Reference Number Title/Description

Back

2002 AGMA ANSI/AGMA 6000-B96 Specification for Measurement of Linear Vibration on

Gear Units

2003 API ANSI/API std 541-2003 Form-Wound Squirrel-Cage Induction Motors 500 hp

and Larger

1997 API API STD 546, second edition Brushless Synchronous Machines, 500 kVA and Larger

2004 API API STD 610/ISO 13709,

ninth edition

Centrifugal Pumps for Petroleum, Petrochemical and

Natural Gas Industries

1997 API API STD 611, fourth edition General Purpose Steam Turbines for Petroleum,

Chemical and Gas Industry Services

2005 API API STD 612/ISO 10437,

sixth edition

Petroleum, Petrochemical and Natural Gas Industries –

Steam Turbines – Special-Purpose Applications

2003 API API STD 613, fifth edition Special Purpose Gear Units for Petroleum,

Chemical and Gas Industry Services

1998 API API STD 616, fourth edition Gas Turbines for the Petroleum, Chemical, and Gas

Industry Services

2002 API API STD 617, seventh edition Axial and Centrifugal Compressors and

Expander-compressors for Petroleum, Chemical and

Gas Industry Services

2000 API API STD 670, fourth edition Mechanical Protection Systems

2004 API API STD 672, fourth edition Packaged Integrally Geared, Centrifugal Air Compressors

for Petroleum, Chemical, and Gas Industry Services

2001 API API STD 673, second edition Special Purpose Fans

1997 API API STD 677, second edition General Purpose Gear Units for Petroleum, Chemical,

and Gas Industry Services

1996 API API STD 681, first edition Liquid Ring Vacuum Pumps for Petroleum, Chemical,

and Gas Industry Services

2000 API API STD 685, first edition Sealless Centrifugal Pumps for Petroleum, Heavy-Duty

Chemical, and Gas Industry Services

1965 BDS BDS 5626-65 Measurement of Vibration on Electrical Rotating Machines

1964 Blake, M.P. Hydrocarbon Processing,

January 1964

New Vibration Standards for Maintenance

1963 CAGI In-Service Standards for Centrifugal Compressors

1975 CAGI Standard for Centrifugal Air Compressors

1971 CSN CSN 011410 Permitted Limits for Unbalanced Solid Machine Elements

1968 Dresser

Industrial

General Guidelines for Vibration on Clark Centrifugal

Compressors

1966 Gosstandart GOST 12379-66 Measurement of Vibration on Electrical Rotating

Machines

2002 HI ANSI/HI 9.6.4 Centrifugal and Vertical Pumps — Vibration

Measurement and Allowable Values

1996 IEC IEC 60034-14 Rotating Electrical Machines, Part 14: Mechanical

Vibrations of Certain Machines with Shaft Heights

56 mm and Higher — Measurement, Evaluation and

Limits of Vibration

1964 IRD IRD #305D General Machinery Vibration Severity Chart

1995 – 2001 ISO Mechanical Vibration — Evaluation of Machine

Vibration by Measurements on Nonrotating Parts:

ISO 10816-1:1995 Part 1: General Guidelines

ISO 10816-2:2001 Part 2: Land-Based Steam Turbines and Generators in

Excess of 50 MW with Normal Operating Speeds

of 1500, 1800, 3000 and 3600 rpm

ISO 10816-3:1998 Part 3: Industrial Machines with Nominal Power above

15 kW and Nominal Speeds between 120 and

15,000 rpm when Measured In Situ

ISO 10816-4:1998 Part 4: Gas Turbine Driven Sets Excluding Aircraft

Derivations

34-42 Vibration and Shock Handbook

© 2005 by Taylor & Francis Group, LLC

TABLE 34.5 (continued)

