13.1

CHAPTER 13 CENTRIFUGAL MACHINES: PUMPS, FANS, BLOWERS AND COMPRESSORS

The purpose of the pump and fan is to transport fluids by converting mechanical work

into energy of the fluid in the form of pressure and velocity. The compressor is used to increase

the energy of the compressed fluid in the form of pressure. All units are driven by diesel en-

gines, electric motors or turbines (gas and steam). These units are divided into radial flow (cen-

trifugal) and axial flow types depending on the motion of the flow as it passes through the impel-

ler. In a pump the working fluid is a liquid, whereas in fans and compressors the working fluid is

a gas. Fans are distinguished from compressors by the density change (compression) in the

moving fluid induced by the compressor. The fluid moved by a fan incurs little compression. All

centrifugal and axial machines have performance characteristics that relate the head (pressure)

as well as the efficiency and horsepower to the fluid flow rate. Figure 13.1 shows characteristic

curves for centrifugal machines with different blade curvatures. For stable, efficient operation,

operation must be on the negative slope of the curve or the result will be unstable flow causing

excessive hydraulically and aerodynamically induced vibrations.

Figure 13.1. Characteristic Curves for Centrifugal Machines

Stable Operation

13.2

PUMPS

PUMP DESIGN AND FUNCTION

The two types of pumps are centrifugal and axial (13.1). The centrifugal pump consists

of a rotating element (shaft and impeller) and a stationary element (casing, bearings and stuffing

boxes). The centrifugal pump (Figure 13.2) uses radial centrifugal action to force the flow from

the inlet (suction) to the outlet (diffuser or volute) at a higher pressure and velocity. The centrif-

ugal pump can be oriented horizontal or vertical. To create a larger discharge head (pressure

and velocity), multistage pumps (Figure 13.2c) are used. The velocity energy is converted to

pressure by either volutes (Figure 13.2a) or stationary diffusers (Figure 13.2b). The volute is an

ever widening spiral from the cutwater to the discharge opening. The axial pump — sometimes

called a turbine pump (Figure 13.2d), has the suction and discharge parallel to the impellers’

axis of rotation. The flow is both axial and rotational due to propeller like action. Diffuser vanes

(stationary) are used to straighten the flow. However, it is common to prerotate the flow with

diffusers prior to entering the impellers. This type impeller is common in vertical pumps.

Figure 13.2a. Volute Pump Figure 13.2b. Diffuser Pump

13.3

Figure 13.2c. Multistage Pump (Courtesy Worthington Pumps Inc.)

Figure 13.2d. Turbine Multistage Pump

GENERAL PERFORMANCE CHARACTERISTICS

Important operating characteristics of pumps include:

Capacity, Q gallons per minute (gpm)

Head, H feet (ft.)

Power, horsepower (HP)

Efficiency,

Speed, RPM

Impeller diameter, D (feet)

Specific weight, (lb/ft3)

13.4

Average Fluid velocity, sec/ftA

QV

Elevation, Z (ft above datum) ft

Static pressure, p (lb/ft2)

Acceleration due to gravity, g 32.174 ft/sec2

The total pump head, H represents the net work done on a unit weight of fluid in passing from

the inlet or suction, s flange to the discharge flange d. It is given by Bernouli’s equation

s

2

d

2

Zg2

VpZ

g2

VpH

Unlike fans, the fluids that pumps transport are non-compressible — resulting in poten-

tially large interactive forces being transmitted between the rotating and stationary components

(Figure 10.2). In addition, under certain conditions, the liquid can vaporize, then collapse back

into the liquid state causing shock waves that can destroy the impeller of the pump. The pres-

ence of abnormal interactive forces and the above described problem called “cavitation” are a

function of where the pump is operating relative to design conditions (Figure 13.3).

The terms

p, called the pressure head, represents the work required to move a unit

weight of fluid in the flow direction, V against the pressure p. g2

V 2

is the velocity head and rep-

resents the kinetic energy of a unit weight of fluid moving with velocity, V. The term Z, called

the elevation head represents the potential energy of a unit weight of fluid with respect to a ref-

erence.

The horsepower required for the pump is

3960

).gr.sp(QHHP

Q = gpm

H = ft

The efficiency is the horsepower HP divided by the power input to the pump shaft.

