25.11 Fault Diagnostics

Back

Depending on the type of equipment being monitored and the maintenance strategy being followed,

once a faulty condition has been detected and the severity of the fault assessed, repair work or

replacement will be scheduled. However, in many situations, the maintenance strategy involves further

analysis of the vibration signal to determine the actual type of fault present. This information then allows

for a more accurate estimation of the remaining life, the replacement parts that are needed, and the

maintenance tools, personnel, and time required to repair the machinery. For these reasons, and many

more, it is often advantageous to have some idea of the fault type that exists before decisions regarding

maintenance actions are made.

There are obviously a large number of potential different fault types. The description of these faults can

be systemized somewhat by considering the type of characteristic defect frequencies generated

(synchronous to rotating speed, subsynchronous, harmonics related to rotating speed, nonsynchronous

harmonics, etc.). Such a systemization requires a focus on frequency-domain analysis tools (primarily

frequency spectra). While this organization strategy is effective, it inherently leaves out potentially

valuable information from other display formats. For this reason, the various faults that usually develop

in machinery are listed here in terms of the forcing functions that cause them and specific machine types.

In this way, a diagnostic template can be developed for the different types of faults that are common in a

given facility or plant. Further reading on machinery diagnostics can be found in these references:

Wowk (1991), Taylor (1994), Eisenmann and Eisenmann (1998), Goldman (1999), and Reeves (1999).

25.11.1 Forcing Functions

Listed and described below are a variety of forcing functions that can result in accelerated deterioration of

machinery or are the result of damaged or worn mechanical components. The list is not meant to be

exhaustive and is in no particular order.

* Fault detection can be defined as the departure of a measurement parameter from a range

that is known to represent normal operation. Such a departure signals the existence of a

faulty condition.

* ISO Technical Committee 108, Sub-Committee 5 is responsible for standards for Condition

Monitoring and Diagnostics of Machines.

* Standards are based on machinery type or vibration severity.

* The development and use of acceptance limits that are close to normal operating values for

specific machinery will detect even slight changes in condition.

* Statistical acceptance limits are set using statistical information calculated from

the vibration signals measured from the equipment that the limits will ultimately be

used with.

* Judging vibration characteristics within the frequency spectra is sometimes a more accurate

method of the early detecting and trending of fault conditions because the frequency

domain is generally more sensitive to changes in the vibration signal that result from

changes in machine condition.

* Frequency domain limits include limited band monitoring, constant bandwidth limits, and

constant percentage bandwidth limits.

Machine Condition Monitoring and Fault Diagnostics 25-25

© 2005 by Taylor & Francis Group, LLC

25.11.1.1 Unbalance

Unbalance (also referred to as imbalance) exists when the center of mass of a rotating component is not

coincident with the center of rotation. It is practically impossible to fabricate a component that is

perfectly balanced; hence, unbalance is a relatively common condition in a rotor or other rotating

component (flywheel, fan, gear, etc.). The degree to which an unbalance affects the operation of

machinery dictates whether or not it is a problem.

The causes of unbalance include excess mass on one side of the rotor, low tolerances during fabrication

(casting, machining, assembly), variation within materials (voids, porosity, inclusions, etc.), nonsymmetry

of design, aerodynamic forces, and temperature changes. The vector sum of all the different sources

of unbalance can be combined into a single vector. This vector then represents an imaginary heavy spot

on the rotor. If this heavy spot can be located and the unbalance force quantified, then placing an

appropriate weight 1808 from the heavy spot will counteract the original unbalance. If left uncorrected,

unbalance can result in excessive bearing wear, fatigue in support structures, decreased product quality,

power losses, and disturbed adjacent machinery.

Unbalance results in a periodic vibration signal with the same amplitude each shaft rotation (3608).

A strong radial vibration at the fundamental frequency, 1X, (1 £ rotational speed) is the characteristic

diagnostic symptom. If the rotor is overhung, there will also be a strong axial vibration at 1X. The

amplitude of the response is related to the square of the rotational speed, making unbalance a

dangerous condition in machinery that runs at high rotational speeds. In variable speed machines

(or machines that must be run-up to speed gradually), the effects of unbalance will vary with the shaft

rotational speed. At low speeds, the high spot (location of maximum displacement of the shaft) will be

at the same location as the unbalance. At increased speeds, the high spot will lag behind the unbalance

location. At the shaft first critical speed (the first resonance), the lag reaches 908, and at the second

critical and above, the lag reaches 1808.

