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Basics of Vibration
Vibration theory & analysis
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What is Vibration?
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Vibration Terms
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Time Waveform Analysis
complex time waveform
individual vibration signalscombine to form a complextime waveform showing overallvibration
f r e q u e
n c y
l o w f r e
q .
h i g h
f r e q .
t i m e overall vibration
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Scale Factors When comparing overall vibration signals, it is
imperative that both signals be measured on thesame frequency range and with the samescale factors .
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Measurements & Units
Displacement (Distance )mils or micrometer, mm
Velocity (Speed - Rate of change of displcmt)in/sec or mm/secAcceleration (Rate of change of velocity)
Gs or in/sec 2 or mm/sec 2
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10 100 1,000 10,000
Frequency (Hz)
10
1.0
0.1
1
0.01
100
Displacement (microns)Acceleration(g's - 9,81m/sec2 )
Velocity (mm/sec)
Common MachineryOperating Range
Amplitude(microns,
mm/sec, gs
Sensor Relationships
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Multi-Parameter MonitoringSame Data in Velocity and Acceleration
VelocitySpectrum
AccelerationSpectrum
On the same bearing cap,low freq. events (imbalance,misalignment, etc.) showbest in the velocityspectrum; while high freq.events (bearing faults ,gearmesh) show best in theacceleration spectrum
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Accelerometers
Rugged Devices Operate in Wide Frequency
Range (Near 0 to above 40 kHz)
Good High Frequency Response
Some Models Suitable For High Temperature
Require Additional Electronics(may be built into the sensor housing)
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What is vibration?Complex signal?
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FFT Signal Processing
T i m e
A m p l i t u d e
T i m e
A m p l i t u d e
F r e q u
e n c y A m p l i t u d e
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Narrow Bands with trend
T re nd of Balance
Alarm
A m p l i t u d e
S u b -H armonic 1X 2X B earing B earing G e ars B earing
1x 2x
.3in/sec
.1in/sec T ime(Days) T ime(Days)
T re nd o f Bearings
10x
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Alarm Types Narrow Bands A2 - 8.2.4. BPFI Pomp PNV
P1/K10 -PNV POMP NIET-KOPP VERTIKAA L
Label: BPFI with 1xrPM modulatio ns.
Route Spectru m
30-jan-96 15:14:51
OVERALL= 13.52 V-DGRMS = 13.46LOAD = 100.0RPM = 2987. (49.78 Hz)
0 500 1000 1500 2000 2500
0
2
4
6
8
10
12
14
Frequency in Hz
R M S V e
l o c
i t y
i n m m
/ S e c
Fault Limit
Freq:Ordr:Spec:
475.009.542
.06356
I m b a
l a n c e
M i s a
l i g n m e n
t
L o o s e n e s s
B e a r i n g
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Overall Vibration The total vibration energy
measured within a specificfrequency range.
includes a combination of allvibration signals withinmeasured frequency range
does not include vibrationsignals outside measuredfrequency range
produces a numerical value
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Alarm Types Overall Alarm
Look to the global vibration level
A2 - 8.2.4. BPFI Pom p PNVP1/K10 -PNV POMP NIET-KOPP VERTIKAAL
Label: BPFI with 1xrPM modu lation s.
Route Spectrum30-jan-96 15:14:51
OVERALL= 13.52 V-DGRMS = 13.46LOAD = 100.0RPM = 2987. (49.78 Hz)
0 500 1000 1500 2000 2500
0
3
6
9
12
Frequency in Hz
R M S V e
l o c
i t y
i n m m
/ S e c
Fault Li mit
Freq:Ordr:Spec:
1321.926.55.119
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Analyse of data: Spectra,Waveform and Trends
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Vibration -Imbalance -Misalignment -Looseness -Bearing problems -Belt problems
-Gear problems -Lubrification -Electrical problems -Resonance -Sleeve Bearing problems -Other
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Vibration analysis
"Of all the parameters that can be measured nonintrusively in industry today, the one containing themost information is the vibration signature." ArtCrawford
Vibration Analysis is the foundation of a predictivemaintenance program
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SIGNATURE ANALYSISSIGNATURE ANALYSIS
Which frequencies exist and what are therelationships to the fundamental excitingfrequencies.
What are the amplitudes of each peak How do the peaks relate to each other If there are significant peaks, what are their
source
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Vibration analysis
Unbalance
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COUPLE UNBALANCECOUPLE UNBALANCE
180 0 out of phase on the same shaft 1X RPM always present and normally dominates
Amplitude varies with square of increasing speed Can cause high axial as well as radial amplitudes Balancing requires Correction in two planes at 180 o
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OVERHUNG ROTORUNBALANCE
OVERHUNG ROTORUNBALANCE
1X RPM present in radial and axial directions
Axial readings tend to be in-phase but radial readingsmight be unsteady Overhung rotors often have both force and couple
unbalance each of which may require correction
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Diagnosing UnbalanceDiagnosing Unbalance Vibration frequency equals rotor
speed.
