the balance plane with respect to real axis of rotation is at 294º.
Trial weight (1) was removed and
correction weight (1) was added
Correction weight (1) added
45 gm
Angle (with real axis)
297º
Trial Run 2
Vertical vibration (in rms at 1X):
Phase (with theoretical axis)
9.95 mm/s
147°
Page 99
Vector diagram after Trial Run 2
Index
U
WC1
U+WC
1
Vibration due to
initial unbalance
Vibration due to only
correction weight
Vibration due to both
correction weight and
initial unbalance
So as per the above Vector diagram, the correction weight (1) was
added at 250º from the theoretical axis of reference and the vector
(WT) was 6.88 mm/s.
But the actual correction weight (1) was added at 297º from the mark
on the shaft (real axis of reference) against the direction of rotation.
So there was deviation of 47º between the Theoretical axis of rotation
and Real axis of rotation.
Page 100
So the balance plane of resultant vector (U+WC1) with respect to real
axis of rotation is at 14º.
Correction weight (1) was kept in
place and correction weight (2) was
added
Correction weight (2) added
20 gm
Angle (with real axis)
335º
Trial Run 3
Vertical vibration (in rms at 1X):
Phase (with theoretical axis)
6.32 mm/s
162º
Vector diagram after Trial Run 3
Index
U1
WC2
U1+W
C1
Vibration due to initial
unbalance
and
correction weight (1)
Vibration due to only
correction weight (2)
Vibration due to both
U1
and
correction
weight (2)
Page 101
So as per the above Vector diagram, the correction weight (2) was added at
304º from the theoretical axis of reference and the vector (WT) was 3.99
mm/s.
But the actual correction weight (2) was added at 335º from the mark on the
shaft (real axis of reference) against the direction of rotation.
So there was deviation of 31º between the Theoretical axis of rotation and
Real axis of rotation.
So the balance plane of resultant vector (U+WC1+WC2) with respect to
real axis of rotation is at 13º.
Correction weights (1&2) were kept in
place and correction weight (3) was
added
Correction weight (3) added
Angle (with real axis)
Trial Run 4
Vertical vibration (in rms at 1X):
Phase (with theoretical axis)
20 gm
15º
5.61 mm/s
52º
Vector diagram after Trial Run 4
Page 102
Index
U2
Vibration
due
to
initial unbalance and
correction
weights(1+2)
WC3
Vibration due to only
correction weight (3)
U2+WC3 Vibration due to both
U2 and correction
weight (3)
So after trial run 4 it was seen that the vibration because of unbalance has died
down to 5.61 mm/s (rms) and the phase has shifted by 110º. As the vibration
level was within the accepted level of 2007 and the phase shift has taken place
without much change in the vibration level, it was decided to leave the machine
at this level of vibration.
The overall vibration was 6.38 mm /s (rms) at the non-DE bearing (vertical).
Page 103
PART-7
ANALYSIS OF JOURNAL BEARING 2
OF BBGS UNIT 3
Page 104
I have also analyzed a special problem of high vibration in the journal
bearing 2 of BBGS unit 3. The analysis comprises of general FFT and time
waveform along with the orbit analysis. The following part gives a detail
discussion about the journal bearing, the problem, analysis and possible
solution.
7.1 DETAILS OF JOURNAL BEARING 2 OF BBGS UNIT 3
The journal bearing 2 of unit 3 of BBGS is a combined radial journal and
thrust bearing. The lubrication oil gets into the bearing from the x side
horizontal through a nozzle, which controls the amount of fluid flow. The
bearing is self aligning spherical journal bearing. There is no allowance for
swiveling about vertical or horizontal axis. The bearing has the dimension of
315mm X 350 mm. Inside of the bearing is white metal lining. It is two lobe
bearing. The bearing shell is made of two lobes. The lower part of the bearing is
bolted to the torus, which is seated on the pedestal.
