4. Time Waveform Analysis
complex time waveform
individual vibration signals
combine to form a complex
time waveform showing overall
vibration
frequency
low
freq.
high freq.
time
overall vibration
5. Scale Factors
– When comparing overall vibration signals, it is
imperative that both signals be measured on the
same frequency range and with the same
scale factors.
6. Measurements & Units
Displacement (Distance)
mils or micrometer, mm
Velocity (Speed - Rate of change of displcmt)
in/sec or mm/sec
Acceleration (Rate of change of velocity)
G’s or in/sec2 or mm/sec2
8. Multi-Parameter Monitoring
Same Data in Velocity and Acceleration
Velocity
Spectrum
Acceleration
Spectrum
On the same bearing cap,
low freq. events (imbalance,
misalignment, etc.) show
best in the velocity
spectrum; while high freq.
events (bearing faults,
gearmesh) show best in the
acceleration spectrum
9. 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)
12. Narrow Bands with trend
T re n d o f
B a la n c e
A la rm
Amplitude
S u b -
H a rm o n ic 1 X 2 X B e a rin g B e a rin g G e a rs B e a rin g
1 x 2 x
.3
in /s e c
.1
in /s e cT im e
(D a y s )
T im e
(D a y s )
T re n d o f
B e a rin g s
1 0 x
14. Overall Vibration
• The total vibration energy
measured within a specific
frequency range.
– includes a combination of all
vibration signals within
measured frequency range
– does not include vibration
signals outside measured
frequency range
– produces a numerical value
18. Vibration analysis
• "Of all the parameters that can be measured non
intrusively in industry today, the one containing the
most information is the vibration signature." Art
Crawford
• Vibration Analysis is the foundation of a predictive
maintenance program
19. SIGNATURE ANALYSISSIGNATURE ANALYSIS
• Which frequencies exist and what are the
relationships to the fundamental exciting
frequencies.
• What are the amplitudes of each peak
• How do the peaks relate to each other
• If there are significant peaks, what are their
source
21. COUPLE UNBALANCECOUPLE UNBALANCE
• 1800 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 180o
22. OVERHUNG ROTOR
UNBALANCE
OVERHUNG ROTOR
UNBALANCE
• 1X RPM present in radial and axial directions
• Axial readings tend to be in-phase but radial readings
might be unsteady
• Overhung rotors often have both force and couple
unbalance each of which may require correction
23. Diagnosing UnbalanceDiagnosing Unbalance
• Vibration frequency equals rotor
speed.
• Vibration predominantly RADIAL
in direction.
• Stable vibration phase
measurement.
• Vibration increases as square of
speed.
• Vibration phase shifts in direct
proportion to measurement
direction.
900
900
25. ANGULAR
MISALIGNMENT
ANGULAR
MISALIGNMENT
• Characterized by high axial vibration
• 1800 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
26. PARALLEL
MISALIGNMENT
PARALLEL
MISALIGNMENT
• High radial vibration 1800 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
27. MISALIGNED BEARINGMISALIGNED BEARING
• Vibration symptoms similar to angular
misalignment
• Attempts to realign coupling or balance the rotor
will not alleviate the problem.
• Will cause a twisting motion with approximately
1800 phase shift side to side or top to bottom
28. 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 1800 difference
29. OTHER SOURCES OF HIGH
AXIAL VIBRATION
OTHER SOURCES OF HIGH
AXIAL 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 Gears
g. A Sleeve Bearing Motor Hunting for its Magnetic Center
h. Couple Component of a Dynamic Unbalance
31. MECHANICAL
LOOSENESS (A)
MECHANICAL
LOOSENESS (A)
• Caused by structural looseness of machine feet
• Distortion of the base will cause “soft foot”
problems
• Phase analysis will reveal aprox 1800 phase
shift in the vertical direction between the base
plate components of the machine
35. SLEEVE BEARING
WEAR / CLEARANCE
PROBLEMS
SLEEVE BEARING
WEAR / CLEARANCE
PROBLEMS
• Later stages of sleeve bearing wear will give a
large family of harmonics of running speed
• A minor unbalance or misalignment will cause
high amplitudes when excessive bearing
clearances are present
36. 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
37. OIL WHIP INSTABILITYOIL WHIP INSTABILITY
• Oil whip may occur if a machine is operated at 2X the
rotor critical frequency.
