An Overview o Bearing Vibration Analysis t c a r t s b A
Dr. S. J. Lcy Engineering Manager, Schaefer (UK) Ltd
Vibration produced by rolling bearings can be complex and can result rom geometrical imperections during the manuacturing process, process, deects on the the rolling rolling surace suracess or geometrical errors in associated components. Noise and vibration is becoming more critical in all types o equipment since it is oten perceived to be synonymous with quality and oten used or predictive maintenance. In this article the dierent sources o bearing vibration are considered along with some o the characteristic deect requencies that may be present. Some examples o how vibration analysis can be used to detect deterioration in machine condition are also given.
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INTRODUCTION
R
olling contact bearings are used in almost every type o rotating machinery whose successul and reliable operation is very dependent on the type o bearing selected as well as the precision o all associated components, i.e. shat, housing, spacers, nuts etc. Bearing engineers generally use atigue as the normal ailure mode, on the as sumption sumption that the bearings are properly installed, operated and maintained. Today, because o improvements improvements in manuacturing technology and materials, it is generally the case that bear ing ati gue lie, w hich is rel ated to sub-
surace stresses, i s not the limiting actor and probably accounts or less than 3% o ailures in service. Unortunately Unortunately though, many bear ings ail prematurely in service because o contamination, poor lubrication, temperature extremes, poor tting/ts, unbalance and misalignment. All these actors lead to an increase in bearing vibrat ion and condition monitoring has been used or many years to detect degrading bearings beore they catastrophically ail (with the associated costs o downtime or signicant damage to other parts o the machine).
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Rolling element bearings are oten used in noise sensitive applications, e.g. household appliance electric motors which oten use small to medium size bearings. Bearing v ibration is thereore becoming becoming increasingly important rom both an environmental consideration consideration and because it is sy nonymous with quality. It is now generally accepted that quiet running is sy nonymous with the orm orm and nish o the rolling contact suraces. As a re sult, bea ring manuact m anuact urers ha ve developed developed vibration tests a s an eective method or measuring quality. A common approach is to mount the bearing on a quiet running spindle and measure the radial velocity at a point on the b earing’s ear ing’s outer ring and in three requency bands, viz. 50300, 300-1800 and 1800-10000 1800-10000 H z. The bear ing must meet RMS velocity limits in all three requency bands. Vibration Vibrat ion monitoring ha s now become a well accepted part o many planned maintenance regimes and relies on the well known characteristic vibration signatures which rolli ng bear ings ex hibit as t he rolling suraces degrade. However, in most situations bearing vibration cannot be meas ured direct ly and so the be aring ari ng vibration vibrat ion signat ure is modi ed by the machine structure, this situation being urther complicated by vibration rom other equipment on the machine, i.e. electric motors, motors, gears, belts, hydraulics, str uctural resonances etc. This oten makes the interpretation interpretation o v ibration data dicult other than by a trained specialist and can in some situations lead to a mis-diagnosis, resulting in unnecessary machine downtime and costs. In this paper the sources o beari ng vibration vibrat ion are dis cussed cuss ed along wit h the characteristic vibration requencies that are likely to be generated.
Imprving th Rlibility Cl Fird Pr Plnt using SaP – RewoP Intrc
SOURCES OF VIBRATION Rolling contact bearings represents a complex vibration system whose components – i.e. rolling elements, inner raceway, outer raceway and cage – interact to generate complex vibration signatures. Although rolling bearings are ma nuactured using high precision machine tools and under strict cleanliness and quality controls, like any other manuactured part they will have degrees o imperection and generate vibrat ion as the sur aces i nteract through a combination o rolling and sliding. Nowadays, although the amplitudes o surace imperections are in t he order o nanometres, signicant vibrations can still be produced in the ent ire audible requency range (20 Hz – 20 kHz). The level o the vibration will depend upon many actors, including the energy o the impact, the point at which the vibration is measured and the construction o the bear ing.
