vibration monitoring and its features for corelation

Fundamentals of Machinery Condition Monitoring,

Vibration Analysis

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Introduction to Machinery Condition Monitoring

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Introduction to Condition Monitoring

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Maintenance Philosophy

For many years maintenance was simply based on keeping the plant running. There was no real planning or thought involved; it was simply the case that if a machine failed it was repaired or a spare was used. The more later day philosophy on maintenance is based upon:

Optimisation of Production

Optimisation on Plant Availability

No Compromises to Safety

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Maintenance Strategies

There are four main strategies regarding the execution of maintenance activities:

Opportunity Maintenance

Planned Preventative Maintenance

Breakdown Maintenance

Condition Based Maintenance

Reliability based Maintenance

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Adopts the philosophy ‘Fix It When The Opportunity Arises‟. Opportunity Maintenance is carried out, for example, when:

There is a plant shutdown

Other equipment is down for maintenance

A seasonal or weather window is available


Advantages Disadvantages

1. Maintenance does not cause additional plant

downtime.

1. Machine condition may worsen before the

opportunity arises - causing secondary

damage.

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Planned Preventative Maintenance

Adopts the philosophy ‘Fix It Before it Fails‟. Planned Preventative Maintenance is carried out with a high level of planning at specific time intervals.

Based upon generic machinery

reliability data „bath tub‟ effect

Time or running hours based e.g.

after 6 months or 10,000 hours

Preventative Maintenance


1. Maintenance is planned and carried out a

convenient time.

1. Machines repaired when they may not have

a problem.

2. Theoretically fewer catastrophic failures. 2. Repairs can cause more harm than good.

3. Control over storage of spare parts. 3. Unscheduled breakdowns still occur.

4. Not tailored to individual machines.

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Adopts the philosophy ‘Fix It When It Breaks‟.

No element of planning

No forewarning of failure



1. Machines are not over maintained. 1. Unexpected machine downtime - potential loss

of production.

2. No costs associated with preventative or

condition based maintenance. 2. Potential for secondary damage and

catastrophic failure - leading to higher repair

costs.

3. Lack of any form of control or planning

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Adopts the philosophy ‘If It’s Not Broken Don’t Fix It‟. Based upon measured parameters which are sensitive to the development of machinery faults

Maintenance related to machine condition

Predictive - forewarning of failure, diagnosis of faults and root causes



1. Unexpected downtime is reduced. 1. Expense of instrumentation and software,

training personnel and use of specialist contractors.

2. Optimised spares. 2. Does not identify all machine faults e.g. seal

leaks.

3. Causes of failure can be diagnosed.

4. Identifies many common machine faults.

5. Maintenance can be deferred and performed

when opportunity arises.

6. Tailored to individual machines.

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Section 1.2:

Condition Monitoring Strategy

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Condition Monitoring Strategy

A condition monitoring strategy is developed to focus condition monitoring activities on business and safety critical machinery:

A Criticality Assessment is undertaken, based upon the

likelihood of failure and consequences of failure, to

define the equipment for inclusion within the condition

monitoring programme.

Fault Matrices are utilised to identify detectable

machine faults (for each type of critical machinery) and

the parameters that can be monitored which are

sensitive the development of these faults.

The Condition Monitoring Strategy is then defined

based upon the critical machine list and failure modes

which can be realistically identified through condition

monitoring.

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Sample Fault Matrix – Motor Driven Pump (Motor)

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Sample Fault Matrix – Motor Driven Pump (Pump)

The fault matrix is used to define the measurements for inclusion in the condition monitoring database.

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Condition Monitoring Techniques

Vibration Analysis

Lube Oil Analysis

Thermography

Electrical Motor Phase Current Analysis

Rogowski Coil Analysis

Performance Analysis

Condition monitoring techniques include:

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Condition Monitoring vs. Machinery Protection

Condition Monitoring

Trending of „Fault Sensitive‟ machine parameters on a periodic basis, to provide

information regarding current and forecasted machine condition. Allows the

onset of fault conditions to be identified so that maintenance activities can be

planned.

This can prevent unscheduled machinery shutdowns.

Machinery Protection

Acquisition of vibration and temperature values from online systems to initiate

machinery shutdown once a pre-set value has been exceeded.

This avoids catastrophic failure.

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Vibration Theory & Measurement Transducers

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Section 2.1:

Vibration Theory

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What is Vibration?

The number of times a complete

cycle takes place per second is

called the Frequency (measured in

hertz (Hz)).

The motion can be of a single

frequency, as with a tuning fork, or

a number of frequencies such as the

motion of a piston in an internal

combustion engine or a gearbox.

Vibration is defined as an oscillating

motion about an equilibrium point e.g.

a mass on a spring.

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Acceleration, Velocity and Displacement

Acceleration – is a vector quantity (i.e. it has magnitude and

direction) which defines the rate of change of velocity. It is measured in

units of m/s2 or more commonly g, where 1 g = 9.81 m/s2. An acceleration

signal can be integrated to give a velocity signal.

Velocity – is a vector quantity which defines the speed of motion in a

particular direction i.e. the rate of change of displacement. It is normally

measured in units of mm/s (it can be found measured in Imperial units of

„ips‟ (millionths of an inch per second where 1 ips = 25.4 mm/s). A velocity

signal can be integrated to give a displacement signal.

Displacement – is a vector quantity which defines the change of

position from a rest position. It is measured in units of mm or mm

(microns). It can be found measured in Imperial units of „mils‟ where 1 mil

= 25.4 microns.

Vibration is described in one of three terms:

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Simple Harmonic Motion

Where vibration is of a single frequency the motion is sinusoidal and repeats in

an identical pattern over time. This is known as simple harmonic motion.

Vibration is described in terms

of amplitude (the level or

magnitude of the motion) and

frequency (the number of

repetitions of one full cycle

per second).

When viewed in the time

domain, the time waveform

exhibits a pure sine wave.

When viewed in the frequency

domain simple harmonic

motion exhibits a single peak.

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Simple Harmonic Motion – Mass on a Spring Example

The oscillating movement of a mass on a spring exhibits Simple Harmonic Motion :

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Quantifying The Amplitude of Vibration

The amplitude of vibration describes it severity and can be quantified in several

ways: peak-to-peak level, peak level, average level and RMS level.

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Quantifying The Amplitude of Vibration

The Peak-to-Peak value is the maximum excursion (positive to negative motion) of the

time waveform. Useful for measuring direct shaft displacement within a bearing housing

and is often applied to acceleration envelope measurements.

The Peak value (also termed true peak and zero-to-peak) is the absolute value from zero

to the maximum point in the time waveform . Useful for measuring short duration shocks.

The RMS (Root Mean Square) value is directly related to the energy content of the time

waveform and is an indication of the destructive properties of the vibration. Used

predominantly for machine casing (bearing cap) measurements.

The Average value is of limited practical use in describing vibration.

The following is true for purely harmonic

motion (single frequency):

Peak-to-Peak = 2 x Peak

Peak =1.414 x RMS

RMS = 0.707 x Peak

(Average = 0.9 x RMS = 0.637 x Peak)

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Non-Sinusoidal (Complex) Waveforms

The mathematical relationships between peak-to peak, peak, average and RMS

levels are inaccurate once the time waveform is no longer sinusoidal.

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Time vs Frequency Domain

Once more than one frequency is present within a time waveform it can become

very difficult to analyse. A vibration data collector (Fast Fourier Transform

analyser) will convert the time waveform into individual frequency components

for ease of analysis.

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Phase Angle Measurement

Phase measurements are used to assess the relationship between two vibration

signals. It is commonly used in machine balancing and advanced diagnosis of

machinery fitted with proximity probes.