Vibration predominantly RADIALin direction.
Stable vibration phasemeasurement.
Vibration increases as square of speed.
Vibration phase shifts in directproportion to measurementdirection.
90 0
900
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Vibration analysis
Misalignment/Bent shaft
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ANGULARMISALIGNMENT
ANGULARMISALIGNMENT
Characterized by high axial vibration 180 0 phase change across the coupling Typically high 1 and 2 times axial vibration Not unusual for 1, 2 or 3X RPM to dominate Symptoms could indicate coupling problems
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PARALLELMISALIGNMENT
PARALLELMISALIGNMENT
High radial vibration 180 0 out of phase Severe conditions give higher harmonics
2X RPM often larger than 1X RPM Similar symptoms to angular misalignment Coupling design can influence spectrum
shape and amplitude
RadialRadial
1x1x 2x2x
4x4x
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MISALIGNED BEARINGMISALIGNED BEARING
Vibration symptoms similar to angularmisalignment
Attempts to realign coupling or balance the rotorwill not alleviate the problem. Will cause a twisting motion with approximately
180 0 phase shift side to side or top to bottom
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BENT SHAFTBENT SHAFT
Bent shaft problems cause high axial vibration 1X RPM dominant if bend is near shaft center 2X RPM dominant if bend is near shaft ends Phase difference in the axial direction will tend
towards 180 0 difference
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OTHER SOURCES OF HIGHAXIAL VIBRATION
OTHER SOURCES OF HIGHAXIAL VIBRATION
a. Bent Shafts
b. Shafts in Resonant Whirl
c. Bearings Cocked on the Shaft
d. Resonance of Some Component in the Axial Direction
e. Worn Thrust Bearings
f. Worn Helical or Bevel Gearsg. A Sleeve Bearing Motor Hunting for its Magnetic Center
h. Couple Component of a Dynamic Unbalance
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Vibration analysis
Mechanical looseness
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MECHANICALLOOSENESS (A)
MECHANICALLOOSENESS (A)
Caused by structural looseness of machine feet Distortion of the base will cause soft foot
problems Phase analysis will reveal aprox 180 0 phase
shift in the vertical direction between the baseplate components of the machine
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MECHANICALLOOSENESS (B)MECHANICAL
LOOSENESS (B)
Caused by loose pillow block bolts
Can cause 0.5, 1, 2 and 3X RPM Sometimes caused by cracked frame structure
or bearing block
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MECHANICALLOOSENESS (C)
MECHANICALLOOSENESS (C)
Phase is often unstable Will have many harmonics Can be caused by a loose bearing liner, excessive
bearing clearance or a loose impeller on a shaft
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Vibration analysis
Sleeve bearing/Rotor rub
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ROTOR RUBROTOR RUB
Similar spectrum to mechanical looseness Usually generates a series of frequencies which
may excite natural frequencies Sub harmonic frequencies may be present Rub may be partial or through a complete
revolution.
Truncated waveform
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OIL WHIP INSTABILITYOIL WHIP INSTABILITY
Oil whip may occur if a machine is operated at 2X therotor critical frequency.
When the rotor drives up to 2X critical, whirl is closeto critical and excessive vibration will stop the oil filmfrom supporting the shaft.
Whirl speed will lock onto rotor critical. If the speed isincreased the whip frequency will not increase.