• When the rotor drives up to 2X critical, whirl is close
to critical and excessive vibration will stop the oil film
from supporting the shaft.
• Whirl speed will lock onto rotor critical. If the speed is
increased the whip frequency will not increase.
oil whirl
oil whip
38. OIL WHIRL
INSTABILITY
OIL 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
40. RESONANCERESONANCE
• Resonance occurs when the Forcing
Frequency coincides with a Natural
Frequency
• 1800 phase change occurs when shaft speed
passes through resonance
• High amplitudes of vibration will be present
when a system is in resonance
41. 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 FREQUENCY
HARMONICS
42. BELT PROBLEMS (D)BELT PROBLEMS (D)
• High amplitudes can be present if the belt natural
frequency 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 natural
frequency 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
43. HYDRAULIC AND
AERODYNAMIC FORCES
HYDRAULIC AND
AERODYNAMIC FORCES
• If gap between vanes and casing is not equal, Blade Pass
Frequency 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 Pass
Frequency 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 PASS
FREQUENCY
44. HYDRAULIC AND
AERODYNAMIC FORCES
HYDRAULIC AND
AERODYNAMIC FORCES
• Flow turbulence often occurs in blowers due to variations
in 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 variations
in pressure or velocity of air in ducts
• Random low frequency vibration will be generated,
possibly in the 50 - 2000 CPM range
FLOW TURBULENCEFLOW TURBULENCE
45. HYDRAULIC AND
AERODYNAMIC FORCES
HYDRAULIC AND
AERODYNAMIC FORCES
• Cavitations will generate random, high frequency
broadband 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 frequency
broadband 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
46. BEAT VIBRATIONBEAT VIBRATION
• A beat is the result of two closely spaced frequencies going
into and out of phase
• The wideband spectrum will show one peak pulsating up
and down
• The difference between the peaks is the beat frequency
which itself will be present in the wideband spectrum
• A beat is the result of two closely spaced frequencies going
into and out of phase
• The wideband spectrum will show one peak pulsating up
and down
• The difference between the peaks is the beat frequency
which itself will be present in the wideband spectrum
WIDEBAND SPECTRUM
ZOOM
SPECTRUM
F1 F2
48. ELECTRICAL PROBLEMSELECTRICAL PROBLEMS
• Stator problems generate high amplitudes at 2FL (2X line
frequency )
• Stator eccentricity produces uneven stationary air gap, vibration
is very directional
• Soft foot can produce an eccentric stator
STATOR ECCENTRICITY, SHORTED LAMINATIONSSTATOR ECCENTRICITY, SHORTED LAMINATIONS
AND LOOSE IRONAND LOOSE IRON
49. • 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 ( FS )= Synchronous speed – Rotor rpm.
• Pole pass frequency (FP )= Slip Frequency x No of Poles.
•• Electrical line frequency.(Electrical line frequency.(FLFL) =) = 50Hz = 3000 cpm.50Hz = 3000 cpm.
60HZ = 36060HZ = 3600 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 ( FFSS )=)= 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
50. ELECTRICAL PROBLEMSELECTRICAL PROBLEMS
• Loose stator coils in synchronous motors generate high
amplitude at Coil Pass Frequency
• The coil pass frequency will be surrounded by 1X RPM
sidebands
SYNCHRONOUS MOTORSYNCHRONOUS MOTOR
(Loose Stator Coils)(Loose Stator Coils)
51. ELECTRICAL PROBLEMSELECTRICAL PROBLEMS
• Phasing problems can cause excessive vibration at 2FL
with 1/3 FL sidebands
• Levels at 2FL 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)
52. ELECTRICAL PROBLEMSELECTRICAL PROBLEMS
• 1X, 2X, 3X, RPM with pole pass frequency sidebands
indicates 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 frequency
and only low level at 1X rotor bar pass frequency.
