Shaft Orbits

What Are Shaft Orbits Anyway?

The importance, and usefulness of shaft orbits andthe average shaft position as they relate to malfunctiondiagnostics, provides vital vibration information whichwould otherwise go unchecked.

A distinct advantage is gained by making use ofthese valuable tools in the problem solving processregarding today’s complex machinerymalfunctions.

Introduction

A myriad of vibration transducers are availablein today’s marketplace, but choosing the correcttransducer for a specific application is not only crucialfor accurate vibration monitoring, but for advanceddiagnostic capabilities also. The technical elementspresented here first intend to establish a viable workingrelationship between proximity probe generatedorbits, and shaft average position information. Then,further develop an understanding of these data formatsas they relate to the field of advanced machinerydiagnostics. A very common question often askedby our customers is... “Just what are shaft orbitsanyway, and furthermore how can they help mesolve my machinery problems?”

First, there are many ways to observe signalsgenerated by the non-contacting proximity probe,the most common of which is the bode and polarplot formats. These plots establish a rotor’sfrequency filtered amplitude and phase components,usually through transient and steady state operations.

WHAT ARE SHAFT ORBITS ANYWAY?by

MARK A. JORDANSr. Rotating Equipment Vibration Engineer

Industrial Machinery Diagsnotics, LLC

www.imd.us

About the Author

Mark Jordan is a Senior Field Vibration Engineer in the fieldof rotating equipment As a professional industry consultant,he is responsible for the diagnosis of mechanical & processmalfunctions in all types of industrial machinery for bothpetrochemical and power generation industries.Additioanlly, Mark is a qualified shaft alignment engineer,and uses optical surveing and lasers to rectify alignmentproblems on all variants of machinery. Until 2001 he was aSr. Field Engineer for Bently Nevada Corporation's MDS(Machinery Daignostic Services) group for 22 years beforethe company was purchased by General Electric PowerSystems (GEPS).

He received an Associate in Applied Science degree inElectronics Technology in 1984, and a Baccalaureate inMechanical Engineering from the University of Nevada-Renoin 1989.

Abstract

This paper discusses basic theory concepts coveringconsiderations associated with observing static anddynamic motion from non-contacting proximityvibration probes. Dynamic shaft displacementinformation is available from the proximity probes,but very often is not used to identify potentially harmfulmechanical malfunctions.

Using this proven diagnostic tool, it is possibleto isolate and identify potentially harmfulmachinery malfunctions by observing shaft orbitsin conjunction with the average rotor position within agiven bearing clearance.


in coplanar fashion resulting in what is known as shaftabsolute motion. The term “absolute motion” wasused historically because antique shaft riders originallyyielded this reading. Unfortunately, shaft riders dohave problems of reliability, and furthermore have nocapability to provide slow roll data. This severelylimits their use in machinery diagnostics, and evenmore so in balancing.

When non-contacting eddy current probes andProximitors® are used to monitor lateral shaftmotion, this transducer system provides thefollowing individual signal components:

1. A DC (Direct Current) signal whichmonitors the shaft average positionrelative to the probe mounting.

2. An AC signal (in this case, negativelyfluctuating) which monitors shaft dynamicmotion relative to the probe mounting.

In most plant applications, transducer signalsare usually processed by radial vibration monitors;these values are typically the machineryinformation that is displayed as the amount ofDIRECT (or overall) machine vibration in mils peakto peak (pp). Use of proximity probes are primarilyapplicable to those machines using fluid filmlubricated bearings such as seen in turbines, motors,pumps and compressors. Although a variety ofdifferent bearing types do exist, the use of proximityprobes is universally ideal, and the diagnostic capabilityis afforded equally to all. As with so many otherdiagnostic applications, the comparison of whatmachinery condition is normal or ideal, to whatactually exists is appropriate by observing proximityprobes that measure the machine’s dynamic rotormotion.

An arrangement for adequate machinerymonitoring (and protection) is to install orthogonal(X&Y) proximity probes mounted at each bearing.This provides the required AC/DC signals for on linemonitoring and diagnostics. When used in conjunction

However, an understanding of orbit-timebase andaverage shaft centerline position can prepare thereader to comprehend how the dynamics ofmachinery malfunctions take place, and moreimportantly, how these problems can be accuratelyidentified before a failure occurs.

Therefore, by monitoring shaft orbits and averageshaft centerline position within the bearings,important and relevant information to rapidlychanging machinery conditions is made available.

These concepts, illustrated in the followingtheoretical discussion, are followed by a casehistory in which shaft orbits and average shaftposition information become the key componentsin identifying and solving a serious and very realvibration problem.

Discussion

Diagnostic conclusions concerning rotating equipmentusing fliud film bearings are generally based upon thequality of data acquired from a machine’s transducersystem. Failure to observe proximity probe data in itssimplest form can lead to gross shortcomings in theinterpretation of the data obtained.

Bearing cap vibration information alone cannottruly indicate the rotor’s dynamic response whilein a state of malfunction. Casing absolutemeasurements acquired by seismic transducers(either velocity or accelerometer) can sometimesbe grossly inaccurate in the lower frequencyranges. Therefore, by utilizing a case mountedtransducer system by itself can only be viewedas an indirect method of quantifying a machine’soperational condition.

Conversely, proximity probes can measure thedirect relative response of the rotor to thestationary bearing housing. And for those machinesthat possess high bearing cap activity, both theproximity probe and casing transducer may be used


As machine speed increases (or decreases) duringstartup/shutdown, measurements are made ofchanging gap voltages from the two orthogonallymounted probes that indicate the amount traveledinside that bearing clearance by the rotor. At runningspeed, the rotor’s average position within the bearingis then easily identified when the referenced zero speedgap values are used. Values are calculated bysubtracting the running speed gap from the referencegap, then dividing through by the transducer’s scalefactor. By analyzing average shaft position within theknown diametral bearing clearance, valuableinformation regarding coupling and bearing alignment,overall bearing condition, oil film thickness and shaftradial loading etc., becomes available.