Variable compliance Under radial and misaligning loads bear ing vibr ation is an inherent eature o rolling bearings even i the bearing is geometrically perect and is not thereore indicative o poor quality. This type o vibrat ion is oten reerr ed to as variable compliance and occurs because the external load is supported by a discrete number o rolling elements whose position with respect to the line o action o the load continually changes with time ( see Figure 1 ). As the bearing rotates, indiv idual ball loads, hence elastic defections at the rolling element raceway contacts, change
Ra
to produce relative movement between the inner and outer rings. The movement takes the orm o a locus which under radial load is two dimensional a nd contained in a radial plane, whilst under misalig nment it is t hree-dimensional. The movement is also periodic with bas e requency equal to the rate at which the rolling elements pass through the load zone. Frequency ana lysis o the movement yields the base requency and a series o har monics. For a single row radial ball bearing with an inner ring speed o 1800 rev/min a typical ball pas s rate is 100 Hz and signicant ha rmonics to more than 500 Hz ca n be generated. Variable compliance v ibration is heavily dependent on the number o rolling elements supporting the externally applied load; the greater the number o loaded rolling elements, the less the v ibration. For radially loaded or misaligned bearings ‘running clearance’ determines the extent o the load region, and hence, in general, variable compliance increases with clearance. Running clearance should not be conused σraceway with radial internal clearance (RIC), the ormer normally being lower than the R IC due to intererence t o the rings and dierential thermal expan sion o the inner and outer rings duri ng operation. Variable compliance v ibration levels can be higher than those produced by roughnes s and wav iness o the rolling suraces. However, in applications where vibration is critical it can be reduced to a negligible level by using ball bear ings with the correct level o axial pre-load.
al Load
Geometrical imperections
Figur 1 Simple bearing model
Because o the very nature o the manuacturing processes used to produce bearing components geometrical imperections will alway s be present to varying degrees depending on the accuracy
class o the bearing. For axially loaded ball bear ings operat ing under moderate s peeds the orm and surace nish o the cr itical rolling suraces are generally the largest source o noise and vibration. Controlling component waviness and surace nish during the manuacturing process is thereore critical since it may not only have a signicant eect on vibration but also may aect bearing lie. It is convenient to consider geometrical imperections in term s o wavelength compared with the width o the rolling element-raceway contacts. Surace eatures o wavelength o the order o the contact width or less are termed roughness, whereas longer wavelength eatures are termed waviness (see Figure 2 ).
Width of Contact
ball
σ
h
Figur 2 Waviness and roughness o rolling suraces
SURFACE ROUGHNESS Surace roughness is a signicant source o vibration when its level is high compared with the lubricant lm thickness generated between the rolling elementraceway contacts (see Figure 2 ). Under this condition surace asperities can break through the lubricant lm and interact with the opposing surace, resulting in metalto-metal contact. The resulting vibration consists o a random sequence o small impulses which excite all the natural modes o the bearing and supporting structure. Surace roughness produces vibration predominantly at requencies above six ty times the rotational speed o the bearing. Thus the high requency part o the spectrum usually appears as a ser ies o resonances.
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100 80 m l i F 60 t n e c r e P 40
Region of lubricated related surface distress
Region of possible surface distress with severe sliding
Operating region for most industrial applcations
Region of increased life
20
0.4 0.6
1
2
4
6
10
Λ Figure 3 Percent flm versus Λ (unction o flm thickness and surace roughness)
A common parameter used to estimate the degree o asperity interaction is the lambda ratio (Λ ). This is the ratio o lubricant lm thickness to composite surace roughness and is given by the expression Λ = h (σЪ2 + σr2)0.5 where Λ = degree o asperit y interacti on h = the lubricant flm thickness σЪ = RMS roughness o the ball σr = RMS roughness o the raceway
I we assume that the surace nish o the raceway is tw ice that o rolling element, then or a typical lubricant lm thickness o 0.3µm surace nishes better than 0.06 µm are required to achieve a Λ value o three and a low incidence o asperity interaction. For a lubricant lm thickness o 0.1_m surace nishes better than 0.025 _m are required to achieve Λ=3. The eect o Λ on bear ing lie is shown in Figure 3. I Λ is less than unity it is unlikely that the bearing will attain its estimated design lie because o surace distress, which can lead to a rapid at igue ailure o the rolling suraces. In general, Λ ratios greater than t hree indicate complete surace separation. A transition rom ull EHL (elastohydrodynamic lubrication) to mi xed lubrication (partial EHL lm w ith some asperity contact) occurs in the Λ range between 1 a nd 3.
