Machinery Vibration Diagnostics 1

UNBALANCE

Force Unbalance

Typical Spectrum Phase Relationship
forcespec.gif
force.gif

Force Unbalance will be in-phase and steady. Amplitude due to unbalance will increase by the square of speed (3x speed increase = 9x higher vibration. 1x RPM always present and normally dominates the spectrum. Can be corrected by placement of only one balance weight in one plane at rotor centre of gravity (CG).

Couple Unbalance
Typical Spectrum Phase Relationship
forcespec.gif couple.gif

Couple Unbalance tends toward 180° Out-of-phase on same shaft. 1x always present and normally dominates the spectrum. Amplitude varies with square of increasing speed. May cause high axial vibrations as well as radial. Correction requires placement of balance weights in at least 2 planes. Note that approx. 180° phase difference should exist between Outboard and Inboard horizontals as well as Outboard and Inboard verticals.

Overhung Rotor Unbalance
Typical Spectrum Phase Relationship
overhungspec.gif overhungdraw.gif

Overhung Rotor Unbalance causes high 1x vibration in both Axial and Radial directions. Axial readings might be unsteady. Overhung rotors often have both force and couple unbalance, each of which will likely require correction.

Eccentric Rotor
Typical Spectrum Phase Relationship
eccentric.gif eccendraw.gif

Eccentricity occurs when the centre of rotation is offset from the geometric centreline of a sheave, gear, bearing, motor armature, etc. The largest vibration occurs at 1x RPM of eccentric component in a direction through the centres of the two rotors. Comparative horizontal and vertical phase readings usually differ either by  0° or by 180° (each of which indicate straight line motion). Attempts to balance an eccentric rotor often results in reducing the vibration in one direction, but increasing it in the other radial direction (depending on the amount of eccentricity).

Bent Shaft
Typical Spectrum Phase Relationship
bentshaftspec.gif bentdraw.gif

Bent Shaft problems cause high axial vibration with axial phase differences tending toward 180° on the same machine component. The dominant vibration ins normally at 1x if bent near the shaft centre, but at 2x if bent near the coupling. (Be careful to account for the transducer orientation for each axial measurement if you reverse probe direction).

MISALIGNMENT

Angular Misalignment

Typical Spectrum Phase Relationship
Spectra Drawing

Angular Misalignment is characterised by high axial vibration, 180° Out-of-phase across the coupling. Typically will have high axial vibration with both 1x and 2x rpm. However, not unusual for either 1x, 2x or 3x to dominate. These symptoms may also indicate coupling problems as well.

Parallel Misalignment

Typical Spectrum Phase Relationship
Spectra Drawing

Offset Misalignment has similar vibration symptoms to Angular, but shows high radial vibration which approaches 180° Out-of-phase across the coupling. 2x often larger than 1x, but its height relative to 1x is often dictated by coupling type and construction. When either Angular or Radial Misalignment becomes sever, it can generate either high amplitude peaks at much higher harmonics (4x – 8x) or even a whole series of high frequency harmonics similar in appearance to mechanical looseness. Coupling construction will often greatly influence the shape of the spectrum when misalignment is severe.

Misaligned Bearing Cocked On Shaft

Typical Spectrum Phase Relationship
Spectra bearing

Cocked Bearing will generate considerable axial vibration. Will cause twisting motion with approximately 180° phase shift top to bottom and/or side to side as measured in the axial direction of the same bearing housing. Attempts to align the coupling or balance the rotor will not alleviate the problem. The bearing must be removed and correctly installed.

MECHANICAL LOOSENESS

Mechanical Looseness

Typical Spectrum Phase Relationship
Type ‘A’
forcespec.gif
typeadraw.gif
Type ‘B’
typeb.gif typebdraw.gif
Type ‘C’
typec.gif typecdraw.gif