Phase information is taken in the time domain (carrying out a FFT of a time

waveform looses the phase information).

Phase Angle is the timing relationship between two signals of identical

frequencies. Phase is normally measured in degrees and can be either relative

or absolute:

Relative Phase requires two vibration signals of the same frequency.

Absolute Phase requires a vibration signal and a synchronous reference pulse

e.g. Keyphasor or optical tachometer.

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Relative Phase

Relative Phase is the timing relationship between two signals measured in

degrees. It is measured from a point on one signal to the nearest corresponding

point on another signal.

Rules to follow when measuring

relative phase:

1. Two vibration signals are

required

2. Same frequency

3. Same units (e.g. mm/s, mm)

4. Either signal is chosen as the

reference signal

5. Relative phase is measured

as either lead or lag from 0 to

180 degrees

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Absolute Phase

Absolute Phase is the number of degrees of a vibration cycle following the

triggering of a once-per revolution pulse (e.g. keyphasor, optical pickup, strobe

light or magnetic pickup).

Rules to follow when measuring

absolute phase:

1. Two signals are required (one

vibration signal and one reference

signal)

2. Measured from reference signal

thus always a phase lag angle from

0 to 360 degrees

3. 0o location is the point at which

the reference signal triggers

4.Vibration signal to be filtered to a

single frequency which is a integer

multiple of reference signal

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Phase Relationships of Acceleration/Velocity/Displacement

It is very important to measure in consistent units when measuring phase. When

a signal is integrated from acceleration to velocity or from velocity to

displacement its phase angle changes. It is also essential that vibration

transducer orientation is taken into account.

Velocity leads displacement by a phase angle of 90o.

Acceleration leads velocity by a phase angle of 90o.

Acceleration leads displacement by a phase angle of 180o.

Period T = 360o

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Uses of Phase Angle Measurements

Examples of how phase angle measurements can be used:

Shaft Balancing – vibration measurements are related to the phase reference

to calculate the placement of balance weights

Shaft Crack Detection

Shaft/structural resonance detection

Shaft mode shapes

Direction of shaft precession

Confirming force or couple imbalance

Confirming misalignment

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Shaft Orbits

When two XY vibration signals are added together the resultant signal shows a

two dimensional picture of the vibration motion. This is known as an orbit.

If the two vibration signals are from casing mounted transducers (e.g.

accelerometers) the orbit reveals the casing motion.

If the two vibration signals are from proximity probes the orbit reveals the

actual shaft motion within the bearing clearance.

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Shaft Orbit Shape

Shaft orbits must be taken from XY vibration transducers mounted orthogonally

(90o apart). They do not need to be true vertical and true horizontal but must be

90o apart in the radial plane.

Figure 1 shows a typical orbit. As machines are typically more stiff vertically

than horizontally the orbit is elliptical in shape.

Figure 2 shows a circular orbit. A circular orbit is normally indicative of an

imbalance condition.

Figure 3 shows a figure of eight orbit. This shape of orbit is characteristic of

misalignment.

1. Typical Orbit 2. Orbit Indicating

Imbalance

3. Orbit Indicating

Misalignment

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Resonance

When a tuning fork is struck it rings at a single frequency. This frequency is

known as its resonant frequency and is the same every time the fork is struck.

The natural frequency of any system is a function of its stiffness and mass.

Mass

Stiffness

2

1

nFFrequencyNatural

Resonance occurs when the

forcing frequency is equal to

the natural frequency.

At resonance very little

excitation is required to

produce a large response. The

same excitation above or

below resonance will produce

a greatly reduced response.

Increasing or decreasing the

system‟s stiffness or mass will

change the resonant

frequency.

[Note: the above equation is true for an undamped single degree of

freedom system] Boben Anto C

Why Machines Vibrate?

All machines have the properties of mass and stiffness and therefore possess the

ability to vibrate.

Perfect engineering would produce machines with no vibration, however, in

reality all machines are built to tolerances and as such any rotating system will

have an element of unbalance. This unbalance force will produce vibration when

the machine rotates.

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Rigid vs Flexible Rotors

1st Critical

2nd Critical

3rd Critical

Rigid Support and Flexible Rotor

A machine is said to have a Rigid

Rotor if the rotating elements‟

natural frequency is above the

running speed of the unit.

If a machine runs at a speed above

the rotating elements‟ natural

frequency it is said to have a

Flexible Rotor. On run up and

shutdown the machine will pass

through resonance (a critical speed).

A Critical Speed is a natural

frequency of the rotating element

and its support system including

bearings and lubricating film.

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Critical Speeds

It is important to know a machine‟s

critical speeds so that they do not

coincide with normal operating

speeds.

The response of a rotor at critical

speed will give an indication of the

system‟s damping and hence the

condition of the journal bearings.

The synchronous amplification factor

between operating speed vibration

and resonance should be determined

during commissioning to calculate

machine protection system alarms.

Critical speeds of rotating machinery are speeds which correspond with the

systems resonant frequencies.

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Synchronous Amplification Factor

When a unit is running at a speed above resonance (shaft critical speed) the

maximum allowable vibration should be governed by the level the vibration will

reach when it passes through resonance upon shutdown. The simplest

calculation of SAF is the Peak Ratio Method where the level of vibration at

resonance is divided by the steady state running speed vibration. This is often

used to calculate machinery trip levels in order that shutdown occurs at such a

level of vibration to avoid damage on rundown.

Example:

If the level of running vibration is 100

microns (pk-pk) and the bearing oil

clearance is 200 microns. The unit will

only safely run down if the SAF is

significantly less than 2.

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Section 2.2:

Vibration Transducers

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Vibration Transducers

Accelerometer – measures acceleration from machine casing (the signal

can be integrated to measure velocity or displacement).

Velocity Transducer – measures velocity from machine casing.

Proximity Probe (Eddy Current Probe) – non-contact transducer which

measures shaft relative vibration.

There are three main types of vibration transducer:

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Transducer Sensitivities

Accelerometer sensitivities are defined in units of mV/m/s2 (millivolts per

metre per second squared) or more commonly mV/g (where g is gravity).

Velocity Transducer sensitivities are defined in units of mV/ips (millivolts

per inches per second) or mV/mm/s (millivolts per millimetre per second).

Proximity Probe (Eddy Current Probe) sensitivities are defined in units of

mV/mil (millivolts per millionth of an inch (1x10-6 inches)) or mV/mm

(millivolts per micron or micrometre (1x10-6 metres).

The sensitivity of a transducer describes its electrical output per unit of

mechanical input. The higher the electrical output per unit input, the more

sensitive the transducer.

Sensitivities can be described in metric or imperial units. The nominal sensitivity

of a transducer is normally displayed on its casing. For vibration transducers

sensitivities are normally given in mV/EU (millivolts per engineering unit).

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Sensitivity Conversion Factors

The following are the most common sensitivity conversion factors for Imperial to

metric units. When setting up a condition monitoring database it is good

practice to convert all sensitivities into metric units for ease of analysis:

Sensitivity Conversion

Divide by to Obtain

mV/m/s2 9.81 mV/g

mV/ips 25.4 mV/mm/s

mV/mil 25.4 mV/mm

Unit Conversion:

1 g = 9.81 m/s2 (meters per second squared)

1 ips (inches per second) = 25.4 mm/s (millimeters per second)

1 mil (millionth of an inch) = 25.4 mm (micrometers or microns)

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Transducer Frequency Response

Accelerometer – useable range 1 to

20,000 Hz (dependent upon mounting

arrangement), specialist

accelerometers exist for very low and

very high frequency applications.

Velocity Transducer – useable range

10 to 1,500 Hz for electromechanical

type and 1 to 2,000 Hz for

piezoelectric type.