elements ollowing the surace contours. The relationship between surace geometry and vibration level is complex, being dependent upon the bearing and contact geometry as well as conditions o load and speed. Waviness can produce vibration at requencies up to approximately three hundred times rotational speed but is usually predominant at requencies below sixty t imes rotational speed. The upper limit is attributed to the nite area o the rolling element raceway contacts which average out the shorter wavelength eatures. In the direction o rolling, elastic deormation at the contact attenuates simple harmonic waveorms over the contact width (see Figure 4 ). The level o attenuation increases as wavelength decrea ses unti l, in the limit, or
Discrete deects Whereas surace roughness a nd waviness result di rectly rom the bear ing component manuacturing processes, discrete deects reer to damage o the rolling suraces due to assembly, contamination, operation, mounting, poor maintenance etc. These deects can be extremely small and dicult to detect and yet can have a signicant impact on vibration-critical equipment or can result in reduced bearing
Contact Width
Ball
Raceway Waviness
Waviness For longer wavelength surace eatures, peak curvatures a re low compared with that o the Hertzian contacts and rolling motion is continuous with the rolling
a wavelength equal to the contact width, waviness a mplitude is theoret ically z ero. The contact length also attenuates short wavelengt h surace eature s. Genera lly poor correlation can ex ist between parallel surace height proles taken at di erent points across the tracks and this averages measured waviness amplitudes to a low level. For typical bearing surace s poor correlation o parallel surace heights proles only exists at shorter wavelengths. Even with modern precision machining technology wavi ness cannot be eliminated completely and an element o waviness will alway s exi st albeit at relat ively low levels. As well as the bearing itsel, the quality o t he associated components can also aect beari ng vibration and any geometrical errors on the outside diameter o the shat or bore o the housing can be refected on the bearing raceways w ith the associated increase in v ibration. Thereore, careul attention is required to the orm and precision o all associated bearing components.
Figure 4
Attenuation due to contact width
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Attention due to Elastic Deformation
Figure 5(a) Signal rom a good bearing
lie. This type o deect can take a variety o orms, viz. indentations, scratches along and across the rolling suraces, pits, debris and particles in the lubricant. Bearing ma nuacturers have adopted simple vibration measurements on the nished product to detect such deects but these tend to be limited by the type and size o bearing. An ex ample o this type o measurement is shown in Figures 5(a) and 5(b) where, compared to a good b earing, the discrete damage on a beari ng outer ring raceway has produced a characteristically impulsive vibration which has a high peak/ RMS ratio. Where a large number o deects occ urs individual peaks a re not so clearly dened but the RMS vibration level is several ti mes greater than that normally associated with a bear ing in good condit ion.
Bearing characteristic requencies Although the undamental requencies generated by rolling bearings are ex pressed by relatively simple ormula s they cover a wide requency range and can interact to give very complex signals. This is oten urther complicated by the presence on the equipment o other sources o mechanical,
Figure 6
Figure 5(b) Signal rom a damaged bearing
structural or electro-mechanical vibration. For a stationary outer ring and rotating inner ring, the undamental requencies are derived rom the bearing geometry as ollows – c/o = r/2 [1 – d/D Cos α ] c/i = r/2 [1 + d/D Cos α ] b/o = Z c/o b/i = Z c/i b = D/2d r [1 – (d/D Cos α)2] where
r = inner ring rotational requenc y c/o = undamental train (cage) requency relative to outer ring c/i = undamental train requency relative to inner ring b/o = ball pass requenc y o outer ring b/i = ball pass requenc y o inner ring b = rolling element spin requenc y D = Pitch circle diameter d = Diameter o roller elements Z = Number o rolling elements α = Contact angle
The bearing equations assume that there is no sliding and that the rolling elements roll over the raceway suraces. However, in practice this is rarely the case and due
Axial vibration acceleration spectrum on end cap o a 250 kW electric motor
to a number o actors the rolling elements undergo a combination o rolling and sliding. As a consequence, the actua l characteristic deect requencies may dier slightly rom those predicted, but this is very dependent on the type o bearing, operating conditions and ts. Generally the bear ing characteristic requencies will not be integer multiples o the inner ring rotational requency which helps to distinguish them rom other sources o vibration. Since most vibration requencies are proportional to speed it is important when comparing vibration signatures that data is obtained at identical speeds. Speed changes will cause shits in the requency spectr um causing inaccuracies in both the amplitude and requency mea surement. Sometimes, in vari able speed equipment spectral orders may be used where all the requencies are normalized relative to the undamental rotational speed. This is generally called ‘order normalisation’, where the undamental requency o rotation is called the frst order. The bearing speed ratio (ball pass requency divided by the shat rotational requency) is a unction o the beari ng loads and clearances and can thereore give some indication o the bearing operating perormance. I the bearing speed ratio is below predicted values it may indicate insucient loading, excessive lubrication or insucient bearing radial internal clearance, which could result in higher operating temperatures and premature ailure. Likewise, a higher than predicted bear ing speed ratio may indicate exce ssive loading, excessive bearing radial internal clearance or insucient lubrication. A good exa mple o how the bear ing speed ratio can be used to identiy a potential problem is shown in Figure 6, which shows a vibration acceleration spectrum measured axia lly on the end cap o a 250 kW electric motor. In this ca se the Type 6217 radial ball bear ings were e xperiencing a high axial load as a result o the non-locating bearing ailing to slide in the housing (thermal
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Imperections on the sur ace o raceways and rolling elements, as a result o the manuacturing process, interact to produce other discrete requencies and sidebands (summarised in Table 1 ).