13.5

The specific speed is given by

43s

H

QNN

where: Q = gpm

H = ft.

N = RPM

The specific speed is used to judge efficiency.

PUMP OPERATION

Figure 13.3 shows a typical pump curve that is a plot of total pump head, efficiency, and

power against flow. The curve shows the best efficiency point (BEP) in terms of flow, head, and

power components (Figure 13.3. These interactive forces can be minimized by proper selection

of the number of vanes and diffusers (Figure 13.3). Under certain operating conditions, when

the local fluid pressure is below the vapor pressure of the fluid, bubbles and/or vapor filled cavi-

ties can develop. These bubbles will collapse and cause pulse like random forces on the pump

casing when higher pressure is encountered. In turn, these interactive forces cause damage

and noise (like stones in the fluid) to the impeller and casing. This phenomena is called cavita-

tion and is encountered when the net positive suction head, NPSH, is below the liquid vapor

pressure. Net positive suction head is determined in an installation by considering the atmos-

pheric pressure pa, gage suction pressure, ps, vapor pressure, pop, average fluid velocity, V and

static head Zps.

g2

VZ

pppNPSH

2

pspsa

Therefore cavitation occurs as a function of where the pump is operating relative to design con-

ditions (Figure 13.4b).

13.6

Figure 13.3. Pump Characteristics for Backward-Curved Vanes

At best efficiency design point, fluid discharge angle matches angle of diffuser and flow is

smooth with minimal disturbances. If flow is decreased (too much back pressure) or is in-

creased (too little back pressure), the fluid angle no longer matches diffuser angle resulting in

recirculation and higher vibration and a loss in efficiency. The dynamic head exceeds the cen-

trifugal head generated by the impeller. As back pressure is decreased, flow through pump in-

creases and fluid discharge angle increases causing cavitation.

Both phenomena give symptoms of noisy and erratic operation as well as damage to impeller

vanes, pump casings and bearings because of eddies and turbulence

13.7

Figure 13.4a. Flow Path Characteristics (after Baxter13.2)

Figure 13.4b. Pump Head versus Flow Curve (After Baxter13.2)

It is therefore apparent that the vibration level that is measured on a pump is highly de-

pendent upon the operating conditions. Important parameters to monitor are:

1. Back (discharge) pressure

2. Suction (inlet) pressure

3. Fluid temperature

4. Speed

13.8

PUMP FAULT ANALYSIS

A commonly occurring problem with pumps that is often present on vertical pumps is

called rocking mode resonance of the casing. This problem is the result of the first reed (canti-

lever design) natural frequency of the vertical pump matching the pump’s running speed, result-

ing in high vibration at the running speed frequency of the pump. This problem can easily be

confirmed with a resonance test – Chapter 6. Table 13.1 contains some common problems

unique to pumps.

Table 13.1. Pump Fault Analysis*

FAULT FREQUENCY SPECTRUM/

TIME DOMAIN CORRECTION EXAMPLE

Recirculation 1x, vane pass noisy time waveform and spectrum — elevated baseline

increase flow through pump

13.6

Cavitation 1x, vane pass noisy data and acoustics

increase head on the pump

13.7

Incorrect Assembly vane pass elevated vane pass frequency

reassemble casing 13.10

Structural Resonance

1x, 2x or vane pass focused energy at resonant frequency

alter natural frequency

13.14

Shaft Resonance (Critical Speeds)

1x, 2x or vane pass focused energy at shaft critical speed

alter natural fre-quency or fine tune balance

--

Impeller Deflection 1x focused energy at 1x

increase size of shaft or bearing stiffness

13.19

TorsionaL Resonance

vane pass, resonance

focused energy at vane pass fre-quency

change natural or forcing frequency

13.21

Trapped Foreign Material

1x increased vibration at 1x

unclog pump 13.23

Excessive Wear Ring Clearance

1x, critical speed focused energy at 1x due to critical speed

return internal components to specification

--

Impeller-Diffuser Gaps

1x, vane pass vibration at 1x, vane pass, and random noise

alter gaps to opti-mize efficiency and vibration

--

Improper Inlet Conditions

1x, random noise noisy time and frequency

change inlet to in-crease straight section length — 10 pipe diameters

--

Piping Structural Resonance

1x, 2x, vane pass focused energy change natural or forcing frequencies

--

Acoustical Resonance (Piping)

1x, or vane pass focused energy at natural frequency

Change piping path or vane pass fre-quency

--

Foundations 1x, 2x, vane pass 1x and orders improve foundation design or construc-tion

--

*Faults other than common problems such as mass unbalance, misalignment, looseness or rolling ele-ment bearing faults.