A special form of unbalance is caused by a bent shaft or bowed rotor. These two conditions are

essentially the same; only the location distinguishes them. A bent shaft is located outside the machine

housing, while a bowed rotor is inside the machine housing. This condition is seen on large machines

(with heavy rotors) that have been allowed to sit idle for a long time. Gravity and time cause the natural

sag in the rotor to become permanent.

The vibration spectrum from a machine with a bent shaft or bowed rotor is identical to

unbalance, largely because it is an unbalanced condition. Bent shafts and bowed rotors are difficult

to correct (straighten), so they need to be balanced by adding counterweights as described above.

The best way to avoid this condition is to keep the shaft/rotor rotating slowly when the machine is

not in use.

25.11.1.2 Misalignment

While misalignment can occur in several different places (between shafts and bearings, between gears,

etc.), the most common form is when two machines are coupled together. In this case, there are two

main categories of misalignment: (1) parallel misalignment (also known as offset) and (2) angular

misalignment. Parallel misalignment occurs when shaft centerlines are parallel but offset from one

another in the horizontal or vertical direction, or a combination of both. Angular misalignment

occurs when the shaft centerlines meet at an angle. The intersection may be at the driver or driven end,

between the coupled units or behind one of the coupled units. Most misalignment is a combination of

these two types.

Misalignment is another major cause of excessive machinery vibration. It is usually caused

by improper machine installation. Flexible couplings can tolerate some shaft misalignment, but

misalignment should always be minimized.

The vibration caused by misalignment results in excessive radial loads on bearings, which in turn

causes premature bearing failure. Elevated 1X vibrations with harmonics (usually up to the third, but

sometimes up to the sixth) in the frequency spectrum are the usual diagnostic signatures. The harmonics

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allow misalignment to be distinguished from unbalance. High horizontal relative to vertical vibration

amplitude ratios (greater than 3:1) may also indicate misalignment.

One final note regarding misalignment is that the heat of operation causes metal to expand resulting in

thermal growth. Vibration readings should be taken when the equipment is cold and again after normal

operating temperature has been reached. The changes in alignment due to thermal growth may be

minimal, but should always be measured since they can lead to significant vibration levels.

Because unbalance and misalignment are perhaps the two most common causes of excessive

machinery vibrations and they have similar characteristic indicators, Table 25.1 has been included to help

distinguish between them.

25.11.1.3 Mechanical Looseness

While there are many ways in which mechanical looseness may appear, there are two main types: (1) a

bearing loose on a shaft and (2) a bearing loose in a housing. A bearing that is loose on a shaft will display

a modulated time signal with many harmonics. The time period of modulation will vary and the time

signal will also be truncated (clipped). A bearing that is loose in its housing will display a strong fourth

harmonic, which can sometimes be mistaken for the blade-pass frequency on a four-blade fan. These

faults may also look like rolling-element-bearing characteristic defect frequencies, but always contain a

significant amount of wideband noise.

Another way to diagnose mechanical looseness is by tracking the changes in the vibrations signal as

the condition worsens. In the early stages, mechanical looseness generates a strong 1X response in the

frequency spectrum along with some harmonics. At this stage, the condition could be mistaken for

unbalance. As the looseness worsens, the amplitude of the harmonics will increase relative to the 1X

response (which may actually decrease). The overall RMS value of the time waveform may also decrease.

Further deterioration of the condition results in fractional harmonics

􀀄

1

2

; 1

3

; 1 1

2

; 2 1

2

􀀅

increasing in

amplitude. These harmonics are most visible in signals taken when the machine is only lightly loaded.

These harmonics show up because of the clipping described above.

25.11.1.4 Soft Foot

Another condition that is in fact a type of mechanical looseness, but often masquerades as misalignment,

unbalance, or a bent shaft, is soft foot. Soft foot occurs when one of a machine’s hold-down bolts is not

tight enough to resist the dynamic forces exerted by the machine. That part of the machine will lift off and

set back down as a function of the cyclical forces acting on it. All the diagnostic signs associated with

mechanical looseness will be present in the vibration signal.