Figur 7 Photograph o Type 6217 inner ring showing running path oset rom centre o raceway
loading). For a nominal shat speed o 3000 rev/min the estimated outer ring ball pass requency, b/o, was 228.8 Hz giving a bearing speed ratio o 4.576. The actual outer ring ball pass requency was 233.5 Hz giving a ball speed ratio o 4.67, an increas e o 2%. A photograph o the inner ring is shown in Figure 7, showing the ball running path oset rom the centre o the raceway towards the shoulder. Eventually this motor ailed catastrophically and thermal loading (cross location) o the bearings was conrmed. A number o har monics and su m and dierence requencies are also ev ident in the spectrum. Ball pass requencies can be generated as a result o ela stic properties o the raceway materials due to vari able compliance or as the rolling elements pass over a deect on the raceways. The requency generated at the outer and inner ring raceway can be estimated roughly as 40% (0.4) and 60% (0.6) o the inner ring speed times the number o rolling elements respectively. Unortunately, beari ng vibration signals are rarely straightorward and are urther complicated by the interaction o the various component part s, but this ca n be oten used to advantage in order to detect a deterioration or damage to the rolling suraces.
which is a mplitude modulated at in ner ring rotational requency. In the requency domain this not only gives rise to a discrete peak at the carrier requency (ball pass requency) but also a pair o sidebands spaced either side o the carr ier Surace Deect requency by an amount equal to the Frquncy modulating requency (inner ring Comp onent I mper ec tion rotational requency) (see Figure 8 ). Innr Rcy eccntricity r Generally, as the level o amplitude modulation increases so wvinss nZ c/i ± r wil l the sidebands. As the deect Discrt Dct nZ c/i ±r increases in si ze more sidebands are generated and at some point the outr wvinss nZ c/ Rcy ball pa ss requenc y may no longer Discrt Dct nZ c/± r; be generated, but in stead a ser ies o nZ c/± c/ peaks w ill be generated spaced at the Z c/ Rlling Dimtr inner ring rotational requency. elmnt Vritin A discrete ault on the outer wvinss 2nb ± c/ raceway will generate a series o high energy pulses at a rate equal Discrt Dct 2nb ± c/ to the ball pass requency relative to the outer ring. Because the outer ring is Tbl 1 Frequencies related to surace stationary the amplitude o the pulse will imperections remain theoretically the same a nd hence will appear a s a single discrete peak within Ana lysis o bearing vibration signals the requency domain. is usual ly complex and the requencies An unba lanced rotor w ill produce generated will add and subtract and a re a rotating load, so as with a n inner ring almost always present in beari ng vibration deect, the resulting vibration signal can spectra. This is particularly tr ue where be amplitude modulate d at inner ri ng multiple deects are present. However, rotational requency. depending upon the dynamic range o Likewise the ball pass requency can the equipment, background noise levels also be modulated at the undamental and other sources o vibration bearing train requency. I a rolling element has requencies can be dicult to detect in a deect it will enter and leave the load the early stages o a deect. However, zone at the undamental train requency over the years a number o diagnostic causing amplitude modulation and result in sidebands around the ball pass requency. algorithms have been developed to detect bear ing aults by measuring the vibration Amplitude modulat ion at the unda mental signatures on the bearing housing. Usually, train requency can also occur i the cage is located radially on the inner or outer ring. these methods take advantage o both the characteristic requencies and the ‘ringing Although deects on the inner and requencies’ (i.e. natural requencies) o the outer raceways tend to behave in a similar manner, or a given size deect the amplitude bear ing (see later). o the spectrum o an inner raceway deect Raceway deect is generally much less. The reasons or this might be that a deect on the inner ring A discrete deect on the inner ra ceway will generate a ser ies o high energy puls es raceway only comes into the load zone once per revolution and the signal must travel at a rate equal to the ball pas s requency through more structural interaces beore relative to the inner raceway. Because the inner ring is rotating, the deect will enter reaching the transducer location, i.e. rolling element, across an oil lm, through the outer and leave the load zone causing a variation ring and through the bearing housing, to in the rolling element-raceway contact orce, hence defections. While in the load the transducer position. The more dicult zone the amplitudes o the pulses w ill transmission path or an inner raceway ault probably explains why a ault on the outer be highest but then reduce as the deec t leaves the load zone, resulting in a signal raceway tends to be easier to detect.