ROTOR PROBLEMSROTOR PROBLEMS
54. CALCULATION OF GEAR
MESH FREQUENCIES
CALCULATION OF GEAR
MESH 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
55. GEARS
NORMAL SPECTRUM
GEARS
NORMAL SPECTRUM
• Normal spectrum shows 1X and 2X and gear mesh
frequency 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
56. • 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
57. GEARS
TOOTH WEAR
GEARS
TOOTH WEAR
• Wear is indicated by excitation of natural frequencies
along with sidebands of 1X RPM of the bad gear
• Sidebands are a better wear indicator than the GMF
• GMF may not change in amplitude when wear occurs
14 teeth
1500 rpm
8 teeth
2625 rpm
GMF = 21k CPM
58. GEARS
GEAR ECCENTRICITY AND BACKLASH
GEARS
GEAR ECCENTRICITY AND BACKLASH
• Fairly high amplitude sidebands around GMF suggest
eccentricity, backlash or non parallel shafts
• The problem gear will modulate the sidebands
• Incorrect backlash normally excites gear natural
frequency
59. GEARS
GEAR MISALIGNMENT
GEARS
GEAR MISALIGNMENT
• Gear misalignment almost always excites second order or
higher harmonics with sidebands of running speed
• Small amplitude at 1X GMF but higher levels at 2X
and 3X GMF
• Important to set Fmax high enough to capture at least
2X GMF
60. GEARS
CRACKED / BROKEN TOOTH
GEARS
CRACKED / 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 be
sidebanded 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
61. GEARS
HUNTING TOOTH
GEARS
HUNTING TOOTH
• Vibration is at low frequency and due to this can often be
missed
• 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
fHt = (GMF)Na
(TGEAR)(TPINION)
63. D0
D1DB
Note : shaft turning outer race fixed
F = frequency in cpm
N = number of balls
BPFI = Nb/2 · (1+(Bd/Pd)cosӨ) · RPM
BPFO = Nb/2 · (1-(Bd/Pd)cosӨ) · RPM
BSF = Pd/2Bd · (1-((Bd/Pd)cosӨ)2) · RPM
FTF = ½ (1-((Bd/Pd)cosӨ)) · RPM
64. ROLLING ELEMENT
BEARINGS STAGE 1 FAILURE MODE
ROLLING ELEMENT
BEARINGS STAGE 1 FAILURE MODE
• Earliest indications in the ultrasonic range
• These frequencies evaluated by Spike EnergyTM 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
65. ROLLING ELEMENT
BEARINGS STAGE 2 FAILURE MODE
ROLLING ELEMENT
BEARINGS 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 A
ZONE B ZONE C ZONE D
gSE
66. ROLLING ELEMENT
BEARINGS STAGE 3 FAILURE MODE
ROLLING ELEMENT
BEARINGS STAGE 3 FAILURE MODE
• Bearing defect frequencies and harmonics appear
• Many defect frequency harmonics appear with wear the
number of sidebands grow
• Wear is now visible and may extend around the periphery of
the bearing
• Spike Energy increases to between 0.5 -1.0 gSE
ZONE A ZONE B ZONE C ZONE D
gSE
69. LFPS 1 024
Route Spectrum
28-JUL-06 21:56:44
OVRALL= 2.79 V-DG
RMS = 2.76
LOAD = 100.0
RPM= 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
FrequencyinHz
RMSVelocityinmm/Sec
Freq:
Ordr:
Spec:
Dfrq:
142.28
93.24
.186
1.534
•• Sideband activity around theSideband activity around the
troubled frequency (140 Hz)troubled frequency (140 Hz)
•• The modulation/sidebandThe modulation/sideband
activity tells us that theactivity tells us that the
troubled frequency is workingtroubled frequency is working
along with the rpm of thealong with the rpm of the
shaft.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 when
measurements were taken.measurements were taken.
Singing Propeller
Describing the frequency spectra
70. Singing Propeller
Conclusion
After thorough measurements/analysis our conclusion is that the port side propeller suffers
from 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 confirmed
by single propeller transit on both starboard and port side.
• The ~140 Hz frequency measured has sideband (modulation) which is directly connected
to the speed of the port side shaft. This indicates that the troubled frequency is situated
somewhere along this shaft.
• There is no other “rpm independent” component along port side shaft line that can be a
source to this frequency. The size and weight to the propeller can possibly fit to the
“singing” frequency.
Recommendation
Grinding an anti singing edge on the propeller.
Result: The grinding of the propeller blades were carried out and the singing tone
disappeared
71. Bearing damage
SF8000.182 645 AKSEL REIMHJUL 1. LAGER RADIELL
Route Spectrum
10-MAY-05 12:07:36
OVRALL= 10.23 V-DG
RMS = 1.71
LOAD = 100.0
RPM= 2937.