Shaft Eccentricity Ratio ε , is a dimensionlessquantity representing the average shaft position withina bearing (or seal). The average eccentricity ratio,obtained by dividing the distance between the averageposition of the shaft centerline and the bearing (orseal) centerline by the radial clearance, can varybetween zero and one. Zero represents shaftconcentricity with the bearing, and one represents theshaft in contact with the bearing. A trend ofdecreasing eccentricity ratio (ε→0) indicates apotential stability problem. Conversely, an increasingeccentricity ratio (ε→1) suggests the rotor isapproaching the constraints of the bearing wall.

Another property available from the shaft averagecenterline plot is the Rotor Position Angle Ψ. It isdefined as the angle between a vertically oriented linedrawn through the center of a bearing (horizontalmachines), and the line connecting the bearing andshaft centers, measured in the direction of rotation(positive angles). This parameter helps identify rotorpreloads.

AC Component - Orbits

The AC component of a transducer’s signal producesa periodic waveform from each probe in theorthogonal probe pair. A typical output waveform isshown in Figure 2 below.

with a once-per-turn reference probe (also known asthe Keyphasor®), the diagnostic capability is evenmore pronounced. It is from this transducerarrangement that machinery information such as shaftorbits and average shaft position becomes available.

DC Component-Average Shaft Position

For identification of a rotor’s average positionwithin a bearing clearance, voltage fluctuationsare seen by the proximity probes as relativedistance changes caused by dynamic rotor motionunder operating conditions. A method ofobtaining accurate shaft centerline data requiresthat one be cognizant a 'zero machine speed' gapvoltage reference is required (at rest or <100 rpm).For horizontal machines, this reference is generallyobtained with the rotor at rest, on turning gear or avery very slow rotative speed.

Figure 1: Shows the shaft average centerline positionwithin a 0.007” diametral bearing clearance using twoorthogonally mounted displacement probes. Note the ‘REF’values, these are the ‘at rest’ gap measurements.

In this condition, it is assumed the rotor is at rest inthe bottom of its bearings; therefore, all subsequentgap voltage changes are REFerenced to thisstarting position.

0.476

0.159

(-1.10, 2.26) 2990

POINT: BRG 1 Vertical

POINT: BRG 1 Horizontal

45 Left

45 Right

REF: -9.14 Volts

REF: -9.01 Volts

MACHINE: Turbine

From 06JUN95 16:44:30 To 06JUN95 16:53:05 Startup

0.5 mil/div

0

2

4

6

-4 -2 0 2 4

203200590

1240

1830

2340

2930

(not orbit or polar plot) UP


Figure 2: The upper waveform represents the rotor’s synchronous (filtered to 1X) lateral vibration response, while thelower is a representation of DIRECT vibration.

Figure 3: The graphical result of plotting Equations 1 and 2 from time T1 to time T2. To the right of the waveforms, theassociated shaft orbit plotted as amplitude vs. amplitude. Numerical points (1, 2, 3 etc,.) along the timebase waveformscorrespond to specific points on the orbit precession. The same inputs are used for Figure 4.

180

270 -0.5

180

-1

+0.5

2

90

0+1

X AXIS

Y AXIS

HORIZONTAL

INPUT

VERTICAL

INPUT

VERTICAL TIMEBASE WAVEFORM

AMPLITUDE VS. TIME

HORIZONTAL TIMEBASE WAVEFORM

TIMEBASE

AMPLITUDE VS. AMPLITUDE

ORBIT

1

3

5

4

6

7

8

9

ROTATION=XtoY (CCW)

450 90 135 315225 270 360

450 90 135

T1

-1

-0.5

1

0.51

3

4

2

T1

-1

1

0

12 3

4

180 315 225 270 360

7

5 6

T28

9

5

76

T2

89

P OINT: B RG 3 V e r tic a l 60 Lef t V E C TOR:0.63 2 m il p p 8 9 S R: 0.1 08 21 5M A C HINE : C om p resso r10 M A R9 5 22 :04 :5 7S ta r tup 1 X C OM P

0 .0 5 m il/d iv C W ROTA TIO N 5 m sec /d iv 27 92 rpm

0 .4

0 .2

0 .0

0 .2

0 .4

0 2 0 4 0 6 0 8 0 10 0

P OINT: B RG 3 V e r tic a l 60 L ef t D IR A M P T: 0 .69 9 m il ppM A C HINE : C om p resso r10 M A R9 5 22 :04 :3 9S ta r tup D IRE C T

0 .0 5 m il/d iv C W ROTA TIO N 5 m sec /d iv 27 90 rpm

0 .4

0 .2

0 .0

0 .2

0 .4

0 2 0 4 0 6 0 8 0 10 0


Note that in Figure 2, two separate waveforms areshown; the upper waveform, filtered to running speed(1X), shows a smooth sinusoid while the lowerrepresents the unfiltered DIRECT vibration. Note,when observing DIRECT data, unless it has beenspecifically compensated, this data includes everycomponent between 0 (DC) and 10 KHz (600,000cpm) including the operating parameters such as the1X, 2X and nX components plus scratches, shaftovality and electrical/mechanical runout.

To establish how shaft orbits are formed, we mustfirst consider that the waveform produced by eachtransducer is an individually processed vibration signal.This signal, generated at a specific angular locationon the rotor, relates the amount of shaft lateral motionin that plane. When two probes are mountedorthogonally (XY configuration, separated by 90°),the two individual signals (waveforms) arerepresentative of shaft peak to peak displacement intheir respective angular planes, and are plotted asamplitude vs. time (Again, refer to Figures 2 and 3).