Fig: BEARING 2 OF BBGS UNIT 3
Page 105
7.2 PROBLEM
The bearing stated here is showing high shaft vibration. The overall
vibration is 96.99 in X direction. The convention for X & Y direction can be
explained by the following figure.
Standing in front of the HP turbine and looking at it, the left hand side
denotes X and the right hand side denotes Y. both are at 45 degrees above the
horizontal.
7.3 HISTORY
The history shows the increase is not sudden. There is gradual increase in
the amplitude in X direction, where as in the Y direction it has been relatively
constant over the same period of time.
Fig: History of 2X
Page 106
Fig: History of 2Y
Fig: History of 3Y
From the history it is also clear that the vibration at bearing 2 has been
significantly higher for a long time.
7.4 ANALYSIS
The FFT is given below.
Page 107
Fig: 2X shaft vibration displacement peak to peak
From the FFT it is clear that the vibration at 1x is 76.56 which is about
80% of the total vibration. So we have identified 1x as the predominant
vibration frequency. But 1x vibration can be due to a lot of reasons. One of the
causes is high bearing clearance. As the vibration in Y direction is constant over
the time, I assumed that the bearing clearance has increased in the X direction.I
took the phase of the vibrations and those are given below
Fig: 3x phase
Fig: 7X phase
Page 108
These phases do not say much about the cause of the vibration. But from
the history we can see that there is gradual shift in phase.
For pin pointing the cause we took orbit of the shaft vibration of bearing 2,
3, 6& 7.
7.4.1 THE THEORY OF ORBIT ANALYSIS
Orbit is the trajectory of the motion of the shaft. The shaft motion is
measured by proximity probe. When non-contacting eddy current probes and
proximitors are used to monitor lateral shaft motion, this transducer system
provides the following individual signal components:
1. A DC (Direct Current) signal whichmonitors the shaft average position
relative to the probe mounting.
2. An AC signal (in this case, negativelyfluctuating) which monitors shaft
dynamicmotion relative to the probe mounting.
Page 109
The AC signal is the measurement of vibration. The AC component of a
transducer’s signal produces a periodic waveform from each probe in the
orthogonal probe pair. A typical output waveform is shown in Figure below.
Fig: the upper waveform is the synchronous (1x) vibration and the lower
waveform is the direct unfiltered vibration.
Page 110
Fig: Orbit and the time waveform
Let the two waveforms above be denoting the vibration in two mutually
perpendicular axes. Both of them can be expressed as the function of time. Now
if we eliminate the time parameter from both the waves we get one vibration as
the function of another. The plot of this function is the orbit. Therefore, the orbit
simply represents the PATH of the shaft centerline at the lateral position of the
proximity probe.
The orbit analysis is an important tool with machine with dynamic fluid film
bearing. With the following examples, it will be clear how it helps in deciding
the source problem for high 1x vibration.
For a machine running continuously for a period of time there will be a
vibration. But that should be within the specified limit. In case of journal
bearing there is a high pressure zone (fig). The load is acting here. The stiffness
in this direction is high. Therefore, in general there will be higher vibration
perpendicular this direction.
Fig: Journal bearing
Page 111
Fig: Acceptable condition
But whenever there is some kind of problem the direction of orbit, the
amplitude or both changes.
Like when there is high loading, the shaft moves nearer to the journal bearing,
which can be seen by the following figure.
For whirl the orbit will be like
Page 112
The two dots signify the vibration frequency to be ½ X. When there is rotor rub
the fig will look something like this
If the orbit becomes more circular there is a problem in the bearing. There can
be a lot of other problems like starvation, high clearance etc. All of these
problems can only be diagnosed by orbit analysis.