13.9

Recirculation

Recirculation (Figure 13.4) is an excellent example of what can happen to a pump that

operates against too much back pressure. For a fixed speed pump there is only one back pres-

sure for which the flow angle of the fluid coming off the impeller matches the diffuser angle.

Operation at any other point can result in inefficient operation and excessive vibration. A pump

can experience inlet or discharge recirculation resulting in a loud crackling noise around the

pump suction.

Case history on recirculation. In a pump undergoing recirculation, antifriction bearings

were failing at 6 week intervals on a horizontal split case pump, 2250 gpm capacity at 300 ft. of

total developed head. Visual observation of the pump showed that the rotor was moving in the

axial direction at a low frequency. A random crackling noise was present.

A head curve (Figure 13.5) for the pump was requested to determine where the pump

was operating relative to its best efficiency point. Comparison of the discharge pressure from a

gauge showed that the pump was being operated at 150 psi back pressure versus 125 psi re-

quired causing a low flow far to the left hand side of the curve.

The bypass orifice was stamped indicating that it has a 2 inch opening. The pump called

for a 3 inch orifice to ensure the correct minimum flow. It was therefore recommended that the

orifice be replaced. When the orifice was removed, it was found that the hole was actually only

1 inch in diameter. Since the pump was forced to operate as low as 800 gpm, the recirculation

line needed to carry 1450 gpm even though the orifice restricted it to 150 gpm.

Pumps that are forced to operate at drastically reduced flow have pressure that builds up

on one side of the rotor, then the other, due to recirculation. This results in slowly oscillating

axial forces and vibration (Figure 13.6). This can cause rapid failure of anti-friction bearings that

were not designed to take the extra axial loading. It is recommended that all pumps which have

axial shuttling of the rotor be checked to determine if they are operating against excessive back

pressure. The pump in the above case was filling a tank several floors above. When the tank

reached a predetermined level, a control valve closed. This meant that the only outlet for the

pump was the recirculation line. Since the orifice in the recirculation line was too small, the

pump operated against too much head. The piping for the recirculation line had a flow orifice to

restrict flow to 150 gpm. This was removed, and the isolation valve in the line was used to set

the recirculation line flow at a better point until the system could be redesigned. Review of the

design calculations indicated that a flow control valve was needed in the recirculation line to

regulate the flow allowing the pump to run closer to its design point.

13.10

Figure 13.5. Pump Curve (13.2)

Figure 13.6. Vibration on 3BCCW Inboard Pump, Axial

2400 2800

13.11

Cavitation

Cavitation occurs when a pump is operated with insufficient back pressure causing large

flow and a steady low amplitude crackling noise. During baseline vibration monitoring, high vi-

bration levels were discovered on the circulating water pumps at a utility. The vibration was de-

tected in both the horizontal direction on the inboard motor bearing and in the axial direction of

the outboard motor bearing. The spectrum was broad band in nature with no mechanically re-

lated identifiable frequencies being observed. Figure 13.7 shows the spectrum of the vibration

on the inboard motor bearing. Mechanically related frequencies can be separated from flow

noise by synchronous time averaging.

Figure 13.7. Pump Vibration Spectrum

(Courtesy of Nelson Baxter13.2)

Investigation showed that some of the pumps were only operating against 10 ft. of back

pressure. This seemed quite low so a copy of the pump head capacity curve was obtained

(Figure 13.8). The design point of the circulating water pumps was 156,000 gpm at 38 ft. of

head. The head flow curve ended at 15 ft. of back pressure indicating that operating with only

10 ft. of back pressure was not even considered by the manufacturer. An estimated flow of

200,000 gallons/minute was obtained by projecting the head capacity curve to the 10 ft. dis-

charge pressure level.

Vane Pass

13.12

Figure 13.8. Pump Curves (Baxter13.2)

To verify this theory, the discharge valves on the condenser were partially closed to in-

crease the back pressure to a level nearer to the design point. When the valves were partially

closed, the vibration dropped off to an acceptable level.