If the foundation (hold-down points) of a machine does not form a plane, then tightening the holddown

bolts will cause the casing and/or rotor to be distorted. This distortion is what leads to the

misalignment, unbalance, and bent shaft vibration signatures. In order to check for a soft foot,

TABLE 25.1 Characteristics that Can Help Distinguish between Unbalance and

Misalignment

Unbalance Misalignment

High 1X response in frequency spectra High harmonics of 1X relative to 1X

Low axial vibration levels High axial vibration levels

Measurements at different locations

are in phase

Measurements at different locations are

1808 out of phase

Vibration levels are independent of

temperature

Vibration levels are dependent on

temperature (change during warm-up)

Vibration level at 1X increases with

rotational speed. Centrifugal

force increases as the square

of the shaft rotational speed

Vibration level does not change with

rotational speed. Forces due to

misalignment remain relatively

constant with changes in shaft

rotational speed

Machine Condition Monitoring and Fault Diagnostics 25-27

© 2005 by Taylor & Francis Group, LLC

the vibration level must be monitored while each hold-down bolt is loosened and then retightened.

The appearance and/or disappearance of the diagnostic indicators mentioned above will determine if soft

foot is the problem. When a machine’s vibration levels cannot be reduced by realignment or balancing,

soft foot could well be the cause.

25.11.1.5 Rubs

Rubs are caused by excessive mechanical looseness or oil whirl. The result is that moving parts come into

contact with stationary parts. The vibration signal generated may be similar to that of looseness, but is

usually clouded with high levels of wideband noise. This noise is due to the impacts. If the impacts are

repetitive, such as occurring each time a fan blade passes, there may be strong spectral responses at the

striking frequency.

In many cases, rubs are the result of a rotor pressing too hard against a seal. In these cases, the rotor will

heat up unsymmetrically and develop a bowed shape. Subsequently, a vibration signal will be generated

that shows unbalance. To diagnose this condition, it will be noted that the unbalance is absent until the

machine comes up to normal operating temperature.

25.11.1.6 Resonances

The analysis of resonance problems is beyond the scope of this chapter. However, some basic description

is provided here because of the high likelihood that at some time a resonance will be excited by repetitive

or cyclic forces acting on or nearby a machine. A resonance is the so-called “natural frequency” at which

all things tend to vibrate. A machine’s natural resonant frequency is dictated by the relationship vn ¼

ðk=mÞ1=2; where vn is the natural frequency, k is the spring stiffness, and m is the mass. Most systems will

have more than one resonance frequency. These resonances (also called modes) can be excited by any

forcing function that is at or close to that frequency. The response amplitude can be 10 to 100 times that

of the forcing function. The term “critical speed” is also used to refer to resonances when the machine

rotating speed equals the natural frequency.

The amount of response amplification depends on the damping in the system. A highly damped

system will not show signs of resonance excitation, while a lightly damped system will be prone to

resonance excitations. Resonances can be diagnosed by monitoring the vibration level while the speed of

rotation of the machine is changed. A resonance will cause a dramatic increase in the 1X vibration levels

as the speed is slowly changed. Most machines are designed to operate well away from known resonance

frequencies, but changes to the machine (support structure, piping connections, etc.) and proximity to

other machines may excite a resonance.

25.11.1.7 Oil Whirl

Oil whirl occurs when the fluid in a lightly loaded journal bearing does not exert a constant force on the

shaft that is being supported and a stable operating position is not maintained. In most journal bearing

designs, this situation is prevented by using pressure dams or tilt pads to insure that the shaft rides on an

oil pressure gradient that is sufficient to support it. During oil whirl, the shaft pushes a wedge of oil in

front of itself and the shaft then migrates in a circular fashion within the bearing clearance at just less than

one half the shaft rotational speed. The rotor is actually revolving around inside the bearing in the

opposite direction from shaft rotation.

Because of the inherent instability of oil whirl, in many situations where oil whirl occurs, the time

waveform will show intermittent whirl events. The shaft makes a few revolutions while whirl is present

and then a few revolutions where the whirl is not present. This “beating” effect is often evident in the time

waveform and can be used as a diagnostic indicator.

