Diagnostic Chart

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IRD Mechanalysis (UK) Ltd. VIBRATION FREQUENCIES AND THE LIKELY CAUSES Frequency In Other Possible Causes & Remarks Terms of RPM Most Likely Causes 1 x RPM Unbalance 1) Eccentric journals, gears or pulieys 2) Misalignment or bent shaft - if high axial vibration 3) Bad belts if RPM of belt 4) Resonance 5) Reciprocating forces 6) Electrical problems 2x RPM Mechanical 1) Misalignment if high axial vibration Looseness 2) Reciprocating forces 3) Resonance 4) Bad belts if 2 x RPM of belt 3 x RPM Misalignment Usually a combination of misalignment and excessive axial clearances (looseness). Less than Oil Whirl (Less than ¥2 RPM) 1) Bad drive belts 1 x RPM 2) Background vibration 3) Sub-harmonic resonance 4) "Beat" Vibration Synchronous Electrical Problems Common electrical problems include broken rotor bars, eccentric (A.C. Line rotor, unbalanced phases in poly-phase systems, unequal air gap. Frequency) 2 x Synch. Torque Pulses Rare as a problem unless resonance is excited Frequency Many Times RPM {Harmonically Related Freq). Bad Gears Aerodynamic Forces Hydraulic Forces Mechanical Looseness Reciprocating Forces Gear teeth times RPM if bad gear Number of fan blades times RPM Number of impeller vanes times RPM May occur at 2,3,4 and sometimes higher harmonics if severe looseness High Frequency (Not Harmonically Related) Bad Anti-Friction Bearings 1) Bearing vibration may be unsteady - amplitude and frequency 2) Cavitation, recirculation and flow turbulence cause random, high frequency vibration 3) Improper lubrication of journal bearings (Friction excited vibration) 4) Rubbing 2 IRDIVA1/95 FORM VIBFREQ

VIBRATION IDENTIFICATION CAUSE AMPLITUDE FREQUENCY PHASE REMARKS ! Unbalance Proportional to 1 x RPM Single Most common cause of Vibration unbalance. Largestin Reference radial direction Mark-stable repeatable Misalignment Large in axial 1 x RPM usual | Single Best found by appearance of large Couplings or direction 50% or 2& 3 xRPM Double or axial vibration. Use dial indicators or Bearings and more of radial sometimes Triple other method for positive diagnosis. If Bent Shaft vibration sleeve bearing machine and no coupling misalignment, balance the rotor. Bad Bearings Unsteady-use Very high Erratic - Bearing responsible most likely the Anti-Friction velocity, acceleration, | Several times Multiple one nearest point of largest high Type and Spike Energy RPM Marks frequency vibration. Spike Energy Measurements measurement recommended when analysing bearing failures “ccentric Usually not large 1 x RPM Single If on gears largest vibration in line with wournals Mark gear centers. If on motor or generator vibration disappears when power is turned off. If on pump or blower attempt to balance Bad Gears or Low-Use Very High Erratic - Velocity, Acceleration and Spike Gear Noise Velocity, Acceleration, | Gear Teeth Multiple Energy measurements recommended and Spike Energy times RPM Marks when analysing gear problems. Measurements Analyse higher orders and sideband frequencies Mechanical Sometimes 2 x RPM Two Usually accompanied by unbalance Looserness erratic Reference and/or misalignment Marks, Slightly Erratic Drive Erratic or 1,2,3%4 x RPM | One or Two | Strobe Light best tool to freeze faulty Belts Pulsing of Belts Depending Belt. on Frequency, i Usually Unsteady Electrical Disappears when 1 x RPM Single or If vibration amplitude drops off power is turned off ortor2 Rotating instantly when power is turned off X synchronous Double Mark | cause is electrical. Mechanical and frequency electrical problems will produce "beats". Aerodynamic or Can be large in 1 x RPMor Multiple Rare as a cause of trouble except in Hydraulic Forces the axial direction Number of Marks cases of resonance. blades on fan or impeller x RPM Reciprocating Higher in line 1, 2, & higher Multiple Inherent in reciprocating chhines, Forces with motion orders x RPM Marks can only be reduced by design changes or isolation © 1987 IRD Mechanalysis Inc.

