casos de vibracion

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DESBALANCE Desbalance.- Usualmente es el problema más simple para diagnosticar (y uno de los más comunes). El Desbalance es una fuerza centrífuga. Considere el siguiente caso alrededor de 3 pies (.91 metros) diámetro un ventilador que rota a 2000rpm: Circunferencia = 3 x 3.14159 = 9.42 pies (2.87m). 2000 rpm = 120,000 rev/hora (rph) 120,000 rph x 9.42 ft/rev = 1,130,400 ft/hour (344,546 m/hour) 1,130,400 ft/hr / 5280 ft/mile = 214.1 mph ó 344.6 km/hr Un desbalance de la masa (cualquier desbalance que sea) en el borde de un ventilador que viaja cerca de su maxima velocidad de un carro de carrera de la INDY. Adicionalmete recordemos: Fuerza = Masa x Velocidad cuadrado Plano Simple de Desbalance Figure 1 - Typical Radial FFT Generated By Unbalance En ausencuia de otros problemas, el desbalance causa una sinusoide pura (uno de los unicos problemas que no deforma la forma de alguna manera) y por lo tanto denera un pico de 1x rpm. Single-Plane Unbalance Symptoms: Radial vibration @ 1x rpm.

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Page 1: Casos de vibracion

DESBALANCE

Desbalance.- Usualmente es el problema más simple para diagnosticar (y uno de los más comunes). El Desbalance es una fuerza centrífuga. Considere el siguiente caso alrededor de 3 pies (.91 metros) diámetro un ventilador que rota a 2000rpm: Circunferencia = 3 x 3.14159 = 9.42 pies (2.87m). 2000 rpm = 120,000 rev/hora (rph) 120,000 rph x 9.42 ft/rev = 1,130,400 ft/hour (344,546 m/hour) 1,130,400 ft/hr / 5280 ft/mile = 214.1 mph ó 344.6 km/hr

Un desbalance de la masa (cualquier desbalance que sea) en el borde de un ventilador que viaja cerca de su maxima velocidad de un carro de carrera de la INDY. Adicionalmete recordemos:

Fuerza = Masa x Velocidad cuadrado

Plano Simple de Desbalance

Figure 1 - Typical Radial FFT Generated By Unbalance

En ausencuia de otros problemas, el desbalance causa una sinusoide pura (uno de los unicos problemas que no deforma la forma de alguna manera) y por lo tanto denera un pico de 1x rpm.

Single-Plane Unbalance Symptoms:

Radial vibration @ 1x rpm. Phase around bearing shifts with transducer shift - 90°

transducer shift causes 90° phase shift.

Little or no phase shift across or "between" bearings [bearings vibrating "in-phase"]· Síntomas de Desequilibrio Solos planos: · vibración Radial 1x revoluciones por minuto. · la Fase alrededor del porte de cambios(movimientos) con el cambio(movimiento) de transductor - 90 cambio(movimiento) de transductor ° causa 90 cambio de fase(desfasamiento) °. Poco o ningún cambio de fase(desfasamiento) a través "o entre" portes [portes que vibran "en fase"]

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Fig. 1. Simple Plane unbalanceObserve que al medida que la masa de desbalance se mueva se generara una fuerza en esa dirección. Moviendose los dos apoyos de rodajes en la misma dirección en el mismo tiempo.

TWO-PLANE UNBALANCE

Figure 1 - Typical Radial FFT Generated By Unbalance

Two-Plane Unbalance Symptoms:

Radial vibration @ 1x rpm. Phase around bearing shifts with transducer shift - 90°

transducer shift causes 90° phase shift.

Significant phase shift (> 60°) across or "between" bearings [bearings vibrating "out-of-phase"]

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Figuere 2 – Two-Plane UnbalanceEn este tipo de desbalance los dos apoyos de rodaje no se moveran en la misma dirección en un tiempo, sino que se geneararan fuerzas en direcciones no iguales en le tiempo.

OVERHUNG ROTOR UNBALANCE

Figure 1 - Typical Axial FFT Generated By Unbalance

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Figure 1 - Typical Radial FFT Generated By Unbalance

Overhung Rotor Unbalance Symptoms:

Radial vibration @ 1x rpm. Axial vibration @ 1x rpm.

Phase around bearing shifts with transducer shift - 90° transducer shift causes 90° phase shift.

Axial phase readings usually in-phase.

Radial phase readings may be out-of-phase.

Balancing may require use of axial phase readings.

Figure 2 - Overhung Rotor UnbalanceEn este tipo de desbalance los dos apoyos de rodaje no se moveran en la misma dirección en un tiempo, sino que se geneararan fuerzas en direcciones no iguales en le tiempo

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MISALIGNMENT

Misalignment - The most common vibration problem. Unlike unbalance, does not have a single vibration symptom. As a result, it should always be considered as a possibility.

Definition of Perfect alignment - Shaft centerlines are parallel and intersect.

Types of misalignment: Angular - Shaft centerlines intersect but are not parallel Offset - Shaft centerlines are parallel but do not intersect.

It is extremely unlikely that you will encounter a case of either pure angular or pure offset misalignment - it will always be a combination. That results in the wide variety of vibration symptoms.

Angular Misalignment

Figure 1 - Typical FFT Generated By Angular Misalignment

Definition: Shaft Centerlines Intersect But Are Not Parallel

Angular Misalignment Symptoms:

High axial vibration @ 1x rpm, possible harmonics at 2x & 3x.

2x rpm axial component may be as high or even higher than 1x component.

Radial vibration, probably lower amplitude than the axial, at 1x, 2x and 3x.

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Radial vibration will depend on where the shaft centerlines intersect the assembly centerline.

Axial phase across coupling shifts significantly (> 60°).

Figure 2 - Shaft Centerlines Intersect @ The Coupling. Note The Absence Of Coupling Movement And The High Radial & Axial Bearing Movement.

Figure 3 - Shaft Centerlines Intersect @ The Bearings. Note The High Radial Coupling Movement, The Low Radial & The High Axial Bearing Movement.

Offset Misalignment

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Figure 1 - Typical FFT Generated By Offset Misalignment Definition: Shaft Centerlines Are Parallel But Do Not Intersect

Offset Misalignment Symptoms:

High radiual vibration @ 1x rpm, harmonics at 2x & 3x.

2x rpm axial component may be as high or even higher than 1x component.

Axial vibration, probably lower amplitude than the axial, at 1x, 2x and 3x.

Radial phase across coupling shifts significantly (> 60°).

Axial phase across coupling shifts significantly (> 60°).

Figure 2 - Shaft Centerlines Do Not Intersect. Note The High Radial & Axial Bearing Movement

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Figure 3 - Shaft Centerlines Do Not Intersect. Note The Even Higher Radial & Axial Bearing Movement

Cocked Bearing / Shaft Bent Through Bearing

Cocked Bearing / Shaft Bent Through Bearing - Creates similar or even identical vibration symptoms (with the exception of phase) to misalignment - primarily angular misalignment (axial vibration). Must be diagnosed with axial phase analysis or inspected for.

Figure 1 - Typical FFT Generated By Cocked Bearing

Cocked Bearing Symptoms:

Vibration symptoms very similar to direct drive angular misalignment.

High axial vibration @ 1x rpm, harmonics at 2x & 3x.

2x rpm radial component often as high or higher than 1x component.

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Axial phase shift around the face of the bearing equal to change in transducer location.

Figure 2 - Cocked BearingSe produce un tambaleo

Bent Shaft @ Bearing

Figure 1 - Typical FFT Generated By Shaft Bent Through The Bearing

Shaft Bent Through Bearing Symptoms:

Vibration symptoms very similar to direct drive angular misalignment.

High axial vibration @ 1x & 2x rpm.

2x rpm radial component often as high or higher than 1x component.

Axial phase shift around the face of the bearing equal to change in transducer location (twisting action).

Radial phase shifts significantly on either side of bearing (> 60°). This can best be seen in Figure 2 below. When the shaft just to the right of the bearing is moving up, the

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shaft just to the left of the bearing is moving down and vice versa. This measurement, of course, requires a direct shaft reading with an attachment such as a shaft stick.

Figure 2 - Note The Axial, Twisting Action Of The Bearing

Figure 3 - Note The Axial, Twisting Action Of The Bearing

Looseness

Looseness - Not a vibration source but an amplifier. That means that when a component is loose, whatever forces are present will be able to move the affected components much more easily. If there are little or no forces present, however, vibration may only increase a very small amount. To understand this, imagine a perfect machine - no mechnical imperfections to cause any vibration. Now loosen the bolts holding down the feet and . . . nothing happens because there are no forces attempting to lift it off of its base.

Looseness can occur at a number of locations that affect the vibration measurements. They are:

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Bearing / Shaft (Bearing Looseness) Bearing / Housing (Bearing Looseness) Internal bearing clearances (Bearing Looseness) Adjacent, fastened surfaces (Structural) Areas of the base (Structural) Each, however, gives different likely symptoms.