7.4.2 ORBIT ANALYSIS OF BEARING # 2 OF BBGS # 3
The orbits we got for the shafts here are given below
Page 113
Fig: Shaft orbit at bearing 6
Fig: Orbit of shaft at bearing 3
Both the above figures are showing the orbits of a shaft vibration which is in the
allowable range. The direction of high vibration is also in the perpendicular
Page 114
direction of the high pressure and load zone.We know the high pressure region
will have higher stiffness, so the vibration in that direction will be much smaller
that the direction perpendicular to it. The orbit of Bearing 3 complies the theory
and can be declared as the bearing is running in good condition.
Fig: Orbit at bearing 2
Which has high amplitude and also it is showing higher vibration in the
direction of pressure zone. There is lower vibration in the perpendicular
direction. The orbit is completely different from that of bearing 3. So this
bearing has some problem.
7.5 DIAGNOSIS
From the FFT we took 1x as the predominant frequency of vibration. For better
analysis, we took FFT at 3 bearings. Among those two are showing low and
stable vibration without any anomaly. But bearing 2 has high vibration in the
direction of higher stiffness.
There is not more than one dot, so there is no rubbing.
We are not considering one of the two possible causes i.e. lower clearance.
Then the other remaining cause is high clearance. If there is high clearance, the
pressure distribution is not uniform overall. The pressure fluctuation decreases
the stiffness in the specific direction and increases the vibration. Analysing the
Page 115
orbits, it can be said that there is high clearance with pressure fluctuations and
the high pressure zone may have shifted too. There is high vibration in the X
direction and it is showing and increasing trend over the last few months.
Page 116
PART-8
EFFECT OF PRESSURE AND
TEMPERATURE OF THE
LUBRICATION OIL
Page 117
8.1 BASIC THEORY
Any journal bearing is constructed on the principle of fluid film pressure
development due to wedge mechanism. The lubrication oil develops a wedge
shape due to the rotation of the shaft and creates high pressure zone below the
shaft which carries the load of the shaft. The theory of hydrodynamic
lubrication is based on a differential equation derived by Osborne Reynold
assuming ideal conditions.
The equation is written as
߲ ଷ ߲‫݌‬
߲‫݌‬
߲
߲ℎ
൤ℎ
൨ + ൤ℎଷ ൨ = 6ߤܷ( )
߲‫ݔ‬
߲‫ݖ‬
߲‫ݔ‬
߲‫ݖ‬
߲‫ݔ‬
Where,
h = fluid film thickness
p = pressure developed
µ = viscosity of the lubricant
U = surface velocity of the shaft
This equation cannot be solved analytically. However Raimondi and Boyd
solved this equation on computer using iteration technique. The results of that
are available as chart and tables. They use basically 2 variables to define a
particular bearing. One of them is length to diameter ration and another being
the Sommerfeld number. Sommerfeld number is
S=
࢘૛ ૄ‫ܖ‬
ࢉ૛ ࢖
Where
r = shaft diameter
µ = viscosity of the lubricant
n = journal speed in rev/s
p = unit bearing pressure, i.e., load per unit of the projected area.
Page 118
The Sommerfeld number contains all the variables, which are controlled by the
designer.
From the design point of view, the parameters stated above are chosen so that
thick fluid film develops under the shaft. However there are certain other
parameters that have to considered namely minimum clearance, highest
temperature of the lubricant, delivery pressure and volume in case of forced
lubrication.
8.2 OPERATING CONDITION
The bearing in consideration here is having higher clearance in the high
pressure zone which is analysed. For the stability of the bearing we can change
the pressure or the temperature of the lubricant to reduce the vibration of the
shaft.
The variation of viscosity with temperature of the ISO VG 46 oil is given below
in a chart. The trend shows the viscosity decreases with temperature. We can
change the temperature by changing the settings of the cooler. But we cannot
change the viscosity as much as we want as the temperature of oil will increase
for 6 other bearings of the turbo-generator set, which may cause instability in
those bearings.
The pressure can be varied by changing flow-rate of oil. Again oil from a single
pump and a single main line is fed into all the bearings, so we have to adjust the
main oil line settings as well.
8.3 DIFFERENT PARAMETERS
The relation between the stability of the shaft and pressure and temperature is
discussed below.