TABLE 6.0 1LLUSTRATED. VIBRATION DIAGNOSTIC CHART PROBLEM TYPICAL PHASE SOURCE SPECTRUM RELATIONSHIP REMARKS X " MASS UNBALANCE RADIAL Force Un?iatmce wiibybe in-phase and|st | Amplitude due to un- balance will increase by the square of X speedincrease = 89X A. FORCE UNBALANCE higher vibration). 1X RPM ahways present and normally dommnates ( 5&‘\*‘\ q spectrum. Can be corrected by placement of only one balance weight S in one plane at Rotor center of gravity (CG). 8. COUPLE UNBALANCE 21X Couple Unbalance tends toward 180° out-of-phase on same shah 1X RADIAL = always present and normally dominates spectrum. Amplitude varies with square of increasing speed. May cause high axial vibrations as well as radial. Correction requires placement of balance weights in at 1= least 2 planes. Note that approx. 180" phase difference should exist between OB & IB horizontals as well as OB & 1B venticals. C. OVERHUNG ROTOR UNBALANCE Qverhung Rotor Unbalance causes high 1X RPM in both Axial and Radial directions. Axial readings tend to be in-phase whereas radial phase readings might be unsteady. Overhung rotors often have both force and couple unbalance, each of which will likely require correction. ECCENTRIC ROTOR 1X FAN 1X MOTOR 'BENT SHAFT Eccentricity occurs when center of rotation is offset from geometric centeriine of a sheave, gear, bearing, motor amature, etc. Largest vibration occurs at 1X RPM of eccentric componentin a direction thru centers of the two rotors. Comparative horizontal and vertical phase readings usualty differ either by 0° or by 180° (each of which indicate straight-ine motion). Attempts to balance eccentric rotor often resutt in reducing vibration in one direction, butincreasingit in the other radial direction (depending on S rl\t ?fﬂeccentricity). _WM T [ Bent Shaft problems cause high axial vibration with axial phase differences tending toward 180° on the same machine component. Dominant vibration normally at 1X if bent near shaft center, but at 2Xif bentnear the coupking. (Be careful to account for transducer orientation for each axial measurement if you reverse probe direction.) MISALIGNMENT A. ANGULAR MISALIGNMENT Anguiar Misalignmentis characterized by high axial vibration, 180° out- . of-phase across the coupling. Typically will have high axial vibration with both 1X anid 2X RPM. However, not unusual for either 1X, 2X or 3X to dominate. These symptoms may also indicate coupling problems as well o&w’( B. PARALLEL MISALIGNMENT x 2X h !‘l 3X RADIAL Offset Misalignment has simitar vibration symptoms to Angular, but shows high radial vibration which approaches 180° out-of-phase across coupling. 2X often larger than 1X, but its height relative to 1Xis often dictated by coupling type and construction. When either Anguiar or Radial Misaignment becomes severe, can generate either high amplitude peaks at much higher harmonics (4X-8X) or even a whole series of high frequency harmonics similar in appearance to mechanical looseness. construction will often greatly influence shape of spectrum when misalignment is severe. _C. MISALIGNED BEARING COCKED ON SHAFT 1X2X X AXIAL Cocked Bearing wil generate considerabie axial vibration. Will cause Twisting Motion with approximatedy 180° phase shift top to bottom and/or side to side as measured in axial direction of same bearing housing. Attempts to align coupling or balance the rotor will not alleviate problem. Bearing must be removed and comrectly installed. RESONANCE Y cloesnt Cree g ¥ mant 1} ]\10 d(red{o\«& w oane (X9 Resonance occurs when a Forcing Frequency coincides with a System Natural Frequency, and can cause dramatic ampitude ampiification, which