RPS = 48.95
0 20 40 60 80 100
0
0.2
0.4
0.6
0.8
1.0
Frequencyin Order
RMSAccelerationinG-s
Ordr:
Freq:
Spec:
5.436
266.08
.03517
>FAG 6322
F=BPFI : 5.44
F F F F F F F F F F
SF8000.182 645 AKSEL REIMHJUL 1. LAGER RADIELL
Route Spectrum
10-MAY-05 12:07:36
OVRALL= 10.23 V-DG
RMS = 1.71
LOAD = 100.0
RPM= 2937.
RPS = 48.95
0 20 40 60 80 100
0
0.2
0.4
0.6
0.8
1.0
Frequencyin Order
RMSAccelerationinG-s
Ordr:
Freq:
Spec:
3.540
173.27
.01331
>FAG 6322
E=BPFO : 3.56
E E E E E E E E E E
Observing frequencies that
matches ball pass frequencies
inner race (fault frequencies
BPFI) on bearing FAG 6322
Observing frequencies that
matches ball pass frequencies
outer race (fault frequencies
BPFO) on bearing FAG 6322
72. Bearing damage
Trend Display
of
1. - 20. kHz
-- Baseline --
Value: 1.143
Date: 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
5.0
Days: 10-JAN-03 To 10-MAY-05
RMSAccelerationinG-s
Date:
Time:
Ampl:
10-MAY-05
12:07:40
4.281
OFF ROUTE ORP OFF ROUTE MEASUREMENTPOINTDATA
Label: WF 63 1RER-1 /
Route Spectrum
10-MAY-05 12:09:49
(Demod-HP 1000 Hz)
OVRALL= 1.49 A-DG
RMS = 1.50
LOAD = 100.0
RPM= 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
Frequencyin Order
RMSAccelerationinG-s
Ordr:
Freq:
Spec:
5.433
265.94
.715
>FAG 6322
F=BPFI : 5.44
F F F
Observing powerful
increasement in the area 1-20
kHz (which represents the are
of bearing noise) This supports
the assumption of a bearing
damage under development
Also the demodulated
measurement indicates fault
frequencies from the bearing
inner ring on bearing FAG 6322
74. Bearing damage
Outer ring
SF8000.129 716 AKSELREIMHJUL 2. LAGER RADIELL
Route Spectrum
01-MAR-05 09:47:29
OVRALL= 15.10 V-DG
RMS = 4.14
LOAD = 100.0
RPM= 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
FrequencyinHz
RMSAccelerationinG-s
Freq:
Ordr:
Spec:
300.17
6.869
.00788
>SKF NU2224
E=BPFO : 299.6
E E E E E E E E E E
SF8000.129 716 AKSELREIMHJUL 2. LAGER RADIELL
Trend Display
of
1. -20. kHz
-- Baseline --
Value: .986
Date: 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
RMSAccelerationinG-s
ALERT
FAULT
Date:
Time:
Ampl:
01-MAR-05
09:47:37
5.531
Observing powerful
increasement in the area 1-20
kHz (which represents the are
of bearing noise) This supports
the assumption of a bearing
damage under development
Observing frequencies that
matches ball pass frequencies
outer race (fault frequencies
BPFO) on bearing SKF
NU2224
75. Bearing damage
Outer ring
Observing powerful
increasement in the area 1-20
kHz (which represents the are
of bearing noise) This supports
the assumption of a bearing
damage under development
Observing frequencies that
matches ball pass frequencies
outer race (fault frequencies
BPFO) on bearing TMK
HH840200 (HH840249/210)
003 - GEAR SN: 61.88.6032.01.01
G0008 -086 GEAR,INNG.AKS 1.LAGER RADIAL
Route Spectrum
06-JUN-05 21:04:14
OVRALL= 21.82 V-DG
RMS = 6.58
LOAD =1550.0
RPM= 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
Frequencyin Hz
RMSAccelerationinG-s
Freq:
Ordr:
Spec:
255.02
10.17
.102
>TMK HH840210/249
E=BPFO : 256.5
E E E E E E E E
003 - GEAR SN: 61.88.6032.01.01
G0008 -086 GEAR,INNG.AKS 1.LAGER RADIAL
Trend Display
of
1. - 20. kHz
-- Baseline --
Value: 2.937
Date: 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
RMSAccelerationinG-s
Date:
Time:
Ampl:
06-JUN-05
21:04:15
6.656
76. Bearing damage
Outer ring (large transmission)
Observing increasement in the area
1-20 kHz (which represents the are of
bearing noise) This supports the
assumption of a bearing damage