Proximity Probe (Eddy Current

Probe) – stated useable range 0 to

10,000 Hz, typically used in the

frequency range 0 to 2,000 Hz as

higher frequencies can be influenced

by shaft surface imperfections.

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Accelerometer

An accelerometer converts acceleration into an electrical output.

Typical Sensitivities are 25 mV/g, 50 mV/g and most commonly 100 mV/g.

Very sensitive accelerometers for low frequency, low amplitude applications

can be as high as 1000 mV/g (1000 V/g).

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Accelerometer Mounting

An accelerometer‟s frequency response

is highly dependent upon how it is

mounted to the machine surface.

Stud mounted measurements provide

the best frequency response and

repeatability for machine condition

monitoring. This is critical for rolling

element bearing and gearbox vibration

analysis.

Magnetic mounted measurements have

reduced frequency response.

Hand Held measurements provide the

poorest repeatability and frequency

response (should only be used when no

other alternative e.g. very high

machine casing temperature).

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Velocity Transducer

A velocity transducer measures the rate of change of displacement and is

traditionally an electromechanical device.

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Proximity Probe (Eddy Current Probe) A proximity probe is a non-contact electromagnetic sensor which

converts displacement (distance) to voltage. The DC component

of the signal measures the average distance from the shaft

whereas the AC component measures the dynamic fluctuation in

displacement i.e. the vibration.

Typical Sensitivities are 3.94 mV/mm (100 mV/mil) and most commonly 7.87

mV/mm (200 mV/mil). [Note: 1 mil = 25.4 mm]. Boben Anto C

Proximity Probe - Mounting Pairs of radial probes are orientated 90o apart and referred to as „X‟ and „Y‟.

Processing these two signals together produces a shaft orbit.

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Proximity Probe – Gap Voltage The DC component of proximity probe is known as its gap voltage and accurately

measures the distance of the probe tip from the shaft. A probe has a typical

range of 0 to -18 Vdc (~ 2.3 mm). The probes response is linear across a large

proportion of this range. Proximity probe gap voltages are

normally set up at – 9 Vdc.

As a rough guide if the gap voltage is

between -6 Vdc and – 12Vdc it is well

within its linear operating range. If

the gap voltage is not between -3

Vdc to -15 Vdc it is potentially

outside its linear range.

If a gap voltage is close to zero it is

short circuited or to close to the

shaft. If a gap voltage is -18 Vdc it is

open circuit or pointing into space.

Note: Gap voltages by convention are

always negative. Thus the more

„positive‟ the gap voltage the closer

you are to the shaft.

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Proximity Probe – Run out

On high speed machines (>2,500 RPM) run out is measured as part of machine

commissioning during slow roll tests (typically 300 to 600 RPM).

API standards set limits for acceptable levels run out.

As a guide 6 mm (microns) or 10% of the overall vibration signal is acceptable.

Shaft surface imperfections (e.g. scratches, dents, irregular conductivity or

permeability) are indistinguishable from vibration to a proximity probe. This

additional „signal‟ is known as run out.

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Proximity Probe – Used as a Keyphasor

Shaft

Probe

-18V or -24V Proximitor

Out

Shaft

Probe

-18V or -24V Proximitor

Out

Keyway Projection

(20)

(15)

(10)

(5)

0

-Volts

(20)

(15)

(10)

(5)

0 -Volts

A Keyphasor provides a once-per-revolution pulse used as a reference to

measure absolute phase. [Note: Keyphasor is a Bently Nevada trade name]

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Proximity Probe – Used as a Speed Reference

A Keyphasor provides a once

per revolution pulse which

will give an indication of

machine running speed.

To use a proximity probe for

precise speed control requires

a pulse multiple times per

revolution.

For precision speed control a toothed wheel (also known as a Phonic Wheel) is

targeted by a proximity probe. The signal is processed to give a highly accurate

measurement of shaft speed which is updated multiple times per revolution.

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Vibration Transducer Comparison - Advantages

Advantages

Accelerometer Velocity Transducer Proximity Probe

1. Surface Mounted.

2. Small, Portable and

Robust.

3. Large Dynamic Frequency

Range.

4. Relatively Inexpensive.

5. Signal can be integrated

to measure velocity or

displacement.

1. Surface mounted and

portable.

2. Self-generating no

complex signal

conditioning.

1. Direct measurement of

shaft of shaft

motion/position within

journal bearing.

2. Very sensitive to low

frequencies down to DC.

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Vibration Transducer Comparison - Disadvantages

Disadvantages


1. Requires amplifier

electronics. 1. Bulky.

2. Limited Frequency Range

(<1.5 kHz).

3. Moving parts potentially

wear over time.

1. Limited frequency range

(0 to 10 kHz). Practical

range 0 to 2 kHz.

2. Permanently mounted

(not portable) often

difficult to replace.

3. Conditioning electronics

required and interface

panel must be housed in

non-hazardous area.

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Vibration Transducer Comparison - Applications

Applications


1. Machines with rolling

element bearings.

2. Gearbox Fault Diagnosis.

3. Heavy rigid rotors with

light casing/foundations.

4. Highly utilised with

portable handheld data

collectors.

1. Portable transducer for

measurement of low speed

machines.

1. Machines with journal

bearings.

2. Machines with lightweight

high speed rotors in

heavy

casing/foundations.

3. Measurement of radial

shaft vibration and axial

shaft position.

4. Keyphasor (phase

reference device).

5. Speed reference.

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ISO & API Standards

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Alarm Setting Guidelines

American Petroleum Institute (API) and International Standards Organisation

(ISO) have produced guidelines as to acceptable levels of vibration based upon

generic machine types.

These guidelines are used in acceptance testing, e.g. during commissioning, to

ascertain if a new machine is fit for purpose. They are also useful for

reference purposes but it should be noted that alarm settings for condition

monitoring purposes should be set on a machine by machine basis taking into

account historical data.

The rate at which vibration levels and characteristics deteriorate over time is

as important as the magnitude of vibration i.e. a high level of vibration that

remains stable over time may be less cause for concern than a lower level of

vibration which shows a deteriorating trend or changing characteristics.

In routine condition monitoring we are looking for deterioration in vibrations

levels not just absolute values.

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ISO-10816-1 Vibration Standard

ISO-10816-1 – Mechanical Vibration – Evaluation of Machinery Vibration by

Measurements on Non-Rotating Parts is the most commonly referenced standard

in routine condition monitoring for the evaluation of overall vibration levels taken

on machine casings.

The standard provides general guidelines for the severity of overall casing (i.e.

bearing cap) vibration levels based upon Machine Classes I, II, III and IV. These

classes are define by power rating and stiffness of the mounting arrangement

(i.e. rigid or relatively soft mounts).

The severity of vibration is classified into Evaluation Zones A, B, C and D to

quantify if the level of vibration is acceptable for long term operation of the

machine.

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Extract from ISO-10816-1 – Machine Classifications

The machine classifications are as follows:

Class I: Individual parts of engines and machines, integrally connected to the

complete machine in its normal operating condition. (Production electrical

motors up to 15 kW are typical examples of machines in this category).

Class II: Medium sized machines (typically electrical motors with 15 kW to 75 kW

output) without special foundations, rigidly mounted engines or machines (up to

300 kW) on special foundations.

Class III: Large prime-movers and other large machines with rotating masses

mounted on rigid and heavy foundations which are relatively stiff in the direction

of vibration measurements.

Class IV: Large prime-movers and other large machines with rotating masses

mounted on foundations which are relatively soft in the direction of vibration

measurements (for example, turbo generator sets and gas turbines with outputs

greater than 10 MW).