Year Author/

Organization

Reference Number Title/Description

ISO 10816-5: 2000 Part 5: Machine Sets in Hydraulic Power Generating

and Pumping Plants

2002 ISO Mechanical Vibration — Vibration of Active Magnetic

Bearing Equipped Rotating Machinery

ISO 14839-1: 2002 Part 1: Vocabulary

ISO/CD 14839-2:2004 Part 2: Evaluation of Vibration

1996 – 2001 ISO Mechanical Vibrations of Nonreciprocating Machines —

Measurement on Rotating Shafts and Evaluation Criteria

ISO 7919-1: 1996 Part 1 (1996): General Guidelines

ISO 7919-2: 2001 Part 2 (2001): Land-Based Steam Turbines and

Generators in Excess of 50 MW with Normal

Operating Speeds of 1500, 1800, 3000 and 3600 rpm

ISO 7919-3: 1996 Part 3 (1996): Coupled Industrial Machines

ISO 7919-4: 1996 Part 4 (1996): Gas Turbine Sets

ISO 7919-5: 1997 Part 5 (1997): Machine Sets in Hydraulic Power

Generating and Pumping Plants

1993 ISO ISO 8579-2 Acceptance Code for Gears, Part 2: Determination of

Mechanical Vibration of Gear Units During Acceptance

Testing

2004 ISO Rolling Bearings — Measuring Methods for Vibration

ISO 15242-1:2004 Part 1: Fundamentals

ISO 15242-2:2004 Part 2: Radial Ball Bearings with Cylindrical Bore and

Outside surface

ISO/CD 15242-3 Part 3: Spherical and Taper Radial Roller Bearings with

Cylindrical Bore and Outside Diameter

1959 Kruglov, N.V. Teplonerg, 8 (85), 1959 Turbomachine Vibration Standards

1967 Maten, S Hydrocarbon Processing,

January 1967

New Vibration Velocity Standards

1983 McHugh, J.D. J. Lub. Tech., Trans. ASME,

1983, 105

Estimating the Severity of Shaft Vibration within

Fluid Film Journal Bearings

1974 MIL MIL-STD-167-1 Mechanical Vibration of Shipboard Equipment, Type I:

Environmental, Type II: Internally Excited

2003 NEMA NEMA MG 1-2003 Motors and Generators, Part 7 — Mechanical Vibration —

Measurement, Evaluation and Limits

1991 NEMA NEMA SM 23-1991 Steam Turbines for Mechanical Drive Service

1991 NEMA NEMA SM 24-1991 Land Based Steam Turbine Generator Sets 0 to 33,000 kW

1965 PKN PN-65/E-04255 Measurement of Vibration of Electrical Rotating

Machines

1939 Rathbone, T.C. Power Plant Engineering,

November 1939

Vibration Tolerances

1964 VDI VDI 2056 Evaluation Criteria for Mechanical Vibrations in

Machines

1982 VDI VDI 2059 P1 Shaft Vibrations of Turbosets Principles for Measurement

and Evaluation

1990 VDI VDI 2059 P2 Shaft Vibrations of Steam Turbosets for Power Station

Measurement and Evaluation

1985 VDI VDI 2059 P3 Shaft Vibrations of Industrial Turbosets Measurement

and Evaluation

1981 VDI VDI 2059 P4 Shaft Vibrations of Gas Turbosets Measurement and

Evaluation

1982 VDI VDI 2059 P5 Shaft Vibrations of Hydraulic Machinesets Measurement

and Evaluation

1949 Yates, H.G. Trans. N.E. Coast Inst. Engrs

Ship Builders, Vol. 65, 1949

Vibration Diagnosis of Marine Geared Turbines

Vibration in Rotating Machinery 34-43

© 2005 by Taylor & Francis Group, LLC

FIGURE 34.14 (a) Measuring points; (b) measuring points for vertical machine sets. (Source: ISO 10816-3,

1998-05-15. With permission.)

34-44 Vibration and Shock Handbook

© 2005 by Taylor & Francis Group, LLC

FIGURE 34.14 (continued )

Vibration in Rotating Machinery 34-45

© 2005 by Taylor & Francis Group, LLC

resulting in wasted energy and premature failure of components due to high vibration. The current

practice to obtain flow changes in the pump is by means of speed change. This eliminates flow throttling

and allows the pump to operate close to its best efficiency point, where energy is not wasted and

vibrations are a minimum. However, as illustrated below, variable speed operation of a pump-motor set

over a wide speed range could pose several challenging problems.