Persistent oil whirl usually requires a replacement of the bearing. However, temporary measures to

mitigate the detrimental effects include changing the oil viscosity (changing the operating temperature or

the oil), running the machine in a more heavily loaded manner, or introducing a misalignment that will

load the bearing asymmetrically. This last course of action is of course not recommended for more than

relatively short-term relief.

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© 2005 by Taylor & Francis Group, LLC

25.11.1.8 Oil Whip

Oil whip occurs when a subsynchronous instability (oil whirl) excites a critical speed (resonance), which

then remains at a constant frequency regardless of speed changes. Oil whip often occurs at two times the

critical speed because, at that speed, oil whirl matches the critical speed. Figure 25.15 shows a waterfall

(cascade) plot of a mass unbalance that excites oil whirl and oil whip. Note how the oil whip “locks on” to

the critical speed resonance.

25.11.1.9 Structural Vibrations

Structural vibrations can range dramatically in amplitude and frequency. Large-amplitude, lowfrequency

vibrations can be excited in multistory buildings during an earthquake or by the wind. These

vibrations are usually the result of a building resonance being excited. While these sources of structural

vibration are important, the source that we are concerned with here is that of machinery operating as part

of a building’s utility system, as part of the production plant, or construction equipment close-by. Fans,

blowers, compressors, piping systems, elevators, and other building service machines all produce

vibrations in a building and, if they are not properly isolated they can cause disruption and/or damage to

other machines or processes operating close-by. The same is true of heavy machinery operating within a

plant (stamping machines, presses, forges, etc.) and construction equipment. High-impact and repetitive

vibrations can excite resonances large distances from the source of the excitation.

25.11.1.10 Foundation Problems

Machine foundations provide rigidity and inertia so that the machine stays in alignment. The energy

generated by a machine in the form of vibrations is transmitted, reflected, or absorbed by the foundation.

Especially on larger machines, the foundation is paramount to successful dynamic behavior. Maximum

energy is transmitted through the foundation to the earth when the mechanical impedance of the

foundation is well matched to that of the source of vibration. That is, the source of vibration and the

foundation should have the same natural frequency. If this is the case, all frequencies of vibration below

Rotor Speed Oil Whirl OIL WHIP Mass Unbalance

Critical Speed Frequency

FIGURE 25.15 Waterfall (cascade) plot of a mass unbalance that excites oil whirl and oil whip.

Machine Condition Monitoring and Fault Diagnostics 25-29

© 2005 by Taylor & Francis Group, LLC

the natural frequency will be transmitted by the foundation to earth. A poor match will mean that more

energy is reflected or absorbed by the foundation, which could effect the operation of the machine

attached. Changing foundations can grossly affect amplitude and phase measurements, which means that

vibration measurements can be used to easily detect a changing foundation or hold-down system.

25.11.2 Specific Machine Components

25.11.2.1 Damaged or Worn Rolling-Element Bearings

Rolling-element bearings produce very little vibration (low level random signal) when they are fault free,

and have very distinctive characteristic defect frequency responses (see Eschman, 1985, for the equations

for calculation of defect frequencies) when faults develop. This, and the fact that most damage in rollingelement

bearings occurs and worsens gradually, makes fault detection and diagnosis on this component

relatively straightforward. Faults due to normal use usually begin as a single defect caused by metal

fatigue in one of the raceways or on a rolling element. The vibration signature of a damaged bearing is

dominated by impulsive events at the ball or roller passing frequency. Figure 25.16 shows the

characteristic time waveform and frequency spectra at various stages of damage. As the damage worsens,

there is a gradual increase in the characteristic defect frequencies followed by a drop in these amplitudes

and an increase in the broadband noise. In machines where there is little other vibration that would

contaminate or mask the bearing vibration signal, the gradual deterioration of rolling-element bearings

can be monitored by using the crest factor or the kurtosis measure (see above for definitions).

A key factor in being able to accurately detect and diagnose rolling-element-bearing defects is the

placement of the vibration sensor. Because of the relatively high frequencies involved, accelerometers

should be used and placed on the bearing housing as close as possible to, or within, the load zone of the

stationary outer race.