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Figur 8
and this limits the eectiveness o the envelope spectrum (see later). In the case o cage ailure the signature is likely to have random bursts o vibration as the balls slide and the cage starts to wear or deorm and a wide band o requencies is likely to occur. As a cage starts to deteriorate, or example rom inadequate lubrication, wear can start to occur on the sliding suraces, i.e. in the cage pocket or in the case o a ring guided cage on the cage guiding surace. This may gives ris e to a less stable rotation o the cage or a greater excursion o the rolling elements, resulting in increased sideband activity around the other bearing undamental requencies, e.g. the ball spin requency. Excessive clearance can cause vibration at the undamental train requency (FTF) as the rolling elements accelerate and decelerate through the load zone, which can result in large impact orces between the rolling elements and cage pockets. Also, outer race deects and roller deects can be modulated with the FTF undamental requency.
Amplit ude mod ulation (AM) (a) Ampl itude m odulated time sig nal
Amplitude
Other sources o vibration
Ac /2
f c - f m Am /4
f c + f m
f c Figur 8
frequency
Amplit ude mod ulation (AM) (b) Spec trum o ampli tude mo dulated signal
Rolling element deect Deects on the rolling elements can generate a requency at twice ba ll spin requency, and als o harmonics and t he undamental train requency. Twice the rolling element spin requency can be generated when the deect strikes both raceways, but sometimes the requency may not be this high because the bal l is not always in the load zone when the deect strikes and energy is lost as the signa l passes th rough other structural interaces as it strikes the inner raceway. Also, when a deect on a ball is orientated in the axia l direction it will not always contact the inner and outer raceway and thereore may be
dicult to detect. When more than one rolling element is deective, sums o t he ball spin requency can be generated. I these deects are large enough then vibration at undamental train requency can be generated.
Cage deect As already show n, the cage tend s to rotate at typically 0.4 times inner ri ng speed, generally has a low mass and thereore, unless there is deect rom the manuacturing process, is generally not visible. Unlike raceway deects, cage ailures do not usually excite specic ringing requencies
Contamination is a very common source o bearing deterioration and premature ailure and is due to the ingress o oreign particles, either as a result o poor handling or during operation. By its very nature the magnitude o the v ibration caused by contamination will vary and in the early stages may be dicult to detect, but this depends very much on the type and nature o the contaminants. Contamination can cause wear and damage to the rolling contact suraces and generate vibration across a broad requency range. In the early stages the crest actor o the time signal w ill increase, but it is unlikely that this will be detected in the presence o other sources o vibration. With grease lubr icated bearings, vibration may be initially high as the bear ing ‘works’ and d istr ibutes the gre ase. The vibration will generally be irregular but will di sappear with r unning time and generally, or most applications, doesn’t present a problem. For noise-critical applications special low-noise-producing greases a re oten used.
VIBRATION MEASUREMENT Vibration measurement can be genera lly characterised as alling into one o three categories, viz. detection, diagnosis and
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prognosis. Detection generally use s the most basic orm o vibration measurement, where the overa ll vibration level is measured on a broadband basis in a range, say, o 10 to 1000 Hz, or 10 to 10000 Hz. In machines where there is little vibration other than rom the bearings, the spikiness o the vibration signal indicated by the Crest Factor (the ratio Peak/ RMS) may imply incipient deects, whereas the high energy level given by the RMS level may indicate severe deects. Generally, other than to the experienced operator this type o measurement gives limited inormation, but can be useul when used or trending, where an increasing vibration level is an indicator o a deteriorating machine condition. Trend analysis involves plotting the v ibration level as a unction o time and using this to predict when the machine must be taken out o service or repair. Another way o using the measurement is to compare the levels with published vibration criteria or dierent types o equipment. Although broadb and vibration measurements may provide a good starting point or ault detection it has limited diagnostic capability, and although a ault may be identied it may not give a reliable indication o where the ault is, i.e. bearing deterioration or damage, unbalance, misalignment etc. Where an improved diagnostic capability is required requency analysis is normally employed, which usually gives a much earlier indication o the development o a ault and, secondly, the source o the ault. Having detected and diagnosed a ault the prognosis – i.e. what the remaining useul lie and possible ailure mode o the machine or equipment are likely to be – is much more di cult and oten rel ies on the continued monitoring o the ault to determine a suitable time when the equipment can be taken out o service, or relies on known experience with similar problems. Generally, rolling bearings produce very little vibration when they are ault ree and have distinctive characteristic requencies when aults develop. A ault that begins as a single deect, e.g. a spall on the raceway, is normally dominated by impulsive events at the raceway pass requency resulting in a narrow band requency spectrum. As the damage worsens there is likely to be an increase in the characteristic deect
requencies and sidebands ollowed by a drop in these amplitudes and an increase in the broadband noise with considerable vibration at shat rotational requency. Where machine speeds are very low, the bearings generate low energy signals, which again may be dicult to detect. Also, bearings located within a gearbox can be dicult to monitor because o the high energy at the gear meshing requencies, which can mask the bearing deect requencies.