under development
,
Trend Display
of
2. -20. kHz
-- Baseline --
Value: .00000
Date: 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
4.0
Days: 09-JAN-02 To 03-JAN-05
RMSAccelerationinG-s
ALERT
FAULT
Date:
Time:
Ampl:
09-JAN-02
11:03:24
.340
,
Trend Display
of
2. -20. kHz
-- Baseline --
Value: .00000
Date: 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
4.0
Days: 09-JAN-02 To 03-JAN-05
RMSAccelerationinG-s
ALERT
FAULT
Date:
Time:
Ampl:
03-JAN-05
14:04:35
.551
Observing powerful increasement
in the area 1-20 kHz (which
represents the are of bearing noise)
This supports the assumption of a
bearing damage under development
Input shaft motor side Input shaft drive side
77. ,
Trend Display
of
2. -20. kHz
-- Baseline --
Value: .00000
Date: 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
4.0
Days: 09-JAN-02 To 03-JAN-05
RMSAccelerationinG-s
ALERT
FAULT
Date:
Time:
Ampl:
09-JAN-02
11:03:24
.340
Bearing damage
Outer ring (large transmission)
Points of observedPoints of observed
damages on same type ofdamages 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 particularlar
shaft, our conclusion is that there is a bearing damage.shaft, our conclusion is that there is a bearing damage.
78. Bearing damage on inner race motor sideBearing damage on inner race motor side
79. Bearing damage on inner race drive sideBearing damage on inner race drive side
80. Bearing damage
Outer ring (thrust bearing)
Observing increasement in the area
1-20 kHz (which represents the are of
bearing noise) This supports the
assumption of a bearing damage
under development
Route Spectrum
03-NOV-*3 14:37
OVRALL= 18.24 V-DG
RMS = 2.30
LOAD = 100.0
RPM = 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
2.4
Frequency in Hz
RMSAccelerationinG-s
Freq:
Ordr:
Spec:
247.50
9.900
1.047
>SKF NU1026
E=BPFO
E E E E E E E E
Trend Display
of
1. - 20. kHz
-- Baseline --
Value: .00000
Date: 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
RMSAccelerationinG-s
ALERT
FAULT
Date:
Time:
Ampl:
03-NOV-*3
14:37:54
9.625
Observing frequencies that
matches ball pass frequencies
outer race (fault frequencies
BPFO) on bearing SKF NU1026
81. Gear damage
Input crown wheel
Time-waveform indicates that there is
a pulsation on time per revolution.
This supports the assumption of a
gear damage. Possible broken tooth.
Observing harmonic rpm
frequencies on the input shaft of
this gear
,
Route Spectrum
03-FEB-04 14:37:03
OVRALL= 3.31 V-DG
RMS = .4406
LOAD = 100.0
RPM= 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
FrequencyinHz
RMSAccelerationinG-s
Freq:
Ordr:
Spec:
25.19
5.437
.02161 Time in mSecs
AccelerationinG-s
0 40 80 120 160 200 240
Plot
Span
-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
82. Gear damage
Intermediate shaft
,
WaveformDisplay
07-OCT-*3 13:16
RMS = .1089
LOAD = 100.0
RPM = 296.
RPS = 4.94
PK(+) = .7243
PK(-) = .8067
CRESTF= 7.41
0 100 200 300 400 500 600
-1.0
-0.8
-0.6
-0.4
-0.2
-0.0
0.2
0.4
0.6
0.8
1.0
Time in mSecs
AccelerationinG-s
Time:
Ampl:
Dtim;
Freq:
240.57
.559
195.61
5.112
,
Route Spectrum
07-OCT-*3 13:20
(Demod- HP 500 Hz)
OVRALL= .0701 A-DG
RMS = .0700
LOAD = 100.0
RPM = 76.
RPS = 1.27
0 5 10 15 20 25 30 35 40 45 50
0
0.01
0.02
0.03
0.04
Frequencyin Hz
RMSAccelerationinG-s
Freq:
Ordr:
Spec:
5.094
3.998
.02843
Time-waveform indicates that
there is a pulsation on time per
revolution. This supports the
assumption of a gear damage.
Demodulated measurement shows that
there is a harmonic frequency of 5.094 Hz.