The origin of a shaft orbit begins when two XYwaveforms are paired together so that the timeelement is removed, leaving the X (horizontal plane)amplitude component vs. Y (vertical plane) amplitudecomponent, plotted in what is commonly known asthe Cartesian Coordinate system (or polarcoordinate system). To better illustrate this concept,take a pair of XY timebase waveforms which areseparated by a phase difference of 90°, and whosewaveform amplitudes are 2.00 mil pp (from the verticalprobe), and 1.00 mil pp (from the horizontal probe).

Subsequently, these two signals are best describedby the following equations:

X(θ)horizontal = 1.00 Cos (θ) (Eq. 1) Y(θ)vertical = 2.00 Sin (θ) (Eq. 2)

Where θ=ωt (ω=rotational rotor frequency, t=time)representing the angle over one shaft revolution (T1 to T2)in radians and the numerical values (2.00 mil & 1.00 mil) arethe amplitudes of lateral shaft vibration.

In Figure 3, equations 1 and 2 are plotted in theamplitude vs. time domain (waveform). The sameresult is achieved using a machine’s 1X filtered XYwaveform pair, (or unfiltered waveform pair) per eachbearing, resulting in a 1X filtered orbit, (or a DIRECTvibration orbit). Waveforms and DIRECT orbitprecessions are typically available for display on atwo channel oscilloscope. It is important to note thatthe scope should have a third channel, a “Z” channelfor a Keyphasor® signal input. Refer to Figure 4.

T1 T2

T1 T2

VERTICAL

HORIZONTAL

T1 T2

1 REVOLUTION

SCOPE SCALE: 0.500 mil/DIV

2mil pp.

1mil pp.

1mil

2 m

il

KEYPHASOR

TRIGGER

INTENSITY

Figure 4: Shows two individual time base signals beingfed to an oscilloscope along with a Keyphasor® signal. Theorbit amplitude spans 2.00 mil pp vertically (Y direction), and1.00 mil pp horizontally (X direction) with the XY channelamplitude scales set at 0.500 mil per division.


Y: BRG 1 VerticalX: BRG 1 Horizontal

45 Left45 Right

VECTOR:0.360 mil pp 87VECTOR:0.211 mil pp 174

SR: 0.73 26 (man)SR: 0.76 117 (man)

MACHINE: Turbine13JUL95 10:52:22Startup 1X COMP

0.05 mil/div ROTATION: Y TO X (CW) 2 msec/div 3770 rpm

UP Y: BRG 1 Vertical

X: BRG 1 Horizontal

0.4

0.2

0.0

0.2

0.4

0.4

0.2

0.0

0.2

0.4

0 20 40 60

Figure 5: Shows four revolutions of the rotor with the 1Xtimebase waveforms to the right, and the associated 1X shaftorbit to the left. Shaft rotation is Y to X (CW).

Understanding the Data

Machinery diagnostics depends on knowledge of amachine’s specific bearing parameters. Elements suchas diametral clearance and specific bearing type arehelpful when applying the diagnostic techniquesdiscussed here.

But larger benefits are realized when the vibration datais slow roll compensated. Slow roll vectors areobtained during a period of low rotative speed whendynamic motion effects from forces such as unbalanceare negligible. At this speed (usually 100-400 rpm),shaft bow and electrical-mechanical runout can bemeasured, and subtracted. Typically, slow roll speedshould be below 10% of the first balanceresonance.Let us now review a simple example ofslow roll compensation.

When two vibration signals are placed into a dualchannel oscilloscope (or data acqusition device) andobserved on its display, the amount of vibration canbe displayed in timebase (sinusoid waveform, Figure2) or in orbital form (as shown in Figure 4).In its orbitmode, the oscilloscope places the Vertical (Y) andHorizontal (X) signals along their respective axis tocreate a display of amplitude vs. amplitude. The formin which this takes place is governed by the followingequations:

X(r,θ)horizontal axis ≡ r Cos (θ) (Eq. 3) Y(r,θ)vertical axis ≡ r Sin (θ) (Eq. 4)

Where θ=ωt (ω=rotational rotor frequency, t=time) represents one shaft revolution

(in radians), and r denotes lateral shaft amplitude.

Simply put, an orbit pattern as seen on an oscilloscopeis nothing more than a light beam dot moving veryrapidly such that to the eye it appears on the screenas a continuous line. This rapidly moving dotrepresents the centerline motion of the shaft as seenby the proximity probes. Therefore, the orbit simplyrepresents the PATH of the rotor centerline atthe lateral position of the proximity probes.

The Keyphasor® pulse, when fed to a “Z” intensityinput of an oscilloscope, intensifies the dot at the instantin time when the keyway (once-per-turn event) ispassing under the Keyphasor® probe. TheKeyphasor® dot on an orbit (or waveform)represents the centerline location of the shaft inits path of travel (or high spot) at the instant that thekey way is in front of the Keyphasor® probe. Thisallows identification of a fixed physical reference tothe shaft. This arrangement produces not only peakto peak amplitude, but important phase informationthat is used for diagnostics. Figure 5 below showsactual field data processed by a vibration diagnosticssoftware package. By taking the orbit/shaft centerlineexperience one step further, it is rotor loading and/ordifferences in dynamic stiffness at a bearing locationthat is indicated by the average shaft position withinthe bearing clearance, and the orbit’s elliptical shape.

Notice the orbit representation in Figure 5 is slightlyelliptical. This data suggests the rotor is in goodoperational position with normal minor influentialelements present, such as gravity, fluidic, and bearingload forces.