Overall vibration level This is the simplest way o measuring vibrat ion and usua lly consi sts o measuring the Root Mean Square (RM S) vibration o the bearing housing, or o some other point on the machine, with t he transducer located as close to the bearing as possible. This technique involves measuri ng the vibration over a wide requency range, e.g. 10-1000 Hz or 10-10000 H z. The measurements can be trended over ti me and compared w ith known levels o vibration, or pre-alarm and alarm levels ca n be set to indicate a change in the machine condition. Alternatively, measurements can be compared with general standards. Although this method represents a quick and low cost method o vibrat ion monitoring, it is le ss sensitive to incipient deects, i.e. it detects deects in the advanced condition and has a limited diagnostic capability. Also, it is easily infuenced by other sources o vibration, e.g. unbalance, misalignment, looseness, electromagnetic vibration etc. In some situations, the Crest Factor (the ratio Peak/ RMS) o the vibration is capable o giving an earlier warning o bear ing deects. The development o a local ault produces short bursts o high energy which increase the peak level o the vibrat ion signal, but have little i nfuence on the overall RMS level. A s the ault progresses, more peaks w ill be generated until nally the Crest Factor will reduce but the RMS vibration will increase. The main disadvantage o this method is that in the early stages o a bear ing deect the vibration is normally low compared with other sources o vibration present and is thereore easily infuenced, so any changes in bearing condition may be dicult to detect.
Frequency spectrum Frequency analysis plays an important part in the detection and diagnosis o machine aults. In the time domain the
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individual contributions, e.g. unbalance, to the overall machine vibration are dicult to identiy. In the requency domain they become much easier to identiy a nd can thereore be much more easily related to individual sources o vibration. As we have already discussed, a ault developing in a bear ing wi ll show up as incr easi ng vibrat ion at requencies related to the bearing characteristic requencies, making detection possible at a much earlier stage than with overall v ibration.
Envelope spectrum When a bearing starts to deteriorate the resulting time signal oten exhibits characteristic eatures which can be used to detect a ault. Also, bearing condition can rapidly progress rom a very small deect to complete ailure in a relatively short period o time, so early detection requires sensitivity to very small changes in the vibration signature. As we ha ve already discussed, the vibration signal rom the early stage o a deective bearing may be masked by machine noise mak ing it dicult to detect the ault by spectrum analysis alone. The main advantage o envelope analysis is its ability to extract the periodic impacts rom the modulated random noise o a deteriorating rolling bearing. This is even possible when the signal rom the rolling bearing is relatively low in energy and ‘buried’ within other vibration rom the machine. Like any other structure with mass and stiness the beari ng inner and outer rings have their own natural requencies which are oten in the kilohertz range. However, it is more likely that the natural requency o the outer ring will be detected due to the small intererence or clearance t in the housing. I we consider a ault on the outer ring: as the rolling element hits the ault the natural requency o the ring will be excited and will result in a high requency burst o energ y which decay s and then i s excited again as the next rolling element hits the deect. In other words, the resulting time signal will contain a high requency component amplitude-modulated at the ball pass requency o the outer ring. In practice, this vibration will be very small a nd almost impossible to detect in a raw spectrum, so a method to enhance the signal is required. By removing the low requency components through a suitable high pass
peaks spaced at cage speed, 2.93 Hz, which again is consistent with deterioration in the condition o the rollers. The 374.4 Hz component is related to the gear mesh requency, w ith sidebands at rotational speed, 6.56 Hz. As prev iously mentioned, bea ring deects normally produce a signal which is amplitude modulated, so by demodulating the signal and a nalysing the envelope provides a useul technique or early ault detection. Figure 10 shows the envelope spectrum, where discrete peaks are present at 62.5 Hz, and its harmonics which correspond with the roller deect requency, clearly showing how demodulation can be used, in some circumstances, to provide a convenient and early detection o deterioration in rolling bea rings.