can resultin premature, or even catastrophic falure. This may be a natural frequency of the rotor, but can often originate from support frame, founda- tion, gearbox or even drive beits. If a rotor is at or near resonance, it will be almost impossible to balance due to the great phase shift it expeniences (90° at resonance; nearly 180° when passes thru). Often requires changing natural frequency location. Natural Frequencies do notchange with a change in speed which helps facilitate their identification. MECHANICAL LOOSENESS — we Rave o idesty N ¢ c.d\c\ Qw(m;ni@ . Yrwncaked u Ao miwaed \AQOAO\ i | generatycausedbybosepiowbbekboks.cradshtheharneswcue or bearing pedestal. Type C is normally generated by improper fit between component parts which will cause many harmonics due to nonlinear response of loose parts to dynamic forces from rotor. Causes atruncation of time waveform. Type C is often caused by a bearing iner loose inits cap, excessive clearance n either a sleeve or roling element bearing, or aloose impetier on ashatt, Type C Phaseis often unstable and may vary widely from one measurement to next, particularly if rotor shifts position on shatt from one startup to next. Mechanical Looseness is often highly directional and may cause noticeably differentreadings if compare levels at 30° increments in radial direction all the way around one bearing housing. Also, note that looseness will often cause subharmonic multiples atexactly 1/2 or /3 X RPM (.5X, 1.5X, 2.5X, etc). Pg. 1 of 4 ©COPYRIGHT 1990 - TECHNICAL ASSOCIATES OF CHARLOTTE, INC. R-0782-3

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R AT g, TABLE 6.0 iLLU TRATLD vionanon DIAGNOSTIC CHART REMARKS Rotor Rub produces simiar spectra to Mechanical Looseness when rotating parts contact stationary components. Rub may be either partial or throughout the whole revolution. Usually generates a seres of frequencias, often exciting one or more resonances. Often excites integer fraction subharmonics of running speed (1/2, 1/3, 1/4 1/5,..1/n), depending onlocation of rotor natural frequencies. Rotor rub can excite many high frequencies (similar to wide-band noise when chalk is drug a blackbosrd). It ¢an be very sefious ahd of short duration if caused by shaft contacting bearing babbitt; but less serious when shaft rubbing & s8al, an agitator blade rubbing the wall of a tank, or a couping querd pressing against a shaft. PROBLEM TYPICAL SOURCE SPECTRUM ROTOR RUB RADIAL % 5T .5 B @ Ml l j /A TRUNCATED FLATTENED WAVEFORM SLEEVE BEARINGS o A. WEAR / CLEARANCE | = 5 & PROBLEMS A 1 }t Faoy A L 13 Latter stages of sleeve bearing wear are normally evidenced by presence of whote series of running speed harmonics (up to 10 or 20). Wiped slgeve bearings often will allow high vertical amplitudes compared 1o horizontal. Sleeve bearings with excessive clearance may aliow a minor unbalance andior misalignment 10 cause high vibration which would be much lower if bearing clearances were to spec. B. OIL WHIRL INSTABILITY (.42 -.48X RPM) X RADIAL Oil Whirl instability occurs at 42 - .48 X APM and is often quite severe. Considered excessive when amplitude exceeds 50% of bearing clearances. Oit Whirl is an oil film excited vibration where deviatons in normal operating conditions (attitude angle and eccentricity ratio) cause oil wedge to “'push” shaft around within bearing. Destabilizing force in direction of rotation resutts in a whirl (or precession). Whirl is inherentty unstable since it increases centritugal forces which increase whirt forces. Can cause oil to no longer support shaft, or can become unstable when whirl frequency coincides with a rotor