7-1 Machine Diagnosis

8/8/2019 7-1 Machine Diagnosis

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E X A M P L E S O N M A C H I N E R Y F A I L U R E

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MACHINE DIAGNOSIS1. In t roduct ionAfter the vibration signatures are verified as to validity and the spectral peaks,especially the 1X components positively identified, can the diagnosis of machineproblems begin. The following section discusses a variety of machine problems andillustrates them with their typical vibration signatures.

In analyzing vibration spectra from rotating machines, it is important to note thatindividual faults are seldom seen by themselves. Care must be taken in theinterpretation of vibration signatures since different faults can cause spectralcomponents at the same frequencies.Imbalance

2.Calcu la t ing the Im balance Force,

where F = the imbalance force, Im = the mass, r = its distance from the pivot, andw (omega) is the angular frequency, equal to 2p times the frequency in Hz..

Fig.1

From this, it is seen that the force on the pivot is proportional to its distance fromthe center of rotation and to the speed squared.A rotor containing a heavy spot is not exactly equivalent to the stone on a string. Inthe case of the stone, the center of gravity of the system is the center of the stoneitself,whereas the CG of a rotor with imbalance is outside the imbalance mass andis near the axis of rotation of the rotor.

Fig.2

If the structure holding the bearings in such a system is infinitely rigid, the center ofrotation is constrained from moving, and the centripetal force resulting from theimbalance mass can be found from the above formula. This force is borne by the

bearings. Now, consider a hypothetical machine where the bearings are not rigidlysupported, but are suspended on springs.


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Fig.3

Under these conditions the shaft centerline is not constrained, and the rotor willrotate around its center of gravity. The 1 x RPM force on the bearings will be verysmall because it is only required to accelerate the bearings to the above mentionedamplitude. The double amplitude of vibration of the bearings will be equal to twicethe distance between the CG and the centerline of the rotor. Moreover, theamplitude of bearing vibration is constant regardless of the rotor speed, provided

the speed is higher than the natural frequency of the spring-rotor system. I t is seenhere that the vibration amplitude has nothing to do with the above centripetal forceformula.At speeds well below the natural frequency, the system is said to be "springcontrolled", and the centripetal force formula holds. Speeds above the naturalfrequency are in the "mass-controlled" region where the amplitude is constant, andthe bearing forces are not so easily predictable, be dependent on the equivalentmass of the bearings and springs.

Fig.4 Static Imbalance


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3.Couple Imbalanc e

Fig.5 Couple Imbalance

With pure imbalance, either static or dynamic, the axial 1X and 2X vibration levelswill be low

4.Sever i ty o f Im balanceThe severity of imbalance depends on both the type and size of the machine as wellas the vibration level. To assess imbalance severity, average 1X levels for healthymachines of the same type should be used as a comparison. I f the second orderpeak is as large as the first order, you should suspect misalignment.

The following levels are guidelines for general use in diagnosing imbalance formachines running at 1800 or 3600 RPM. Very high-speed machines have lowertolerance levels.

1X Vibration Level,

VdB

Diagnosis Repair Priority

Less than 108 VdB

(0.141 ips)

Slight Imbalance No recommendation

108 VdB -- 114 VdB

(0.141 0.282 ips)

Moderate Imbalance Desirable

115 VdB -- 124 VdB

(0.316 0.891 ips)

Severe Imbalance Important

More then 125 VdB

(>1.00 ips)

Extreme Imbalance Mandatory

The measured vibration level at 1X depends on the stiffness of the machinemounting as well as the amount of imbalance, with spring-mounted machinesshowing more 1X than solidly mounted machines for the same degree ofimbalance. The overall size of the machine also affects the allowable 1X level asfollows:

1X Vibration Level,

VdB

Machine Type Repair Priority


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109 VdB (0.158 ips) Small Single-stage

Pump

Desirable

118 VdB (0.447 ips) Large Hydraulic Pump Desirable

116 VdB (0.355 ips) Medium Sized Fan Desirable

The tangential and radial 1X levels should be compared. The more nearly equalthey are, the more likely that imbalance is the cause. In any case, the direction inwhich the machine has the least stiffness will be the direction of the highest 1Xlevel.