oil whirl
oil whip
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OIL WHIRLINSTABILITYOIL WHIRL
INSTABILITY
Usually occurs at 42 - 48 % of running speed Vibration amplitudes are sometimes severe Whirl is inherently unstable, since it increases centrifugal
forces therefore increasing whirl forces
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Resonance
typically 10% or greater
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RESONANCERESONANCE
Resonance occurs when the ForcingFrequency coincides with a NaturalFrequency
1800
phase change occurs when shaft speedpasses through resonance High amplitudes of vibration will be present
when a system is in resonance
( )BELT PROBLEMS (A)
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BELT PROBLEMS (A)BELT PROBLEMS (A)
Often 2X RPM is dominant Amplitudes are normally unsteady, sometimes pulsing with either
driver or driven RPM Wear or misalignment in timing belt drives will give high amplitudes
at the timing belt frequency Belt frequencies are below the RPM of either the driver or the driven
Often 2X RPM is dominant Amplitudes are normally unsteady, sometimes pulsing with either
driver or driven RPM Wear or misalignment in timing belt drives will give high amplitudes
at the timing belt frequency Belt frequencies are below the RPM of either the driver or the driven
WORN, LOOSE OR MISMATCHED BELTSWORN, LOOSE OR MISMATCHED BELTS
BELT FREQUENCYHARMONICS
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BELT PROBLEMS (D)BELT PROBLEMS (D)
High amplitudes can be present if the belt naturalfrequency coincides with driver or driven RPM
Belt natural frequency can be changed by altering the belt
tension
High amplitudes can be present if the belt naturalfrequency coincides with driver or driven RPM
Belt natural frequency can be changed by altering the belt
tension
BELT RESONANCEBELT RESONANCE
RADIAL
1X RPM
BELT RESONANCE
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HYDRAULIC ANDAERODYNAMIC FORCES
HYDRAULIC ANDAERODYNAMIC FORCES
If gap between vanes and casing is not equal, Blade PassFrequency may have high amplitude High BPF may be present if impeller wear ring seizes on
shaft Eccentric rotor can cause amplitude at BPF to be
excessive
If gap between vanes and casing is not equal, Blade PassFrequency may have high amplitude High BPF may be present if impeller wear ring seizes on
shaft Eccentric rotor can cause amplitude at BPF to be
excessive
BPF = BLADE PASSFREQUENCY
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HYDRAULIC ANDAERODYNAMIC FORCES
HYDRAULIC ANDAERODYNAMIC FORCES
Flow turbulence often occurs in blowers due to variationsin pressure or velocity of air in ducts
Random low frequency vibration will be generated,possibly in the 50 - 2000 CPM range
Flow turbulence often occurs in blowers due to variationsin pressure or velocity of air in ducts
Random low frequency vibration will be generated,possibly in the 50 - 2000 CPM range
FLOW TURBULENCEFLOW TURBULENCE
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HYDRAULIC ANDAERODYNAMIC FORCES
HYDRAULIC ANDAERODYNAMIC FORCES
Cavitations will generate random, high frequencybroadband energy superimposed with BPF harmonics
Normally indicates inadequate suction pressure Erosion of impeller vanes and pump casings may occur if
left unchecked
Sounds like gravel passing through pump
Cavitations will generate random, high frequencybroadband energy superimposed with BPF harmonics
Normally indicates inadequate suction pressure Erosion of impeller vanes and pump casings may occur if
left unchecked
Sounds like gravel passing through pump
CAVITATIONCAVITATION
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BEAT VIBRATIONBEAT VIBRATION
A beat is the result of two closely spaced frequencies goinginto and out of phase
The wideband spectrum will show one peak pulsating upand down
The difference between the peaks is the beat frequencywhich itself will be present in the wideband spectrum
A beat is the result of two closely spaced frequencies goinginto and out of phase
The wideband spectrum will show one peak pulsating upand down
The difference between the peaks is the beat frequencywhich itself will be present in the wideband spectrum
WIDEBAND SPECTRUM
ZOOMSPECTRUM
F1 F2
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Vibration analysis
Electrical
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FREQUENCIES PRODUCED BYFREQUENCIES PRODUCED BY
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Electrical line frequency.(FL) = 50Hz = 3000 cpm.60HZ = 3600 cpm
No of poles. (P)
Rotor Bar Pass Frequency (Fb) = No of rotor bars x Rotor rpm.
Synchronous speed ( Ns) = 2xFL)
Slip frequency ( F S )= Synchronous speed Rotor rpm.
Pole pass frequency (F P )=Slip Frequency x No of Poles.
Electrical line frequency.(Electrical line frequency.( FLFL) =) = 50Hz = 3000 cpm.50Hz = 3000 cpm.60HZ = 36060HZ = 360 0 cpm0 cpm
No of poles.No of poles. ((PP ))
Rotor Bar Pass Frequency (Rotor Bar Pass Frequency ( FbFb ) =) = No of rotor bars x Rotor rpm.No of rotor bars x Rotor rpm.
Synchronous speed (Synchronous speed ( NsNs )) == 2xFL2xFL ))
Slip frequency (Slip frequency ( FF SS )=)= Synchronous speedSynchronous speed Rotor rpm.Rotor rpm.
Pole pass frequency (Pole pass frequency ( FFPP )=)= Slip Frequency x No of Poles.Slip Frequency x No of Poles.
FREQUENCIES PRODUCED BY
ELECTRICAL MOTORS.
FREQUENCIES PRODUCED BY
ELECTRICAL MOTORS.
PP
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ELECTRICAL PROBLEMSELECTRICAL PROBLEMS
Loose stator coils in synchronous motors generate highamplitude at Coil Pass Frequency
The coil pass frequency will be surrounded by 1X RPMsidebands
SYNCHRONOUS MOTORSYNCHRONOUS MOTOR(Loose Stator Coils)(Loose Stator Coils)
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ELECTRICAL PROBLEMSELECTRICAL PROBLEMS
Phasing problems can cause excessive vibration at 2F Lwith 1/3 F L sidebands Levels at 2F L can exceed 25 mm/sec if left uncorrected Particular problem if the defective connector is only
occasionally making contact
POWER SUPPLY PHASE PROBLEMSPOWER SUPPLY PHASE PROBLEMS(Loose Connector)(Loose Connector)
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ELECTRICAL PROBLEMSELECTRICAL PROBLEMS
1X, 2X, 3X, RPM with pole pass frequency sidebandsindicates rotor bar problems.