[Note: soft or special foundations refer to anti-vibration mounts]

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Extract from ISO-10816-1 - Evaluation Zones

The following typical evaluation zones are defined to permit a qualitative

assessment of the vibration on a given machine and to provide guidelines on

possible actions:

Zone A: The vibration of newly commissioned machines would fall within this

zone.

Zone B: Machines with vibration within this zone are normally considered

acceptable for unrestricted long-term operation.

Zone C: Machines with vibration within this zone are normally considered

unsatisfactory for long-term continuous operation. Generally, the machine may

be operated for a limited period in this condition until a suitable opportunity

arises for remedial action.

Zone D: Vibration values within this zone are normally considered to be of

sufficient severity to cause damage to the machine.

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Extract from ISO-10816-1 – Typical Zone Boundary Limits

ISO-10816-1 - Typical Zone Boundary Limits

Vibration Velocity

mm/s (RMS) Class I Class II Class III Class IV

0.28

A A

A A

0.45

0.71

1.12 B

1.8 B

2.8 C B

4.5 C B

7.1

D

C 11.2

D

C 18

D 28 D

45

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API Standards

The American Petroleum Institute (API) has produced numerous standards to

satisfy the specific needs of the petroleum, chemical and gas industries. These

standards closely specify the detailed design, inspection and testing of generic

machine types, for example:

API 610 – Centrifugal Pumps

API 611 – General Purpose Steam Turbines

API 612 – Special Purpose Steam Turbines

API 613 – Special Purpose Gear Units

API 616 – Gas Turbines

API 617 – Axial and Centrifugal and Expander Compressors

API 618 – Reciprocating Compressors

API 619 – Rotary-Type Positive Displacement Compressors

API 670 - Vibration, Axial Position and Bearing Temperature Monitoring

Systems

API 674 – Positive Displacement Pumps – Reciprocating

API 676 – Positive Displacement Pumps – Rotary

API 677 – General Purpose Gear Units

API 681 – Liquid Ring Pumps

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API Standards – Guidelines for Vibration

API 670 „Vibration, Axial Position and Bearing Temperature Monitoring Systems‟

specifies requirements for the supply, installation and calibration of radial shaft

vibration and axial-position transducers and bearing temperature sensors for

online machinery protection systems. The various machine specific standards give

guidance as to acceptable levels of vibration.

The following equation is common to several of the API standards and defines the

maximum allowable level of relative shaft vibration (i.e. vibration measured using

a proximity probe):

NA

000,124.25

Where:

A = amplitude of unfiltered vibration, in micrometers true peak-to-peak

N = maximum continues speed, in resolutions per minute

e.g. Gearbox vibration shall not exceed 50 micrometers (peak-to-peak) or that

defined by the above equation, whichever is less [API 613].

e.g. Centrifugal Compressor vibration shall not exceed 25 micrometers (peak-to-

peak) or that defined by the above equation, whichever is less [API 617]. Boben Anto C

Machinery Fault Diagnosis

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Section 4.1:

Imbalance

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Imbalance

Imbalance is one of the most common causes of excessive vibration in

rotating machinery.

It is always characterised as radial vibration at the 1X running speed

(rotational speed) of the shaft.

Dependent upon the relative support stiffness, radial vibration may be

more prominent in the horizontal or vertical axis.

It is often misdiagnosed as many faults exhibit 1X running speed

characteristics - other vibration symptoms should be investigated before

balancing is attempted.

Imbalance (also known as unbalance) occurs when there is a deviation between

the geometric centre of a rotor and its centre of mass. Or put

more simply – when there is a heavy spot on the shaft.

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Static (Force) Imbalance

Is characterised by a dominant

1X running speed component.

Is measured in the radial

direction and is in-phase.

Is corrected by one balance

weight in one plane at Rotor

centre of gravity (single plane

balance).

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Is again characterised by a

dominant 1X running speed

component.

Will be 180o out of phase on

same shaft and can exhibit both

high axial and radial vibration.

Is corrected by applying balance

weights in more than one plane

(multi-plane balance).

Couple Imbalance

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Dynamic Imbalance

A rotor with static imbalance can be diagnosed when the machine is not running.

This is carried out by placing the rotor in frictionless bearings. If the rotor has a

heavy spot it will rotate within the bearings until the heavy spot is at the

bottom.

Conversely a rotor with pure coupled imbalanced will not rotate when placed in

frictionless bearings and will only manifest itself when the machine is running.

In practice a rotor is likely to have a combination of static and couple imbalance

which is collectively referred to as Dynamic Imbalance.

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Is again characterised by a

dominant 1X running speed

component.

Will show a high axial 1X

running speed component.

Axial vibration readings tend

to be in-phase whereas radial

readings may be unsteady.

Overhung Rotor Imbalance

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Characterised by a dominant 1X running

speed radial vibration.

Will show highest vibration at the motor

non-drive end irrespective of source.

Vertical Rotor Imbalance

Note: Vertically mounted pumps will often show large 1X running speed vibration at the

motor non-drive end for a number of faults (e.g. pump bush wear, flow turbulence). Try to

isolate the problem by uncoupling the motor/pump. Carry out measurements on motor

whilst uncoupled. If 1X vibration is still relatively high the motor is at fault if not it is the

pump.

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Fan and Overhung Imbalance

Fan imbalance is characterised by a

dominant 1X running speed component

in the radial direction.

Overhung fan imbalance is

characterised by a dominant 1X

running speed component in the axial

direction.

Ensure the fan blades are clean and

show no signs of damage. Often the

cause of imbalance can be a build up

of deposits on the fan blades.

Fan balancing can often be carried out

in situ. The impeller can normally be

balanced by applying a single weight.

Imbalance can be very common in fans but should not be mistaken for belt drive

problems which can also reveal 1X running speed vibration.

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Imbalance Severity

Manufacturer‟s and API Standards may impose acceptable levels of

imbalance for specific machine types. The following is a rough guideline to the

severity of imbalance relating to the 1X running speed vibration component of

vibration (for machines running between 1800 and 3600 RPM).

Very high speed machines will have lower tolerance levels as the forces

generated by imbalance, increase with machine speed.

Severity of Imbalance Guidelines for Machines Running at 1800 to 3600 RPM

1X Vibration Level Diagnosis Repair Priority

VdB (US) re 10E-8 m/s (rms) Equivalent mm/s (rms)

<108 0 to 2.5 Slight Imbalance No Recommendation

108 to 114 2.5 to 5.0 Moderate Imbalance Desirable

114 to 124 5.0 to 15.8 Serious Imbalance Important

>124 >15.8 Extreme Imbalance Mandatory

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Effects of Imbalance

All machines inherently have some form of residual imbalance.

Some slight imbalance will have little effect on the operating lifespan of a

machine or its components.

An unacceptable level of imbalance can severely reduce the lifespan of

bearings and seals.

A high level of imbalance can have catastrophic effects for large machinery

with flexible rotors (running above shaft critical speeds).

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Misalignment

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Misalignment

Misalignment can be parallel (offset) or angular which will be diagnosed by

whether the vibration characteristics are dominant in the radial or axial

planes respectively. It can, however, be a combination of both.

Misalignment introduces a static preload force into the coupled shafts.

It is typically characterised by 1X, 2X and 3X running speed vibration

components.

Misalignment characteristics may also indicate coupling problems.

Misalignment will be effected by thermal and dynamic growth and may

manifest itself more prominently once the machine reaches its steady state

operating condition.

Misalignment occurs when the axis's of coupled machine components are not

collinear.

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Parallel (Offset) Misalignment

Parallel misalignment is characterised by dominant 1X and 2X and to a lesser

extent 3X running speed components of vibration in the radial direction.

Approaches 180o out of phase across the coupling in the radial direction.