Structural Looseness

Structural Looseness Symptoms:

High radial vibration @ 1x, 2x rpm (often higher at 2x) and possibly 3x (lower).

Amplitude may be extremely high in direction of looseness only (vertical or horizontal) - far higher than in the perpendicular radial direction.

Found easily with background vibration checks of adjacent surfaces.

Slow Motion Study can be a very useful tool in diagnosing this condition.

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Figure 2 - Looseness Allows Movement In The Direction Of The Looseness

Animation simulates a loose motor foot moving vertically under a slow motion study.

Note how vertical amplitudes in this case are far higher than horizontal amplitudes would be.

Foot lifts and drops once per shaft revolution. 2x rpm component may appear due to the bounce (shape of the time

domain plot - see time domain section for more information on that). Additional harmonics can be created due to the shape of the time

domain signal. As it changes from a sinusoid towards a square wave, more harmonics appear.

Bearing Looseness

Figure 1 - Typical Radial FFT Generated By Bearing Looseness

Structural Looseness Symptoms:

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High radial vibration harmonics of 1x.

Harmonics can stretch all the way across the spectrum in cases of severe looseness and can even generate half-harmonics in extreme cases (1.5, 2.5, 3.5, etc.).

Figure 2 - Bearing Looseness Generates More Of A "Square" Wave Than A Sinusoid. That Shape Creates Harmonics

Housing Distortion (Soft Foot, Pipe Stress, etc.)

Figure 1 - Typical Axial FFT Generated By Housing Distortion

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Figure 2 - Typical Radial FFT Generated By Housing Distortion

Figure 3 - Soft Foot Or Other Housing Distortion Such As Pipe Stress Can Cause Bearings Within A Component To Misalign And Can Throw Off Normal Clearances

Housing Distortion Symptoms:

High axial vibration @ 1x & possibly 2x rpm.

Axial phase analysis may show phase shift across bearings within component.

Axial phase analysis may show twisting bearing (like cocked bearing).

2x Line frequency on motors due to air gap variation, especially radially.

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Pumps / fans may develop clearance problems (vane or blade pass).

High axial vibration on non-direct drive components (belt drives, integral fans, etc.).

Pipe stress can develop similar symptoms on pumps, compressors & fans/bl

Resonance

Resonance is simply the natural frequency of a component or combination of components (assembly). All structures have a resonant frequency. If you impact the structure with enough force to make it move, it will vibrate briefly at its natural frequency. A structure will have a resonant frequency in each of its 3 directional planes (x, y and z, or as we call them, horizontal, vertical and axial). Resonance serves to amplify the vibration due to whatever vibration force is present at (or near) that resonant frequency. It is important to note that resonance does not cause vibration - it amplifies it. Resonance problems occur in two primary forms. They are:Critical speeds – occurs when a component rotates at its own natural frequency.

A "critical speed" is simply when the rotational speed (rpm) coincides with the natural frequency of the rotor (cpm).

The tiniest amount of residual unbalance (something that is always present) is enough to cause huge amounts of vibration when rotating at a critical.

Rotors that are sped up or slowed down slowly are susceptible to this (i.e. turbines). In these cases, the critical speed is usually well known.

The most common problem related to unknown critical speeds is probably belts. Belts rotating at their resonant frequency (or having a nearby source of excitation of that resonant frequency) can vibrate excessively and cause other problems. For example, if the natural frequency of the belts coincides with the rpm of the fan, the belts will vibrate at their natural frequency.

2nd and 3rd criticals also may occur if the rotor speed gets high enough.

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Structural resonances - This is far more common than a critical speed problem. It becomes a problem when some forcing frequency comes close (+/- 10%) to the resonant (natural) frequency of a structure.

The structure can be the machine housing itself or some nearby structure such as a hand rail or I-beam.

A common example of this is a vertical pump. Due to the lack of a support at the top of the unit, these typically have very low resonant frequencies (~ 300 cpm). While running, this is not a problem but during start-up or coast-down, the unit experiences a "shudder" as it passes through the structural resonance (this is not a critical speed - it is a structural resonant frequency).

The structure itself will vibrate excessively - do not confuse with a critical speed.

The "shape" of the structure's vibration is an important clue and is known as a "mode shape".

Testing for the structure's natural frequency is crucial (required) to confirming a resonance problem.

Resonance, once diagnosed, can be simple to correct. It can also be extremely complex and difficult

to correct. The trick is in the diagnosis. But how do you diagnose it ? One method for determining a critical speed is a "Coast Down/Start Up Plot". This plot consists of the 1x vibration amplitude being collected simultaneously with a 1x rpm phase reading as the machine coasts to a stop or goes from stopped to full running speed. This test requires a 1x rpm reference (from a photoeye or some other speed tracking signal) in order to track the amplitude and phase at that frequency. Two things are observed as the rotor passes through a critical:

The 1x rpm amplitude will increase until the rotor reaches it's critical and then decrease to the normal level as the speed continues to change.

Phase will shift 180° as the rotor passes through the critical. This is due to the rotor changing from a rigid rotor (while operating below it's critical) to a flexible rotor (while operating above it's critical). It practical terms, on a rigid rotor, the heavy spot pulls the rotor around as it rotates. On

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a flexible rotor, the heavy spot pushes the rotor around as it rotates.

Structural resonances can be first suspected by several characteristics:

Disproportionately high amplitude at a single frequency (the resonant frequency) in the direction in which the resonant frequency is being excited.

A "mode shape" analysis shows the structure vibrating in a way that models resonance. Those models are covered on the next page.

Neither of those characteristics confirms resonance as a problem. A test must be performed that actually determines the natural frequency of the structure in question - a "bump test". Although there are high-tech methods available for this test (and some work very well), this test can be as simple as bumping the structure (causing it to vibrate) while it is not running and measuring the response (i.e. the frequency it vibrates at). A simple method for doing this involves collecting a 2 second sample (time domain plot) while bumping the structure, measuring the period of one cycle and converting it to a frequency. The time sample may have to be adjusted depending on the resonant frequency being measured (longer sample for very low resonant frequencies, shorter sample for high frequencies).If the measured response of the structure (i.e. it's resonant frequency) is within about 10% of the forcing frequency (i.e. the rpm of the machine although it can be at any frequency), resonance should be considered a problem. The closer the two frequencies are, the more of a problem it is.To correct a resonance problem, there are 4 methods:

Stiffen the structure - This method raises the resonant frequency of the structure.

Add mass to the structure - This method lowers the resonant frequency.

Change exciting frequency - Change the speed of the machine.

Add a dynamic absorber to the structure - This method attaches the equivalent of a tuning fork to the structure. This attachment is tuned to have the same resonant

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frequency as the structure and sets up an out-of-phase signal that has the effect of cancelling out (reducing) the signal being generated by the structure. The dynamic absorber must be properly sized to handle the forces being generated.

Structural Resonance

Figure 1 - Relatively High Amplitudes Will Be Generated. The Closer The Exciting Frequency Is To The Structure's Resonant Frequency,

The Higher The Amplitude Will Be.

Structural Resonance Symptoms:

High (at times, extremely high) vibration in one direction. This is an important symptom - the vibration in one

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direction will be disproportionately high compared to the other directions.

The structure shape, mass and rigidity will determine what is proportionate and what is disproportionate. It could be as low as 2 or 3:1 or as high as 10 or 20:1.

The structure itself will also determine whether or not the vibration is high in more than one direction (i.e. vertical pumps tend to have very similar resonant frequencies in all radial directions all raidal directions have the same mass and structural stiffness).

Similar (identical) machines exhibiting similar vibration symptoms (as described above).

Shape analysis can be initially used to see if the shape fits one of the models shown above. This test simply involves plotting amplitude values taken along the structure to determine the "shape" in which it is moving, or vibrating. This does not confirm resonance.

Some test (i.e. bump test) must used to determine the actual structural resonant frequency(s). The existence of the above symptoms does not prove resonance, it only makes it one of the strong possibilities (looseness, for instance, can cause disproportionately high vibration in the direction of the looseness).

Shape analysis should be performed before attempting to stiffen, or brace, the structure to correct the problem.

In the case of a nearby structure (i.e. an I-beam), a clue will be that the structure will often be vibrating more than the machine itself at the resonant frequency.

Critical Speeds

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Structural Resonance Symptoms:

Radial vibration @ 1x rpm. Phase will shift 180° once range has been completely

passed through.

Vibration usually satisfactory when rotating sufficiently above or below critical - only when rotating near the critical is there a problem.

One test used for determining the critical speed of a rotor is the test shown below. By measuring the amplitude @ 1x rpm simultaneously with the phase at 1x rpm as speed is increased ("Start-Up" plot) or decreased ("Coast Down" plot), the critical speed can be determined. The amplitude spike accompanied by the phase shift indicates a critical speed. Click here to see an example of this test.