When one of the circulating water pumps was removed for repair, it was found that there

was serious damage to the suction bell. Based upon the above, it was theorized that this dam-

age was due to cavitation.

A review of the above problem found that the low discharge condition was the result of

operating with only one pump in the header rather than two. However, two pumps produced too

much flow. This mode of operation occurred when the cooling water temperature was low

enough to allow one pump to supply enough water to the condenser to satisfy the back pressure

requirements of the turbine. The unfortunate result was the one pump in operation would cavi-

tate because the discharge head was too low.

13.13

Improper Pump Assembly

Operations complained that vibration was high on a 3600 rpm split case ash sluice pump

motor. Analysis on the pump and motor confirmed this observation. The motor and pump vi-

bration signatures (Figures 13.9 and 13.10) clearly showed that the predominant frequency was

at five times running speed. This frequency was found to be the pump blade pass frequency.

Therefore, the problem was not in the motor but was coming from the pump. The motor 1x vi-

bration was only .1 inches per second.

Figure 13.9. Sluice Pump Motor Vibrations (Courtesy of Nelson Baxter13.2)

Figure 13.10. Sluice Pump Vibrations (Courtesy of Nelson Baxter13.2)

13.14

Since the pump blade pass frequency was at a level of .661 inches per second, it was

recommended that the pump be disassembled. It was found that the casing had been offset

during reassembly after an overhaul, causing the high blade pass frequency.

Structural Resonance

Vertical pumps are known for their casing structural resonance problems — the natural

frequency of the unit, foundations, and piping is equal to pump speed or vane pass frequency.

Shafting critical speeds can also be a problem — the natural frequency of the shaft on its bear-

ings is equal to the operating speed or vane pass frequency.

This case history on a vertical firewater pump concerns the excessive vibrations the

pump casing, Figure 13.11. From steady state and transient test data, it was determined that a

severe resonance problem existed (natural frequencies on both flanks of the operating speed

(Figure 13.12). A computer model was used to determine means of altering natural frequencies

through addition of internal and external stiffening.

Figure 13.11. Schematic Design – Fire Water Pump

H

A

13.15

Figure 13.12. Impact Test Data

Using all available test data, Figure 13.13 shows the operating deflection shape of the

pump housing for once-per-rev pump vibration (1750 cpm) plotted in velocity (in./sec.) versus

pump elevation. The data show an operating deflection shape with high vibration at the suction

and drive ends of the pump with restraint where it is attached to the floor. Data taken at the

gearbox input shaft housing in the north/south direction, Figure 13.14, show predominately

pump generated vibration. However, the 6x engine vibration level is above good practice — 6x

is a major engine order. Figure 13.15 taken in the north/south direction at the top of the gear-

box, shows the severity of the problem — 1.64 in./sec. rms (17.9 mils) at pump rotational speed.

A computer model of the gearbox and pump was constructed to determine the effect of

various alterations on the vibration characteristics of the pump. The effect of changing mass

and/or stiffness of the various components or adding stiffeners at various locations was studied.

13.16

Figure 13.13. Pump Operating Deflection Shapes

13.17

Figure 13.14. Horizontal at Gearbox Split Case – North/South

Figure 13.15. Horizontal at Top of Gearbox – North/South

13.18

The model shown in Figure 13.16 was formed from the best available information. The

pump model was verified with test data. Impact tests were used to simulate the natural fre-

quencies, and operating data were used to simulate the forces on the pump and to establish an

operating mode shape. This model was then modified in various ways to obtain the modifica-

tions to the pump that would have the best chance of solving the problem.

Figure 13.16. Pump Model

13.19

The cause of the excessive vibration on the fire water pump as shown in Figure 13.12 is

resonance. Natural frequencies exist at 1545 cpm and 1830 cpm with the pump operating at

1750 rpm. The test data and model show that the principal vibration at operating speed is am-

plified by the 1830 cpm resonance. Thus, lowering the operating speed of the pump would im-

prove the vibration levels.

Means of altering these natural frequencies to eliminate the excessive vibrations were

studied using the computer simulation model. Also, the use of a dynamic vibration absorber

was evaluated. It would appear that the addition of mass to lower the 1830 cpm natural fre-

quency would be the easiest solution. The following actions were considered.