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Angular Misalignment

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Cocked Bearing

A cocked bearing is a form of misalignment which can generate high axial

vibration, characterised by high 1X, 2X and 3X running speed characteristics.

Will show 180o phase shifts (in

the axial plane) top/bottom

and/or left/right on the same

bearing housing.

Realignment or balancing will

not cure the problem - the

bearing will need to be removed

and reinstalled correctly.

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Mechanical Looseness and Rotor Rub

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Mechanical Looseness

Mechanical looseness is a common cause of high vibration and is by far one of the

easiest problems to check.

Mechanical looseness should ALWAYS be checked before more intrusive

maintenance activities are considered.

Mechanical looseness often reveals high 1X vibration.

Mechanical looseness can exhibit high 1X, 2X and 3X times running speed

components of vibration which can be misinterpreted as misalignment.

A raised noise floor or ½ x multiples of running speed harmonics of

vibration can sometimes be associated with looseness.

The Technical Associates of Charlotte define mechanical looseness in three

categories; A: Structural, B: Fasteners and C: Component Fits.

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Mechanical Looseness – Type A: Structural

“Type A is caused by Structural looseness/weakness of machine feet, base plate or

foundation; also by deteriorated grouting, loose hold-down bolts at the base; and

distortion of the frame or base (i.e., soft foot).

Phase analysis may reveal approximately 90° - 180° phase difference between

vertical measurements on bolt, machine foot, base plate, or base itself.” -

Technical Associates of Charlotte, Inc.

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Mechanical Looseness – Type B: Fasteners

“Type B is generally caused by loose pillowblock bolts, cracks in frame structure or

in bearing pedestal.” - Technical Associates of Charlotte, Inc.

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Mechanical Looseness – Type C: Component Fits “Type C is normally generated by improper fit between component parts. Causes a

truncation of time waveform and a raised noise floor in the spectrum. Type C is often

caused by a bearing liner loose in its cap, a bearing loose turning on its shaft, excessive

clearance in either a sleeve or rolling element bearing, or a loose impeller on a shaft, etc.

Type C Phase is often unstable and may vary widely from one measurement to next,

particularly if rotor shifts position on shaft from one startup to next. Mechanical Looseness

is often highly directional and may cause noticeably different readings comparing levels at

30° increments in radial direction all the way around one bearing housing. Also, note that

looseness will often cause subharmonic multiples at exactly 1/2 or 1/3 RPM (.5X, 1.5X, 2.5X,

etc.). .” - Technical Associates of Charlotte, Inc.

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Rotor Rub “Rotor Rub produces similar spectra to Mechanical Looseness when rotating parts contact

stationary components. Rub may be either partial or throughout the entire shaft revolution.

Usually generates a series of frequencies, often exciting one or more resonances. Often

excites integer fraction subharmonics of running speed (1/2, 1/3, 1/4, 1/5,...1/n),

depending on location of rotor natural frequencies. Rotor rub can excite many high

frequencies (similar to wide-band noise when chalk is drug along a blackboard).

It can be very serious and of short duration if caused by shaft contacting bearing babbitt. A

full annular rub throughout an entire shaft revolution can induce "reverse precession" with

the rotor whirling at critical speed in a direction opposite shaft rotation (inherently unstable

which can lead to catastrophic failure).

- Technical Associates of Charlotte, Inc.

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Rolling Element Bearings

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Rolling Element (Frictionless) Bearing Faults

Rolling element bearing faults are characterised by discrete (non-synchronous)

frequency vibration components which are dependent upon the bearing

construction.

Prism4 includes a database of the most common bearing tags and can compute

these frequencies through it‟s Frequency Analysis Module (FAM).

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Deep Groove Ball Bearing Components

Seal Rolling elements Inner ring

Outer ring Cage Seal

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Rolling Element Bearing Stages of Failure

The very early stages of bearing faults are detected at very high frequencies using

spike energy, shock pulse or HFD (High Frequency Detection) techniques.

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Rolling Element Bearing Stages of Failure

The development of bearing faults can be tracked using acceleration enveloping

techniques.

Final stages of bearing faults will become evident in the vibration velocity spectra.

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Acceleration Enveloping – How it Works?

All low frequency vibration components which are attributed to mechanical

faults such as imbalance, misalignment, mechanical looseness etc. are filtered

out.

Bearing fault frequencies are non-synchronous components of vibration i.e.

they are not exactly 1X, 2X etc. running speed multiples. They may however

be very close to a multiple of running speed (e.g. 4.9X running speed) and thus

difficult to separate out in an unfiltered signal. Enveloping inherently

provides this filtering.

Very sensitive to the onset of bearing faults. By the time a fault is visible in a

vibration spectra, the bearing is already significantly worn.

The principles behind acceleration enveloping:

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Time waveform illustrating dominant 1X running speed vibration with frequent

impacts (the transient superimposed on the waveform is like the ringing of a bell):

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Firstly the low frequency components of vibration are removed using a high pass

filter:


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The resulting filtered signal contains only the high frequencies with the lower

frequencies removed:


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The time waveform would now only show the bearing transient impacts (note the

1X running speed waveform has been filtered out):


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The signal is now demodulated so that the high frequencies are flipped over into

the baseband of the frequency scale:


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The negative components of the signal waveform are flipped over to the positive

portion of the signal:


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A low pass filter is now applied to remove any unwanted signals from other sources

of modulation:


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The high frequency components are now removed. This is the enveloped signal:


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Acceleration Enveloping - Analysis

The analysis on the remaining spectra is based upon trending of the frequency

peaks against the noise floor:

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Acceleration Enveloping - Considerations

Acceleration enveloping has become the favoured technique for assessing the

onset and deterioration of rolling element bearing faults. It is both highly

sensitive and easy to trend, however, great care should be taken in interpreting

the results as:

The measurement location/technique needs to be highly repeatable. It is

advised to stud mount the accelerometer. Swapping between stud and

magnetic mounted or handheld measurements will produce highly spurious

results/trends.

The measured characteristics are analysed as deterioration relative to the

signature of a new bearing. Starting measurements well into a bearings

lifespan provides a less reliable reference point.

Acceleration enveloping can give an indication of insufficient lubrication which

could be misinterpreted as poor bearing condition.

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Acceleration Enveloping - Guidelines

There are basic guidelines to unacceptable levels of acceleration enveloping

(gE); based upon machine speed and shaft diameter.

These should be used with caution as it is the relationship between the

discrete bearing fault frequencies and the noise floor (often termed „carpet

level‟) which provides the best indication of bearing condition.

This relationship will differ dependent on the stage of bearing wear which is

why it is always important to gather baseline data when a new bearing is

fitted.

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Acceleration Enveloping Guidance Levels:

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Section 4.5:

Journal Bearings

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Journal Bearing Faults

Journal Bearings also known as Sleeve, Plain, Fluid Film or White Metal Bearings

show very different fault characteristics to rolling element bearings.

The most common Journal bearing faults are:

Wear and Clearance Problems

Due to improved tolerances in Journal Bearing

design the following faults are nowadays less

common but can still cause catastrophic effects:

Oil Whirl Instability

Oil Whip Instability

NOTES:

1.Wear, Clearance and Oil Whirl problems can be detected in steady state vibration

spectra, whereas Oil Whip is more likely to occur during machine start-up, which requires

more advanced transient analysis.

2. Acceleration Envelope and HFD/Spike Energy/Shock Pulse measurement techniques are

NOT applicable to Journal Bearings.

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Journal Bearing Wear and Clearance Problems

Journal bearing wear and clearance problems show very similar symptoms to

mechanical looseness and are identified by strong running speed harmonics:

Wiped journal bearings will often show high vertical vibration compared with the

horizontal measurement.