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Figure 2 - Shape Of A Supported Rotor Running At Its 1st Critical

Figure 3 - Shape Of A Supported Rotor Running At Its 2nd Critical

Figure 4 - Shape Of A Overhung Rotor Running At Its 1st Critical

Figure 5 - Shape Of A Overhung Rotor

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Critical Speed Test

Critical Speed - Rotating a rotor assembly at its resonant frequency. The closer the rotational speed is to the exact structural resonant frequency, the more the rotor will vibrate. When operating below a critical speed, an unbalance mass on the rotor will pull the rotor towards it. This is defined as being a 'rigid' rotor. When operating above a critical speed, an unbalance mass will push the rotor away from it. This is defined as being a 'flexible' rotor. This phenomenon is why, when passing through a critical, the rotor will experience a 180° phase shift (since the deflection is in the opposite direction) and is an important characteristic that helps identify a rotor's critical speed.

Plot Taken As Unit "Coasts-Down". Note The Amplitude Spike & Phase Shift

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Belt-Drive Problems

Belt-drive problems, which include shaft misalignment, pulley misalignment, belt wear, belt resonance, belts too tight, belts too loose, pulley eccentricity and bent shafts, can be relatively straight forward to detect but can be far more difficult to specifically diagnose and correct. That is mainly due to the wide variety of problems that can occur in the installation and assembling of the belt drive, the difficulty of doing field testing on belts and the possibility of other influences (i.e. the base) having some effect. It is important to realize that some of the belt-drive vibration problems listed above do NOT cause vibration at belt related frequencies. Problems due to the shafts or pulleys (misalignment, eccentricity, etc.) cause vibration at 1x rpm of the component with the problem (i.e. eccentric pulley on the fan causes vibration at 1x rpm of the fan). Worn belts, on the other hand, will cause vibration at harmonics of belt running speed.The good news, especially in the case of component (belt and pulley) wear, is that belts and pulleys are typically relatively easy to inspect and inexpensive to replace. The bad news is that outside of that, they're often difficult to correct. One positive development in recent years has been the availability of laser alignment units for belt drives for a moderate price. Unfortunately, in more cases than not the old string & straight edge is still the alignment method used for belt drives. The first step to identifying a belt problem is to determine the belt speed.Determining the Belt Speed:Obtaining belt speed can be a bit difficult but there are a few tricks. Some methods are listed here:Calculate it.

It can be calculated mathematically if you know some of the variables: belt length, pitch diameters, center distances, etc.; but usually that is not the case. The formulas are listed below.

Measure it.

Detecting it with a strobe light is very difficult since it is usually a slow flash rate and the mark used may be unreliable (lettering on the belt, etc.).

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A photoeye will be very accurate but will require proper setup and a mark applied to the belts.

A "lasertach" would be the best option for an accurate belt rpm since it does not require a traditional "mark" - a good one will operate on pattern recognition.

Estimate it. With a bit a practice and understanding of a simple technique, an analyst can actually extract the probable belt rpm from the spectrum. One important requirement for this technique to be successful - there must be at least some vibration at belt-related frequencies. The following steps should be used:

First, identify any driver and driven related peaks (1x rpm and harmonics). Label them or make a mental note of which peaks they are.

Second, imagine cutting the belt in half and wrapping it around one of the pulleys. How many times will it wrap around - twice ? three times ? This will give you a very rough estimate in your mind of the belt speed (if it wraps 3 times, the speed would be 1/3 of that pulley's speed).

Finally, do the following:

Display your velocity spectrum on a logarithmic scale.

Move your spectrum cursor to your estimated belt rpm and turn on the harmonics.

Move your cursor left and right in the smallest increments possible (some software allows movement of 1/10th of a line of resolution - this helps with identifying harmonics) and try to get the harmonics to line up on top of any significant but previously unidentified amplitude peaks.

If there are significant belt related peaks on the spectrum, you should be able to get them lined up at some point.

If you cannot find any pattern of previously unidentified peaks of significant amplitude, that means one of two things:

Either you do not have the spectrum resolution necessary, or;

There is no significant belt vibration (in which case, why do we need to know the belt rpm ?).

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Formulas for Calculating Belt Frequencies:

If you know:

The Belt Length

- and - Either of the Pulley RPMs

and Its Diameter:

Driver Pulley Diameter and Speed

orDriven Pulley Diameter and Speed

You can calculate belt RPM with the following:

3.14 x PS1 x PD1/BL = Belt RPM

- or -

3.14 x PS2 x PD2/BL = Belt RPM Variable Definitions: PS = Pulley rpm (PS1 = Driver Pulley Speed, PS2 = Driven Pulley Speed) PD = Pulley diameter (PD1 = Driver Pulley Dia., PD2 = Driven Pulley Dia) SD = Distance between shaft centers BL = Belt Length

- OR -If you only know the pulley sizes and diameters, you can roughly calculate belt length and plug it into the formula above by using the following:

Belt Length = 1.57 x (PD1 + PD2) + 2(SD)In other words, 2x the center to center distance plus 1/2 the circumference of each pulley will provide the belt length.

Belt-Drive Problems

Pulley Misalignment

Figure 1 - FFT Typical Of Pulley Misalignment (Which Can Also Be Caused By Shaft Misalignment As Shown In Figures 3 & 4). This Condition Often Results In High Axial Vibration At Both Components 1x RPM. This Is Due To The Axial "Pulling" Force Generated As The Belts Ride Up The Side Of The Pulleys In An Effort To Properly Align Themselves.

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Pulley Misalignment Symptoms:

High axial vibration @ 1x rpm on component A at component B frequency.

Uneven wear axially on pulleys and belts.

Belt-Drive Problems

Belt/Pulley Wear, Improper Tension & Belt Resonance

Figure 1 - Typical FFT Showing Belt/Pulley Wear Problems; Resonance Can Also Be A Problem If The Belt Resonance Coincides With One Of The Forcing Frequencies (Driver, Driven, Belt RPMs)

Belt / Pulley Wear, Belt Resonance Symptoms:

High radial vibration @ 2x, 3x, 4x & 5x belt rpm.

Excessive belt "flap" can often be seen.

Belts and/or pulleys will show excessive wear patterns, cracking, etc. if wear is the problem.

Belt tension may be a problem - belts shouldn't be too loose or too tight.

Belt resonant frequency can be checked by placing transducer on bearing (radially) and "twanging" the belt like a guitar string while collecting a time domain or spectrum

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Belt-Drive Problems

Pulley Eccentricity / Bent Shaft (Near Pulley)

Figure 1 - Typical FFT Showing Pulley Eccentricity / Bent Shaft Near Pulley

Eccentric Pulley / Bent Shaft Near Pulley Symptoms:

High radial vibration @ 1x on both components - can easily be misdiagnosed as unbalance.

Belts act as rubber bands being stretched and relaxed - "reaction" forces - cannot be corrected through balancing of the component.

Directional vibration far higher parallel to belts than perpendicular to belts.

Phase will show 0° or 180° phase shift around bearing

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Figure 2 - Eccentricity Causes High Vibration At 1x RPM Of The Problem Component. Bent Shaft Near Pulley Causes Same Symptom

Sleeve Bearing Problems

Sleeve Bearing Problems – Sleeve bearings are in some ways much more forgiving and easier to analyze than rolling element bearings since there are no fundamental defect frequencies and the like to analyze. However, sleeve bearings also demand different techniques and insights that do not apply to rolling element bearings. For instance:

Measuring vibration on the housing of a sleeve bearing is unreliable since the housing moves only a small fraction (perhaps 10% or even less) of what the shaft is moving.

Vibration is due to mechanical forces being generated by the machine's rotation. In the absence of such forces (slow rotational speeds combined with excellent alignment and balance, for example), extensive wear can take place with absolutely no indication on a vibration spectrum - especially if the readings are taken on the housing.

Unlike greased bearings, sleeve bearings usually have an oil system. If the oil flow stops or the oil becomes severely contaminated, failure can occur very quickly.

What should be done with sleeve bearings to alleviate these concerns ?

Oil analysis - This will monitor bearing condition far more accurately than vibration analysis will.

Direct Shaft Vibration Readings - Although sometimes impractical or impossible, taking readings with a proximity probe, a shaft stick or shaft rider will give far more useful

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vibration data than readings taken on the housing since these techniques measure what the shaft is doing - not the housing.

Time Domain - Looking at the raw time signals will give information on exactly how the shaft is moving and give visual notice of problems such as rubs that spectra will not give.

Sleeve Bearing Problems:

Bearing Wear (Looseness)

Figure 1 - FFT Showing Sleeve Bearing Looseness

Sleeve Bearing Looseness Symptoms

The precise symptoms detected and amplitudes recorded on a spectrum will depend on the amount of force being generated by the shaft's rotation, where we are taking the readings and other variables.

Even if direct shaft readings are taken, if there is not enough force being generated to cause the shaft to throw itself around as in Figure 2, the shaft will simply spin as the bearing continues to wear and the clearances continue to increase. In this case, vibration symptoms of the problem will be minimal or even non-existent.

If the readings are taken on the housing instead of the shaft, you may be measuring only 10% or so of shaft movement and the chances are even greater that

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vibration symptoms of bearing wear will not be generated. Other factors now involved include the relative masses of the rotor and bearing housing / structure (how much can the relatively lightweight shaft move the massive housing ?).