2X line frequency sidebands on rotor bar pass frequency(RBPF) indicates loose rotor bars.
Often high levels at 2X & 3X rotor bar pass frequencyand only low level at 1X rotor bar pass frequency.
ROTOR PROBLEMSROTOR PROBLEMS
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Vibration analysis
Gear
CALCULATION OF GEARCALCULATION OF GEAR
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CALCULATION OF GEARMESH FREQUENCIES
CALCULATION OF GEARMESH FREQUENCIES
20 TEETH20 TEETH
51 TEETH51 TEETH
1700 RPM1700 RPM
31 TEETH31 TEETH
HOW MANY TEETH ON THIS GEAR?HOW MANY TEETH ON THIS GEAR?
8959 RPM8959 RPM
GEARSGEARS
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GEARSNORMAL SPECTRUM
GEARSNORMAL SPECTRUM
Normal spectrum shows 1X and 2X and gear meshfrequency GMF GMF commonly will have sidebands of running speed All peaks are of low amplitude and no natural frequencies
are present
14 teeth
8 teeth GMF= 21k CPM
2625 rpm
1500 rpm
GEARSGEARS
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Gear Mesh Frequencies are often sensitive to load High GMF amplitudes do not necessarily indicate a
problem Each analysis should be performed with the system at
maximum load
GEARS TOOTH LOAD
GEARS TOOTH LOAD
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GEARSGEAR ECCENTRICITY AND BACKLASHGEARSGEAR ECCENTRICITY AND BACKLASH
Fairly high amplitude sidebands around GMF suggesteccentricity, backlash or non parallel shafts
The problem gear will modulate the sidebands Incorrect backlash normally excites gear natural
frequency
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GEARSCRACKED / BROKEN TOOTHGEARSCRACKED / BROKEN TOOTH
A cracked or broken tooth will generate a high amplitude at
1X RPM of the gear It will excite the gear natural frequency which will besidebanded by the running speed fundamental
Best detected using the time waveform Time interval between impacts will be the reciprocal of the
1X RPM
TIME WAVEFORM
GEARSGEARS
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HUNTING TOOTHHUNTING TOOTH
Vibration is at low frequency and due to this can often bemissed
Synonymous with a growling sound The effect occurs when the faulty pinion and gear teeth
both enter mesh at the same time
Faults may be due to faulty manufacture or mishandling
f Ht = (GMF)Na(TGEAR)( TPINION)
Vib i l i
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Vibration analysisBearings
Outer Race(BPFO)
Inner Race
(BPFI)
Ball Spin(BSF)
Cage or Train FTF
D0
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D1DB
Note : shaft turning outer race fixedF = frequency in cpmN = number of balls
BPFI = Nb/2 (1+(Bd/Pd)cos ) RPM
BPFO = Nb/2 (1-(Bd/Pd)cos ) RPM
BSF = Pd/2Bd (1-((Bd/Pd)cos )2) RPM
FTF = (1-((Bd/Pd)cos )) RPM
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ROLLING ELEMENTBEARINGS STAGE 1 FAILURE MODEROLLING ELEMENTBEARINGS STAGE 1 FAILURE MODE
Earliest indications in the ultrasonic range These frequencies evaluated by Spike Energy TM gSE,
HFD(g) and Shock Pulse Spike Energy may first appear at about 0.25 gSE for this
first stage
gSE
ZONE BZONE A ZONE C ZONE D
ROLLING ELEMENTROLLING ELEMENT
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ROLLING ELEMENTBEARINGS STAGE 2 FAILURE MODE
ROLLING ELEMENTBEARINGS STAGE 2 FAILURE MODE
Slight defects begin to ring bearing component natural
frequencies These frequencies occur in the range of 30k-120k CPM At the end of Stage 2, sideband frequencies appear above
and below natural frequency
Spike Energy grows e.g. 0.25-0.50gSE
ZONE AZONE B ZONE C ZONE D
gSE
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Examples
Singing Propeller
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Singing Propeller
0 50 100 150 200 250 300 350 40
0
0.06
0.12
0.18
0.24
0.30
0.36
Frequency in Hz
R M S V e
l o c
i t y
i n m m
/ S e c
0 50 100 150 200 250 300 350 400
0
0.3
0.6
0.9
1.2
1.5
1.8
2.1
.