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Oil Whirl Instability

Oil whirl can be characterised by vibration frequencies just below, but never equal

to, half times running speed:

Changes in lube oil viscosity, lube oil pressure and external preloads can all

effect oil whirl.

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Oil Whip Instability

“Oil Whip may occur if machine operated at or above 2X rotor critical frequency.

When rotor brought up to twice critical speed, whirl will be very close to rotor

critical and may cause excessive vibration that oil film may no longer be capable

of supporting. Whirl speed will actually "lock onto" rotor critical and this peak will

not pass through it even if machine is brought to higher and higher speeds.

Produces a lateral forward processional subharmonic vibration at rotor critical

frequency. Inherently unstable which can lead to catastrophic failure.” – Technical

Associates of Charlotte, Inc.

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Journal Bearings

Journal bearings are often fitted to large machinery with online protection

systems.

Alert and trip setting will be set for vibration, axial displacement and bearing

temperatures.

When a journal bearing wipes both the vibration and temperature will

increase instantaneously, most likely tripping the machine.

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Gear Analysis

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Gear Faults

When analysing gears (e.g. helical, spur, worm, bevel, epicyclic) gear mesh

frequencies can be calculated from:

Input & Output Shaft Speed

Number of Teeth on Pinion & Wheel

(Note: For a two stage gearbox the shaft speed

and teeth of the intermediate gears would also

be required).

When collecting vibration data on gearboxes,

where possible, time waveform data should

be captured along with the FFT Spectra, in

at least one axis.

Theoretically Acceleration Envelope techniques can be applied to gear analysis,

they should be utilised with caution however, as the transmission path between

the meshing gears and vibration measurement location is often indirect

(especially on large gearboxes).

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Gear Mesh Frequency

Gear mesh frequency is defined as the number of teeth on a gear multiplied by

its shaft rotating frequency:

Gear mesh frequency = (Low Speed Shaft RPM / 60) x number of teeth on wheel

Or (High Speed Shaft RPM / 60) x number of teeth on pinion

[NOTE: The RPM has been divided by 60 to convert it into Hertz]

When analysing gearboxes it is essential that data is captured using a suitable

frequency range to capture up to 3.25 x gear mesh frequency in order to assess

gear misalignment issues.

We would normally capture a vibration velocity measurement with a frequency

range up to 5 kHz to capture gear natural frequencies and an acceleration

reading up to 20 kHz to capture harmonics of gear mesh frequency.

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Gear Mesh Characteristics

A typical gearbox vibration spectrum will show low speed and high speed shaft

running speed components accompanied by low amplitude gear mesh frequency

with shaft running speed sidebands.

The highest level of vibration will be

either radial or axial dependent upon

the type of gear e.g. spur or helical

gear.

The time waveform should show

evenly spaced impulses of similar

amplitude for a healthy gearbox. A

pulse is produced as each tooth

meshes.

The time waveform is often easier to

analyse than the vibration spectra in

the diagnosis of gear faults.

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Gear Tooth Wear

When gear teeth start to wear the sidebands of gear mesh frequency become more

pronounced – the amplitude and number of sidebands will increase. Gear natural

frequencies will also be excited.

The gear mesh frequency sidebands

will correspond to the gear with the

wear e.g. if the sidebands are equal

to the high speed shaft running speed

it will be the pinion gear teeth which

exhibit wear.

The gear natural frequencies are

lower than gear mesh frequency and

will also exhibit sidebands relating to

the bad gear.

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Gear Tooth Load

Gear mesh frequencies can be very sensitive to load. High gear mesh frequencies

do not necessarily indicate a problem provided sideband frequencies remain at low

amplitudes and gear natural frequencies are not excited.

In order to trend gear mesh activity,

vibration measurements should be

recorded with the machine operating

at the same load each survey

(wherever possible).

Machine load should be recorded

each survey as a manual entry

reading.

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Gear Eccentricity and Backlash

Relatively high sidebands around gear mesh frequency can indicate eccentricity,

backlash or non-parallel shafts. The bad gear will be indicated by the spacing of

the sideband frequencies.

Eccentricity will normally show a high

1X running speed component of

vibration.

Improper backlash often excites gear

mesh harmonics and gear natural

frequencies.

Gear mesh frequency amplitudes will

often reduce with increasing load if

backlash is the problem.

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Gear Tooth Misalignment

Gear tooth misalignment is diagnosed in a similar manner to angular or parallel

misalignment except it is 1X, 2X and 3X gear mesh frequency which reveals the

symptoms.

In order to assess for gear tooth

misalignment vibration measurements

must be taken with a frequency range

> 3.25 x gear mesh frequency.

Gear tooth misalignment will cause

uneven tooth wear.

NOTE: A loose fit journal bearing can also exhibit high 1X, 2X and 3X times gear

mesh frequency vibration.

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Gear Cracked or Broken Tooth

A cracked or broken tooth is best diagnosed in the time waveform which will show

a large impulse every time the problem tooth tries to mesh with the teeth on the

mating gear.

The frequency spectra will reveal

gear natural frequencies.

The time waveform will reveal high

amplitude 1X running speed

component of the problem gear.

These spikes will reveal themselves in

the time waveform up to 10 to 20

times higher than in the frequency

spectrum.

In the example opposite (gear with 12

teeth) the time waveform reveals a

very large impulse for the

cracked/broken tooth.

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Gear Hunting Tooth

Hunting tooth problems occur due to faults on both the gear and pinion created

during manufacture or improper handling. A hunting tooth problem can often be

overlooked as it is revealed at very low frequencies, often less than 10 Hz.

A gearbox with a hunting tooth

problem may emit a „growling‟ sound.

The effect is at its worst when the

faulty gear and pinion teeth try to

mesh at the same time. This may

only occur once every 10 to 20

revolutions.

The number of teeth on a gear are

often a prime number to avoid

hunting tooth problems i.e. two

imperfect teeth will not repeatedly

mesh.

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Hydraulic and Aerodynamic Faults

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Hydraulic and Aerodynamic Faults

Hydraulic and Aerodynamic forces in pumps, fans and compressors will produce

vibration at Blade Pass or Vane Pass frequency. Blade pass frequency can be

calculated by multiplying the number of blades or vanes by the shaft rotational

speed:

Blade Pass Frequency = Number of Blades x (RPM / 60) Hz

Vane Pass Frequency = Number of Vanes x (RPM / 60) Hz

Blade pass frequency is an inherent characteristic of the machine which will vary

with process conditions and does not normally cause a problem.

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Blade Vane Faults

Large blade pass frequencies are generated if the gap between the rotating vanes

and stationary diffusers is not equal all the way round i.e. it is eccentric.

High blade pass frequencies, with 1X

running speed sidebands, can be

generated if an impeller wear ring seizes

on the shaft or if welds which fasten

impeller vanes fail.

High blade pass frequencies can also be

generated by flow disturbance or the

eccentric positioning of the pump or fan

rotor within its housing.

NOTE: Blade pass frequency fluctuates significantly with process condition so a

deteriorating trend must be established before intrusive maintenance is

considered. Closely monitor for increases in the blade pass frequency sidebands.

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Flow Turbulence

Flow turbulence is caused by variations in the pressure or velocity of air passing

through a fan or blower.

Flow turbulence will normally exhibit

sub-synchronous (below 1X running

speed) random noise in the vibration

spectra.

Excessive turbulence can also exhibit

broadband high frequency vibration i.e.

an increase in noise floor above the

blade pass frequency.

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Cavitation

Cavitation is normally caused by insufficient suction pressure (starvation) or inlet

flow and normally generates random high frequency broadband vibration which

appears as a raised noise floor above the blade pass frequency.