TABLE 34.6A Acceptable Vibration Levels for Rotating Machinery Measured on Nonrotating Parts

Machinery Type Power Level Speed Range (rpm) Applicable Vibration Level

Rigid Support Flexible Support

Steam turbines 15 # P # 300 kW 120 # N # 15,000 V1 and D3 V3 and D7

300 kW # P # 50 MW 120 # N # 15,000 V3 and D5 V6 and D8

P . 50 MW N , 1,500 or N .3,600 V3 and D5 V6 and D8

P . 50 MW N ¼1,500 or 1,800 V5 V5

P . 50 MW N ¼ 3,000 or 3,600 V7 V7

Gas turbines 15 # P # 300 kW 120 # N # 15,000 V1 and D3 V3 and D7

300 kW # P # 3 MW 120 # N # 15,000 V3 and D5 V6 and D8

P . 3 MW 3,000 # N # 20,000 V8 V8

Hydraulic turbines and

pump turbine

Horizontal machines

P . 1 MW 60 # N # 300 N/A V4

P . 1 MW 300 , N # 1,800 V2 and D6 N/A

Vertical machines P . 1 MW 60 , N # 1,800 V2 and D6 N/A

P . 1 MW 60 , N # 1,000 V2 and D6 V4 and D9

Centrifugal pumps

Separate driver P . 15 kW 120 # N # 15,000 V3 and D2 V6 and D4

Integral driver P . 15 kW 120 # N # 15,000 V1 and D1 V3 and D2

Electric motors

Shaft height H $ 315 mm P . 15 kW 120 # N # 15,000 V3 and D5 V6 and D8

Shaft height 160 # H , 315 mm P . 15 kW 120 # N # 15,000 V1 and D3 V3 and D7

Generators, excluding those used

in hydraulic power generation

15 # P # 300 kW 120 # N # 15,000 V1 and D3 V3 and D7

300 kW # P # 50 MW 120 # N # 15,000 V3 and D5 V6 and D8

P . 50 MW N , 1,500 or N . 3,600 V3 and D5 V6 and D8

P . 50 MW N ¼ 1,500 or 1,800 V5 V5

P . 50 MW N ¼ 3,000 or 3,600 V7 V7

Generators and motors used in

hydraulic power generation

Horizontal machines

P . 1 MW 60 # N # 300 N/A V4

P . 1 MW 300 , N # 1,800 V2 and D6 N/A

Vertical machines P . 1 MW 60 , N # 1,800 V2 and D6 N/A

P . 1 MW 60 , N # 1,000 V2 and D6 V4 and D9

Compressors, rotary, blowers, 15 # P # 300 kW 120 # N # 15,000 V1 and D3 V3 and D7

and fans 300 kW # P # 50 MW 120 # N # 15,000 V3 and D5 V6 and D8

TABLE 34.6B Maximum Vibration Velocity Limits for Different Levels (mm/sec, RMS)

Vibration Level ZoneA ZoneB ZoneC Alarm Trip

V1 1.4 2.8 4.5 3.5 5.6

V2 1.6 2.5 4.0 3.1 5.0

V3 2.3 4.5 7.1 5.6 8.9

V4 2.5 4.0 6.4 5.0 8.0

V5 2.8 5.3 8.5 6.6 10.6

V6 3.5 7.1 11.0 8.9 13.8

V7 3.8 7.5 11.8 9.4 14.8

V8 4.5 9.3 14.7 11.6 18.4

34-46 Vibration and Shock Handbook

© 2005 by Taylor & Francis Group, LLC

Description. The following case study is taken from a petroleum pipeline pump application where

pump-motor sets with VFDs were installed in a new pipeline starting in Alberta, Canada and terminating

in Minnesota, USA. VFDs are frequently used in the pipeline industry to power high horsepower pumps

to eliminate power wasted by throttling, reduce inrush current at motor startup, and to provide greater

operating flexibility. However, variable speed operation can cause vibration problems in the pump,

motor, and the couplings that are not normally experienced with fixed speed pumps. Unexpected high

torsional and lateral vibrations were experienced with these pumps and motors at certain

operating speeds. A rotor torsional resonance, motor housing resonance, acoustic resonance in the

internals of the pump, and discharge piping were identified to be the causes of the high vibration in the

pump-motor set. Details on diagnosing the problems and the corrective measures taken to resolve them

are given below:

Pump type: The pump was a centrifugal, two-stage, double volute horizontal pump, with six vane

impellers, normally designed to operate at a fixed speed. Generation of pressure pulsations at the vane

passing frequencies of 6 £ and 12 £ rotational speed is normally expected.

TABLE 34.6C Maximum Vibration Displacement Limits for Different Levels (mm, RMS)

Vibration Level ZoneA ZoneB ZoneC Alarm Trip

D1 11 22 36 28 45

D2 18 36 56 45 70

D3 22 45 71 56 89

D4 28 56 90 70 113

D5 29 57 90 71 113

D6 30 50 80 63 100

D7 37 71 113 89 141

D8 45 90 140 113 175

D9 65 100 160 125 200

Zone A: Newly commissioned machines should fall within this zone. Zone B: Machines with vibrations within this

zone are considered acceptable for long-term operation. Zone C: Machines with vibrations within this zone are normally

considered unsatisfactory for long-term operation. Such a machine may be operated for a short period in this condition.

Alarm: The values chosen will normally be set relative to a baseline value determined from experience. However, it is

recommended that the alarm value shall not exceed those given herein. Trip: The values will generally relate to the

mechanical integrity of the machine. They will generally be the same for all machines with similar design. It is

recommended that the trip value shall not exceed those given herein. Notes: (1) The measured vibration is broadband,

and the frequency range will depend on the type of machine being considered. A range from 2 to 1000 Hz is typical

except for in high-speed machines, . 10,000 rpm, where the upper limit should at least be six times the rotational

frequency. (2) It is common practice to evaluate rotating machinery based on the broadband RMS vibration velocity,

since it can be related to the vibration energy levels. However, other quantities such as vibration displacement or

acceleration may be preferred. Especially low speed machines can have unacceptably large vibration displacements when

the 1 £ rpm component is dominant. Therefore, where specified, both the velocity and displacement criteria are met. (3)

Since typical vibration waveforms measured on rotating machinery are complex in nature, there is no simple relationship

between broadband velocity, displacement, and acceleration. (4) Vibration measurements shall be taken on bearing

support housings, or other structural components, which adequately respond to the dynamic forces of the machine.

Recommended locations for bearing housings are shown in Figure 34.14. (5) For certain types of machines, the axial

vibration limits may differ from those for radial directions. Also, within the same machine set, in particular hydraulic

power-generating sets, the applicable level may differ from bearing to bearing depending on its classification as a rigid or

flexible support. (6) Above vibration limits apply to steady-state/normal operating conditions of the machine. If the

vibration levels are sensitive to the operational conditions, then evaluation of the machine for operating conditions

outside steady-state conditions will have to be based on different criteria. (7) The vibration limits specified herein should

not be used to assess the condition of rolling element type bearings although it encompasses machines that may have

these types of bearings. (8) It must be recognized that the vibration measurement on nonrotating parts alone does not

form the only basis for judging the condition of a machine. In certain types of machines, it is common practice also to

judge the vibration based on measurements taken on rotating shafts. (9) A support may be considered as rigid in a

specific direction only if its natural frequency in that direction exceeds the main excitation frequency by at least 25%,

otherwise it is considered to be flexible. In some cases, a support may be rigid in one direction and flexible in another.

(10) In the case of hydraulic machine sets, major differences in radial bearing support arrangement can occur. For

evaluation of the support type it is recommended that the reader refer to ISO 10816-5.