Frequency in Hz
R M S V e
l o c
i t y
i n m m
/ S e c
Starboard side Port side
Singing Propeller
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Route Spectrum28-JUL-06 21:56: 44
OVRALL= 2.79 V-DGRMS = 2.76LOAD = 100.0RPM= 92.RPS = 1.53
80 100 120 140 160 180 200
0
0.3
0.6
0.9
1.2
1.5
1.8
2.1
2.4
Frequency in Hz
R M S V e
l o c
i t y
i n m m
/ S e c
Freq:Ordr:Spec:Dfrq:
142.2893.24.186
1.534
Sideband activity around theSideband activity around thetroubled frequency (140 Hz)troubled frequency (140 Hz)
The modulation/sideband The modulation/sidebandactivity tells us that theactivity tells us that thetroubled frequency is workingtroubled frequency is workingalong with the rpm of thealong with the rpm of theshaft.shaft.
Dfrq (Delta frequency) =Dfrq (Delta frequency) =1.534 Hz (*60sec)= 92 RPM1.534 Hz (*60sec)= 92 RPM
92 rpm = shaft speed when92 rpm = shaft speed whenmeasurements were taken.measurements were taken.
Singing PropellerDescribing the frequency spectra
Singing Propeller
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g g pConclusionAfter thorough measurements/analysis our conclusion is that the port side propeller suffersfrom a phenomenon called a singing propeller. The conclusion is justified by:
A frequency of approximately 140 Hz is causing the noise/vibration.
This frequency is independent from rpm within the troubled range of propeller revolution(60-105 rpm).The ~140 Hz frequency only appears on the port side propeller shaft. This was confirmedby single propeller transit on both starboard and port side.The ~140 Hz frequency measured has sideband (modulation) which is directly connectedto the speed of the port side shaft. This indicates that the troubled frequency is situatedsomewhere along this shaft.There is no other rpm independentcomponent along port side shaft line that can be asource to this frequency. The size and weight to the propeller can possibly fit to thesingingfrequency.
RecommendationGrinding an anti singing edge on the propeller.Result: The grinding of the propeller blades were carried out and the singing tonedisappeared
Bearing damage
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Bearing damage
Route Spectrum10-MAY-05 12: 07:36
OVRALL= 10.23 V-DGRMS = 1 .71LOAD = 100.0RPM = 2937.RPS = 48.95
0 20 40 60 80 100
0
0.2
0.4
0.6
0.8
1.0
Frequency in Order
R M S A c c e
l e r a
t i o n
i n G - s
Ordr:Freq:Spec:
5.436266.08.03517
>FAG 6322F=BPFI : 5.44
F F F F F F F F F F
Route Spectrum10-MAY-05 12: 07:36
OVRALL= 10.23 V-DGRMS = 1.71LOAD = 100.0RPM = 2937.RPS = 48.95
0 20 40 60 80 100
0
0.2
0.4
0.6
0.8
1.0
Frequency in Order
R M S A c c e
l e r a
t i o n
i n G - s
Ordr:Freq:Spec:
3.540173.27.01331
>FAG 6322E=BPFO : 3.56
E E E E E E E E E E
Observing frequencies thatmatches ball pass frequenciesinner race (fault frequenciesBPFI) on bearing FAG 6322
Observing frequencies thatmatches ball pass frequenciesouter race (fault frequenciesBPFO) on bearing FAG 6322
Bearing damage
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Bearing damageTrend Display
of 1. - 20. kHz
-- Baseline --Value: 1.143Date: 26-FEB-03
0 200 400 600 800 1000
0
0.5
1.0
1.5
2.0
2.5
3.0
3.5
4.0
4.5
5.
Days: 10-JAN-03 To 10-MAY-05
R M S A c c e
l e r a
t i o n
i n G - s
Date:Time:
Ampl:
10-MAY-0512:07:40
4.281
Label: WF 63 1RER-1 /
Route Spectr um10-MAY-05 12:09: 49
(Demod-HP 1000 Hz)
OVRALL= 1.49 A-DGRMS = 1.50LOAD = 100.0RPM= 2937.RPS = 48.95
0 2 4 6 8 10 12 14 16 18 20 22
0
0.2
0.4
0.6
0.8
1.0
Frequency in Order
R M S A c c e
l e r a
t i o n
i n G - s
Ordr:Freq:Spec:
5.433265.94
.715
>FAG 6322
F=BPFI : 5.44
F F F
Observing powerfulincreasement in the area 1-20kHz (which represents the areof bearing noise) This supportsthe assumption of a bearing
damage under development
Also the demodulatedmeasurement indicates faultfrequencies from the bearinginner ring on bearing FAG 6322
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FAG6322 (outer race)FAG6322 (outer race)
Bearing damage
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Outer ring. .