Specific applications can also pose significant challenges to fault diagnosis. Very low-speed machines

have bearings that generate low energy signals and require special processing to extract useful bearing

condition indications (Mechefske and Mathew, 1992a). Machines that operate at varying speeds also

pose a problem because the characteristic defect frequencies are continuously changing (Mechefske and

Liu, 2001). Bearings located close to, or within, gearboxes are also difficult to monitor because the high

energy at the gear meshing frequencies masks the bearing defect frequencies (Randall, 2001).

Time Domain

Frequency Domain

Amplitude Amplitude

FIGURE 25.16 Characteristic time waveform and frequency spectra at various stages of damage in a rolling-element

bearing.

25-30 Vibration and Shock Handbook

© 2005 by Taylor & Francis Group, LLC

25.11.2.2 Damaged or Worn Gears

Because gears transmit power from one rotating

shaft to another, significant forces are present

within the mating teeth. While gears are designed

for robustness, the teeth do deflect under load and

then rebound when unloaded. The local stresses are

high at the tooth interface and root, which leads to

fatigue damage. Proper design and perfect fabrication

of gears (with perfect form and no defects)

would result in relatively low vibration levels and a

long life. However, the presence of nonperfect gears

gives rise to excessive vibration (Smith, 1983).

The time waveform, the frequency spectral, and

the cepstral patterns generated by gear vibrations

all contain critical information needed to diagnose

defects (see Figure 25.17). In relatively simple

gearboxes, the time waveform can be used to

distinguish impacts due to cracked, chipped, or

missing teeth (McFadden and Smith, 1984, 1985).

The frequency spectra and cepstra are powerful

tools when the gearbox contains several sets of

mating gears, which is most often the case.

Even a significant defect on one tooth (or even a missing tooth) often does not produce an abnormally

strong frequency spectral response at 1X. However, the defect will modulate the gear mesh frequency

(number of teeth times the shaft rotational speed) and appear as 1X sidebands of the gear mesh

frequency. That is, smaller spectral responses that appear a distance of 1X (and multiples of 1X for more

severe gear faults) above and below the gear mesh frequency. Because these sidebands occur at multiples

of 1X and a spectral plot can become quite cluttered with response lines, cepstral analysis is well suited to

distinguish the frequency components that are strong fault indicators. Often, a change in the response at

two times the gear mesh frequency is a good indicator of developing gear problems. The amplitude of the

gear mesh frequency, and its multiples, vary with load. This makes it important to sample the vibration

signal at the same load conditions. When unloaded, excessive gear backlash may also cause an increase in

the amplitude of the gear mesh frequency.

Because each gear tooth meshes with an impact, structural resonances may be excited in the gears,

shafts, and housing. Proper design of a gearbox will minimize this effect, but resonances in gearboxes

cause accelerated gear wear and should be monitored.

Gears provide an excellent example of how machines must wear-in during early use. New gears

will have defects that are quickly worn away in the machine’s early life. Vibration levels will become steady

and only increase gradually later in the machines life as the gears wearout. These gradual increases in

vibration level are normal. Sudden changes in vibration levels (at gear mesh frequency, two times gear

mesh frequency, or sidebands), especially decreases, are very significant. A drop in the vibration level

usually means a decrease in stiffness, and that more of the transmission forces are being absorbed due

to bending of the gear teeth. Catastrophic failure is imminent. Premature gear failures are usually

a symptom of other problems such as unbalance, misalignment, bent shaft, looseness, improper

lubrication, or contaminated lubrication.

25.11.3 Specific Machine Types

25.11.3.1 Pumps

There are two principal types of pumps: (1) centrifugal pumps and (2) reciprocating pumps.

Reciprocating pumps will be discussed in a later section. The sources of vibration in pumps are widely

Amplitude

Quefrency

Rahmonics

Amplitude

Frequency

Harmonics

Sidebands

FIGURE 25.17 Frequency and quefrency plots

(damaged gear).

Machine Condition Monitoring and Fault Diagnostics 25-31

© 2005 by Taylor & Francis Group, LLC

varied. In addition to the standard mechanical problems (unbalance, misalignment, worn bearings, etc.),

problems that are particular to pumps include vane-pass frequency generating conditions (starvation,

impeller loose on the shaft, impeller hitting something) and cavitation.