Vibration in Rotating Machinery 34-47

© 2005 by Taylor & Francis Group, LLC

Motor: The motor was a 3000 hp, two pole horizontal induction motor, designed to operate at

3600 rpm.

Supply: The supply was a VFD of the current source inverter type. These drives are known to generate

an oscillatory torque at 6 £ and 12 £ the operating frequency.

Coupling: Flexible disc type coupling with a spacer was used. These couplings have very little torsional

damping capacity.

Speed range: The speed range was from 1440 rpm (24 Hz) to 3900 rpm (65 Hz).

Reference: Refer to Figure 34.15 to Figure 34.18.

As for the resonance at second and third torsional critical speeds (Figure 34.15) the second and third

torsional modes are excited when the 6 £ component of rotational speed corresponds to the critical

speeds of 92 and 268 Hz, respectively. The 6 £ rpm torsional excitation is caused by the pressure

pulsations in the pump. The waterfall plot (Figure 34.15a) was taken during a run down of the set with

the power to the motor turned off.

A similar plot taken during run up of the motor (Figure 34.15b) shows excitations at the same

frequencies but having different amplitude. Since both the pump and motor generate 6 £ excitation, it

suggests a phase difference between the excitation torques.

It is important to note that the conventional vibration monitoring devices cannot detect the torsional

resonance problem. The only indication of a problem was the unusual chattering noise emitted by the

coupling. Special techniques to measure dynamic torque using strain gauges had to be used to detect the

torsional vibrations.

As for the motor housing resonance (Figure 34.16), the 2 £ rotational speed vibration of the motor is

dominant and peaks at 118 Hz corresponding to a natural frequency of the motor frame. In the waterfall

plot, the natural frequencies corresponds to excitations that are parallel to the axis. Excitation at

harmonics, including the 6 £ component, is present but is not dominant.

As for the pump vibrations at the vane passing frequency (Figure 34.17), the 6 £ vane passing

frequency is dominant at all operating speeds. It peaks at 238 Hz, possibly due to an acoustic resonance in

TABLE 34.7A Acceptable Vibration Levels for Rotating Machinery, Measured on Rotating Shafts

Machinery Type Power Level Speed Range (RPM) Applicable Vibration Level

Relative Displacement Absolute Displacement

Steam turbines P # 50 MW 1,000 # N # 30,000 D8 —

P . 50 MW N ¼ 1,500 D5 D7

P . 50 MW N ¼ 1; 800 D4 D6

P . 50 MW N ¼ 3,000 D2 D5

P . 50 MW N ¼ 3,600 D1 D3

Gas turbines P . 3 MW 3,000 # N # 30,000 D8 —

P # 3 MW 1,000 # N # 30,000 D8 —

Hydraulic turbines and pumps

used in hydraulic power

generation and pumping plants

P . 1 MW 60 # N # 1,800 D9 D9

Centrifugal pumps All 1,000 # N # 30,000 D8 —

Electric motors All 1,000 # N # 30,000 D8 —

Generators, excluding those used P # 50 MW 1,000 # N # 30,000 D8 —

in hydraulic power generation P . 50 MW N ¼ 1,500 D5 D7

P . 50 MW N ¼ 1,800 D4 D6

P . 50 MW N ¼ 3,000 D2 D5

P . 50 MW N ¼ 3,600 D1 D3

Generators and motors used P . 1 MW 60 # N # 1,000 D9 D9

in hydraulic power generation P . 1 MW 1,000 , N # 1,800 D8 —

Compressors, rotary, blowers,

and fans

All 1,000 # N # 30,000 D8 —

34-48 Vibration and Shock Handbook

© 2005 by Taylor & Francis Group, LLC

the discharge pipe. It does not seem to correspond to a structural natural frequency due to the absence of

excitations at 238 Hz at all speeds.