It is important to understand that vibration is not monitoring bearing condition as it is with rolling element bearings. It is monitoring a result of the bearing wear - looseness - that does not cause vibration. Looseness merely allows the forces present to have more of an effect than they would if everything was properly fastened in place. If there are insufficient forces to throw the rotor around, vibration symptoms are not generated.

Oil analysis - which monitors oil properties, contaminants and wear metals, is the best predictive tool to use for sleeve bearing systems.

Most Common Symptoms:

1. High radial vibration @ 1x and numerous harmonics of rpm - like bearing looseness. In severe cases, peaks may appear at 1/2 harmonics (0.5 x rpm, 1.5 x rpm, etc.).

Figure 2 - Looseness Allows Signal Shape To Become More Of A Square Wave. This Causes Harmonics On The FFT

Sleeve Bearing Problems:

Oil Whirl

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Figure 1 - FFT Resulting From Oil Whirl (0.42-0.48xRPM Peak Indicates Oil Whirl)

Oil Whirl Symptoms

High Radial Vibration at 0.42 - 0.48 x RPM.

Oil Whirl, although unusual, can occur when clearances become excessive. An oil wedge is formed that is held in place by the rotation of the shaft. The friction of the shaft against the wedge then pushes the shaft around the housing. Fortunately (for the analyst), it occurs in a very precise sub-synchronous frequency range. Note that in the animation above the shaft is rotating at a different frequency than it is moving around the bearing sleeve

Figure 2 - Note Shaft Is

Spinning At Different Rate Than It Is Rotating Around Bearing Sleeve

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Sleeve Bearing Problems:

Oil Whip

Figure 1 - Oil Whirl Is Present, Rotor Passes Through Its 1st Critical

Figure 2 - Rotor Speed Continues To Increase Until Its Critical Speed Coincides With The Oil Whirl Frequency (i.e. Its 1st Critical Equals About 0.45x Current RPM)

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Figure 3 - Rotor Speed Continues To Increase But Destructively High Vibration Remains At Frequency Of Rotor's 1st Critical Speed

Oil Whip Symptoms

Oil whirl is present (bearing clearances are excessive).

Problem develops when rotor is running at 2.1-2.4x critical speed (at this speed, the frequency of the rotor's 1st critical is between 0.42-0.48xRPM - the oil whirl range)

High vibration develops at frequency of rotor's critical speed. This occurs when the vibration due to the oil whirl condition acts to excite the resonant frequency of the rotor.

High vibration remains at frequency of 1st critical even as rotor speed continues to increase.

Figure 2 - Oil Whirl. Note Shaft Is Spinning At Different Rate Than It Is Rotating Around Bearing Sleeve

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Rubs

Figure 1 - One Possible FFT Resulting From A Rub. Unpredictable Plot Due To Wave Shape; Time Domain Plots Essential)

Rub Symptoms:

Time domain the easiest way to diagnose - a 'truncated' signal is produced (see animation).

Readings should be taken at several radial positions (would you see the above signal shape with horizontal readings ?).

FFT can produce numerous harmonics of rpm (like bearing looseness) but also sub-harmonics at 1/2 x RPM in severe cases due to the wave shape (unpredictable results).

With a short enough time period collected (~2 shaft rotations), the length of the rub can be estimated.

Highest amplitude harmonic may be one that is the closest to one of the component's resonanc

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Figure 2 - Rotor Is Striking Something (i.e. The Housing) In The Vertical Direction. Note The Signal Shape. How Will The FFT Treat This Signal Shape ?

Rub Symptoms:

Time domain the easiest way to diagnose - a 'truncated' signal is produced (see animation).

Readings should be taken at several radial positions (would you see the above signal shape with horizontal readings ?).

FFT can produce numerous harmonics of rpm (like bearing looseness) but also sub-harmonics at 1/2 x RPM in severe cases due to the wave shape (unpredictable results).

With a short enough time period collected (~2 shaft rotations), the length of the rub can be estimated.

Highest amplitude harmonic may be one that is the closest to one of the component's resonance

Rolling Element Bearing Problems

Assessing the condition of rolling element bearings is arguably the single most important job vibration analysts have. Unfortunately, the vibration symptoms generated by a bearing going bad can vary greatly. However, bearings usually undergo a fairly predictable series of symptoms as they deteriorate. Considering the importance of the task and to enhance the analyst's chances of catching a bad bearing, it is important to use all of the tools at your disposal. These include:

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Velocity or, preferably, acceleration spectra that cover the frequency range between 30,000 and 120,000 cpm.

Enveloped spectra such as ESP, gSE, HFD, etc. These spectra are sensitive to the impact energy a developing bearing defect generates (a ball or roller striking a defect is similar to a car hitting a pothole in the road – impact energy is created).

Time domain will show the impacts better than the spectrum - especially on slow speed equipment.

Since most analysts use velocity spectrums to analyze data, we will focus on the 'normal' progression that occurs on velocity spectra. The advantage of using acceleration units is that the specific frequencies in question show up more clearly (at higher amplitudes relative to low-mid frequency range amplitudes). The absolute minimum for analyzing bearings, however, should include the use of enveloped spectra. The corresponding development for those will also be covered

Rolling Element Bearings

Earlier Failure Stage Symptoms

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Figure 1 - Defect Causes Impacts At A Frequency Equal To The Component Multiplier x RPM

Early Stage Symptoms Of Rolling Element Bearing Defects:

Even in the earliest stages of a bearing defect, it generates frictional or impact-related high frequency vibration.

Due to those impacts occurring, bearing defects show up earliest on the enveloping spectra (Fig 3).

Enveloping signals include gSE (IRD), HFD (SKF) , ESP (DI), Peakvue (CSI), Shock Pulse, and more.

Signal being collected will contain the impact spikes showing up at an interval equal to the defect frequency.

At early stages, the time domain plot will be a better analysis tool than either a velocity or an acceleration spectrum.

Figures 3 and 4 show two what might be considered "typical" plots (envelope in Fig. 3, velocity in Fig. 5) showing symptoms of an early stage bearing defect

The impact frequency is displayed by the envelope plot in Fig. 3. The highest peak: 1x defect frequency.

Figure 4 - the velocity FFT - shows the bearing "condition" (i.e. how bad is the bearing ?).

Figure 2: Two Frequencies Are Produced. The Frequency Of The Bearing Assembly Resonance Affects The FFT Plot While The Frequency Of The Impacts Affects The Enveloping Plot

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The defect frequency harmonics can be very low amplitude (not even noticeable) in the early stages.

There will typically be no peak at 1x defect frequency on the velocity or acceleration FFTs in the defect's early stages.

To analyze, place cursor on bearing defect frequency (as observed on the envelope plot) and attempt to use harmonics to line up higher frequency peaks

Figure 5 and 6 show actual examples of the two types of plots discussed in the relatively early stages.

Velocity FFT provides bearing condition (how bad is the bearing).

A measure of impact intensity (how quickly the bearing will deteriorate) can be assessed by re-scaling the envelope plot for dB (a logarithmic measure) and comparing the bearing defect peak amplitude to the surrounding carpet level (Figure 7)

Figure 3 Typical Enveloping Plot Showing Impacts At Bearing Defect Frequency

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Figure 4 Typical Velocity FFT Showing Early Stage Bearing Defect. Amplitudes Can Be Very Low In Early Stages. It Should Be Noted That The Acceleration Spectrum Will Show The High Frequency Peaks Far More Clearly Than The Velocity Spectrum

Figure 7 - Compare Peak (125dB) To Carpet Level (~102dB). For Difference Above 12dB, Bearing Should Be Watched To Help Determine The Bearing's Rate Of Deterioration. Note Scale (dB) In Upper Left Hand Corner

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Rolling Element Bearings

Later Failure Stage Symptoms

Figure 1 Typical Enveloping Plot Showing Impacts At Bearing Defect Frequency. Ampltiudes May Actually Decrease As Bearings Continue To Worsen

Figure 2 Typical Velocity FFT Showing Early Stage Bearing Defect. Amplitudes Can Be Very Low In Early Stages. It Should Be Noted That The Acceleration Spectrum Will Show The High Frequency Peaks Far More Clearly Than The Velocity Sp

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Advanced RE Bearing Defects:

In the later stages of a bearing defect, the sharp spikes (impacts) occurring usually diminish in intensity due to the bearings components wearing.

Figure 2 shows the velocity spectrum showing the advancing symptoms of the bearing defect.

1x defect frequency eventually becomes visible on the velocity spectrum.

Noise surrounding bearing related peaks continues to increase as the signal shape continues to further distort due to the bearing wearing and the components approaching failure.

The velocity FFT continues to show the bearing "condition". How bad is the bearing ? Much worse now

Figure 3 and 4 show the same two readings as shown on the previous page only one month later.