Route Spectrum01-MAR-05 09:47: 29
OVRALL= 15.10 V-DGRMS = 4.14LOAD = 100.0RPM= 2622.RPS = 43.70
0 1000 2000 3000 4000
0
0.3
0.6
0.9
1.2
1.5
1.8
2.1
2.4
Frequency in Hz
R M S A c c e
l e r a
t i o n
i n G
- s
Freq:Ordr:Spec:
300.176.869
.00788
>SKF NU2224E=BPFO : 299.6
E E E E E E E E E E
Trend Display
of 1. - 20. kHz
-- Baseline --Value: .986Date: 03-FEB-03
0 100 200 300 400 500 600 700 800
0
1
2
3
4
5
6
7
Days: 03-FEB-03 To 01-MAR-05
R M S A c c e
l e r a
t i o n
i n G
- s
ALERT
FAULT
Date:Time:
Ampl:
01-MAR-0509:47:37
5.531
Observing powerfulincreasement in the area 1-20kHz (which represents the areof bearing noise) This supportsthe assumption of a bearing
damage under development
Observing frequencies thatmatches ball pass frequenciesouter race (fault frequenciesBPFO) on bearing SKFNU2224
Bearing damage
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Outer ring
Observing powerfulincreasement in the area 1-20kHz (which represents the areof bearing noise) This supportsthe assumption of a bearing
damage under development
Observing frequencies thatmatches ball pass frequenciesouter race (fault frequenciesBPFO) on bearing TMK HH840200 (HH840249/210)
003 - GEAR SN: 61.88.6032.01.01G0008 -086 GEAR,INNG.AKS 1.LAGER RADIAL
Route Spectr um06-JUN-05 21:04:14
OVRALL= 21.82 V-DGRMS = 6.58LOAD =1550.0RPM= 1505.RPS = 25.09
0 1000 2000 3000 4000
0
0.3
0.6
0.9
1.2
1.5
1.8
2.1
2.4
2.7
Frequency in Hz
R M S A c c e
l e r a
t i o n i n
G - s
Freq:Ordr:Spec:
255.0210.17.102
>TMK HH840210/249E=BPFO : 256.5
E E E E E E E E
003 - GEAR SN: 61.88.6032.01.01G0008 -086 GEAR,INNG.AKS 1.LAGER RADIAL
Trend Displayof
1. - 20. kHz
-- Baseline --Value: 2.937Date: 12-MAR-03
0 200 400 600 800 1000
0
1
2
3
4
5
6
7
8
Days: 09-JAN-03 To 06-JUN-05
R M S A c c e
l e r a
t i o n i n
G - s
Date:Time:
Ampl:
06-JUN-0521:04:15
6.656
Bearing damage
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Outer ring (large transmission)
Observing increasement in the area1-20 kHz (which represents the are of bearing noise) This supports theassumption of a bearing damageunder development
Trend Displ ayof
2. - 20. kHz
-- Baseline --Value: .00000Date: 28-MAY-98
0 200 400 600 800 1000 1200
0
0.5
1.0
1.5
2.0
2.5
3.0
3.5
.
Days: 09-JAN-02 To 03-JAN-05
R M S A c c e
l e r a
t i o n
i n G - s
ALERT
FAULT
Date:Time:
Ampl:
09-JAN-0211: 03:24
.340
Trend Displayof
2. - 20. kHz
-- Baseline --Value: .00000Date: 28-MAY-98
0 200 400 600 800 1000 1200
0
0.5
1.0
1.5
2.0
2.5
3.0
3.5
.
Days: 09-JAN-02 To 03-JAN-05
R M S A c c e
l e r a
t i o n
i n G - s
ALERT
FAULT
Date:Time:
Ampl:
03-JAN-0514: 04:35
.551
Observing powerful increasementin the area 1-20 kHz (whichrepresents the are of bearing noise)
This supports the assumption of abearing damage under development
Input shaft motor side Input shaft drive side
Bearing damage
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Trend Displayof
2. - 20. kHz
-- Baseline --Value: .00000Date: 28-MAY-98
0 200 400 600 800 1000 1200
0
0.5
1.0
1.5
2.0
2.5
3.0
3.5
.
Days: 09-JAN-02 To 03-JAN-05
R M S A c c e
l e r a
t i o n
i n G - s
ALERT
FAULT
Date:Time:
Ampl:
09-JAN-0211: 03: 24
.340
Outer ring (large transmission)
Points of observedPoints of observeddamages on same type of damages on same type of bearingbearing
Due to earlier observation in this trending tool on this particuDue to earlier observation in this trending tool on this particu lar lar shaft, our conclusion is that there is a bearing damage.shaft, our conclusion is that there is a bearing damage.