natural frequen- ¢cy. Changes in oil viscosity, lube pressure and external preloads can affect oil whirl. ROTOR SPEED MASS oL whe | On wHIP WMNCLE A soqcmafvm C. OIL WHIP ‘ INSTABILITY 2 = e Hawlet! Packard “Applcations Note 24317, Fig. 4.3-2, Page 30 Qil Whip may occur if machine operated at or above 2X rotor critical frequency. When rotor brought up o twice critical speed, whirt will be very close to rotor critical and may cause excessive vibration that oll fim may nio longer be capable of supporting. Whirt speed will actually “lock onto” rotor critical and this peak wil not pass thru it even if machine broughtto higher and higher speeds. DOMINANT FAILURE SCENARIO ROLLING ELEMENT BEARINGS (4 Failure Phases) ZONE B ZONE C 20NE BEARING DEFECT BEARING COMPOMN. o FREQ REGION NATURAL FREQ. SPIKE ENERGY IOKE A REQION b3 8| smoen 4 1i; } BEARING DEFECT FREQUENCIES: g apm-_’;n.(‘ +%cose)xm W_%D_Q-%cosgxw 2 2 asp-zgé_f 1y (ccse;]xw m-%_q—%mse)xm g g g § L‘L— - W ale sl i | |, = Natural Frequencies of N7 instaked Bearing Components STAGE 2. Np = Number of Balls or Rolers snaes 3 g kLs”-fn‘}?un“"o"' > By = Bal/Aoker Diameter (in or mm) RANDOM HIGH 1 ese Py = Bearng Pitch Diameter (in or mm) FREQ. VIBRATION 72 8 = Contact Angio (Sogrees) [y W l { 1132000K AR ™ {4\ CPM 4 ROLLING ELEMENT BEARING FAILURE STAGES: STAGE 1: Earfiest indibations of bearing problems ar in witrasonic frequencies ranging from approximately 20,000-60,000 Hz (1,200,000-3,600,000 CPM). These are frequencies evaluated by Spike Energy (gSE), HFD{g) and Shock Pulse (dB). For exampie, spke energy may first appear at about .25 gSE in Stage 1 (actual value depending on measurement jocation and machine speed). STAGE 2: Skght bearing defects begin to “ring" bearing component natural frequenciesifn) which predominantly occur in 30K-120K CPM range. Sideband frequencies appear above and below natural frequency peak at end of Stage 2. Spike energy grows {for exampie, from .25 to .50 gSE). = § ] = SKEBANG l 1\ _FREQ STAGE 3: Bearing defect frequencies and harmonics appesr (see page entitled "Rotiing Element Bearing Defect Frequencies'). When wear progresses, more defect frequency harmonics appear and number of sidebands grow, both around these and around bearing natwal fre- quencies. Spike energy continues to increase (for exampie, from Sto over 1 gSE). Wear is now usually visible and may extend throughout periphery of beanng, particutarty when wed tormed sidebands accomp- any bearing defect frequency hammonics. Replace bearings now. STAGE 4: Towards the end, ampiitude of 1X RPM is even effected. it Grows, and nomally causes growth of many running speed harmonics. Discrete bearing defect and component natural frequencies actually begin to “disappear’ and are replaced by random, broad band high frequency “noise fioor”". In addition, ampktudes of both high frequency noise fioor and spike energy may in fact decrease; but jst prior to failure, spike energy will usually grow to excessive ampiitudes. HYDRAULIC AND BPF = # Blades X RPM AERODYNAMIC FORCES A. BLADE PASS & e VANE PASS = RANDOM B. FLOW TURBULENCE 3 n 8""§§§§%2’és | | C. CAVITATION A —1 IM—..