P O IN T: B R G 3 V e r tic a l 4 5 L e f t V E C TO R :0 .1 7 2 m il p p 3 1 1M A C H IN E : C o m p r e s so r0 6 J U N 9 5 1 7 :0 8 :0 9S ta r tu p 1 X U N C O M P

0 .0 2 m il/ d iv C W R O TA TIO N 2 m se c /d iv 2 9 9 0 r p m

0 .2

0 .1

0 .0

0 .1

0 .2

0 1 0 2 0 3 0 4 0 5 0

P O IN T: B R G 3 V e r tic a l 4 5 L e f t D IR A M P T: 0 .6 2 6 m il p pM A C H IN E : C o m p r e s so r0 6 J U N 9 5 1 6 :5 3 :0 5S ta r tu p D IR E C T

0 .0 2 m il/ d iv C W R O TA TIO N 2 m se c /d iv 2 9 9 0 r p m

0 .2

0 .1

0 .0

0 .1

0 .2

0 1 0 2 0 3 0 4 0 5 0

Figure 6a: Uncompensated 1X and DIRECT data.

P O IN T: B R G 3 V e r tic a l 4 5 L e f t V E C TO R :0 .0 6 9 m il p p 3 2 4 S R : 0 .1 0 5 3 0 3M A C H IN E : C o m p re ss o r0 6 J U N 9 5 1 7 :0 8 :0 9S ta r tu p 1 X C O M P

0 .0 2 m il/d iv C W R O TA TIO N 2 m s e c /d iv 2 9 9 0 r p m

0 .2

0 .1

0 .0

0 .1

0 .2

0 1 0 2 0 3 0 4 0 5 0

P O IN T: B R G 3 V e r tic a l 4 5 L e f t D IR A M P T: 0 .2 3 9 m il p pM A C H IN E : C o m p re ss o r0 6 J U N 9 5 1 6 :5 3 :0 5S ta r tu p D IR E C T C O M P

0 .0 2 m il/d iv C W R O TA TIO N 2 m s e c /d iv 2 9 9 0 r p m

0 .2

0 .1

0 .0

0 .1

0 .2

0 1 0 2 0 3 0 4 0 5 0

Figure 6b: This figure serves to illustrate the marked difference in the vibration waveforms when slow roll compensationis employed. This data is the compensated 1X and DIRECT waveforms as shown in Figure 6a. ‘SR’ indicates the slow rollvector.

By subtracting these slow roll ‘vectors’ from a given data set, the true vibration response is achieved. Slow rollcompensation can be performed on 1X, 2X, nX and DIRECT data sets. Figure 6 strongly illustrates thisconcept.

Three simple cases exemplifying changing machinery conditions are now presented.

Each case shows a relative shaft position within the bearing clearance, along with an associated orbit shape.These examples progress from normal operation, to a state of high preload malfunction. They are illustrationstaken from actual data sets.


Case 2. Increasing Preloading

The next data set (Case 2) demonstrates a malfunction,such as misalignment between two machines, thataffect both the shaft position, and orbit display as aresult of induced shaft preloading. A change of shapein the shaft orbit can, for example, provide anindication to changing preloads (i.e. misalignment) thatmay be acting on the rotor. A preload is typicallydefined as unidirectional, axial or radial (side) loaddue to external or internal mechanisms. It can alsoact to stabilize or destabilize the dynamic condition of

Y: IB-YX: IB-X

090 Right

DIR AMPT: 0.735 mil ppDIR AMPT: 1.40 mil pp

MACHINE: Rotor Kit08MAY96 08:28:32Startup DIRECT

0.1 mil/div ROTATION: Y TO X (CW) 1840 rpm

UP

Y: IB-YX: IB-X

090 Right


SR: 0.319 0 (man)SR: 0.411 187 (man)

MACHINE: Rotor Kit08MAY96 08:28:32Startup 1X COMP


UP

0.403

-0.183

(-0.916, 2.01)

POINT: IB-Y

POINT: IB-X

0

90 Right

REF: -10.21 Volts

REF: -9.79 Volts

MACHINE: Rotor Kit

From 08MAY96 08:28:23 To 08MAY96 08:29:17 Startup

0.5 mil/div

0

2

4

6

8

-4 -2 0 2 4


DIRECT ORBIT 1X ORBIT SHAFT CENTERLINE

Case 1. Acceptable Operation

Orbit-timebase and shaft centerline data shown inFigure 7 (Case 1) represents a horizontally mountedmachine in good operating condition, with a shaftrotation of Y to X (clockwise).

This machine operates using a plain sleeve journal

bearing design. The shaft centerline data shows therotor to be in the lower left quadrant (ε ≅ 0.5), whichis acceptable and normal for the Y to X (clockwise)shaft rotation. The orbit shape can also be consideredacceptable under the aforementioned parameters.

the machine. If the restraining forces (the dynamicstiffness) in the machine are equal in all radialdirections, with the only force acting on the rotor is itsresidual imbalance, then the orbit, theoretically, shouldbe completely circular. Dynamic stiffness refers to thespring stiffness of the mechanical system,complemented by the dynamic effects of mass anddamping.

Dynamic stiffness is a characteristic of the mechanicalsystem, and it opposes an applied dynamic force tolimit vibration response. A common (and mistakenly

Figure 7: An acceptable unfiltered orbit at 3600 rpm along with its associated shaft centerline travel in the bearingduring start up. Notice the machine rotation is Y to X (clockwise), as indicated by the shaft position in the LOWER LEFTQUADRANT of the shaft centerline plot. (for CCW rotation, this shaft position would be in the lower right quadrant,and the orbit shape would be a mirror image of the one shown). Eccentricity ratio shown is approximately 0.5.


Case 3. Heavy Preloading

Illustrated in Case 3, the condition of progressivepreload may result in the orbit shape transgressing toa shape resembling a “figure 8”. The associated shaftaverage position is shown to be adjacent to the bearingwall in the shaft centerline plot.