Cage damage
Figur 9 Spectra obtained rom the housing o a taper roller bearing
lter, rectiying and then using a low pass lter the envelope o the signal is let, the requency o which corresponds to the repetition rate o the deect. This technique is oten used to detect early damage in rolling element bearings and is also oten reerred to as the High Frequency Resonance Technique (HFRT) or Envelope Spectrum.
roller, plus a number o harmonics, i.e. 186.5 (x 3), 497 (× 8), 560 (× 9), 748 (× 12), 873 (× 14) and 936 Hz (× 15). This would suggest some deterioration in the condition o the roller(s), which was conrmed upon examination o the bear ing. The spect rum al so shows di screte
The vibration spectrum shown in Figure 11 was measured on the spindle housing o an internal grinding machine which was g rinding the raceways o bearing outer rings. Although the machine wa s producing work to the required qu ality the routine vibration measurement immediately rais ed some concerns regarding the condition o the spindle. The spindle was rotating at 19,200 rev/min (320 Hz) and the most unusual aspect o the spectrum is the presence o a large number o discrete peaks spaced at 140 Hz, which related to the undamental train requency o the angular contact ball bear ings which had a plastic cage and were lubricated with oil mist. Upon examination o the bearing the
Examples o vibration spectra Roller deterioration Figure 9 is an ex ample o spectra obtained rom a taper roller bearing with a 432 mm diameter bore rotating at 394 rev/min. The shat was gear driven with a drive shat speed o 936 rev/min (2.375 reduction) giving a theoretical gear mesh requency o 374.4 Hz. Vibration at shat speed 6.56 Hz is clearly ev ident along with its harmonics. Also ev ident in the spectra is vibration at 62.4 Hz, which corresponds with t wice the rotational requency o the
Figur 10 Envelope spectrum rom the housing o a taper roller bearing
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cage outer diameter showed clear signs o damage with some ragments o plastic material which had broken away, but was still attached to the outer diameter. As a result, the spectrum had sum and dierence requencies related to the shat (r) and cage (c), e.g. 1740 Hz (5r+c). As already discussed, the deterioration o rolling element bearings will not necessa rily show at t he bear ing characteristic requencies, but the v ibration signals are complex and produce sum and dierence requencies which are almost always present in the spectra. Figure 11
Vibration acceleration measured on the spindle housing o an internal grinding machine.
Raceway damage High axial load
An ex ample o a vibrat ion spectrum measured axial ly on the drive side end cap o a 250 kW electric motor is shown in Figure 12. The rotational speed was approximately 3000 rev/min (50 Hz) and the rotor was supported by two type-6217- C4 (85 mm bore) radial ball beari ngs, gre ase lubric ated. The vibration spectrum shows dominant peaks between 1 kHz a nd 1.5 kHz, which can be related to the outer raceway ball pass requency. The calculated outer raceway ball pa ss requency, b/o, is 229 Hz and the requency o 1142 Hz relates to 5 b/o with Figure 12 Vibration acceleration measured axially on the DE o a 250 kW electric motor. a number o sidebands at rotational requency, r. When the bea rings were removed rom the motor and exa mined the ball running path was oset rom the centre o the raceways towards the shoulders o the both the inner and outer rings, indicative o high axia l loads. The cause o the ailure was thermal pre-loading as a result o the non-locating bearing not sliding in the housing to compensate or ax ial thermal expansion o the shat; this is oten reerred to as ‘cross location’. The non-drive end bearing had severe damage to the raceways and the rolling elements which wa s Figure 13 Vibration acceleration measured radially on the housing o a Type 23036 spherical roller bearing
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Also a number o the rollers h ad black corrosion stains, which was consistent with the vibration at cage rotational requency, c=4 Hz, in the envelope spectrum ( see Figure 15 ). The modulation o the time signal at cage rotational requency can be clearly s een in the time signal, Figure 16.