\ 120K CPM Blade Pass Frequency (BPF) = No. of Blades (or Vanes) X RPM. This frequency is inherent in pumps, fans and compressars, and normally does not present a probiem. However, large amplitude BPF (and har- monics) can be generated in pump if gap between rotating vanes and stationary diffusers is not kept equal all way around. Also, BPF (or har- monic) somatimes can coincide with a system natural frequency caus- ing high vibration. High BPF can be generated if impeller wear ring seizes on shaft or if welds fastening diffusers fail. Also, high BPF can be caused by abrupt bends in pipe (or duct), obstructions which disturb tlow, o if pump or fan rotor is positioned eccentrically within housing. Flow Turbulence often occurs in blowers due to variations in pressure or velocity of the air passing thru the fan or connected ductwork. This flow disruption causes turbulence which will generate random, low fre- quency vibration, typically in the range of 50 to 2000 CPM. Cavitation normally generates random, higher frequency broadband energy which is sometimes superimposed with blade pass frequency harmonics. Normally indicates insufficient suction pressure {starvation). Cavitation can be quite destructive to pump internals if left uncorrected. It can particularly erode impeller vanes. When present, if often sounds as if “'gravel” is passing thru pump. Pg. 2 of 4 ©COPYRIGHT 1990 - TECHNICAL ASSOCIATES OF CHARLOTTE, INC. R-0792-3

TABLE 6.0 ILLUSTRATED VIBRATI1ON DIAGNOSTIC CHART PROBLEM TYPICAL. SOURCE SPECTRUM REMARKS G EA RS E RADIAL-SPUR {ALL GEAR SPECTRA] FREQ. Normat Spectrum shows 1X and 2X RPM, along with Gear Mesh Fre- quency (GMF). GMF commoniy will have runming speed sidebands GMF | .soeeano around it. All peaks are of low amplitude, and no natural frequencies . J j 1\ _FREQ of gears are excited. AXIAL-HELICAL GMF = GEAR MESH A. NORMAL SPECTRUM . 1X PINION GMF = 8 Teath X RPM Ltk Key indicator of Tooth Wear is excitation of Gear Natural Frequency, along with sidebands around it spaced at the running speed of the bad . B. TOOTH WEAR % R J gear. Gear Mesh Frequency (GMF) may or may not change in amplitude, In « GEAR NAT'L although high amplitude sidebands surrounding GMF usually occur FRE when wear is noticeabls. Sidebands may be a better wear indicator than GMF frequencies themselves. Gear Mesh Frequencies are often very sensitive 10 load. High GMF ampiitudes do not necessarily indicate a problem, particulariy if side- band trequencies remain low level and no gear natura! frequencies are L) | , excitqu Each Anatysis should be performed with system at maximum 1 AT operating load. C. TOOTH LOAD E Fairly high amplitude sidebands around GMF often suggest gear @ eccentricity, backiash, or non-parallel shafts which allow the rotation A of one gear to ‘modulate’ the running speed of the other. The gear D. GEAR ECCENTRICITY AND BACKLASH fn i with the problem is indicated by the spacing of the sideband 1 frequencies. improper backlash normally excites GMF and Gear Natural Frequency, both of which will be sidebanded at 1X RPM GMF amplitudes will often decrease with increasing load if backliash is the problem. - - - £ Ea Gear Misalignment almost always excites second order or higher GMF E. GEAR MISALIGNMENT g A harmonics which are sidebanded at running speed. Often will show only 1X GEAR small amplitude 1X GMF, but much higher levels at 2X or 3X GMF Important to set FpmAx high encugh to capture atleast 2 GMF harmonics if transducer system has the capability. P = -, - = GMF F. CRACKED / , . ) o e BROKEN TOOTH A Cracked or Broken Tooth will generate a high amplitude a1 1X RFM TIME of this gear, plus it will excite gear natural frequency (f,) sidebanded WAVEFORM at its running speed. It is best detected in Time Wavetorm which will e o T e | show a pronounced spike every time the problem tooth tries 10 mash A i A & D with teeth on the mating gear. Time between impacts (A) will a.fi- OF GEAR WITH BROXEN OR correspond to 1/speed of ?ear with the problem. Ampiitudes of M CRACKED TOOTH impact Spikes in Time Wavelorm often will be much higher than that 1X GEAR APM of 1X Gear RPM in FFT. G. HUNTING TOOTH Hunting