All are indicators of excessive rotor loading that maybe the result of misalignment coupled with excessivepipe strain and/or severe bearing problems. Amachinery train experiencing such problems shouldbe considered a candidate for immediate analysis andphysical inspection

8, and 9.so) notion is to think that machines in good workingorder should possess perfectly round orbits. This issimply not true, especially for machines with plainsleeve type bearings. Other forces, or unequalrestraints cause the rotor to respond in various non-circular shapes such as those illustrated in Figures 7,

Y: IB-YX: IB-X

090 Right


MACHINE: Rotor Kit08MAY96 09:57:30 Steady State DIRECT COMP


UP

Y: IB-YX: IB-X

090 Right

VECTOR: 2.59 mil pp 153VECTOR: 0.740 mil pp 224

SR: 0.452 129 (man)SR: 0.380 221 (man)

MACHINE: Rotor Kit08MAY96 09:57:30 Steady State 1X COMP


UP


Figure 8: A condition of lateral loading malfunction evidenced by shaft centerline travel into the upper left quad-rant, and an elliptical orbit shape caused by a heavy resultant force acting on the rotor. Conditions such as this aretypically seen on the inboard bearings of turbines/compressors when misalignment is present. Additionally, this mayalso be seen on the number one bearing of steam turbines, mainly due to the summation of dynamic forces present i.e.,steam flow, gravity, bearing preload, fluid properties, etc.. Eccentricity ratio is very high, ε = .70-.80.

1.28

-0.500

(-2.50, 6.41) 3890

POINT: IB-Y

POINT: IB-X

0

90 Right

REF: -10.38 Volts

REF: -9.94 Volts

MACHINE: Rotor Kit

From 08MAY96 09:54:29 To 08MAY96 09:59:11 Shutdown

1 mil/div

0

5

10

-5 0 5

410420500

610

780

1484

1486

1486

1483

1485


The DIRECT orbit shows a reverse vibrationprecession, physically caused by the ‘figure 8’ orbitprecession. Also shown is the 1X vibration precessionwhich shows a forward vibration precession. Theshaft average centerline depicts that a high eccentricityratio exists, and this value can be determined inapproximately the 0.95 region (rotor close to thebearing wall).


Y: IB-YX: IB-X

090 Right


MACHINE: Turbine08MAY96 09:07:47 Steady State DIRECT COMP


UP

Y: IB-YX: IB-X

090 Right

VECTOR: 5.13 mil pp 138VECTOR: 1.75 mil pp 225

MACHINE: Turbine08MAY96 09:07:47 Steady State 1X UNCOMP


UP

1.10

-0.977

(-4.88, 5.49) 3840

POINT: IB-Y

POINT: IB-X

0

90 Right

REF: -10.44 Volts

REF: -9.48 Volts

MACHINE: Turbine

From 08MAY96 09:04:50 To 08MAY96 09:08:09 Shutdown

1 mil/div

0

5

10

-5 0 5

2703602390

2660

2870

3080

3210

3320

3660



Figure 9: Shows a condition of heavy loadingevidenced by shaft centerline travel against the bearingwall. A Figure “8” DIRECT orbit shape suggests reversevibration precession, indicating the preload condition issevere. Eccentricity ratio very high, ε = 0.95. The 1X orbitshows a planar type orbit (highly restricted).

Preloads - A General Summary

The three cases illustrated here show how shaftcenterline position and orbit shape might respond withtime to an increasing steady state unidirectionalpreload, progressing from normal conditions to heavypreloading, resulting in the classical Figure “8” pattern.

Therefore, by observing orbital patterns over time,the degree and plane of a preload condition can bedetermined and tracked.

Typically, a heavy preload is not a perfectly shapedFigure “8”, but is more prone to displaying loops ofdifferent sizes as shown in case 3 above. The casehistory following this discussion illustrates this pointprecisely. Preloads affecting a rotor system can fallinto several different categories. Radial preloads canbe caused by gravity, fluidic forces, abnormal bearingloads (especially internally adjustable types), seals,and the effects of pipe strain on the machine caseitself.

Preloads can be categorized into two types:

Soft Preloads: Soft preloads include the effects ofgravity, thermal misalignment and process changes thatcan act on the rotor system.

Hard Preloads: Due to high eccentricity ratiosresulting in increased direct stiffness in the bearing.Probable causes include gross shaft to shaftmisalignment, and piping forces acting on themachinery casing. In either case, preloads can bedamaging if the forces involved are strong enough tocreate fatigue. The results of this can be realized byexcessive bearing wear, shaft fatigue, and shaft cracksetc.. By observing orbit shapes, increasing rotorpreloads will force the rotor’s precession into a moreelliptical shape. An initial response is a change in the1X and DIRECT amplitudes followed by an increasein other frequencies such as 2X, indicating that severemisalignment, or other malfunctions are present.

Either case can present harmful effects on theoperational condition of the machine which can leadto permature part failure or even a cracked rotor.


this powerful diagnostic tool.

-Acceptable

Figure 11 illustrates a rotor with minimal preload (otherthan normal loading forces, gravity, fluidic etc.), whosepredominant frequency is 1 times running speed.

Orbit Direction & Vibration Precession

It is important to recognize when the vibrationprecession is reversed. Figure 10 shows two orbitsof the same size and shape, with different orientationsand precessions. By simple observation of the orbits,the differences are clear. Orbit “A” might beconsidered normal for a bearing with the journalrotating X to Y (counterclockwise), but abnormal forY to X (clockwise) rotation. The converse is true fororbit “B” - fairly normal for CW rotation, andabnormal for CCW rotation.