As for the acoustic resonance in the pump cross-over pipe (Figure 34.18), dynamic pressure pulsation

measurements made on the pump cross-over pipe from the first stage discharge to the second stage

suction show an acoustic resonance at 540 Hz. The consistent presence of some excitation at 540 Hz at all

speeds confirms that it is an acoustic natural frequency of the cross-over pipe. When the 6 £ rpm

pressure pulsation frequency coincides with the acoustic natural frequency, a resonance condition occurs

and the magnitude of the pressure pulsation increases by almost a factor of 30.

As for corrective action, for a pump that has to operate over a wide speed range, totally eliminating the

coincidence of all the frequencies of exciting forces with the system natural frequencies is impractical.

Therefore, the system has to be designed such that the resulting magnitudes of the forces are controlled to

within tolerable levels so that safe and reliable operation can take place. This can be accomplished by a

TABLE 34.7B Maximum Vibration Displacement Sp – p Limits (mm) Peak-to-Peak Limits for Different Levels

ZoneA ZoneB Zone C

D1 75 150 240

D2 80 165 260

D3 90 180 290

D4 90 185 290

D5 100 200 320

D6 110 220 350

D7 120 240 385

D8 4800=

ffiffi

pn 9000=

ffiffi

pn 13,200/

ffiffi

pn

D9 10ð2:338120:0704 log nÞ 10ð2:559920:0704 log nÞ 10ð2:860920:0704 log nÞ

Zone A: Newly commissioned machines should fall within this zone. Zone B: Machines with vibrations within this zone

are considered acceptable for long-term operation. Zone C: Machines with vibrations within this zone are normally

considered unsatisfactory for long-term operation. Such a machine may be operated for a short period in this condition.

Alarm: The values chosen will normally be set relative to a baseline value determined from experience. However, it is

recommended that the alarm value shall not exceed those given herein. Trip: The values will generally relate to the mechanical

integrity of the machine. They will generally be the same for all machines with similar design. It is recommended that the trip

value shall not exceed those given herein. Notes: (1) The measured vibration is broadband and is shaft vibration displacement

peak to peak. Where applicable, vibration limits for both absolute and relative radial shaft vibrations are given in certain

cases. (2) Relative displacement is the vibratory displacement between the shaft and an appropriate structural component

such as the bearing housing. Absolute displacement is the vibratory displacement of the shaft with reference to an inertial

frame of reference. (3) Relative measurements are carried out with a noncontacting transducer. Absolute readings are

obtained by one of the following methods: by a shaft riding probe on which a seismic transducer is mounted so that it

measures absolute shaft displacement directly, or with the combination of a noncontacting transducer which measures

relative shaft displacement and a seismic transducer which measures support vibration. Their conditioned outputs are

vectorially added to provide a measure of the absolute shaft motion. (4) The vibration evaluation criteria are dependent upon

a variety of factors and the criteria adopted will vary for different types of machines. Some of these factors are the

bearing type, clearance, and diameter. The adopted criteria have to be compared with the bearing diametral clearance

ðCÞ and adjusted to suit. Typical values are: Zone A #0.4C; Zone B #0.6C; and Zone C #0.7C. (5) Above vibration limits

apply to steady state/normal operating conditions of the machine. If the vibration levels are sensitive to the operational

conditions then evaluation of the machine for operating conditions outside steady-state conditions will have to be based on

different criteria. (6) It is recommended that vibration readings at each location be made with a pair of transducers and that

the transducers are mounted perpendicular to the shaft axis and they are at an angle of 908 to one another. The vibration

limits apply to each measured direction. (7) The mechanical and electrical run-out at each measurement location must be

assessed and should be , 25% of the allowable limit or 6 mm, whichever is greater. (8) It must be recognized that the

vibration measurement on rotating shafts does not form the only basis for judging the condition of a machine. In certain

types of machine, it is common practice also to judge the vibration based on measurements taken on nonrotating parts.

(9) ALARM levels should be set relative to a baseline value determined from experience for the measurement position,

direction and type of machine. It must provide a warning that a defined value, which is significantly above the baseline value,

has been reached. The maximum ALARM setting should be # 0.75C. (10) The TRIP values should be based on protecting the

mechanical integrity of the machine. Consideration of damage to bearings is typical; therefore, maximum TRIP setting

should be # 0.9C.