Tooth Frequency (fiyT) is particularly effective for detecting faults PROBLEMS on both the gear and pinion that might have occurred during the 1 »_(GMFYN3) manufacturing process or due to mishandling. It can cause quite high T searXT piNiON? vibration, but since it occurs at low frequencies predominately less than 600 CPM, it is often missed. A gear set with this tooth repeat problem normally emits a ‘'growling’” sound from the drive. The maximum ef- fect occurs when the faulty pinion and gear testh both enter mesh at the same time (on somsa drives, this may occur only 1 of every 10 to 20 revolutions, depending on the fy1 formula). Note that TG AR and TRINION refer to number of teeth on gear and pinion, respectively. Na = number of unique assembly phases for a given tooth combina- tion which equals the product of prime factors common 10 the number of teeth on each gear. b 14T b— 2lyy Ewees 1X RPM - 2X APM ELECTRICAL PROBLEMS Stator problams generata high vibration at 2X line frequency (2F). Stator . RADIAL 2F F, = LINE FREQ. 6ccentricity produces uneven stationary air gap between rolor and stator A ngFO(PEgCLCAEP:JILTTCIgg’S t t (3800 CPM) which produces very directional vibration. Differential Air Gap should not exceed 5% for induction motors and 10% for synchronous motors. AND LOOSE {RON 2x Soft foot and warped bases can produce an eccentric stator. Loose iron is due to stator support weakness or looseness. Shorted stator lamina- 12K CPM tions can cause uneven, localized heating which can bow a motor shaft. Produces thermally-induced vibration which can significantly grow with operating time. P— 1X B. ECCENTRIC ROTOR RADIAL e Eccentric Rotors produce a rotating variable air gap between rotor and (Variable Air Gap) stator which induces pulsating vibration (normally betwesn 2F| and P Line /_D_z* g0l | & x Fo SIDEBANDS ND 2F closest running speed harmonic ). Oftgnreq&eg“zoom"spectmmto NS - sm:m“w ."fm o :- P AROUI L separate 2F|_andrunning speedharmoni. Eccentric rotors generate 2F_ : =t Lrp0HE ] ‘L 4 surrounded by Pole Pass frequency sidebands (Fp), as well as Fp side- £g - Swp Freq. = Ng - RPM 12K CPM bands around running speed. Fp appears itself at low frequency. (Fole Fp « Pole Pass Freq. = Fg X P - Pass Frequency = Slip Frequency X # Poles). Common values of Fp P = & Poles range from approximately 20 to 120 CPM (.30 - 2.0 Hz). . RADIAL £, SIDEBANDS AROUND 1X. 2X, 3x.. Broken or Cracked rotor bars or shorting rings, bad joints betweer rotor C. ROTOR PROBLEMS oo 2x ax bars and shorting rings, or shorted rotor laminations will produce high 1X running speed vibration with pole pass frequency sidebands (Fp). In ek addition, cracked rotor bars often wil generate Fp sidebands around the ROTOR . 2F_ SIDEBANDS AROUND RBPF third, fourth and fifth running speed harmonics. Loose rotor bars are AIR-GAP P ABPF indicated by 2X line frequency (2F( ) sidebands surrounding Rotor Bar Comk—x— 1X RBPF = ROTOR.BAR Pass Frequency (RBPF) and/or its harmonics (RBPF = Number of Bars MAGHETIC HELDL T : 2% PASS FREQ. X RPM). Often will cause high levels at 2X RBPF, with only a small = # BARS X RPM amplitude at 1X RBPF. 2F Phasing problems due 1o 10038 or broken CONNECIOrs Can causa ex- D. PHASING PROBLEM RADIAL t cessive vibration at 2X Line Frequency (2F{ ) which will have sidebands (Loose Connector) 113 F| SIDEBANDS around it at one-third Line Frequency (/3 Fi). Levels at 2F; can ex- LN S -4 AROUND 2F( cead 1.0 infsec if left uncorrected. This is particularly a problem if the 1 [ 4 11 defective connector is only sporadically making contact and periodicai- ly not. COIL PASS FREQ. Loose stator coils in synchronous motors will generate fairly high vibra- E. SYNCHRONOUS MOTORS 1X RPM tion at Coil Pass Frequency (CPF) which equals numbar ot stator coils {Loose Stator Coils) ox A SIDEBANDS X RPM (# Stator Coils = # Poles X # Coils/Pole). The Coil Pass Fre- 11 quency will be surrounded by 1X RPM sidebands. 