Note the Keyphasor® dot and the preceding “blank”section in the orbit precession. From this mark, thedirection of vibration precession can be determined.The term blank-bright infers the vibration directiongiven by the Keyphasor® location on the orbitprecession, noting that the machine’s actual rotationaldirection under normal conditions should coincide withvibration direction. This convention is establishedby always viewing the transducer locations fromthe driving element, regardless of shaft rotation.

Common Malfunctions-An Overview

Orbit data presentations hold great value to themachinery vibration specialist. Many different typeof malfunctions can be identified through orbit analysis.A few of these malfunctions are illustrated here togive a flavor of how much can be recognized by using

Figure 11: Representative of a machine with minimalpreload characteristics, and no abnormal problems.Predominant amplitude frequency is 1X. Shaft rotation is Xto Y (counterclockwise).

-Lateral Load Malfunctions

When machine-to-machine misalignment across thecoupling (the most common preload cause) is present,the shaft orbit loading might resemble something similarto that shown in Figure 12. Increased angular and/orparallel offsets between two rotors is the most commoncause of machinery misalignment. Other factorsaffecting the overall alignment condition can beattributed to something more complex such asfoundation degridation when subsidence is detecteddue to elements such as concrete contamination (oil,cracks) or the support capability of the ground itself.By experience, the number of failed foundations thathave been found by this author are many.

Figure 10: Orbit precession (shown inside the bearingclearance). The Keyphasor® blank-bright spot showsthe direction of vibration precession when viewed onan oscilloscope. (Note, this serves to illustrate the rotordirection of two different machine trains).

CCW ROTATION CW ROTATION

ORBIT

0.200 mil/DIV ROT: CCW


-Fluid Induced Instabilities

Figure 13 shows the presence of a subsynchronouscomponent, illustrated by the two Keyphasor® dots.The precession of the whirl vibration component is

For a common plain sleeve fluid film bearing, the selfexcited vibration frequency of a whirl instability istypically within the range of 30% to 48% of machineoperating speed. When fluid whirl is viewed on anoscilloscope, a key identity to look for is that the twoKeyphasor® dots rotate slowly against machinerotation. This indicates that the subsynchronousfrequency is less than 50% of machine speed (usuallyindicating a fluidic malfunction caused by a bearingdeficiency). Conversely, if the dots remain stationary,then the frequency is exactly half (50%) of machinespeed (indicative of a rub). Lastly, if the dots appearto rotate forward, (with machine rotation) then thesubsynchronous frequency is above 50% of runningspeed (usually indicative of a rotor, or structuralresonance excitation).

But relating to this topic further, there is always theage old question bantered about regarding thecapability of a fluidic induced instability occuring intilting pad bearings. In short, the answer to this is adefinitive "NO!" In the bearing itself? - absolutelynot. But, experience has shown that fluid whirl (andeven whip) can be generated on the bearing casingitself, usually on the small lip between the casing andthe shaft (which if the clearnance is too tight acts likea minature bearing). The best way to stop this is tocut small marks into the casing land area to preventcircumferential fluid flow of the oil exiting the bearingbetween the bearing carrier and the shaft.

Normally, fluid whirl is a self excited fluidic malfunctionthat typically occurs in plain sleeve bearings. The mostcommon instance is misalignment (unloading of abearing where the eccentricity ratio approached 0)or a much more common cause is an excessive bearingclearance.

Progressing beyond whirl is a condition known as whipwhich can be extreamly destructive to machinery ifnot properly diagnosed and rectified.

Figure 14 shows a destructive bearing malfunction inprogress. Typically known as fluid whip, its vibration

Figure 12: Representative of a machine with poor alignment,or a “cocked” bearing assembly. Further investigationwarranted. Note, large amplitude present in the major planeof vibration.

Figure 13: Subsynchronous components at approximatelyhalf running speed. Suggests self excited fluid whirl. Note -TWO Keyphasor® dots. Usually accompanied by largervibration amplitudes.

ROT: CCW2.00 mil/DIV

ORBIT

always forward (same direction as machine rotation),and the orbit shape is usually circular.

PRELOAD FORCE

ROT: CW0.500 mil/DIV

ORBIT


Figure 14: Representative of a machine withsubsynchronous excitation at a first balance rotor resonancewhen the machine speed is well above two times thatresonance. Typically referred to as fluid whip. NOTE -Multiple Keyphasor® dots - THIS MALFUNCTION CANBE VERY DESTRUCTIVE.

This condition is accompanied by large vibrationamplitudes that usually traverse the bearing clearance.

-Rotor Contact Malfunctions - Rub

Shaft position and orbit shapes can indicate anotherform of machinery malfunction commonly seen intoday’s applications.

Rubs are encountered when the rotating shaft contactsa stationary part of the machine. Malfunctions includeshaft contact with seals (usually with minimal radialclearances), newly installed steam packing, contactof turbine/compressor blading due to a failed thrust

bearing, abnormal case anomalies due to thermalwarping, etc..

A rub occurs as a secondary effect to a machinerymalfunction, resulting in increased vibration levels, achange of orbit shape and average shaft centerlineposition. Once the rub continues, this malfunctionresponse becomes dominant. Unlike fluid inducedwhirl and whip, a rub can take on many different orbitshapes. From looping figure “8’s”, with increasingamplitudes over time, to a complete circular orbitshape (full annular rub - which would be fullyencompassing the seal or bearing clearance in whichthe rotor is in contact with).

Indications that a machine is in this state of malfunctionmay also be identified by increased heat in the system.Lubrication oil and process temperatures may rise dueto friction as heat is added to the system. This runsconcurrent with increased vibration levels throughoutthe train, most likely due to shaft bow.