Cavitation can be audible – sounding like

gravel is passing through the pump.

Cavitation can cause erosion of pump

internals and impellers if left

uncorrected.

Cavitation may vary from one survey to

the next but can often be overcome by

increasing the pump suction pressure.

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Electrical Faults

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Electrical Faults

Mechanical faults such as imbalance, misalignment and bearing problems are

typically more common in electric motors than electrical faults.

Electrical fault frequencies may be present in a vibration spectra at low levels,

which could be a characteristic of the machine and not likely to cause long term

detrimental effects.

One of the simplest ways to identify whether a vibration component is mechanical

or electrical in origin is to shutdown the power to the machine whilst recording

real time vibration. If the fault frequency immediately disappears when the

power is switched of the problem is electrical but if the faults frequency reduces

with running speed the fault is mechanical in origin.

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AC Induction Motors

3-Phase AC Induction Motors are the most common motors utilised in industrial

applications due to their relatively high efficiencies.

A motor with a 60 Hz line frequency and 2 stator poles will run at a speed

of 3600 RPM (or 3000 RPM with a 50 Hz line frequency). A motor with a

60 Hz line frequency and 4 stator poles will run at a speed of 1800 RPM

(or 1500 RPM with a 50 Hz line frequency).

In an induction motor the motor speed is always slightly less than

synchronous speed. The difference between the actual speed and the

synchronous speed is known as Slip. The difference between the

running speed frequency and synchronous speed frequency is known as

the Slip Frequency.

The greater the slip, the greater the induced current in the rotor bars

and the greater the output torque. This is why the actual speed of an

induction motor will vary slightly with load.

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AC Induction Motors – Stator Eccentricity

Stator problems generate high 2X line frequency components of vibration i.e. 100

Hz or 120 Hz dependent upon whether the line frequency is 50 Hz or 60 Hz

respectively. Stator eccentricity produces a uneven stationary air gap between

the rotor and stator which produces very directional vibration.

Differential air gap should not exceed 5%

for induction motors and 10% for

synchronous motors.

Soft foot and warped bases can produce

eccentric stators.

Shorted stator windings can produce

thermally-induced vibration which can

significantly increase with operating time

causing stator distortion and air gap

problems.

NOTE: Electric motors will have a low level of 2X line frequency vibration as a

normal characteristic.

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AC Induction Motors – Eccentric Rotor

An eccentric rotor will produce a variable air gap between the rotor and stator

producing pulsating vibration. Eccentric rotors produce 2X line frequency

components with pole passing frequency sidebands.

Pole passing frequency = slip frequency X

numbers of poles.

Not to be confused with soft foot or

misalignment which can produce variable

air gaps due to distortion (mechanical

problem not electrical).

Zoom analysis may be required to

separate 2X line frequency from 2X

running speed harmonics.

On high voltage motors; Motor Phase

Current Analysis can be used to assess

rotor eccentricity.

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AC Induction Motors – Rotor Bow

Uneven heating of a rotor due to unbalanced rotor bar current distribution can

cause a rotor to warp or bow. Rotor bow can be misdiagnosed as mechanical

imbalance as it has similar 1X running speed characteristics.

Rotor bow can be distinguished from

imbalance as it will worsen when the

motor is hot and the symptoms will

subside when the motor cools down.

If the local heating effect is very severe

it can cause the offending rotor bar to

melt and lodge within the air gap.

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AC Induction Motors – Broken or Cracked Rotor Bars

Broken/Cracked rotor bars or shorting rings, bad joints between shorting rings and

rotor bars or shorted rotor laminations will produce high levels of 1X running speed

vibration harmonics with pole pass frequency sidebands.

High resolution measurements are

required typically using a 3200 line FFT

spectra.

Running speed harmonics may be notable

to 5X running speed and above.

On high voltage motors; Motor Phase

Current Analysis is often employed to

assess for broken rotor bars.

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AC Induction Motors – Loose Rotor Bars

Loose rotor bars will exhibit a peak at rotor bar passing frequency (the number of

rotor bars times the motor RPM) with 2X line frequency sidebands.

A relatively high frequency measurement

is required to detect loose rotor bars as

the rotor passing frequency is often over

2000 Hz.

Even if the number of rotor bars is

unknown a high frequency vibration

component exhibiting 2X line frequency

sidebands is most likely caused by loose

rotors.

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AC Induction Motors – Phasing Problems

Phasing problems due to loose or broken connectors can result is excessive 2X line

frequency vibration with 1/3 line frequency sidebands.

Very high 2X line frequency vibration

levels in excess of 25 mm/s can result if

the problem is left uncorrected.

The problem is accentuated if the

defective connector makes intermittent

contact.

The loose or broken connector must be

repaired to avoid catastrophic failure.

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Belt Drive Faults

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Belt Drive Faults

Belt drives are an inexpensive means of power transmission. They can however be

prone to a number of faults including:

Belt Wear

Misaligned Sheaves (Pulleys)

Eccentric Sheaves (Pulleys)

Belt Resonance

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Belt Drive Equations

The following equations are useful in determining operating speeds and fault

frequencies for belt drives:

DiameterSheaveDriven

DiameterSheaveDrivingRPMDrivingRPMDriven

__

____

LengthBelt

DiameterSheaveRPMSheavePIFrequencyBelt

_

___

Where PI = 3.1416

TeethBeltofNumberFrequencyBeltFrequencyBeltgTi ______min

SheaveonTeethofNumberRPMSheaveFrequencyBeltgTi _______min

or

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Worn, Loose or Mismatched Belts

Belt frequencies are below both the driving and driven units running speeds. A

worn, loose or mismatched belt will exhibit up to 3X or 4X harmonics of belt

frequency.

A high amplitude of 2X belt frequency is

normally present.

Amplitudes are unsteady and will

sometimes fluctuate between the driving

and driven unit running speed.

Timing belts will exhibit high amplitudes

of timing belt frequency.

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Sheave (Pulley) Misalignment

Misaligned sheaves will produce high vibration at 1X RPM predominantly in the

axial direction. Often with pulley misalignment, the highest axial vibration on the

motor will be at fan RPM, or vice versa.

A high amplitude of 1X belt frequency

will be present in the axial direction.

Harmonics of belt frequency may

sometimes be present in the axial

direction.

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Eccentric Sheaves (Pulleys)

Eccentric sheaves will generate high 1X running speed vibration, especially in the

axis parallel to the direction of the belts.

This condition is very common and can

be misdiagnosed for imbalance.

The 1X running speed of an eccentric

sheave will be evident on both the

driving and driven unit bearings.

Pulley eccentricity can be confirmed by

phase analysis which should reveal

horizontal and vertical phase differences

close to either 0o or 180o.

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Belt Resonance

Belt resonance can reach high amplitudes if the belt natural frequency coincides

with either the driving or driven unit running speed.

This condition can be checked by

tensioning and then releasing the belt

whilst taking vibration readings on the

pulleys or bearings.

The natural frequency of the belt can be

altered, by either changing the belt

length or tension, to correct this

problem.

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Section 4.10:

Machinery Fault Diagnosis Guidelines

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Machinery Fault Diagnosis Guidelines

When carrying out machinery fault diagnosis there is often more than one type of

fault that can be associated with the vibration characteristics measured. The

following guidelines should be taken into account before making

recommendations involving intrusive maintenance:

Vibration levels change with load and process conditions. If a step increase in

1X running speed vibration is evident look for evidence of changes in process

conditions. Process parameters should be recorded wherever possible (as

manual entry readings) for comparison with historical data.

Pump 1X running speed and blade pass frequency vibration can often fluctuate

spuriously due to altering process conditions.