6F, = DC motor problems can be detected by higher than normat amplitudes F. DC MOTOR PROBLEMS L = SCRFIRING FREQ. at SCR Firing Frequency (6F_ ) and harmonics. Thesa problems include X broken fieid windings, bad SCHs and loose connections. Other problems ij including loose or blown fuses and shorted control cards can cause high amplitude peaks at 1X thru 5X Line Frequency (3600 - 18000 CPM). Pg. 3 of 4 ©COPYRIGHT 1990 - TECHNICAL ASSOCIATES OF CHARLOTTE, INC. R-0792-3

TABLE 6.0 ILLUSTRATED VIBRAT!ON DIAGNOSTIC CHART PROBLEM TYPICAL SOURCE SPECTRUM REMARKS BELT DRIVE PROBLEMS T e GaT PO, < 1142 X P A X TN O A. WORN, LOOSE OR MISMATCHED BELTS HARMONICS s FREQUENCY ELT PADIAL IN LINE WITH BELTS TIMING BELT FREQ. » BELT FREQ. X § BELT TEETH = PULLEY RPM X # PULLEY TEET e Belt frequencies are below the RPM of either the motor or the driven e Machine. When mer are worn, loose or mismatched, they normally 6 cause 3 to 4 multiples of belt frequency. Often 2X beft frequency is the dominant peak. Amplitudes are normally unsteady, sometimes pulsing with either driver or driven RPM. On timing belt drives, wear r pulfey misalignment is indicated by high amplitudes at the Timing Beit Frequency. B. BELT / SHEAVE MISALIGNMENT OFF SETiE PIDGEON {| TOE ;E ANGLE AXIAL 1X DRIVER OR DRIVEN Misalignment of sheaves produces high vibaration at 1X RPM predom:- nantly in the axial direction, The ratio of ampktudes of dnver to dnven RPM depends on where the data s taken as well as onrelative mass and frame, stifiness. Often with sheave misalignment, the highest axial vibration on the motor will be at fan RPM. C. ECCENTRIC SHEAVES RADIAL 1X RPM ECCENTRIC SHEAVE Eccentric and/or unbalanced sheaves cause high vibration at 1X RPM of this sheave. The amplitude is normatty highest mn line with the beits, and should show up on both dniver and driven bearngs. it 18 sometimes possible o balance eccentric sheaves by attaching washers to taper Iock bolts. However, even if balanced. the eccentricity will stitinduce vibration and reversible fatigue stresses in the belt D. BELT RESONANCE RADIAL 1X RPM BELT RESONANCE | Belt Resonance can cause high ampiitudes if the belt natural frequency should happen to approach of concide with either the motor or driven RPM Belt natural frequency can be altered by changing etther the belt tension or length. Can be detected by tensioning and then releasing beit while measuring response on sheaves or beanngs. " BEAT VIBRATION Fa-F) PULSATING AMPLITUDES - WIDEBAND SPECTRUM A » BEAT FREQUENCY ZOOM SPECTRUM A Beat Frequency 1s the result of two closely spaced frequencies going into and out of synchronization with one another. The wideband spectrum normalkty will show one peak pulsating up and cown. When you zoominto this peak (lower spectrum, it actualty shows two closely spaced peaks The ditference in these two peaks (F2-F 1} is the beat frequency which itsetf appears in the wideband spectrum. The beat frequency isnotcom- monty seenhnormafreqn.;encyrangemeasuementssnceitisﬂ\efenﬂy low frequency, usually ranging from only approximately 5 to 100 CPM Maximurm vibration will result when the time waveform of one frequency (F41) comes into phase with other frequency {F2). Minimum vibration occurs when waveforms of these two frequencies line up 180° out of phase 6-

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