5.Sources o f Im balanc eThe following machine problems are among the conditions that will create

imbalance: Uneven dirt accumulationon fan rotors

Lack of homogeneity in

cast parts, such as bubbles,

blow-holes, porous

sections

Rotor eccentricity

Roller deflection,

especially in paper

machines

Machining errorsUneven mass distribution

in electric motor rotor bars

or windings

Uneven erosion and

corrosion of pump

impellers

Missing balance weights

Bowed Shaft

6.Misal ignmentMisalignment is a condition where the centerlines of coupled shafts do notcoincide. If the misaligned shaft centerlines are parallel but not coincident, then themisalignment is said to be parallel misalignment. I f the misaligned shafts meet at apoint but are not parallel, then the misalignment is called angular misalignment.Almost all misalignment conditions of machines seen in practice are a combinationof these two basic types.

Parallel Misalignment


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Fig.6 Parallel Misalignment

If the machine speed can be varied, the vibration due to imbalance will vary as thesquare of the speed. I f the speed is doubled, the imbalance component will rise by afactor of four, while misalignment-induced vibration will not change in levelFollowing is a typical vibration spectrum from a misaligned machine.

Fig.7

7.Angular Misal ignm entAngular misalignment produces a bending moment on each shaft, and thisgenerates a strong vibration at 1X and some vibration at 2X in the axial direction atboth bearings, and of the opposite phase. There will also be fairly strong radialand/ or transverse 1X and 2X levels, but in phase.

Fig.8 Angular Misalignment

Misaligned couplings will usually produce fairly high axial 1X levels at the bearingson the other ends of the shafts as well!

8.General Misal ignm entMost cases of misalignment are a combination of the two above described types,and diagnosis is based on stronger 2X peaks than 1X peaks and the existence of 1X


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and 2X axial peaks. Take care that high axial 1X levels are not caused by imbalancein overhung rotors.

Misalignment produces a variety of symptoms on different types of machines, andthe average vibration signatures for healthy machines should be consulted todetermine allowable 1X and 2X levels.

9.Temperature Ef fect s on A l ignmentThe best alignment of any machine will always occur at only one operatingtemperature, and hopefully this will be its normal operating temperature. I t isimperative that the vibration measurements for misalignment diagnosis be madewith the machine at normal operating temperature.

10. Causes of Misal ignmen tMisalignment is typically caused by the following conditions:Inaccurate assembly of components, such as motors, pumps, etc.Relative position of components shifting after assemblyDistortion due to forces exerted by pipingDistortion of flexible supports due to torqueTemperature induced growth of machine structureCoupling face not perpendicular to the shaft axisSoft foot, where the machine shifts when hold down bolts are torqued.

11. Bent Shaf t

Fig.9 Bent Shaft12. J ourna l Bear ingsMost journal-bearing problems will generate spectral peaks at lower frequenciesthan 1X, and these are called sub-synchronous peaks. Sometimes harmonics ofthese sub-synchronous peaks are also created, indicating severe degradation of thebearing. Here are some things to look for in diagnosing journal bearings:

13. Oil WhirlOil Whirl is a condition in which a strong vibration occurs at between 0.38X and0.48X. I t never shows up at precisely 0.5X, but is always a little lower in frequency.


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It is caused by excessive clearance and light radial loading, which results in the oilfilm building up and forcing the journal centerline to migrate around in the bearingopposite the direction of rotation at less than one-half RPM. Oil whirl is a seriouscondition and needs to be corrected when found, for it can deteriorate fairlyquickly to the point where metal-to-metal contact occurs in the bearing.

14. Oil WhipThe solutions for oil whip and oil whirl are suitably small bearing clearances andadequate radial loading. When bringing a large turbine up to speed, it is importantto pass through the critical frequencies very quickly to prevent the buildup of oilwhip.

15. J ourna l Looseness

Fig.10 Journal or bearing housing looseness

One half, one third, and one fourth-order harmonics are sometimes called subharmonics.