1. Lower the pump speed 60 rpm – move forcing frequency.

2. Insert additional bolts or springs between the spacer and/or the pump and between the spacer and the foundation.

3. Add stiffeners at the gearbox and at the water line in the sump.

4. Add mass to the gearbox.

5. Add a dynamic vibration absorber at the top of the gearbox.

The valley between the two natural frequencies shown in Figure 13.12 indicates that

lowering the pump speed by 60 rpm would eliminate some of the amplification of the resonance

by the 1830 cpm natural frequency. However, the forcing frequency would be 9¼% away from

the natural frequencies on each side. This would require good speed control.

The test data and the model suggest flexibility being introduced by the bolted joints at

the spacer would lower the higher natural frequency. This modification by the pump manufac-

turer created an unstable mounting. Additional rigidity would increase the 1830 cpm natural fre-

quency; however, it may also increase the 1545 cpm natural frequency. Probably impossible

because stiffening could occur in the north/south direction.

It appears that stiffeners added between the gearbox base and the concrete foundation

would raise the upper natural frequency but not adversely affect the lower natural frequency in

the east-west direction. The addition of stiffeners at the water line would raise both natural fre-

quencies above the pump operating speed.

The addition of weight to the gearbox will lower both natural frequencies; however, the

added weight must not overstress the pump casing. This was the final fix – 785 lb. Added to the

gearbox split line.

13.20

A dynamic absorber, Figure 13.17, was designed for the pump. The absorber is de-

signed to vibrate while the pump stands still. This fix is dependent on the availability of a suita-

ble attachment point on the top of the gearbox. The dynamic vibration absorber is a mass-

spring system with its natural frequency equal to the pump operating speed. A series of plates

(mass) are adjusted to a position on the threaded pipe (spring) where the natural frequency of

the mass-spring system is equal to the pump speed.

Figure 13.17. Dynamic Vibration Absorber

Shaft Resonance (Critical Speeds)

The motor-pump shaft will have natural frequencies that depend on the shaft sizes and

lengths, couplings, bearings, and casings. These natural frequencies can be determined by an

impact test if the shaft is supported on rolling element bearings. Otherwise natural frequencies

are determined by transient testing (Chapter 6) or computation. Critical speeds are excited by

either pump speed or van pass frequency.

Excessive Pump Shaft Deflection

A lowest bidder supplied centrifugal pump (Figure 13.18) was placed into service to

pump water against 300 ft. of head. It was determined that the head on the pump caused up to

20 mils deflection of the pump shaft. It was anticipated that this excessive deflection would

cause premature bearing failures. Figure 13.19 shows vibration data from the pump that had

retrofitted proximity probes. The data show 4 mils pk to pk vibration; however, a portion of that

response is due to a rough shaft at the probe locations. For a three inch diameter shaft —

roughly 5 to 6 mils clearance, this vibration level is excessive.

13.21

Figure 13.18. Double Suction Pump

Figure 13.19. Proximity Probe Data

13.22

Torsional Vibrations

A torsional vibration analysis was conducted on a vertically mounted horizontal sewage

pump (Figure 13.20) because of a series of motor shaft failures. The 300 hp induction motors

drive 24 x 18 type SSEV pumps through a flywheel. The motor speed (400 rpm to 705 rpm)

was controlled by an adjustable speed pulse width modulated type inverter.

Torsional vibration tests (Figure

13.21) were conducted on a motor over its

operating speed range. These data showed

the presence of excessive torsional vibrations

at speeds between 400 rpm and 525 rpm. It

was found from a computer model developed

for the unit that a condition of torsional reso-

nance exists. It is excited in a range of

speeds by the interference of the pulse width

modulation (PWM) motor excitation at a fre-

quency of 5 times motor speed and the first

mode torsional natural frequency at 38.8 Hz

(Figure 13.21). The computer model, vali-

dated by the test data, showed excessive

stresses sufficient to cause fatigue failure of

the motor shaft in its keyway.

Subsequent design studies were con-

ducted using the computer model to deter-

mine how to lower the system torsional vibra-

tions. It was determined that a soft coupling

of the Holset type would lower the natural

frequency of the system enough to remove the resonant conditions from the pump operating

speed range. The addition of high damping in the coupling will remove the concern of passing

through the system first mode natural frequency during startup and coast down. Further, the

high damping will provide protection from torsional shocks and/or low frequency pulse excita-

tion.