Bearing damage on inner race motor sideBearing damage on inner race motor side
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Bearing damage on inner race drive sideBearing damage on inner race drive side
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Bearing damage
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Outer ring (thrust bearing)
Observing increasement in the area1-20 kHz (which represents the are of bearing noise) This supports theassumption of a bearing damageunder development
Route Spectrum03-NOV-*3 14:37
OVRALL= 18.24 V-DGRMS = 2.30LOAD = 100.0RPM = 1500.RPS = 25.00
0 400 800 1200 1600 2000
0
0.3
0.6
0.9
1.2
1.5
1.8
2.1
.
Frequency in Hz
R M S A c c e
l e r a
t i o n
i n G - s
Freq:Ordr:Spec:
247.509.9001.047
>SKF NU1026E=BPFO
E E E E E E E ETrend Display
of 1. - 20. kHz
-- Baseline --Value: .00000Date: 16-JUL-96
0 100 200 300 400 500 600 700
0
2
4
6
8
10
12
Days: 22-JAN-*2 To 03-NOV-*3
R M S A c c e
l e r a
t i o n
i n G - s
ALERT
FAULT
Date:Time:
Ampl :
03-NOV-*314:37:54
9.625
Observing frequencies thatmatches ball pass frequenciesouter race (fault frequenciesBPFO) on bearing SKF NU1026
Gear damage
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Input crown wheel
Time-waveform indicates that there isa pulsation on time per revolution.
This supports the assumption of agear damage. Possible broken tooth.
Observing harmonic rpmfrequencies on the input shaft of this gear
Route Spectrum03-FEB-04 14:37:03
OVRALL= 3.31 V-DGRMS = .4406LOAD = 100.0RPM = 278.RPS = 4.63
0 100 200 300 400 500 600 700
0
0.05
0.10
0.15
0.20
0.25
0.30
0.35
0.40
Frequency in Hz
R M S A c c e
l e r a
t i o n
i n
G - s
Freq:Ordr:Spec:
25.195.437
.02161 Time in mSecs
A c c e
l e r a
t i o n
i n G - s
0 40 80 120 160 200 240
PlotSpan
-4
4
29-NOV-02 13:34:02
12-JUN-03 12:04:11
12-SEP-03 11:49:13
07-OCT-03 13:09: 16
08-JAN-04 12:22: 13
03-FEB-04 14:26:02
Time: Ampl:
32.15-.906
Gear damage
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Intermediate shaftWaveform Display07-OCT-*3 13:16
RMS = .1089LOAD = 100.0RPM = 296.
RPS = 4.94
PK(+) = .7243PK(-) = .8067CRESTF= 7.41
0 100 200 300 400 500 600
-1.0
-0.8
-0.6
-0.4
-0.2
-0.00.2
0.4
0.6
0.81.0
Time in mSecs
A c c e
l e r a
t i o n
i n G - s
Time: Ampl:Dtim;Freq:
240.57.559
195.615.112
Route Spectrum07-OCT-*3 13:20(Demod- HP 500 Hz)
OVRALL= .0701 A-DGRMS = .0700LOAD = 100.0RPM = 76.RPS = 1.27
0 5 10 15 20 25 30 35 40 45 50
0
0.01
0.02
0.03
.
Frequency in Hz
R M S A c c e
l e r a
t i o n
i n G - s
Freq:Ordr:Spec:
5.0943.998
.02843
Time-waveform indicates thatthere is a pulsation on time perrevolution. This supports theassumption of a gear damage.
Demodulated measurement shows thatthere is a harmonic frequency of 5.094 Hz.5.094 Hz x 60 Hz = ~300 RPM which isclose to the intermediate shaft speed.
Therefore it is likely to believe that there isa tooth damage on this shaft
Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft
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Resonance problem
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CaseOn two main gears several tie-/anchor bolts for the pinion bearings on the first gearstep broken just after a couple of hundred hours, and therefore Maskindynamikk ASwas engaged to identify and analyze the vibration in these two gears. It was soondiscovered to be abnormally high levels of vibration in a specific speed-/load area
around these bolts (close to maximum speed), and these vibrations were amplifiedby the gearmesh frequencies of the input shaft.
This was the first observation that pointed in the direction of a possible resonance problem
Additional examination was therefore carried out to identity this resonance-problem.An element analysis was carried out to sort which of the gear components hadnatural frequencies in this frequency range (resonant area). This was not a easycase as more than one component could be involved in this.
Thru this investigation it was revealed that the bolts had radial natural frequencieswhich were amplified (excited) by 1 st level gearmesh frequency.