Starvation occurs when not enough liquid is not present to fill each vane on the impeller every

revolution of the shaft. Pump starvation can be confused with unbalance (see Chapter 34). However, it

can be distinguished by the varying amplitude 1X vibration at constant speed and the reduced load on

the driving motor.

When the vanes on the impeller are striking something, the vane-pass frequency (the number of vanes

times the rotational speed) is excited. Because the striking causes a force on the shaft, an unbalance is also

present. The frequency spectrum will show a response at 1X and vane-pass frequency. The time waveform

will show a high-frequency response (vane pass) riding on a frequency response at 1X. The vane-pass

frequency is in phase with the shaft speed. If the impeller is loose on the shaft, the vane-pass frequency

will be modulated by the shaft speed.

Cavitation occurs when there is sufficient negative pressure (suction) acting on the liquid in the system

that it becomes a gas (it boils). This usually takes place in localized parts of the system. Cavitation usually

occurs in a pump when the suction intake is restricted and the liquid vaporizes when coming off the

impeller. As the fluid moves past the low pressure region, the gas bubble collapses. If the collapsing

bubble is close to a solid surface, it will aggressively erode the surface. Cavitation may be caused by a local

decrease in atmospheric pressure, an increase in fluid temperature, an increase in fluid velocity, a pipe

obstruction, or abrupt change in direction. The vibration signal that results will have significant vibration

levels at 1X with harmonics and strong spectral responses at vane-pass frequency. High-frequency

broadband noise is also common. An increase in the system pressure can reduce cavitation.

Hydraulic unbalance will result if there has been poor design of suction piping (elbow close to inlet) or

poor impeller design (unsymmetrical). The vibration signal will contain high 1X axial vibration

components. Impeller unbalance is a specific form of mechanical unbalance as discussed above. High 1X

vibration levels will result. Pipe stresses result from inadequate pipe support and cause stress on the pump

casing. This may also cause misalignment. Pipe resonances can also be excited by vane-pass frequency

pressure pulsations.

Diagnosis of pump problems can be improved by installing a pressure transducer in the discharge line

of the pump. The measured pressure fluctuations can be processed in the same way as vibration signals.

The frequencies measured represent the pressure fluctuations and the amplitude is the zero-to-peak

pressure change.

25.11.3.2 Fans

Fans account for a significant number of field vibration problems due to their function and construction.

Fans move air or exhaust gases that are often laden with grease, dust, sand, ash, and other corrosive and

erosive particles (also see Chapter 34). Under these conditions, fans blades gain and lose material

resulting in the need to regularly rebalance. The level of balance must also be relatively fine because fans

often have large fan-blade diameters and operate at relatively high speeds. Fans are usually mounted on

spring/damper systems to help isolate vibrations, but they are also constructed in a relatively flexible

manner, which adds to the demands for fine balancing. Along with fine balancing requirements, typical

problems include looseness, misalignment, bent shaft, and defective bearings.

Fans also generate a strong response at blade-pass frequency (number of blades times the shaft

rotational speed). This frequency response is present during normal operation, but it can become

elevated if the blades are hitting something, the fan housing is excessively flexible, or an acoustical

resonance is present. Acoustical resonances are relatively common where large volumes of air are being

moved through large flexible ducts and/or the fan blades are of an air-foil design.

25.11.3.3 Electric Motors

Electric motors can be divided into two groups: (1) induction motors and (2) synchronous motors. A full

description will not be given here as to the differences. Like any machine, electric motors are subject to a

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full range of mechanical problems, and vibrations signals can be used to detect and diagnose these

problems. Apart from the conditions described elsewhere in this section, there are some problems that

occur only in electric motors. For the sake of brevity, these problems and the vibration signals that

typically accompany them are summarized in Table 25.2.

25.11.3.4 Steam and Gas Turbines

Steam and gas turbines (and high-speed compressors) require special mention because of the high speeds

and temperatures involved. Problems on steam turbines are usually limited to looseness, unbalance,

misalignment, soft foot, resonance, and rubs. As discussed above, each of these conditions has a set of

characteristic vibration responses that allow for relatively straightforward diagnosis. However, because of

the high speeds, this type of machinery is usually designed to be lighter and less rigid than other rotating

machines. Excessive vibration can therefore lead to catastrophic failure very quickly. Because of this,

high-speed turbines and compressors are designed to closer tolerances than other types of machines, and

extra care is taken when balancing rotors. These machines also frequently operate above their first critical

speed and sometimes between their second and third critical speeds. At these speeds, the rotor becomes

quite flexible and the support bearings become very important in that they must provide the appropriate

amount of damping.