Figure 13.20. Pump Configuration

13.23

Figure 13.21. Torsional Peak Hold Test A Koppers Max-C flex-rigid size 4.5 coupling, type WB with SBR high damping blocks

was evaluated within a modified motor (4” diameter shaft)-flywheel-pump configuration. It was

found to have acceptable torsional vibration characteristics.

Foreign Objects

A predictive maintenance engineering assis-

tant noticed abnormal noise coming from an over-

hung high-pressure service water pump (Figure

13.22). The pump delivers untreated lake water at

2,500 gpm (1,780 rpm) (Motor 350 hp).

Aside from the abnormal noise and above

average horizontal pump vibration, all looked good.

Horizontal pump vibration was .40 in./sec. on the

inboard pump bearing and .60 in./sec. on the out-

board pump bearing. This is about twice what the

normal vibration.

Figure 13.22. High-Pressure

Service Water Pump

38.8

042.

6.1133

328Q

13.24

Vibration signatures on the motor showed little vibration. Pump vibration signatures

showed all vibration was the 1x running speed frequency (Figure 13.23). Since there was no

axial or vertical vibration, this indicated some type of balance problem.

Figure 13.23. Pump Vibration

In a similar case history with identical vibration signatures, a piece of metal was lodged

in the impeller. Therefore, it was recommended that the impeller be inspected for foreign ob-

jects. The impeller was removed for inspection and a four inch (4”) piece of round stock was

found in the bottom of the pump casing. This piece of metal fit snugly into a hole in the bottom

of the impeller. The pump was put back together and test run. Vibration levels dropped to nor-

mal limits.

Excessive Wear Ring Clearance

The wear rings in a multistage pump, Figure 13.2c, that prohibit flow from a high pres-

sure stage to a low pressure stage are a seal similar to the plain journal bearing. Even though

the wear ring has greater clearance than the normal bearing they act like bearings. These

“bearings” help to support the multistage pump shaft and raise the natural frequencies. For this

reason, excessive wear ring clearance can cause critical speed problems by lowering the sys-

tem stiffness.

13.25

Impeller-Diffuser (Volute) Gaps/Blade Pass

Figure 13.24 shows the gap of a single stage pump. The size of the gap presents a

tradeoff between high efficiency and excessive blade pass vibration. Also, the number of vanes

on the impeller and diffuser determine what magnitude of blade pass vibration measured on a

pump13.3.

Figure 13.24. Single Stage Pump

Acoustic Resonance and Poor Inlet Conditions

A vertical pump similar in design to that shown in Figure 13.24 had excessive vibration

at the drive bearing. Structural resonance was the suspected reason for the elevated vibration

at 3 times operating speed (vane pass frequency). However, after elaborate impact testing no

natural frequency could be identified. Figure 13.25 shows a transient test results from double

integrated accelerometers and a pressure transducer. The figure shows peaks in the structural

vibrations and pressure pulsations — indicating acoustic and structural natural frequencies. Af-

13.26

ter a thorough investigation, it was found that the excessive vibration came from a combination

of resonances as well as high excitation due to poor inlet conditions. The peak vibration meas-

ured in the pump operating speed range was .37 ips. The acoustic resonance was in the dis-

charge piping which had a length of approximately 130 ft. to the discharge header.

Therefore

VS = 4850 ft./sec. Water: Acoustic Velocity

L = 130 ft.

Fn = Hz65.18130x2

4850

Solutions to solve this problem include changing the natural frequencies, tuned absorb-

ers (13.5), and reduction to excitation.

Figure 13.25. Vibration and Pulsation at Vane Pass Frequency

f nVs

L

2

Speed 270-390 RPM

13.27

FANS AND BLOWERS

Many centrifugal fans use a volute or scroll type casing – the flow enters axially and

leaves tangentially. Blading may be fixed or adjustable (sometimes during operation). A typical

fan performance characteristics is shown in Figure 13.26. The basic curve is fan pressure ver-

sus flow through the system – head or pressure varies as the square of the flow. The fan will

operate satisfactorily at the intersection of the system characteristics and the fan pressure char-

acteristic (Figure 13.26). The system characteristic can be changed with an outlet damper con-

trol. Variable vane, pitch, and speed controls alter the fan characteristics. Characteristics of

fans mounted in series and parallel must be considered as a system to avoid excessive vibra-

tion and premature bearing failure.