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16. J ourna l Thrust Bear ingsWorn thrust bearings usually present strong axial components at the first fewharmonics of 1X. Worn Kingsbury bearings with 6 shoes will generate a peak at6X. This vibration peak is predominantly in the axial direction.

17. Rol l ing Elem ent Bear ingsMany years of experience have shown that in practice, less than 10 % of allbearings will run for their design lifetime. About 40 % of bearing failures areattributed to improper lubrication, and about 30% of failures are from impropermounting, i.e. misalignment or "cocking". About 20 % fail for other reasons, suchas overloading and manufacturing defects, etc.

Fig.11These are the formulas for calculating the frequencies of the bearing tones from thebearing geometry, but they are a little imprecise because the axial loading andslippage affects them in an unpredictable manner.

The number of rollers in most bearings is usually between 8 and 12, but in verylarge diameter bearings, such as the ones found in paper machines, the number ofrollers can be much higher.

18. Rol l ing Element Bear ing WearThe first stages of bearing defects will produce telltale non-synchronous vibrationfrequencies called "bearing tones" and their harmonics. Bearing tones at 0.006inches per second peak (81 VdB) or higher are considered significant. Sometimes anew bearing will produce bearing tones, possibly because of damage duringinstallation, shipping, or defective manufacture.


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Fig.12

If the bearing defect is very small in size, such as a crack in one of the races, thevibration signature will show harmonics of the bearing tone with little or nofundamental frequency present. I f the defect begins as a spall over a larger area ofthe race, the bearing tone fundamental will usually be higher in level than theharmonics. As the defect becomes worse, the overall level of the bearing tones willincrease, as will the overall broadband noise level.

19. SidebandsIf the defect is on the inner race of the bearing, the turning speed will amplitudemodulate the bearing tones, and this will cause sidebands around the bearing tones,spaced apart at 1X, to appear. The amplitude modulation comes from the fact thatthe defect on the innerrace moves in and out of the bearing load zone once perrevolution. While in the load zone, the defect produces vibration at the ball passfrequency, but when it is out of the load zone, very little vibration is produced atthis frequency. This accounts for the amplitude modulation of the bearing tone andthe consequent sidebands. Sidebands spaced at 1X around bearing tones are a suresign of advanced bearing wear. Sometimes, if a rotor is strongly out of balance, aninner-race bearing defect will not produce amplitude modulation or sidebands. This

is because the centrifugal force due to imbalance keeps the inner race loaded at thesame location on its periphery all the time.

Another example of sidebands in bearing spectra involves the Fundamental TrainFrequency (FTF). This is the rate at which the cage holding the rollers rotates inthe bearing. If one roller is spalled, cracked, or worse yet, in several pieces, it willmake a lot of noise when it is in the load zone of the bearing, but will be quietwhen not in the load zone. I t will move in and out of the load zone at the FTF ratebecause it migrates around the bearing with the cage. This causes amplitude

modulation of the bearing tones at the FTF rate, and the result is sidebands aroundthe bearing tones spaced apart by the FTF.


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Fig.13

The final stage of bearing wear is sometimes called the "thermal" stage, where thebearing becomes hot, breaking down the lubricant, leading to catastrophic failurewhich can include melting of the rolling elements and/ or the races.

The key to effective predictive maintenance of bearings is the trending of bearingtone levels over time from their onset. Sometimes a bearing condition will progressfrom a very small defect to complete failure in a relatively short time, so earlydetection requires sensitivity to very small vibration signature components. Theanalyst should be awarethat some types of machines will show bearing tones in theaverage spectra. Diagnosis is made on the basis of signif icant increases from theseaverage values. Any significant bearing tone should be carefully watched for signs

of worsening.

20. Misal igned (" c oc k ed") Rol l ing Elem ent Bear ings

Fig.14 Cocked Bearing


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21. Rol l ing Elem ent Bear ing LoosenessExcessive clearance in a rolling element bearing will produce harmonics of 1X,usually in the range from 2X to 8X. Extreme looseness will commonly produceone-half order components, i.e., components at multiples of 0.5X. Looseness in

other parts of the machine will also produce 1X harmonics and sometimes 0.5Xharmonics, so this is not a conclusive sign of bearing clearance problems.