If vibration levels show a marked increase or deteriorating trend; survey more

often. By surveying weekly instead of monthly the rate of deterioration can be

more easily established. If the vibration levels appear stable revert to

monthly monitoring.

Always check for mechanical looseness if 1X, 2X or 3X vibration is present as it

is one of the most inexpensive and easiest maintenance actions to carry out.

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Machinery Fault Diagnosis Guidelines Continued

High acceleration envelope levels on rolling element bearings can be associated

with lack of lubrication. To check this, grease the bearings and then repeat

vibration measurements. The acceleration envelope readings should notably

reduce. Repeat the vibration measurements after the machine has been

allowed to run for a further 24 hours. If the acceleration envelope readings

remain stable, lack of lubrication is confirmed. If the acceleration envelope

readings start to increase closely monitor for bearing fault frequencies.

Ensure that the vibration spectra reveals good quality data. A ski-slope in the

vibration spectra is an indication of a poorly taken measurement and should be

repeated:

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Machinery Fault Diagnosis Guidelines Continued

Ensure measurements are always taken at the same locations. Stud and mark-

up the measurement locations wherever possible to ensure the best possible

repeatability of measurements.

If a stud should fall off a measurement location immediately replace it.

Acceleration envelope measurements are highly dependent upon a repeatable

mounting arrangement. If the measurement is changed from stud to magnetic

mounting the envelope levels will be inconsistent and the baseline/trend will be

lost.

Look for changes in vibration characteristics i.e. frequency component changes

as well as changes in overall vibration levels. A fault may not manifest itself as

an increase in overall vibration levels.

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Section 6:

Supplementary Condition Monitoring Techniques

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Supplementary Condition Monitoring Techniques

Vibration analysis is a very powerful technique in the diagnosis of many common rotating machinery problems. Additional condition monitoring techniques can be employed in support of vibration analysis and to provide supplementary information regarding both machinery and plant condition. Such techniques include:

Lube Oil Analysis

Thermography

Watch Keeping

Motor Phase Current Analysis


Performance Monitoring

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Lube Oil Analysis

Lube oil analysis serves two main purposes:

To assess oil quality to ensure the oil is fit for further

use i.e. meets the lubricating requirements for the

machine

To assess machine condition by

examining the oil for signs of

mechanical wear

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Lube Oil Sampling Guidelines Care should be taken when collecting lube oil samples to ensure the oil sample is representative of the oil circulating in the machine and to ensure that the sample is not contaminated. As a guide:

Oil samples should only be taken from running machines. Cold samples taken

from non-running machines will allow any particles contained within the oil to

separate and settle. Similarly, when taking samples from piping systems ensure

the sample is taken from a location where oil is circulating e.g. a bend where

flow is more turbulent as opposed to straight pipe where flow is laminar.

The sample must be taken from the same location each survey. Note: placing a

sampling tube too far into oil reservoir is likely to collect the debris at the

bottom of the tank which must be avoided.

Samples must be taken upstream of any filters.

Always use fresh tubing and bottles for each sample and ensure they are

properly stored to avoid contamination.

Reference oil samples should be taken whenever a new batch of oil is utilised.

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Lube Oil Analysis – Laboratory Analysis

Lube oil analysis is carried out on a periodic basis to compare the chemical and elemental properties of a used oil to a baseline of unused oil. Over time a trend is built up to determine the rate of deterioration of oil quality or abnormalities which could be indicative of mechanical wear.

Standard tests include:

Viscosity – too low a viscosity reduces oil film strength, weakening its ability to

prevent metal-to-metal contact.

Spectrographic Analysis – measures the concentration (normally in parts per

million (ppm)) of elements (e.g. lead, copper, sodium etc.) entrained in the oil

to determine wear metals, contaminants and additives.

Total Acid Number (TAN) – is used to measure the acidic content of the oil.

Total Base Number (TBN) – indicates the ability of an oil to neutralise acid.

TBN is an important test for diesel engines. A low TBN can indicate overdue oil

changes and overheating.

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Lube Oil Analysis – Laboratory Analysis Continued

Water Content – is a standard test used to assess the percentage of water

within the oil sample. Water content should not normally exceed 0.1%.

PQ Index (Particle Quotient) – gives an indication of ferrous wear debris in the

oil sample. The PQ Index can be easily trended over time and is normally

carried out as part of a standard analysis. A high PQ Index can indicate that

wear is present but ferrographic analysis is required to identify the type of

wear.

Ferrographic Analysis (also known as Wear Debris Analysis) – is a relatively

expensive test in comparison to standard analysis. It is used to assess the size,

shape and number of wear particles suspended in the oil sample. The size and

shape of the wear particles can be associated to specific wear modes indicative

of mechanical faults.

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Lube Oil Analysis – Example Laboratory Report Part 1

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Thermography

Thermography is a highly utilised technique in the condition monitoring of electrical switch gear but can also be used as an indicator of mechanical wear. It works on the principle that temperature changes occur as the condition of components alter e.g. electrical arcing and bearing wear.

Thermography measures infrared radiation emitted from different materials to allow the remote (non-contact) measurement of temperature and temperature differences.

Special viewing panels are often fitted to electrical switch gear enclosures to allow thermographic survey; as opening the enclosure would let the heat escape.

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Thermography - Applications

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Watch Keeping

Watch keeping is normally carried on a daily basis to record operating and process parameters around the plant. This is an ideal opportunity to walk around any machinery and use your senses to look and listen for any abnormalities in machine condition, for example:

Seal Leaks - vibration analysis cannot diagnose seal leaks however a very quick

visual inspection can.

Bearing wear and lack of lubrication – if a machines‟ bearings are worn or

inadequately lubricated they can emit audible noise. This can be confirmed

by vibration analysis.

Cavitation sounds like gravel is passing through the pump and will emit an

audible noise. This can be confirmed by vibration analysis.

Low pump discharge pressure or high motor current readings can indicate a

pump is not running efficiently.

Use Your Senses!

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Motor Phase Current Analysis

Motor Phase Current analysis is used for rotor fault diagnosis of AC induction

motors to detect broken rotor bars and air gap eccentricity. The technique is

based upon frequency analysis of the phase current supplying the motor.

Motor Phase Current Analysis is

generally applicable to large AC

induction motors.

Measurements can be carried

out using portable equipment

by connecting a current clamp

to the low voltage side of the

power transformer.

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Rogowski coil analysis is used to assess the condition of stator winding insulation in

high voltage electrical machines such as power generators. Assessment of

condition is based upon the examination of high frequency signals caused by

partial discharge activity measured using Rogowski coils.

Rogowski coil analysis is generally undertaken on larger, high voltage machines

such as electrical generators and motors of 6.6 kV or greater.

Rogowski coils can be fitted to machines during manufacture on the client‟s

request.

Specialist data collection equipment is required to interface to the fitted

instrumentation to carry out diagnostics.

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Performance Monitoring Performance monitoring can give an indication of machine condition and running efficiency based upon calculations of measured operating and process parameters.

Performance indicators can be utilised to:

Ensure the efficient running of power turbines and compressors; where

reduced efficiency can result is substantial financial loss.

Indicate when a gas turbine should be water washed.

Indicate deterioration in machine condition based upon a reducing trend in

performance.

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Performance Indicator Example – MOL Pump Calculating the ratio of the energy out of a system to the energy in, will give and indication of the system‟s efficiency. In this example the differential pressure multiplied by the flow rate of a pump was divided by the motor current. Over time it was seen that the pump performance indicator showed a deteriorating trend. The pump was overhauled and found to have badly worn impellers. After overhaul the performance indicator improved.

MOL Pump

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[email protected]

vibration monitoring and its features for corelation

Engineering