Mechanical Looseness is indicated by either type A, B or C spectra. Type ‘A’ is caused by structural looseness/weakness of machine feet, baseplate or foundation, also by deteriorated grouting, loose hold-down bolts at the base and distortion of the frame or base (i.e Soft Foot). Phase analysis may reveal approx. 180° phase difference between vertical measurements on the machine foot, baseplate and base itself. Type ‘B’ is generally caused by loose pillowblock bolts, cracks in the frame structure or bearing pedestal. Type ‘C’ is normally generated by improper fit between component parts which will cause many harmonics due to non linear response of loose parts to dynamic forces from the rotor. Causes a truncation of time waveform. Type ‘C’ is often caused by a bearing liner loose in its cap, excessive clearance in either a sleeve or rolling element bearing or a loose impeller on a shaft. Type ‘C’ phase is often unstable and may vary widely from one measurement to the next, particularly if the rotor shifts position on the shaft from one start-up to the next. Mechanical looseness is often highly directional and may cause noticeably different readings if you compare levels at 30° increments in the radial direction all the way around one bearing housing. Also note that looseness will often cause subharmonic multiples at exactly 1/2 or 1/3 x rpm (.5x, 1.5x, 2.5x etc.)

ROLLING ELEMENT BEARING

Rolling Element Bearings
(4 Failure Phases)

Stages of Progressive Deterioration
spacer20v.gif (51 bytes)
Stage 1
stage1.gif
Stage 1 : Earliest indications of bearing problems appear in ultrasonic frequencies ranging from approximately 20,000 – 60,000 Hz (1,200,000 – 3,600,000 CPM). These are frequencies evaluated by spike energy (gSE), HFD(g) and shock pulse (dB). Foe example, spike energy may first appear at about .25 gSE in stage 1 (actual value depending on measurement location and machine speed.
 


Stage 2stage2.gif Stage 2 : Slight bearing defects begin to “ring” bearing component natural frequencies (fn) which predominantly occur in the 30K – 120K CPM range. Sideband frequencies appear above and below natural frequency peak at end of stage 2. Spike energy grows (for example .25 to .50 gSE).
 


Stage 3stage3.gif Stage 3 : Bearing defect frequencies and harmonics appear when wear progresses. More defect frequency harmonics appear and a number of sidebands grow, both around these and around bearing natural frequencies (see Vibration Case History Number 3 for actual example). Spike energy continues to increase (for example, from .5 to over 1 gSE). Wear is now usually visible and may extend throughout periphery of bearing, particularly when well formed sidebands accompany any bearing defect frequency harmonics. replace the bearings now.

Stage 4stage4.gif Stage 4 :Towards the end, the amplitude of the 1x RPM is even effected. It grows, and normally causes growth of many running speed harmonics. Discrete bearing defect and component natural frequencies actually begin to “disappear” and are replaced by random, broadband high frequency “noise floor”. In addition, amplitudes of both high frequency noise floor and spike energy may in fact decrease, but just prior to failure, spike energy will usually grow to excessive amplitudes.

Formulae to Calculate Specific Bearing defect Types.

 

formula.gif

SLEEVE BEARING

Wear / Clearance Problems

spacer20v.gif (51 bytes)Typical Spectrum
weraspec.gif

Latter stages of sleeve bearing wear are normally evidenced by the presence of whole series of running speed harmonics (up to 10 or 20). Wiped sleeve bearings often allow high vertical amplitudes compared to horizontal. Sleeve bearings with excessive clearance may allow a minor unbalance and/or misalignment to cause high vibration which would be much lower if bearing clearances were to specification.

Oil Whirl Instability
Typical Spectrum Shaft Diagram
oilwhirlspec.gif shaftdraw.gif
Oil Whirl instability occurs at 0.42 – 0.48 x RPM and is often quite severe. Considered excessive when amplitude exceeds 50% of bearing clearances. Oil whirl is an oil film excited vibration where deviations in normal operating conditions (attitude angle and eccentricity ratio) cause oil wedge to “push” the shaft around within the bearing. Destabilising force in the direction of rotation results in a whirl (or precession). Whirl is inherently unstable since it increases centrifugal forces which increase whirl forces. Can cause oil to no longer support the shaft, or can become unstable when whirl frequency coincides with a rotor natural frequency. Changes in oil viscosity, lube pressure and external pre-loads can affect oil whirl. 

Oil Whip Instability
Typical Spectrum
A Spectral Map showing Oil Whirl becoming Oil Whip Instability as shaft speed reaches twice critical. 

oilwhipspec.gif

Oil Whip may occur if a machine is operated at or above 2x rotor critical frequency. When the rotor is brought up to twice critical speed, whirl will be very close to rotor critical and may cause excessive vibration that the oil film may no longer be capable of supporting. Whirl speed will actually “lock onto” rotor critical and this peak will not pass through it even if the machine is brought up to higher and higher speeds.

 

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