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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


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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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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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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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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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 frequencies thatmatches ball pass frequenciesouter race (fault frequenciesBPFO) on bearing SKFNU2224

Bearing damage

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Outer ring



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


0 200 400 600 800 1000 1200

0

0.5

1.0

1.5

2.0

2.5

3.0

3.5

.


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


0 200 400 600 800 1000 1200

0

0.5

1.0

1.5

2.0

2.5

3.0

3.5

.


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


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


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



30

33



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


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

Documents

212292