Vibration in Rotating Machinery 34-49

© 2005 by Taylor & Francis Group, LLC

TABLE 34.8 Vibration Cause Identification

Cause Dominant Frequency Spectrum, Time Domain,

Orbit Shape

Characteristics,

Corrections, Comments

Mass unbalance 1£ High 1 £ with much lower

harmonics; circular or

elliptic orbits

Corrected by shop or

field balancing

Shaft bow 1 £ Run down plot shows

decrease of vibration

at critical speed

The shaft has to

be straightened

using an acceptable

method

Misalignment 1 £ and 2 £ Equally high 1 £ and 2 £ ,

figure 8 orbits

Realign at operating

conditions; loads causing

misalignment, such as

nozzle loads, may have to

be reduced

Worn journal bearings 1 £ , 1/2 £ Equally high 1 £ and 1/2 £ Difficult to balance

Gravity critical 2 £ Run down plot will

show excitation at 1/2

critical speed

Can be corrected by

balancing

Asymmetric shaft 2 £ Run down plot will

show excitation at 1/2

critical speed

Typically occurs on

multistage machines when

all the keyways lie in the

same plane; correct by

staggering them

Shaft crack 1 £ and 2 £ High 1 £ and run down plots

may show excitation at

1/2 critical speed

Confirmation and detection

of location of the crack

may require NDE

techniques

Loose components 1 £ and higher orders plus

fractional subharmonics

High 1 £ with lower level

orders and fractional

subharmonics

Shimming and peening may

be used as temporary

methods to fix the

problem

Coupling lockup 1 £ and 2 £ Equally high 1 £ and 2 £ ,

figure 8 orbits

Stop starts may change

vibration pattern

Thermal instability 1 £ High 1 £ varies with temperature.

Phase angle may change

Proper prewarming or

compromise balancing can

correct the problem

Oil whirl , 1/2 £ , typically 0.35 £

to 0.47 £

Run-up plot will show

1/2 £ increasing and locking

into fixed value , 1/2 £

Temporary problem may be

caused by excess clearances,

oil viscosity, or unloading

of the bearing; if

it is a design

problem, correct by

changing to tilting pad

bearings

Internal rubs 1/4 £ , 1/3 £ , 1/2 £ , 2 £ ,

3 £ , 4 £ , etc.

Run down plots may

show decreasing

amplitudes and

disappearance;

loops in orbits

May get progressively

worse; galling between

contact surfaces or heat

build-up may cause seizure

and shaft failure

Trapped fluids

in rotor

0.8 £ to 0.9 £ Time domain signal will

show beating

Balancing the rotor may

reduce the vibration

Defective rolling

element bearings

At bearing defect frequency Peaks at defect frequencies

in spectrum

Shock pulse measurements

can also be used to

detect problem

Damaged gears Gear mesh frequency High peaks at gear mesh

frequency with side bands.

Time domain may

also show pulses

To determine exact nature

of damage further analysis

may be required

34-50 Vibration and Shock Handbook

© 2005 by Taylor & Francis Group, LLC

direct reduction of the exciting force or by means of increased damping. Based on these guidelines, the

following modifications were proposed to correct the problem:

1. Torsional resonance

* Use an electrometric type coupling that has a high degree of torsional damping to reduce the

magnitude of the torsional excitation forces such that the torsional stresses within the rotors are

within acceptable limits.

* Since both the pump and VFD generate excitation at 6 £ rpm, their effects could be

compounding one another. Introducing either five vane or seven vane impellers into the pump

will eliminate this possibility.

* Additional filters could be introduced into the VFD to reduce the 6 £ and 12 £ component

periodic torsional excitation.

* Consider not operating (lock out) the pump within ^10% of the frequency at which torsional

resonance occurs.