The resolution to the problem was therefore divided in two. First stage involvedredesigning and replacing the bolts with others with lower natural frequencies, andthereafter to change the propeller curve so that we achieve a lower maximum
rpm and a lower maximum gearmesh. In addition to this we also achieved to obtain
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the power by increasing the pitch curve.BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT
Analyze Spectrum08-SEP-07 00:48: 28
RMS = 8.38LOAD = 73.0RPM= 1050.RPS = 17.50
0 400 800 1200 1600
0
2
4
6
8
10
Frequency in Hz
R M S V e
l o c
i t y
i n m m
/ S
e c
Freq:Ordr:Spec:
716.9040.976.594
BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT
Analyze Spectrum08-SEP-07 01:00: 27
RMS = 23.05LOAD = 80.0RPM= 1080.RPS = 18.00
0 400 800 1200 1600
0
2
4
6
8
10
12
14
16
18
Frequency in Hz
R M S V e
l o c
i t y
i n m m
/ S
e c
Freq:Ordr:Spec:
735.7740.8813.37
The two engines is running at 1060rpm which gives a gearmesh of 718Hz with a amplitude of 6.7 mm/s.
This is normal
The two engines is running at 1080 rpm whichgives a gearmesh of 736 Hz with a amplitudeof 13.4 mm/s. An 2.5% increasement on thegearmesh frequency doubles the amplitude,and this clearly indicates a resonance problem
BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT
Analyze Spectrum08-SEP-07 01:41: 06
30
33
BSC - Port-gear-1500hzPort-HF -V05 VERTIKALT
Analyze Spectrum16-SEP-07 10:04: 08
0.7
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RMS = 29.24LOAD = 86.0RPM= 1100.RPS = 18.33
0 400 800 1200 1600 2000
0
3
6
9
12
15
18
21
24
27
Frequency in Hz
R M S
V e
l o c
i t y
i n m m
/ S e c
Freq:Ordr:Spec:
753.5341.1026.38
RMS = 2.59LOAD = 15.0RPM= 600.RPS = 10.00
0 200 400 600 800 1000 1200
0
0.1
0.2
0.3
0.4
0.5
0.6
Frequency in Hz
R M S
V e
l o c
i t y
i n m m
/ S e c
7 2 2
. 5 5
8 3 6 . 4 6
Freq:Ordr:Spec:
837.0083.70.220
The two engines is running at 1100 rpmwhich gives a gearmesh of 753 Hz with aamplitude of 26.4 mm/s. An 5.8%increasement on the gearmesh frequencyincreases the amplitude four times, and thisdefinitely indicates a resonance problem
The two engines is running at low and variable rpmwith 1 st order gearmesh around 350-400 Hz. This givesa 2 nd order gearmesh frequency in the are 700-850 Hz.Also the 2 nd order is strongly amplified somethingwhich confirms our assumption. This proves that thereis a resonance problem in this area (700-800Hz)
The measurement technique which were used her is called rpm sweeping with peak-hold functionwhichmeans that you sweep a frequency area to map possible resonance problems
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Unbalanced flexible coupling
Initial vibration analysis revealed
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mechanical unbalance in the coupling. Unbalance is indicated by a dominating
1.st order frequency amplitude. Unbalance can have different reasons Insuficcient dynamic balancing.
Coupling damages, as here where the stressbetween the rubber elements and the innerring (steel) has excedeeded the force limits
and the rubber elements were damaged afteronly a few months
Generator with a unbalanced/damaged coupling elements
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035 - GENERATOR 2Gen 2 -P05 GENERATOR, DE,VERTIKAL
Route Spectrum20-SEP-07 15:20:28
OVRALL= 22.12 V-DGRMS = 20.32LOAD = 100.0RPM = 1800.RPS = 30.00
0 40 80 120 160 200
0
2
4
6
8
10
12
14
16
18
Frequency in Hz
R M S V e
l o c
i t y
i n m m
/ S e c
Freq:Ordr:Spec:
30.001.00013.10
1.orden
The outer steel ring of the coupling was turned 180 degrees vs. therubber elements - wich in this case was the rebalancing trick to reduce
h 1 d ib i l l f 18 4 /
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the 1.st order vibration levels from 18 to 4 mm/s
Before vs. after dynamic balancing reduced 1.st order035 - GENERATOR 2
Gen 2 -P05 GENERA TOR, DE,VERTIKALRoute Spectrum20-SEP-07 15:20:28
OVRALL = 22.12 V-DG
16
18
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RMS = 20.32LOAD = 100.0RPM = 1800.RPS = 30.00
0 40 80 120 160 200
0
2
4
6
8
10
12
14
Frequency in Hz
R M S V e
l o c
i t y
i n m m
/ S e c
Freq:Ordr:Spec:
30.001.00013.10
035 - GENERATOR 2
Gen 2 -P05 GENERA TOR, DE,VERTIKAL Route Spectrum28-SEP-07 10:54:16
OVRAL L= 10.82 V-DGRMS = 10.49LOAD = 100.0RPM = 1801.RPS = 30.01
0 40 80 120 160 200
0
2
4
6
8
10
Frequency in Hz
R M S
V e
l o c
i t y
i n m m
/ S e c
Freq:Ordr:Spec:
30.001.0004.018
Recommended