Note the improvement in the envelope plot shown in Figure 3. This is simply because the impacts are less intense

Rolling Element Bearings

Typical Symptoms

What the previous two pages have shown you is a typical progression of a bearing defect how it reveals itself on a velocity spectrum and gSE spectrum. There are, however, many ways a bearing defect can develop. There are also a variety of methods that are effective at detecting bearing defects. They include the use of acceleration spectra, time domain plots and ultrasonic noise detection (Shock Pulse, for instance). The analyst must be able to "sense" when a bad bearing is developing and tailor the detection method to meet the need. Some of the velocity spectrum variations include:

1) A "haystack" develops at high frequencies on the velocity and acceleration FFTs. The broad frequency band that is affected can make specific frequency identification

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difficult. Frequency range of haystack development depends on resonant frequency range of bearing assembly.

2) Distinct development of the bearing defect frequency harmonics (3x - 10x defect frequency). Which harmonics develop the most and show the highest amplitudes depends on resonant frequency range of bearing assembly.

3) Lower defect frequency harmonics develop (1x, 2x) with little or even no high frequency symptoms. This can easily be confused with running speed harmonics but is quite unusual - far less common than first two possibilities.

Each of these showed up, to one degree or another, on the previous two pages of examples. There was a haystack (area of ill-defined high frequencies), distinct defect frequency harmonics and some development at 1x the defect frequency (very easily confused with 3x rpm, especially if it were to occur in the absence of the high frequency symptoms - which is possible although very unusual).The corresponding development on an acceleration spectrum will, of course, be more prone to the higher frequency symptoms since low frequency amplitudes do not show up well on an acceleration spectra. This is good news unless you have that unusual set of symptoms discussed in #3 above in which case you may get little or even no indication of the bearing problem

Rolling Element Bearing

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Analysis Techniques

Interpreting Enveloping Spectra - As discussed in the 'Enveloping Spectra' section, the actual amplitude of any peaks on the enveloping spectra is not nearly as important as their amplitude relative to the surrounding noise level on the spectrum. What does this mean ? The ultrasonic noise level being detected by your analyzer will have a large impact on the peak amplitudes you will see. Unlike velocity, gSE (to use a well known unit) will be affected by other conditions such as the loading of the bearing and the lubrication level. Poor lubrication quality or, even more so, a lack of lubrication, will raise the entire floor or carpet level of the spectrum. The author has seen bearing running normally with gSE amplitudes that range from 0.05 gSE to 2 gSE. There is no general rule of thumb as far as amplitude levels go - they will vary from machine to machine, environment to environment. Other analysis methods are required.

There are two main methods for using the enveloping spectra to analyze bearing condition. They are:

1) After establishing the frequency of impacts using the gSE spectrum, check your velocity spectrum for any high frequency peaks (even low amplitude ones). If there are none, move on. If there are some (or a 'haystack'), place your cursor on the impact frequency (even if there is no peak) and try to get harmonics to line up or establish some relationship that you can begin to assess for severity. An acceleration spectrum will work even better than a velocity spectrum for this purpose.

2) Go back to your gSE spectrum and change your amplitude scale to dB. Then compare the defect frequency peak amplitude to an estimate of the nearby (surrounding) carpet level. If the difference is 12 - 18 dB, there is a fairly significant amount of impact energy occurring. If the difference is 18+ dB, there is a large amount of impact energy occurring. The greater the level of impact energy, the faster the bearing will deteriorate.

BE CAREFUL - For two main reasons:

If you are using "overall" or "magnitude" (trend) values without analyzing the enveloping spectrum you must be

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aware that a number of different sources can cause the impact energy that these signals detect and many are not related to bearing problems.

These signals are extremely sensitive and can detect problems that are sub-surface or merely very early symptoms of a problem. Calling for drastic repairs at this stage can destroy a program's credibility with people who do not understand the technology. Assuming sufficient lubrication, bearing replacement at this stage is almost always unnecessary.

If you have questions on these spectra, re-visit the 'Enveloping Spectra' section

Hydraulic and Aerodynamic Problems

Hydraulic or aerodynamic forces - combined into a single category since they are similar in nature and involve moving a fluid. In the case of a fan or lp blower, a very compressible

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"fluid" (a low pressure gas) is being moved. In the case of a compressor or hp blower, a less compressible "fluid" (a high pressure gas) is being moved. In the case of a pump, a non-compressible fluid (an actual "fluid") is being moved. The less compressible the fluid, the more susceptible the component is to flow-related vibration problems. Some of the problems we will discuss include vane or blade pass frequency, cavitation, recirculation, internal clearance problems and flow turbulence / surging. The most common and well known of these frequencies, vane or blade pass frequency - which is simply the number of vanes, blades, lobes, etc. x RPM - is briefly discussed here.

Low Pressure (Centrifugal) Fans & Blowers - Not normally susceptible to these problems to a significant degree.

BPF - Rarely a problem unless the frequency excites a resonant frequency in the downstream ductwork. This normally is translated into a noise problem and, less often, a structural problem. Rarely does it cause a mechanical problem such as accelerated bearing or component wear.

Flow Turbulence - Can cause low frequency, broadband vibration (below or just higher than 1x rpm).

High Pressure Blowers and Compressors - Far more susceptible due to much tighter clearances and much higher pressures.

Pumps – Problems are the most severe since it is a non-compressible fluid that is being moved. Tight clearances and high pressures compound the potential problems. Not only can mechanical clearance problems cause large vibrational problems (wear rings, impeller / housing / diffuser clearances) but operating a pump at different pressures and flow rates than its design point can cause excessive and even destructive vibration.

Hydraulic Problems:

Recirculation & Flow Related Problems

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Figure 1 - Typical Spectrum Showing High Vane Pass Frequency ("VPF" = # of Vanes x RPM). Symptoms normally in the radial directions but may also be seen axially

Recirculation & flow related symptoms (including component problems):

High amplitude VPF or BPF, often accompanied by harmonics of VPF or BPF.

Sidebands may be present around VPF and 2x VPF at 1x RPM. If present, they typically indicate a rotor problem - eccentricity, for instance - that would cause a modulation of the VPF at a rate of 1x rpm.

Important to keep in mind that a certain amount of VPF / BPF is normal.

In cases of extreme flow-related problems, flow instability is created and have been known to generate vibration that excites the resonant frequency of the pump impeller much as oil whip acts on the resonant frequency (critical speed) of a sleeve bearing rotor (turbine).

Recommended Actions:

First step should be a thorough inspection of pump with particular attention paid to proper clearances and integrity of mechanical components.

Second step should be to assess system itself - elbows too close to the discharge, for instance, can cause similar vibration symptoms due to the reflection of fluid waves back into the discharge of the pump.

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Third step should be to assess operational parameters - flow rates and pressure - that can also influence this vibration. Actual flow & pressure should be compared to the pump curve and design point of the pump. Although insufficient flows and/or pressures lead to cavitation (different symptoms - see next page), excessive flows and/or pressures can lead to recirculation and symptoms similar to what is seen here

Hydraulic Problems:

Cavitation

Figure 1 - Typical Spectrum Showing Cavitation (Random, Very Broad Haystack-Like Appearance). Symptoms normally in the radial directions but may also be seen axially. Cavitation - occurs when there is insufficient flow into or pressure out of a pump. This causes the fluid entering to literally be torn apart. Vacuum pockets are created and then implode. This occurs in a random, unpredictable manner and can be extremely destructive to the impeller and internal pump components.

Cavitation symptoms:

High frequency, random vibration. Sounds like the pump is pumping gravel.

Although amplitudes may or may not be high enough to affect bearing life significantly, cavitation causes excessive wear on the impeller and other internal components.

May come and go from one collection to the next as load varies.

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Recommended Actions:

First step should be to assess operational parameters - flow rates and pressure - that can also influence this vibration. Actual flow & pressure should be compared to the pump curve and design point of the pump. Insufficient flows and/or pressures lead to cavitation.

Second step should be an inspection of the internal components for excessive wear with particular attention paid to the impeller vanes.

NOTE: The resonant frequency of a probe or stinger attached to the transducer can be amplified by minor cavitation symptoms and give misleading readings. The 9" probe used for years by IRD, for instance, has a resonant frequency of about 40-50kcpm. Readings taken with this probe on a pump with minor cavitation can cause high amplitudes in the 40-50kcpm range and lead to incorrect diagnosis

Aerodynamic Problems:

Flow Turbulence

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Figure 1 - Typical FFT Showing Flow Turbulence. Occurs In Compressors And High Pressure Blowers When Surging Or Load Variations Occur That The Machine Is Affected By. Often, A Reservoir Or Surge Suppressor Can Be Used To Eliminate This Feedback.

Flow Turbulence Symptoms:

High frequency, random vibration similar to cavitation.

High amplitude blade or lobe (screws) pass frequency (referred to as 'VPF' below).

High amplitude harmonics of VPF.