22. Mechanic a l Looseness

Fig.15 Mechanical Looseness

23. Non-Rota t in g LoosenessLooseness between a machine and its foundation will increase the 1X vibrationcomponent in the direction if the least stiffness. This is usually the horizontaldirection, but it depends on the physical layout of the machine. Low-order 1Xharmonics are also commonly produced if the looseness is severe. It is often hardto tell imbalance from foundation looseness or flexibility, especially in verticalmachines. If 1X tangential is much greater than 1X radial, looseness is suspected. I f1X tangential is lower than or equal to 1X radial, then imbalance is suspected.Foundation flexibility or looseness can be caused by loose bolts, corrosion, or

cracking of mounting hardware.

24. PumpsThere are many types of pumps in common use, and their vibration signatures varyover a wide range. When monitoring pump vibration, it is important that theoperating conditions are uniform from one measurement to the next to assureconsistent signatures. Suction pressure, discharge pressure, and especially airinduction and cavitation will affect the vibration signature.

25. Cent r i fugal Pumps


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The following spectrum, containing broadband high-frequency noise, indicatescavitation in a centrifugal pump due to low inlet pressure.

Fig.16 Cavitation in Centrifugal Pumps

Cavitation produces this type of spectrum at all measurement points of the pumpand the housing

Fig.17

26. Gear Pum psGear pumps are commonly used for pumping lube oil, and they almost always havea strong vibration component at the tooth mesh frequency, which is the number ofteeth on the gear times the RPM. This component will be highly dependent on theoutput pressure of the pump. I f the tooth mesh frequency changes significantly,such as the sudden appearance of harmonics or sidebands in the vibrationspectrum, it could indicate a cracked or otherwise damaged tooth.

Fig.18 Typical Gear Pump Spectrum


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27. Sc rew PumpsThe screw type pump can generate a multitude of frequency components in thevibration spectrum. Thread wear or damage will usually produce strong harmonics

of the thread rate, which is the number of threads times the RPM.

28. FansMost fans are either axial flow propeller-type fans, or are centrifugal. Fans,especially when they are handling particle-laden air or gas, are prone to unevenbuildup of detritus on the blades. This causes imbalance, and should be correctedas soon as it is diagnosed. I f any of the blades become deformed, cracked, orbroken, the blade pass frequencyvibrationpeak will increase in level, and if thereare many blades, sometimes 1X sidebands will appear around the blade pass

frequency.

29. Ax ial Flow Fans

Fig.19

Fig.20

30. Cent r i fugal Fans


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Fig.21

A common problem in centrifugal fans is uneven supply air velocity distributionacross the inlet, and this causes increased vibration levels at the blade pass rate. I fthe fan is out of balance and is overhung, high 1X vibration will occur in axial as

well as both radial directions.Defective blades can also cause 1X sidebands around the blade pass frequency.

31. Coupl ingsCouplings exist in many types and configurations, and defects in them usually causesymptoms similar to misalignment. Frequently coupling problems will producestronger 1X vibration components than simple misalignment does. If the couplingis not true, i.e., has non-parallel flange faces, a vibration similar to angularmisalignment is produced.

Coupling imbalance is also a common problem, and results in high 1X and 2Xradial and tangential components.

Coupling wear can produce all the symptoms of misalignment and looseness.Three-jaw motor couplings that contain spacers of improper length will causestrong axial and radial components at 3 times shaft RPM.

32. Drive Bel t sBelt drives are relatively inexpensive types of power transmissions, but they areprone to many problems. There are many types of drive belts, and all are subject to

wear and damage. Belts should be frequently inspected for damage and should bekept at the proper tension and kept clean.

33. Mismat ched, Worn, or St re tc hed Bel tsMismatched, worn, or stretched belts, especially Vee belts, will generate vibration atthe fundamental belt pass frequency and harmonics of it. Usually the secondharmonic is dominant if there are two sheaves in the system. The fundamental beltfrequency FBF is given by the following formula. I t is always sub-synchronous,meaning it is lower in frequency than 1X.