5.094 Hz x 60 Hz = ~300 RPM which is
close to the intermediate shaft speed.
Therefore it is likely to believe that there is
a tooth damage on this shaft
83. Broken tooth on the intermediate shaftBroken tooth on the intermediate shaft
84. Resonance problem
Case
On two main gears several tie-/anchor bolts for the pinion bearings on the first gear
step broken just after a couple of hundred hours, and therefore Maskindynamikk AS
was engaged to identify and analyze the vibration in these two gears. It was soon
discovered to be abnormally high levels of vibration in a specific speed-/load area
around these bolts (close to maximum speed), and these vibrations were amplified
by 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 had
natural frequencies in this frequency range (resonant area). This was not a easy
case as more than one component could be involved in this.
Thru this investigation it was revealed that the bolts had radial natural frequencies
which were amplified (excited) by 1st level gearmesh frequency.
The resolution to the problem was therefore divided in two. First stage involved
redesigning and replacing the bolts with others with lower natural frequencies, and
thereafter to change the propeller curve so that we achieve a lower maximum
85. rpm and a lower maximum gearmesh. In addition to this we also achieved to obtain
the power by increasing the pitch curve.
BSC - Port-gear-1500hz
Port-HF -V05 VERTIKALT
Analyze Spectrum
08-SEP-07 00:48:28
RMS = 8.38
LOAD = 73.0
RPM= 1050.
RPS = 17.50
0 400 800 1200 1600
0
2
4
6
8
10
FrequencyinHz
RMSVelocityinmm/Sec
Freq:
Ordr:
Spec:
716.90
40.97
6.594
BSC - Port-gear-1500hz
Port-HF -V05 VERTIKALT
Analyze Spectrum
08-SEP-07 01:00:27
RMS = 23.05
LOAD = 80.0
RPM= 1080.
RPS = 18.00
0 400 800 1200 1600
0
2
4
6
8
10
12
14
16
18
FrequencyinHz
RMSVelocityinmm/Sec
Freq:
Ordr:
Spec:
735.77
40.88
13.37
The two engines is running at 1060
rpm which gives a gearmesh of 718
Hz with a amplitude of 6.7 mm/s.
This is normal
The two engines is running at 1080 rpm which
gives a gearmesh of 736 Hz with a amplitude
of 13.4 mm/s. An 2.5% increasement on the
gearmesh frequency doubles the amplitude,
and this clearly indicates a resonance problem
86. BSC - Port-gear-1500hz
Port-HF -V05 VERTIKALT
Analyze Spectrum
08-SEP-07 01:41:06
RMS = 29.24
LOAD = 86.0
RPM= 1100.
RPS = 18.33
0 400 800 1200 1600 2000
0
3
6
9
12
15
18
21
24
27
30
33
FrequencyinHz
RMSVelocityinmm/Sec
Freq:
Ordr:
Spec:
753.53
41.10
26.38
BSC - Port-gear-1500hz
Port-HF -V05 VERTIKALT
Analyze Spectrum
16-SEP-07 10:04:08
RMS = 2.59
LOAD = 15.0
RPM= 600.
RPS = 10.00
0 200 400 600 800 1000 1200
0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
FrequencyinHz
RMSVelocityinmm/Sec
722.55
836.46
Freq:
Ordr:
Spec:
837.00
83.70
.220
The two engines is running at 1100 rpm
which gives a gearmesh of 753 Hz with a
amplitude of 26.4 mm/s. An 5.8%
increasement on the gearmesh frequency
increases the amplitude four times, and this
definitely indicates a resonance problem
The two engines is running at low and variable rpm
with 1st order gearmesh around 350-400 Hz. This gives
a 2nd order gearmesh frequency in the are 700-850 Hz.
Also the 2nd order is strongly amplified something
which confirms our assumption. This proves that there
is a resonance problem in this area (700-800Hz)
The measurement technique which were used her is called “rpm sweeping with peak-hold function” which
means that you sweep a frequency area to map possible resonance problems
88. • Initial vibration analysis revealed
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 stress
between the rubber elements and the inner
ring (steel) has excedeeded the force limits
and the rubber elements were damaged after
only a few months
90. The outer steel ring of the coupling was turned 180 degrees vs. the
rubber elements - wich in this case was the rebalancing trick to reduce
the 1.st order vibration levels from 18 to 4 mm/s