NOTE: Again, it is important to note that VPF on a compressor or blower is a normal, mechanical vibration. It is risky to over-react to initial readings without knowing the normal operating characteristics and vibration levels of the machine. The amplitudes will also be load related and attempts should be made to consistently take the readings under the same load characteristics. If the compressor loads or unloads during a reading, it should be taken over

AC Induction Motor Problems

How To Monitor For Electrical Frequencies

Electrically Generated Vibrations- The supply of AC power to a

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motor generates mechanical vibration. Since AC power is supplied as a sinusoid, each pole of the motor is energized twice - once with a "+" peak and once with a "-" peak - during each cycle. This means that the most common vibration frequency that is generated is NOT line frequency - it is 2x line frequency. This manual section is concerned only with the proper diagnosis of electrically-related problems. A more in-depth look at the sources and reasons for these vibration can be found in the <Field Tests> manual. However, if there is one tip for accurately diagnosing and correcting these types of problems, it is that the only truly reliable test is on-line current analysis. If vibration symptoms develop that lead you to believe you have one or more of these problems, do NOT send the motor to a motor shop. Find someone to perform an on-line current analysis of the motor. This test is more reliable than a motor shop test because it is done with the unit under load and heated up and is less expensive when considering the cost of removing and re-installing the unit.

There are two spectra necessary to detecting electrically-related problems. Each example that follows is taken on one or the other.

High frequency (200 x RPM).

High resolution (12kcpm Fmax w/ 1600 lines is usually sufficient).

There are also certain terms and frequencies which must be defined:

FLine = Electrical line frequency - normally 60 Hz (3600 cpm) or 50 Hz (3000 cpm).

2 x FLine = Torque Pulse Frequency. This is a common frequency found on a high resolution spectrum.

P = # of poles on the motor. The number of poles is how the speed of the motor is controlled. The greater the number of poles, the slower the motor runs. The number of poles is always an even number (2, 4, 6, etc.).

FSynch = Synchronous electrical speed = 2 x FLine / P. This can be confusing because it refers to electrically synchronous, not synchronous to the rpm. It is the speed of the rotating magnetic field that is

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generated and the speed the rotor tries to attain (it will never quite reach that speed).

FSlip = Slip frequency = FSynch - rotor RPM (actual speed)

FPole = Pole pass frequency = P x FSlip

WSPF = # Winding Slots x RPM

RBPF = # Rotor Bars x RPM

The most important thing to look for with electrically related vibration is increasing amplitudes - not just the presence of a peak or pattern of peaks. Whenever a problem is detected vibration-wise, the next step should be increased surveillance to see if the amplitudes are trending up or not. Additional testing can also be performed (surge testing, current testing, etc.) but no action should be taken until you have a better idea of the unit's condition. Vibration is NOT the best way to monitor most electrical problems and that fact must be recognized.

AC Induction Motor Problems: Elliptical Stator, Stator Weakness & Winding Shorts

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Figure 1 - Typical Spectrum Showing Indications Of Variation In Air Gap, Winding Shorts, Stator Weakness

Air Gap Variation, Winding Shorts, Stator Weakness Symptoms:

High amplitudes at 2 x FLine.

Recommended Actions:

Check for soft foot and repair.

Check alignment and repair.

Perform winding tests to assess insulation integrity of the windings.

Live with it or buy a new motor.

NOTE: It is important to realize that vibration at 2 x FLine is a normally occurring vibration. The effect on the bearings is no greater or less than the same amplitude due to unbalance. Do not over-react. NOTE: Do not confuse the presence of a running speed harmonic with a pole pass frequency sideband. By definition, a running speed harmonic will always be separated from 2 x FLine by pole pass frequency

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The Stator- consists of the windings and the metal of the motor housing itself (i.e. the 'iron', or 'core'). The symptom we will see here is related to variation in the air gap between the windings and the rotor. That air gap is not perfectly even all the way around. Since the strength of a magnetic field - which causes the rotation of the rotor - is proportional to the gap (the smaller the gap, the stronger the force), variation in the gap produces vibration at (2 x FLine). The greater the variation, the higher the amplitude. The air gap can also be affected, however, by mechanical problems such as soft foot (which stresses & distorts the housing), stator looseness / weakness (allowing it to be influenced to a greater degree by those magnetic forces) and winding shorts (which cause localized heating and thermal distortion). The only one of the previous problems that is easily tested for and fixed is soft foot

MOTOR CONSTRUCTION WIRING CONSTRUCTION

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AC Induction Motor Problems:

Elliptical Rotor

Figure 1 - Typical Spectrum Showing Indications Of Eccentric Rotor. Similar To Eccentric Stator. Some Cases May Exhibit The Sidebands Seen Here; Others May Propagate Strictly At 2x Line Frequency

Eccentric Rotor Symptoms:

High amplitudes at 2 x FLine.

Possible sidebands around 2x line frequency and/or 1x rpm.

Recommended Actions:

Check for soft foot and repair.

Check alignment and repair.

On-line current analysis to assess condition and determine severity.

Live with it or buy a new motor.

NOTE: It is important to realize that vibration at 2 x FLine is a normally occurring vibration. The effect on the bearings is no greater or less than the same amplitude due to unbalance. Do not over-react. NOTE: Do not confuse the presence of a running speed harmonic with a pole pass frequency sideband. By definition, a running speed harmonic will always be separated from 2 x FLine by pole pass frequency

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AC Induction Motor Problems:

Phasing Problems

Figure 1 - One Possible Spectrum Caused By A Problem With A Short In One Of The Phases Or Feeder Cables

Figure 2 - Another Possible Spectrum Caused By A Problem With A Short In One Of The Phases Or Feeder Cables

The main problem caused by phasing shorts is impeding the free flow of current to the motor. This can cause problems ranging from danger to personnel to heat-related damage to catastrophic motor failure.

Single Phasing Symptoms:

High amplitudes at 2 x FLine (this can be the only symptom).

Sidebands around 2 x FLine at 1/3 FLine (1/3 line frequency).

Recommended Actions:

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This may occur with a sudden, dramatic increase in amplitudes. In that case, a short should be suspected and testing should be performed.

Inspect connections at junction box on motor.

Surge test unit from motor control center. This will detect a problem anywhere in the leads, splices or windings.

If a problem is found, the splices at the motor should be broken and the leads and windings tested separately to isolate the problem.

If nothing found in windings, on-line current analysis should be performed on the motor

AC Induction Motor Problems:

Broken / Cracked Rotor Bars

Figure 1 - FFT Showing Advanced Broken / Cracked Rotor Bar Symptoms

Rotor Bar Problems- The electrical problem that is most effectively diagnosed and monitored through vibration analysis. Broken, cracked rotor bars, bad joints between end rings and rotor bars and end ring problems have unique and easily recognizable symptoms. An accurate assessment of condition and remaining life can also be made with on-line current analysis. The initial diagnoses, however, can easily be made with vibration data. Do NOT send the motor to a motor shop - especially in the early stages. It is doubtful that they will have the equipment to diagnose the problem without running unreliable and potentially destructive tests on the unit. In cast rotors, these symptoms can indicate voids in the casting. Broken Cracked Rotor Bar Symptoms:

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FPole sidebands surrounding running speed harmonics. Advanced problems will exhibit a 'humming' or 'pulsing'

sound and feel.

Significant is the number of and size of the sidebands. They increase as the unit deteriorates.

Amplitude at 1x rpm is relatively unimportant - it will fluctuate greatly as the hot spots being generated cause the rotor to bow unpredictably. It is a result of the problem - not a cause.

Recommended Actions:

On-line current analysis to determine severity.

Limit starts since they are easily the single, most destructive thing you can do to a motor.

NOTE: The development of sidebands of any amplitude should be noted. This may require the use of a logarithmic scale. Again, do not confuse a running speed harmonic adjacent to 2 x FLine with a pole pass frequency sideband - they are by definition separated by that amount and are not sidebands

AC Induction Motor Problems:

Loose Rotor Bars

Figure 1 - Spectrum Showing Pattern Of Peaks Separated By 2xLine Frequency (Sidebands) In High Frequency Range (30-90xRPM)

Loose rotor bars- Extremely unusual and never found in cast rotors. As a loose rotor bar passes a winding slot, the magnetic force causes it to momentarily lift and then drop. The frequency, then, is the number of windings slots x RPM

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(Winding Slot Pass Frequency or WSPF). The number of winding slots will be between about 25 and 100. The bad news is that you will not know the number of winding slots and it is very difficult to find out. The good news is that the vibration is accompanied by a precise sideband - 2 x FLine. Loose Rotor Bar Symptoms:

High amplitude at a very high frequency (WSPF, but we don't know what it is) accompanied by sidebands at 2 x FLine. This symptom is not unusual and at low amplitudes - below 0.1 ips or 2.5 mm/sec - often means little more than an imperfection in the machine. In fact, it usually has more to do with a potential winding problem than rotor bar looseness (see the next page).

Vibration at 2 x WSPF and even 3 x WSPF w/ sidebands at 2 x FLine. These are much more unusual and indicate a much more potentially severe problem.

Symptoms identical to the next problem - looseness in the windings (make sure both are understood before recommending any action for either problem).

Recommended Actions:

On-line current analysis to determine severity. If a healthy rotor is found, it is more likely a potential winding problem (next page).