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Where D = Sheave DiameterL = Belt LengthRPM = Turn speed of sheave D

34.

Ec c ent r ic Sheaves, Sheave RunoutEccentric sheaves will generate strong 1X radial components, especially in thedirection parallel to the belts. This condition is very common, and mimicsimbalance. This can be checked by removing the belts and measuring again. 1Xvibration of an eccentric sheave or a sheave with runout will usually also show upat the other sheave.

Fig.22 Eccentric Sheave

35. Sheave Misal ignm entSheave misalignment will generate strong axial 1X components and axial harmonicsof the fundamental belt frequency.

Fig.23

Belt Resonance, or Belt Slap


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Fig.24 Belt Resonance

36. GearboxesIf the gearbox has multiple shafts, each pair of gears will generate its own toothmesh components.

Different types of gear teeth will generate greatly different levels of vibration. Spurgears are inherently the most noisy, followed by Bevel gears, Hypoid gears, Helicalgears, Herringbone gears, and Worm gears in descending order of vibrationseverity.

37. Hunt ing Tooth Gear Set sThe so-called Hunting Tooth Gear Set is a gear set whose tooth counts arerelatively prime; in other words, they have no common factors. This is the bestconfiguration for gears, since any tooth on either gear will contact every tooth onthe other gear before encountering the same tooth. This spreads the wear evenlyover all the gear teeth, increasing the life of the gearbox.

The hunting tooth frequency of a pair of gears is the gear mesh frequency dividedby the least common multiple of the numbers of teeth on the two gears. The leastcommon multiple is often just the product of the numbers of teeth. In somegearboxes, the hunting tooth frequency will appear in the vibration spectrum, and if

so, it should be trended over time because rapid wear usually results under theseconditions.

38. Ghost ComponentsNew gear sets will sometime exhibit spectral components that are not related to thetooth counts of either gear. These components are sometimes called ghostfrequencies, and usually are the result of irregularities in the tooth spacing of one ofthe gears. The irregularities are the result of machining errors when the gear wasmade.


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Ghost components generally tend to disappear over time since there is nomechanical action that reinforces them.

39. Dam aged Gear Teet h

Fig.25

40. Ec c ent r ic Gears and Bent Shaf tsGeareccentricity will generate one sideband on each side of the tooth mesh spaced

at the gear RPM, rather than the multiple sidebands found with individual damagedteeth.

41. Planet ary GearsPlanetary gear systems are somewhat more complex than standard gear pairs due tothe fact that the planet gear centers rotate around the sun gear at a rate called thetrain frequency. The sun gear RPM, a planet gear RPM or the train frequency canmodulate the tooth mesh frequency. This can produce complex sets of sidebands inthe spectrum, and can be difficult to interpret.


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Below is a typical vibration spectrum from a 6-vaned centrifugal compressor.

Fig.27 Vane Pass Harmonics

43. Rec iprocat ing Mac hinesThe most common types of reciprocating machines are piston pumps andcompressors and internal combustion engines. In all these machines, the piston rate(usually 1X) is dominant, along with the firing rate for 4-cycle engines. Vibrationlevels as high as 125 VdB (1.0 inches per second peak) are not uncommon for

healthy machines such as these. The analyst must judge the machine condition bycomparison to previous levels rather than applying absolute reference levels.

Many reciprocating engines have turbo chargers, and they are diagnosed like otherrotating turbines and compressors. Camshaft gear problems are also common, andcan be seen by looking for the tooth mesh frequency. If the engine has a torsionalvibration damper on the shaft, it can fail, greatly increasing vibration at thefrequency of the first crankshaft tensional vibration mode. This frequency must beobtained from the engine maker.

Variable displacement piston pumps are much smoother than compressors, andlend themselves well to vibration analysis. If harmonics of the piston rate arepresent in significant levels, it usually indicates a piston drive linkage problem. Avery prominent tone at piston frequency fundamental may indicate a worn spot onthe wobble plate.

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7-1 Machine Diagnosis