2. Motor housing resonance at 2 £

* Although the 2 £ vibration is dominant, its magnitude is within tolerable levels. The fact that

some 2 £ vibration is also present in the pump indicates that the 2 £ vibration is perhaps

caused by misalignment between the pump and the motor. This can be corrected by proper

TABLE 34.8 (continued)

Cause Dominant Frequency Spectrum, Time Domain,

Orbit Shape

Characteristics,

Corrections, Comments

Electric motor

problems

1 £ (line frequency),

2 £ (line frequency)

High peaks at 1 £ and 2 £ line

frequency with side bands;

disappears when power

to motor is turned off

In the case of two pole motors,

it can be confused with

mechanical causes as the

rotational speed is the

same as line frequency

Casing distortion 1 £ High 1 £ , may change with

time

Caused by high nozzle

loads, casing not free

to expand, soft foot

or foundation distortion

Piping forces 1£ , 2 £ Equally high 1 £ and 2 £ Causes misalignment

between bearings or

between coupled

equipment

Rotor and bearing

critical

1 £ High 1 £ , on rundown plot

1 £ decreases rapidly, may

also show a large phase

angle change

More common in machines

originally designed for

fixed speed operation, later

converted to variable speed

operation

Structural resonance 1 £ , 2 £ High 1 £ and some 2 £;can be

easily identified on run down

plot

Increase or decrease stiffness

of structure or add

or remove mass to

change natural frequency

Rotor hysteresis 0.65 £ to 0.85 £ Spectrum will show high

magnitudes at 0.65 £

to 0.85 £

Occurs in built up rotors

with transitional fits

Hydraulic causes 1 £ (vane pass frequency),

2 £ (vane pass frequency)

High 1 £ and 2 £ vane

pass frequency

Common in centrifugal

pumps due to flow

recirculation or inadequate

gap between impeller and

casing

Vibration in Rotating Machinery 34-51

© 2005 by Taylor & Francis Group, LLC

FIGURE 34.15 (a) Torsional resonance run-up and run-down plots; (b) torsional resonance run-down plot.

(Source: Private communique, Insight Engineering Services Ltd. Alta., Canada. With permission.)

34-52 Vibration and Shock Handbook

© 2005 by Taylor & Francis Group, LLC

FIGURE 34.16 Motor frame resonance. (Source: Private communique, Insight Engineering Services Ltd., Alta.,

Canada. With permission.)

FIGURE 34.17 Pump bearing housing resonance. (Source: Private communique, Insight Engineering Services Ltd.,

Alta., Canada. With permission.)

Vibration in Rotating Machinery 34-53

© 2005 by Taylor & Francis Group, LLC

alignment and thus reducing the 2 £ excitation forces. In some cases, due to an unequal air gap

between the rotor and stator of the motor, the motor could generate the 2 £ vibration. Under

such conditions, accurate centering of the motor bearings will generally correct the problem.

3. Pump vibrations at vane passing frequency

* Generally, high vibrations at vane passing frequency are caused by pressure pulsations

generated at the discharge of the impeller. There are several hydraulic modifications that

can be made to the pump to reduce the amplitude of these pulsations that occur at vane

passing frequency. The most common method is to increase the gap between the impeller

discharge vanes and diffuser/volute. Also, changing the ratio of the number of impeller

vanes to diffuser/volute vanes can help in reducing vane passing frequency pressure

pulsations and the resulting vibration.

4. Acoustic resonance in the pump cross-over pipe

* Once the pump is constructed, it is not possible to change the acoustic natural frequency

of the cross-over pipe. However, the excitation force, pressure pulsations generated at the

impeller discharge, can be reduced by the methods outlined above.

FIGURE 34.18 Pump cross-over pipe acoustic resonance. (Source: Private communique, Insight Engineering

Services Ltd., Alta., Canada. With permission.)

* The root cause of a vibration problem in a rotating machine can be determined by careful

study and analysis of the vibration signals.

* Industrial and international vibration standards and guidelines have been developed to

ensure safe and reliable operation of rotating machinery.

* Equipment manufacturers, users, insurance companies, and public interest groups use

vibration standards to control vibration to within acceptable levels.

34-54 Vibration and Shock Handbook

© 2005 by Taylor & Francis Group, LLC

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