Winding testing in addition to rotor testing will provide for a comprehensive eletrcial PdM program

AC Induction Motor Problems: Loose in Winding Slots, Iron, End Turns And/Or Connections

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Figure 1 - Velocity FFT Showing Pattern Of Peaks Separated By 2xLine Frequency (Sidebands) In High Frequency Range (30-90xRPM). This Will Be Accompanied By The Symptom Seen In Figure 2:

Figure 2 - Envelope Plot Showing 2xLine Peak And Harmonics. This Indicates Impacts Occurring At 2xLine Frequency

Looseness in the winding slots- are detectable with vibration analysis but cannot be trended towards failure since the problem does not worsen (vibration-wise) prior to winding failure. The problem causes wear of the insluation on the windings and eventually a ground short (catastrophic failure). Only winding testers (surge testing) can trend this problem and assess the severity. It is commonly found and should not be over-reacted to. The symptoms are very similar to loose rotor bars on the velocity / acceleration spectra. Additionally, however, there will be high amplitude peaks on your enveloping spectra (e.g. gSE spectrum) at 2 x FLine and

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harmonics. Each time a rotor bar passes the loose winding, it lifts and then drops back. The vibration frequency, therefore, is the number of rotor bars x RPM (rotor bar pass frequency = RBPF). Like WSPF, it will be surrounded by 2 x FLine sidebands. Like the number of winding slots, we won't know the number of rotor bars but it's not important - the pattern of peaks separated by 2 x FLine is the clue we need. Looseness In The Windings Symptoms:

High amplitude at a very high frequency (RBPF, but we don't know what it is) accompanied by sidebands at 2 x FLine. This is not unusual and often means little. In fact, it often has more to do with a potential winding problem than rotor bar looseness (see the next page).

Amplitude peaks on the enveloping spectra at 2 x FLine and harmonics.

Recommended Actions:

Surge testing to check insulation integrity and test for any wire to wire, turn to turn and phase to phase shorts as well as the integrity of the ground wall insulation

AC Variable Frequency Drives

Variable Frequency Drives - VFD's are AC motors that give the operator the speed control that a DC Drive normally provides at a small fraction of the cost and difficulty in maintenance and troubleshooting. It operates exactly as an AC induction motor does with all of the same electrically generated frequencies. That's the good news. The bad news is that a VFD is the vibration analyst's worst nightmare. For example, the vibration frequencies detectable with both AC induction motors and DC motors are constant - only a couple vary at all and they are part of a easily recognizable pattern. With a VFD, the speed is controlled by modifying the frequency of the power supply. In other words, a motor normally running at 3550 rpm can be slowed down to, say, 1775 rpm (1/2) by reducing line frequency from 60 Hz to 30 Hz. In all likelihood, however, you will not know the exact frequency being supplied to the motor and that is the problem. Some of the effects of this change are:

2 x FLine is unknown. This makes identifying air gap and soft

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foot problems more difficult.

Slip Frequency is unknown. This makes identifying rotor bars problems more difficult.

Sidebands occurring at 2 x FLine are unknown. This not only makes loose rotor bars or windings more difficult but also greatly increases the likelihood of confusing an electrical problem with a bearing problem.

Extreme care must be taken with VFD's and especially with bearing defects. Speed at the time of data collection should always be noted as accurately as possible and the analyst's knowledge of the machine's normal operating characteristics is even more important than usual

DC Drives

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Figure 1 - "Normal" FFT Taken On DC Drive

DC Drives - generate certain electrical frequencies due to the way the drive itself works. Direct current is the flow of electricity in one direction (as opposed to alternating current which changes direction at a rate of 60 times per second). However, a DC drive gets its power supplied by an AC power source. Since AC power is a sinusoid, the drive cuts off the bottom ("-" portion) of the sine wave in order to get a constant "+" voltage. This is done with an SCR - a 'Silicon Controlled Rectifier'. Using a single SCR, however, would result in a '+' peak followed by a period of no current flow since the '-' peak would be cut off. This would be unacceptable as it would lead to a surging, pulsing power supply. A better solution is to have 3 SCR's with the AC signals separated by a 120° phase lag. The following animation shows how that would create a much more constant power supplyThis type of drive arrangement is known as "half-wave rectified". You can see from the animation that FLine is supplied to the drive. However, if the drive is operating properly, what frequency would you see ? That's right - 3x FLine. 3x FLine is a normal vibration frequency to be found on a DC motor. This frequency is known as SCR firing frequency, or FSCR. The amplitude at FSCR can be up to 0.1 ips (2.5 mm/sec) before beginning to cause any notice. There can also be a small amplitude peak at 2x FSCR. There is also another type of drive known as 'full-wave rectified' that uses 6 AC signals

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FSCR on a full-wave rectified drive is, of course, 6x FLine. You can see how a full-wave rectified drive gives better control and a more constant voltage than a half-wave rectified drive does

DC Drive Problems

Figure 1 - Full-Wave Rectified Velocity Spectrum w/ Drive Problems

Figure 2 - Half-Wave Rectified Velocity Spectrum w/ Drive Problems

Figure 3 - Spectrum on DC Motor w/ Speed Fluctuations

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DC Drive Problem Symptoms:

Excessive (or increasing) amplitudes at FLine and/or FSCR. These can indicate tuning problems, grounding problems, winding problems, etc.

Peaks at other FLine harmonics - 2x, 3x, 4x, 5x. With the exception of half-wave rectifiers (where 3x FLine is the SCR firing freq.), these peaks should never be present. For example, consider the full-wave rectified signal below. Imagine having a bad SCR. You would get an amplitude peak at 5 x FLine (instead of 6x) plus an increased peak at 1 x FLine. A bad firing card, which can control 1 SCR (half-wave) or 2 SCRs (full-wave), can cause the loss of 1/3 of the power. This causes peaks at 1/3 x FSCR and 2/3 x FSCR. The exact frequencies will depend on whether the drive is full-wave rectified or half-wave rectified and the FLine supplied.

Sidebands around FSCR. These sidebands typically indicate that the motor speed is fluctuating or 'hunting'. This can be caused by comparitor card problems. A high resolution spectrum (at least 1600 lines, probably 3200 and possibly even 6400 lines - depends on motor speed) may be required to detect these sidebands. The hunting may be most easily detected initially with a strobe light.

Whenever abnormal symptoms develop, the drive itself must be thoroughly analyzed. The exact symptoms, however, should provide important clues as to where to look first.

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DC Drive Problems typically show up on a vibration spectrum as amplitude peaks at multiples of FLine between FLine and FSCR. With a full-wave rectifier, this includes 2 x FLine, 3 x FLine, 4 x FLine and 5 x FLine. With a half-wave rectifier, it means only 2 x FLine. In each case, however, it often means high amplitudes at 1 x FLine and FSCR. It is important that the vibration analyst either have a good understanding or work with someone who has a good understanding of the electronic components in the drive (# of SCRs, # of firing cards, control cards, how to tune the system, etc.)

Gears

Gears & Gear Trains - As with other mechanical influences, gears generate vibration under normal circumstances. The most

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common frequency generated is a the number of teeth x RPM. This is known as 'gear mesh frequency' (GMF). Since the gear rotates at the speed of the shaft, there is often a modulation of the vibration at 1x shaft rpm. To better understand the concept of "amplitude modulation", let's examine some generated signals and how the FFT turns the signal into a spectrum.

The signal below is represents pure sinusoidal motion - the kind of signal you only get on the classroom drawing board or in a manual - rarely (if ever) in real life. There are only, in fact, two sources that create such a signal - unbalance and resonance. The result of performing an FFT on this signal is shown below - a single peak labeled on the spectrum at 605 cpm. Note how each cycle takes 100 msec (0.1 seconds). That equals 10 cycles per second (10 Hz) or 600 cpm.

Figure 1: Time Domain for 360 msec (0.360 seconds)

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Figure 2 - FFT performed on signal in Figure 1

The next signal below also represents sinusoidal motion but at a much higher frequency - 22x the frequency above or 13,200 cpm. The time sample has remained the same - 360 msec. This could be generated by a gear with 22 teeth on it mounted on a shaft running at 600 rpm (just like the shaft in Figures 1 & 2 - what a coincidence !!). Of course, there is no influence at all at 1x rpm on the below signal (we'll get to that next).