The two main types of rub malfunction may beclassified as follows:

Partial Rub: Occurs when the rotor occasionallycontacts a stationary part of the machine. Thefundamental frequency most often at 50% of runningspeed. Therefore, the orbit may resemble a figure“8” , but with two stationary Keyphasor® dotspresent. Due to its non regularity through partial rotorcontact with the stationary part(s), other frequenciesmay also appear in the orbit’s progression. Submultiples of running speed such as 0.25X, 0.32X etc.,may be recognized by additional Keyphasor® dots.For example, a 33% (1/3X) frequency may appearas a twisted looping orbit with three Keyphasor® dotsin its vibration precession as shown in Figure 15.

precession is forward. The period of this self excitedmalfunction may or may not be harmonically relatedto the rotational period of the shaft. When it is not,then the Keyphasor® dots move in a seeminglyrandom pattern; when it is, the multiple Keyphasor®dots will appear to be stationary.

ROT: CW2.00 mil/DIV

ORBIT


rub excitation force, or the machine self destructs. Inany case, with this malfunction present, the machineis in imminent danger of total destruction.

When the rotor and seal are in contact, the two majorregimes of the rotor vibration under steady stateconditions are:

1. A forced synchronous precession due to unbalanceexcitations. The rotor bounces inside the seal, the lowestfrequency being 1X.

2. Forced self-excited circular reverse precession with afrequency corresponding to one of the natural frequenciesof the rotor system.

Therefore, the orbit shape is predominantly circularto the extent of the seal clearance, with a reverseprecession orbit. The only recommended course ofaction under these circumstances is to shut themachinery train down, further inspect the damage, andcorrect the initial problem.

A Case History

A case history is presented in which the usefulness ofshaft centerline and orbit data are clearlydemonstrated.

The subject machine consists of an Elliott P-90overhung centrifugal compressor driven by a six stageElliott 2QNV-6 steam turbine. The range of operationis between 1800 and 3100 rpm. Throughout the lastseveral months, this plant experienced a multitude ofvibration problems during the operation of thismachine.

Bently Nevada’s Machinery Diagnostic group wascontracted to perform an in depth vibration analysisof this unit. As a result, through data acquisition, theanalysis determined that the turbine rotor’s vibrationprecession was severely restricted at the inboardbearing (#2). The XY proximity transducers mountedpermanently at each radial bearing provided therequired vibration data to accurately diagnose thismalfunction.

Figure 15: Representative of a machine with 1/3Xsubsynchronous excitation caused by a partial rub condition.Note the positioning of the Keyphasor® dots.

For ease of rub identification, the following conditionsapply to partial rub analysis:

Common Frequencies Generated by Partial Rub

When Frequencies Generated

ΩΩΩΩΩ < 2 ωωωωωresonance 1X

ΩΩΩΩΩ ≥≥≥≥≥ 2 ωωωωωresonance 1X or ½X

ΩΩΩΩΩ ≥≥≥≥≥ 3 ωωωωωresonance 1X, ½X or_X

ΩΩΩΩΩ ≥≥≥≥≥ 4 ωωωωωresonance 1X, ½X,_X or ¼XWith other ratios possible

Where: Ω is Rotor Speed, and ω is Rotor First BalanceResonance.

Full Annular Rub: Most often encountered inseals, when the seal clearance interferes with therotating element. The precession of vibration in thefull annular rub case is reversed to that of rotordirection. This condition is very destructive, with itsorbit shape traversing the full extent of the bearing orseal clearance. Once initiated, the vibration mayworsen until the system restraint forces overcome the

ORBIT

0.500 mil/DIV

CONTACT

REGION

3

ROT: CW

CW ROTATION

3600 RPM

2

1


Figure 15 Subject machinery arrangement. This machine operates with plain sleeve journal type bearings.


60 Left30 Right


SR: 0.4 147 (man)SR: 0.329 238 (man)

MACHINE: HP Turbine06MAR95 09:53:09Steady State 1X COMP



X: BRG 2 Horizontal

0.5

0.0

0.5

0.5

0.0

0.5

0 10 20 30 40 50

Figure 16: Shows a planar (highly preloaded) reverse orbitvibration precession at the No. 2 bearing.

The data contained in plot 16 shows heavy rotorpreloading at turbine’s No. 2 bearing. Note that thepreload force is predominantly oriented in the verticalplane (as indicated by the upward arrow). This datawas produced once coast down and slow roll vectorswere acquired, which were then subsequently appliedto the steady state data. Similarly, the DIRECT

vibration precession at the same location is shown inFigure 17.


60 Left30 Right


MACHINE: HP Turbine06MAR95 09:53:09Steady State DIRECT COMP



X: BRG 2 Horizontal 1.0

0.5

0.0

0.5

1.0

1.0

0.5

0.0

0.5

1.0

0 10 20 30 40 50

Figure 17: Shows a DIRECT highly preloaded reverseorbit vibration precession at bearing #2. A small ‘figure 8’can be seen on the lower left tip of the orbit.

It is feasible to imagine that the vibration precessionshown in Figure 17 looks as if it is part of a largerradius. The circle in question would be the radius ofthe bearing itself. Shown in Figure 18, the DC shaftaverage centerline plot for bearing No. 2. shows thatthe shaft position appears to be highly restricted in


During that past outage, it was discovered that theturbine had been physically removed from itsfoundation for some off site casing work. Armed withthis historical information, we turned to the vibrationdata to complete the analysis.

Using both plot formats together, the extent of thismalfunction was quickly realized. There are severalelements provided by the machinery vibration data toshow that the malfunction is indeed very real. Toachieve an accurate diagnosis and provide a clearrecommendation to resolve the problem, the vibrationdata was analyzed in detail. We shall call data set #1,the machinery ‘as found’ , or original machinerycondition data, which was acquired on 6 March 1995.Using this data, the following preliminary observationswere made:

1. The rotor’s vibration precession clearly showsseverely restricted orbits which are presented in plots16 & 17. The malfunction state was more pronouncedonce the data set was slow roll compensated.