Because steam and gas turbines are supported on journal bearings, most monitoring and diagnostics

work will be based solely on proximity probe signals. While this is not a problem in and of itself,

accelerometer signals should also be taken in order to cover the higher frequencies, which are excited by

conditions such as looseness and rubs.

25.11.3.5 Compressors

Compressors act in much the same way as pumps, except that they are compressing some type of gas. They

come in many different sizes, but only two principal types: (1) screw-type and (2) reciprocating

compressors. Reciprocating compressors will be discussed in a later section. Screw-type compressors have a

given number of lobes or vanes on a rotor and generate a vane-passing frequency. Screw compressors

with multiple rotors can also generate strong 1X and harmonics up to vane-pass frequency. The close

tolerances involved result in relatively high vibration levels, even when the machine is in good condition. As

with pumps, signals taken from pressure transducers in the discharge line can be useful for diagnostics.

25.11.3.6 Reciprocating Machines

Reciprocating machines (gas and diesel engines, steam engines, compressors, and pumps) all have one

thing in common — a piston that moves in a reciprocating manner. These machines generally have

high overall vibration levels and particularly strong responses at 1X and harmonics, even when in

good condition. The vibrations are caused by compressed gas pressure forces and unbalance.

Vibrations at 1

2 X may be present in four-stroke engines because the camshaft rotates at one half the

crankshaft speed.

TABLE 25.2 Mechanical Problems Particular to Electric Motors

Condition Vibration Indicator

Motor out of magnetic center High spectral response at 60 Hz

Motors with broken rotor bars High spectral response at motor running speed

and/or second harmonic

Motor with turn-to-turn shorts

in the windings

Motor runs at a slower than expected speed

(high slit frequency)

Motor out of magnetic center with

broken rotor bars or turn-to-turn

shorts in the windings

Side bands of slip frequency times the number of

poles centered around the motor running speed

and harmonics of the running speed

Machine Condition Monitoring and Fault Diagnostics 25-33

© 2005 by Taylor & Francis Group, LLC

Many engines operate at variable speeds, which will allow the strong forcing functions to excite

resonances of the components and the mounting structure, if it is not designed in a robust manner.

Excessive vibrations in reciprocating machines usually occur due to operational problems such as

misfiring, piston slap, compression leaks, and faulty fuel injection. These problems result in elevated 1

2 X

vibrations, if only one cylinder is affected, and a decrease in efficiency and power output. Gear and

bearing problems may also occur in reciprocating machines, but the characteristic defect frequencies for

these faults are significantly higher.

25.11.4 Advanced Fault Diagnostic Techniques

Much of the discussion in the previous sections has highlighted the fact that many machine defects

generate distinctive vibration signals. This fact has been exploited recently with the development

of a variety of different automatic fault diagnostics techniques (Mechefske and Mathew, 1992b;

Mechefske, 1995). The details of these systems will not be provided here, but the goal of automatic

diagnostics is to augment and assist, rather than replace, the vibration signal analyst. If characteristic

defect indicators can be detected and extracted from a vibration signal without the

intervention of a signal analyst, the analyst will have more time for other duties and will also have

access to information that may not have been uncovered through normal signal processing and

analysis.

There are, however, still many situations where machine defects do not generate distinctive vibration

signals or when the vibration signals are masked by large amounts of noise or vibrations from other

machinery. In such cases, advanced diagnostic algorithms incorporating new signal processing

techniques are currently being developed and implemented. Artificial neural networks (Timusk and

Mechefske, 2002) have been found to provide an excellent basis for detecting and diagnosing faults.

Wavelet analysis (Lin et al., 2004) and short-time Fourier transforms (STFTs) have also been shown to

effectively allow both time domain and frequency domain information to be displayed on the same plot.

This provides an opportunity to clearly see short duration transient events as well as detect faults in

machinery that is operating in nonsteady-state conditions.

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