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Figure 3 - Time Domain for 360 msec (0.360 seconds)

Figure 4 - FFT performed on signal in Figure 3

Now let's look at how these separate signals can combine

Gears

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What happens if we put the two signals together ? We get what is called a high frequency riding a low frequency. In other words, the high frequency vibration rides along on the low frequency as if on a roller coaster. There is no 'modulation' of amplitude. In other words, the height (amplitude) from the top of any peak to the bottom of the previous or next valley is constant. The same applies to the low frequency - the amplitude remains constant. The FFT will create two peaks - one for the high frequency and one for the low frequency

Now let's consider a situation with 'amplitude modulation'. Consider 2 mating gears where one is eccentric. At one point during that gear's rotation, it will bottom out with the mating gear and the vibration at GMF will be very high. In Figure 3, that occurs at about 100 msec, 200 msec and 300 msec. At the opposite point in its rotation, the teeth will be backed away from one another a maximum amount and the amplitude at GMF may be at a minimum (we'll ignore problems such as gear loading and backlash for the purposes of this example). In Fig. 3, that occurs at about 50 msec, 150 msec and 250 msec

What is happening here is a modulation of the amplitude at gear mesh frequency. Moreover, it is going from its minimum amplitude to its maximum and back again to its minimum at a rate of once per shaft revolution - 1x rpm. What the FFT will generate from this signal is a peak at GMF with sidebands at 1x rpm. This type of modulation is where sidebands come from - they are generated by the FFT process

Some amplitude modulation on a gear train is not unusual and should not cause over-reaction. The number of and size of the sidebands should be closely monitored. Even more significant can be the development of an amplitude peak at the natural frequency of the gear or gears. Wear or impacting due to problems such as backlash can cause the excitation of the natural frequency of a gear. The problem, of course, is that

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you will not know that natural frequency. Its appearance on a spectrum must be noticed and investigated.

Gear Problems:

Normal Gear Spectrum

Figure 1 - "Normal" Spectrum

Normal Gear Drive Symptoms:

Amplitude peaks at 1, 2 and/or 3x GMF.

Low amplitude and few sidebands around 1, 2 and/or 3x GMF at 1x rpm of gear with problem

Gear Problems:

Gear Eccentricity / Gear On Bent Shaft

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Figure 1 - Typical FFT For Eccentric Gear Or Gear On Bent Shaft

Gear Eccentricity or Bent Shaft (@ gear) - Causes modulation of GMF amplitude at 1x rpm of the eccentric gear. Can also cause modulation of the shaft speeds of the gears if the problem is severe enough. If the output gear were eccentric, that gear's 1x rpm peak would be higher and the sidebands would be spaced at that frequency instead of 1x rpm of the pinion.

Eccentric Gear Or Gear On Bent Shaft Symptoms:

Higher amplitudes at 1, 2 and/or 3x GMF.

High amplitude sidebands around 1, 2 and/or 3x GMF at 1x rpm of gear with problem.

Higher amplitudes at 1x rpm of gear with problem and, if the problem is severe, running speed harmonics of that frequency.

Recommended Actions:

Inspect gears for wear patterns and check for proper mesh depth.

Inspect gears for proper backlash (similar symptoms - see next page.

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Gear Problems:

Excessive Backlash

Excessive Backlash - Causes high amplitudes at GMF and harmonics. Also, the impacting excites the natural frequency of the gear(s). This can cause unexplained frequencies to appear - they may be the resonant frequencies of the gear(s). The less loaded the gears are, the more effect the excessive backlash has.

Excessive Backlash Symptoms:

Higher amplitudes at 1, 2 and/or 3x GMF.

High amplitude sidebands around 1, 2 and/or 3x GMF at 1x rpm of one or both of the gears.

Amplitude peak at resonant frequency of the gear(s).

Sidebands at 1x rpm surrounding the resonant frequency.

Recommended Actions:

Inspect gears for proper backlash.

Inspect gears for wear patterns and check for proper mesh depth (similar symptoms - see previous page).

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Gear Problems:

Gear Wear (Tooth Wear)

Figure 1 - Typical FFT Showing Wear On Gear Teeth

Gear Wear - Causes high amplitudes at GMF and harmonics. Also, the rubbing / wearing action excites the natural frequency of the gear(s). This can cause unexplained frequencies to appear - they may be the resonant frequencies of the gear(s). The two key indicators are the appearance of the gear's resonant frequency w/ sidebands and the size and number of sidebands surrounding 1, 2 and/or 3x GMF - not the amplitudes at GMF and harmonics alone (these are better indicators for load and alignment).

Gear Wear Symptoms:

Higher amplitudes at 1, 2 and/or 3x GMF.

High amplitude sidebands around 1, 2 and/or 3x GMF at 1x rpm of the worn gear.

Amplitude peak at resonant frequency of the gear(s).

Sidebands at 1x rpm of the worn gear surrounding the resonant frequency.

Recommended Actions:

Inspect gears for wear patterns and check for proper mesh depth (similar symptoms - see previous page).

Inspect gears for proper backlash.

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Gear Problems:

Gear Load

Figure 1 - Typical FFT Showing Gear Load Problems

Gear Load - Often affects GMF and harmonics more than the running speed sidebands which are low amplitude and relatively few. Increases and decreases in GMF and harmonics alone (without significant change in sidebands) does not necessarily indicate a problem. Even if the load itself is fairly constant, the gear that is carrying the load is constantly changing so this amplitude can change from data collection to data collection without any deterioration of the gear condition whatsoever. A change in the load itself can also occur and have an even more dramatic impact on the spectrum amplitudes without reflecting any problem.

Gear Load Symptoms:

Higher amplitudes at 1, 2 and/or 3x GMF.

Recommended Actions:

None unless there is an increase in sideband activity or the appearance of possible gear resonant frequencies.

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Gear Problems:

Gear Misalignment

Figure 1 - Typical FFT Showing Gear Misali

Gear Misalignment - Makes the natural rotation of the gears more difficult since they must fight their way through an area where the gear teeth are misaligned. The causes a momentary binding (slowing) of the rotation. The FFT turns this phenomenon into amplitude peaks at 2x rotational speeds of the gears and 2x GMF. Each of these symptoms - primarily the 2x GMF - may indicate a gear alignment problem (which may, of course, be induced with poor coupling alignment or other external factors such as soft foot).

Gear Misalignment Symptoms:

Highest amplitudes at 2x GMF.

Amplitude peaks at other GMF harmonics - 1x, 3x, etc.

High amplitude sidebands particularly around 2x GMF at 1x or even 2x rpm.

Shaft running speed harmonics - 2x and even 3x rpm.

Recommended Actions:

Inspect gears for wear patterns misalignment causes uneven wear.

Check for external problems - shaft alignment, soft foot, etc.

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Gear Problems:

Hunting Tooth Frequency

Figure 1 - Typical FFT Showing Hunting Tooth Symptoms

Hunting Tooth Frequency (FHT) - Is the phenomenon in which two teeth - one on each gear - that are damaged contact one another at a particular frequency. In other words, every once in a while during the normal rotation of these gears, those two teeth will enter the mesh area simultaneously and contact one another. You can imagine that this would be a relatively low frequency - lower than the rpm of either gear. It is, in fact, determined by the common factors of the number of teeth on each gear. The best way to explain this unusual frequency is with an example.

A 2000 rpm gear with 24 teeth is driving a gear with 84 teeth. That would make the output speed 2000 x 24/84 = 571.4 rpm. To calculate what the hunting tooth frequency is, you must first determine the common factors of each gear (CF). To do this, simply list all the multiplication possibilities for each tooth number and compare:

24 Tooth Gear 1 x 24

2 x 12

3 x 8

4 x 6

84 Tooth Gear 1 x 84

2 x 42

3 x 28

4 x 21

6 x 14

7 x 12

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What are the common factors ? The numbers that appear in each column are: 1, 2, 3, 4, 6 and 12. The highest common factor is what we are after and that number (in this example) is 12. What if you had 23 teeth on the pinion ? Since the only numbers that can be multiplied to generate 23 are 1x23, the only common frequency would be 1 unless the other gear had a number of teeth that was an exact multiple of 23 (23, 46, 69, 92, etc.).

Next, we use that number in the following formula:

Hunting Tooth Frequency =

(Gear Mesh Freq x Common Freq)(#Teeth Pinion x #Teeth Bull Gear)

In this case, that equals (48000 x 12)/(24 x 84) = 285.7 cpm. That means that 285.7 times per minute, those bad teeth (one on each gear) will enter mesh together and generate a very high vibration pulse. That may seem rather high but usually the CF is not as high as 12 - often it will be 1.

Note that the lower the highest common factor is, the lower the FHT is. If we had 23 teeth on the pinion gear instead of 24 in the previous example, the CF would be 1 and the FHT would be 24.84 cpm.

Note also that you would need extremely good resolution to find any of these frequencies - you certainly won't isolate the FHT with the same spectrum used to check 1, 2 & 3x GMF.

Hunting Tooth Symptoms:

Amplitude peaks at 1 x FHT and possibly 2 x FHT.

Sidebands of FHT around 1x rpm (of each shaft).

Sidebands of FHT around 1x GMF and harmonics.

Pulsing, growling noise coming from gearbox or drive.

NOTE: Spectrum resolution will determine if any of the above symptoms are actually visible on a spectrum. In the above example, for instance, to detect sidebands at FHT (143 cpm) around 1x GMF (48,000 cpm), spectrum resolution of about 45 cpm/line would be required. For a 180,000 Fmax (capturing 1, 2 & 3x GMF), that requires 4000 lines of resolution (3200 would be borderline, 6400 would work fine). Recommended Actions:

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Inspect gears for damage.