2. The average DC shaft position (plot 18) indicatesthe rotor is positioned in the wrong quadrant.

Diagnosis

The rotor progression within the bearing clearance isvery minimal, so much so that it appears that the rotoris ‘hard’ against the bearing wall. Additionally, therotor travel should not be so restricted, rather theeccentricity ratio should be in the region of 0.3 to 0.4.The eccentricity ratio for the rotor position shown inplot 18 is approximately 0.99. Under normal operatingconditions for a machine with Y to X (CW) shaftrotation and one that is equipped with plain sleevetype bearings, the rotor should be in the lower leftquadrant. This evidence of hard preloading of therotor to the bearing corresponds with high bearingmetal temperatures observed by the plant on thisbearing.

The induced preload is severe enough to produce a

the lower right quadrant of the bearing. This wouldbe fundamentally incorrect for a plain sleeve typebearing with a shaft rotation of Y to X (clockwise).

-0.136

0.212

(1.10, 0.549) 2695

POINT: BRG 2 Vertica


60 Left

30 Right

REF: -8.69 Volts

REF: -9.05 Volts

MACHINE: HP Turbine

From 06MAR95 09:53:09 To 06MAR95 10:02:46 Steady State

0.5 mil/div

0

2

4

6

-4 -2 0 2 4

2298


Figure 18: Shows a highly restricted rotor precessionwithin the No. 2 bearing clearance.

As a part of the original analysis, several pertinentfactors were discovered by asking questions pertainingto this machine’s past performance.

• It was discovered that the No. 2 bearingtemperature fluctuated over time, and was initiallythough to be a problematic thermocouple.

• The vibration levels changed unexpectedly atbearing No. 2 as process changes were initiated.However, large amplitudes were not seen, andthe monitor alarms were not initiated

• System performance was under suspicion,basically because steam consumption wasmarginally higher than before the last outage.

• Finally, the major pieces of the puzzle were drawntogether after discovering that the machine hadexperienced mechanical problems only since thelast major overhaul.


reversed orbit vibration precession. This suggests thepresence of very high preload forces, subsequently,bearing damage at No.2 has been predicted. Arecommendation to was made to the plant to removethe machine from service, inspect the bearings andverify the shaft alignment.

The machinery train was subsequently shutdown, andallowed to cool. Using a coupling laser alignmentsystem, the static (cold) shaft offsets were checked.

It was discovered that the turbine elevation was toohigh by as much as 0.010” (10 mil). The lateral(horizontal) offsets appeared to be within the desiredtolerance. The correct static shaft offsets werecalculated, and the turbine elevation was adjustedaccordingly.

As a matter of good investigative procedure, bothturbine bearing caps were removed for physicalinspection. Bearing No.2 (turbine inboard) was founddamaged, and was subsequently replaced. Themachine was prepared for startup, and excerpts fromthat data set (data set #2) are presented here toillustrate the dramatic improvement in the operatingconditions.


60 Left30 Right


SR: 0.272 189SR: 0.293 278

MACHINE: HP Turbine10MAR95 22:00:06Startup 1X COMP



X: BRG 2 Horizontal

0.5

0.0

0.5

0.5

0.0

0.5

0 10 20 30 40 50

Figure 19: Shows the 1X compensated orbit vibrationprecession at the No. 2 bearing after the alignment corrections.Notice the orbit orientation, and correct vibration precession.This plot is scaled for direct comparison to plot 16.

The high preload condition previously illustrated indata set #1, figure 17, was alleviated after specificmachinery alignment corrections. This courseof action resulted in the orbit and timebase datapresented below.


60 Left30 Right


MACHINE: HP Turbine10MAR95 22:00:06Startup DIRECT COMP



X: BRG 2 Horizontal 1.0

0.5

0.0

0.5

1.0

1.0

0.5

0.0

0.5

1.0

0 10 20 30 40 50

Figure 20: Shows the DIRECT response to the correctionof abnormal coupling alignment. This plot is scaledaccordingly with plot 17 for comparison purposes.

0.378

0.403

(-0.610, 2.69) 2650

POINT: BRG 2 Vertical


60 Left

30 Right

REF: -8.96 Volts

REF: -8.89 Volts

MACHINE: HP Turbine

From 10MAR95 20:22:34 To 10MAR95 22:00:06 Startup

0.5 mil/div

0

2

4

6

-4 -2 0 2 4

535650

760

10201145

1155

12351545

1480

2185


Figure 21: Represents a significant improvement in rotortravel. Again, the plot is scaled accordingly with plot 18 forcomparison.


Finally, the shaft average centerline data shows amarked improvement in rotor travel within the bearingclearance, and shaft eccentricity is approximately 0.3.It was discovered that the turbine shims had beenremoved, and 'cleaned' in preparation for the turbine'sreturn from the machine shop.

Apparently, some of the original shims had beenmisplaced, and were replaced by "equivalent" shimplates. This was thought to be the main cause ofmalfunction in this particular case.

Conclusions

In today’s continuously changing machinery needs,there is one thing that remains constant, namely, theneed for a reliable method of accurately monitoringindustrial machinery applications. This conceptbecomes critical when dealing with machinery that relyon fluid film hydrodynamic bearings.

Continuing to observe signals from proximity probesin the simplest form is basic to quality machinerymonitoring and machinery diagnostics. Manyparameters needed in complex machinery malfunctionanalysis cannot be determined accurately by any othermeans. Full comprehension of rotating equipmentproblems and impending failures can best be viewedas insurance policy against costly mechanical failuresthat could have easily been identified and correctedwith minimal financial impact. In our experience, manymachinery failures were allowed to progress becuasethey were not delt with, recognized or identifiedcorrectly leading to extreame financial losses as aresult.

_____________________________________

2003 Mark A. JordanRevision 5

Shaft Orbits

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