Eect o bearing vibration on component quality
Figur 14 Type 23036 spherical roller bearing outer ring raceway showing black corrosion stains
Even low levels o vibration can ha ve a signica nt impact on critical equipment, such as machine tools that are required to produce components whose surace nish and orm are critical. A good exa mple o this is during the manuacture o bearing inner and outer rings. One o the most
Figur 15 Envelope spectrum o the Type 23036 spherical roller bearing
consistent with the highly modulated signal and high amplitude o vibration at 5 b/o. The overall RM S vibration level o the motor increased rom ty pically 0.22g to 1.64g. Another example o a vibration acceleration spectrum obtained rom the housing o a Type 23036 (180 mm bore) spherical roller bearing, located on the main drive shat o an impact crusher, is shown in Figure 13. The spectrum shows a number o harmonics o the outer raceway ball pass requency, 101 Hz, with a dominant peak at 404 Hz (4 b/o ) with sidebands at sh at rotational requency, 9 Hz. When the bea ring was removed rom the machine and examined one part o the outer raceway had black corrosion stain (see Figure 14).
critical operations is grinding o the bearing raceways which have to meet very tight tolerances o roundness and surace nish, and any increase in machine vibration can result in a severe deterioration in workpiece quality. Figure 17, which shows the v ibration acceleration spectrum, 0 -500 Hz, measured on the spindle housing o an extern al shoe centreless grinding machine during the grinding o an inner r ing raceway, where the ty pical values or out-o-roundness and surace roughness were >4 µm a nd 0.3 µmRa respect ively. The most distinct ive eature on the nished raceway was the presence o 21 lobes which, when multiplied by the workpiece rot ational speed (370 rev/ min or 6.2 Hz), corresponded to a requency o 129.5 Hz. This was very close to the 126 Hz component in the spectrum which was associated with the ball pass requency relative to outer raceway o a ball bear ing in the drive head motor. Also present are harmonics at 256 and 380 Hz. The dis crete peaks at 38, 116 and 190 Hz correspond to the spindle rotational speed and its harmonics. Figure 18 shows that ater replacing the motor bearings the v ibration at 126 Hz reduced rom 0.012g to 0.00032g and the associated har monics were no longer dominant. This resulted in a dramatic improvement in workpiece out-o-roundness o <0.4 µm and the surace nish improved to 0.19 µmRa. This demonstrates that with some critical equipment such as machine tools it is possible to assess directly the condition o the machine by mea suring the resultant workpiece quality [2, 3].
Figur 16 Accelera tion tim e signal o the Type 23036 sphe rical ro ller b earing
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Figure 17 Vibration spectrum and roundness beore replacing wheel head drive motor bearings (a) Vibration spectrum on spindle housing (b) Roundness o raceway
Figur 18 Vibration spectrum and roundness ater replacing wheel head drive motor bearings (a) Vibration spectrum measured on spindle housing (b) Roundness o raceway
CONCLUSIONS The various sources o bearing v ibration have been discussed, al so how each such source can generate characteristic vibration requencies which can combine to give complex vibration spectra, which at times may be di cult to interpret other than by the experienced vibration analyst. However, with roll ing bearings, characteristic vibrat ion signatures are oten generated, usually in the orm o modulation o the undamental bearing requencies. This
can be used to advantage, and v ibration conditioning monitoring sotware is oten designed to identiy these characteristic eatures and provide early detection o an impending problem. This usually takes the orm o signal de-modulation and establishment o the envelope spectrum, where the ea rly indications o sideband activity, and hence bea ring deterioration, can be more eas ily detected. As long as there are natura l requencies o the bearing and its nearby structures – which occur in the case o a local ized deect on the outer raceway, or on the inner raceway, or on a rolling element – the envelope spectrum works well. However, cage a ilures do not usually excite specic natural requencies. The ocus o demodulation is on the ‘ringing’ requency (the carrier requency) and the rate it is being excited (the modulating requency). Simple broad band vibration measurements also have their place, but oer a very limited diag nostic capability, and will generally not give an early warning o incipient damage or deterioration.
REFERENCES 1.
Harris T A, Rolling Bearing Analysis (4th Ed), Wiley, New York, 2001
2.
Lacey S J, Vibration monitoring o the internal centreless grinding process, Part 1: mathematical models. Proc Instn Mech Engrs. Vol 24. 1990,
3.
Lacey S J, Vibration monitoring o the internal centreless grinding process, Part 2: expe rimental results. Proc Instn Mech Engrs. Vol 24. 1990
4.
Wardle F P and Lacey S J, Vibration Research in RHP. Acoustics Bulletin. stv.lcy@schfr.cm
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