Although the conditions are not simply same as the assumptions made by
Reynold, the Raimondi & Boyd chart is used for reference and explanation.
Page 119
8.3.1 VISCOSITY VS TEMPERATURE
The temperature dependency of viscosity is given as
The trend can be shown as
400
350
300
250
200
150
100
50
0
y = 2E-08x6 - 5E-06x5 + 0.000x4 - 0.048x3 + 1.935x2 - 45.34x +
534.1
0
20
40
60
80
100
Page 120
Here, x-axis denotes the temperature in 0C and the y-axis denotes viscosity in
centipoises.
The trend shows rapid decrease in viscosity in the temperature range of 0-300C.
But our operating temperature is about 450C, and here the viscosity changes
nominally.
8.4 ANALYSIS
The bearing in consideration has diameter (d) of 315.94 mm and length (l) 290
mm which gives us the l/d ratio to be 0.918. We take this bearing as a short
journal bearing. The solution to Reynold’s equation, contributed by Ocvirk for
short journal bearing can be written as
P=
ஜ୙
௥௖
ଶ
ଶ
మ (݈ − ‫) ݖ‬
ଷ∈௦௜௡ఏ
(ଵାఢ௖௢௦ఏ)య
Here
l = length of the bearing
z = axial distance of the point from end of the bearing
∈ = eccentricity ratio, i.e., ratio of shaft eccentricity to bearing clearance
As the clearance at the high pressure zone increases the eccentricity ratio
decreases. The decrease in ∈ correspondingly change the Sommerfeld number,
friction co-efficient and p/pmax (p=load per unit area and pmax = max fluid film
pressure) increase. That also means the high pressure zone has shifted.
From the Ocvirk equation pressure increases as ∈ decreases. So the pmax is to be
increased to decrease the ratio of p/pmax. The maximum fluid film pressure
depends upon the pressure at which the lubrication oil is delivered.So we have
to increase the pressure of the main lubrication oil line.
Another way to get the stability back is to decrease the pressure itself. The
pressure is directly proportional to the viscosity and viscosity is dependent on
temperature. The viscosity as well as the pressure can be decreased by
increasing the temperature.
Page 121
Earlier studies in the field of fluid film bearing showed that viscosity is itself
dependent on the pressure. So the relationships are not linear.
8.5 CONCLUSION
Although viscosity, clearance, pressure and temperature are dependent on each
other, the shaft stability can be increased by increasing the temperature or the
pressure or both. The temperature variation should be done keeping in mind the
effect on other bearings. Same thing applies in case of increasing oil pressure.
The exact amount of change needed cannot be stated due to non-availability of
design data and the relation between the maximum fluid film pressure and
delivery oil pressure. For this reason we have to follow the standard operating
procedure and increase the pressure and temperature from previous experiences
and from instruction sheets provided by the manufacturer.
Page 122
KEY BENEFITS
Page 123
Condition monitoring is regular monitoring of machineries and predicting
failure. The main benefits of this are
a. Predicting failure
b. Planned outage
c. Cost saving due to reduction in sudden failure and unwanted repairs
The report covers the theory and practical application of all the state-of-the-art
analysis techniques for condition monitoring by vibration analysis.
The report also covers the analysis of vibration of different rotating machines in
CESC Ltd. The problem of a journal bearing is also monitored and analysed.
Certain proactive measures for the journal bearing is also advised which may
decrease the vibration of the shaft for the time being so that unplanned shutdown can be avoided.
From the cost effectiveness point of view, if a journal bearing fails suddenly
there will be severe losses
• Minimum 4 days downtime for maintenance which is huge
• Cost of a bearing Rs. 2 lakhs
• Huge losses if the blades or other parts are affected
If the bearing clearance problem is detected early we can take certain proactive
actions to retard the deterioration process. Thus unplanned shut-down can be
avoided and huge cost can be saved.
Page 124