Figure 13.26. Fan Characteristic – Constant Speed

To ensure stable operation, the slopes of the pressure-flow curves of the fan and system

should be of opposite sign (Figure 13.26). When the slopes of the fan and system characteris-

tics are of opposite sign, any system disturbance tending to produce a temporary decreased in

flow is mollified by the increase in fan pressure. The condition which accompanies unsteady

flow is pulsation that occurs when the operating point of the fan is to the left of the maximum

pressure on the fan curve (the surge point). Inlet dampers normally can be used to position the

fan operation to the right of the surge point. At low capacities, flow reversal or puffing can oc-

cur. Air puffs in and out of the let. If the back pressure is greater than the fan discharge pres-

sure flow reversal will occur. Flow separation in the blade passages of the impeller can cause

unsteady flow and vibration.

13.28

Fans can be plagued by impeller eccentricity, asymmetric supports, loose fan wheels

and structural flaws. Improper isolation mounts often present a problem.

In addition to flow noise generation by size and directional changes in inlet ducting,

acoustic resonance can be set up when a fan vane pass frequency matches the acoustic natu-

ral frequency of the air in the duct work. Fans are subject to critical speed and structural reso-

nance problems because of the nature of their mounting on skids, isolators, and flexible frames.

Ducting and casing resonances can be excited by high frequency variable frequency drive in-

duced excitations.

Table 13.2 shows a number of common fan faults — some of which will be illustrated

with brief case histories.

COMMON MECHANICAL PROBLEMS

Mass unbalance (Figure 13.27) and blade pass (Figure 13.28) are two common fan

problems. Balancing is frequently done with the isolators blocked. However, it is better to not

block the isolators because the goal is to minimize the vibration symptom. It should be noted

that balancing of a resonant fan is difficult and should be avoided when possible. It is best to

deal with the natural frequency prior to attempting balancing. Isolators do deteriorate with time.

If the fan has recently shown high vibration, it may be resonant due to isolator property change.

If signs of looseness are shown in the data, figure 13.29, the isolators may be worn out. Exces-

sive blade pass vibration can come from damper position, resonance, or ducting characteristics.

Fan case histories follow.

13.29

Table 13.2. Fan and Blower Fault Analysis

FAULT FREQUENCY SPECTRUM/TIME CORRECTION

Critical Speeds 1x, BP* Focused energy at fre-quency of critical speed

Change natural fre-quency or operating speed

Resonance, Structural 1x, BP Focused energy at fre-quency of resonance

Change natural fre-quency or operating speed

Resonance, Acoustic 1x, BP Focused high vibration at resonant frequency

Change ducting, fan speed, or no blades

Aerodynamic BP or Random Noise

High amplitude BP or high noise flow

Change damper posi-tion or redesign duct-ing

Isolators 1x and Orders Impact like waveform and spectrum or fo-cused energy

Replace isolators if worn out – redesign if resonant

Impeller Eccentricity 1x High amplitude 1x Replace impeller

Impeller Cracks 1x, 2x Difficult to balance – new lower critical speed, high amplitude 1x critical speed at ½ x non repeatable data

Repair wheel

Rubs Fractional Frequen-cy

1x and Orders

Subharmonics or or-ders of operating speed

Change clearance or reduce source of rub excitation

Surge Natural Frequencies Low frequency pulse with ring down

Change operation to get on fan curve – avoid flow reversal

Belts Belt Frequency and Orders

Vertical and horizontal in phase. Pulses in time waveform when defect passes pulley

Replace belts

Eccentric Pulleys 1x Directional 1x phase in horizontal

Replace pulley

Asymmetric Pedestal Flexibility

1x High amplitude in flex-ible direction

Balance using trans-ducer with highest am-plitude

Blower Pulsation No. of Lobes x RPM

High amplitude of pul-sation frequency and multiples

Add pulsation damper in discharge pipe

Loose Fan Wheel 1x and Multiples

Walking phase Repair keys or shrink fits

13.30

Figure 13.27. Fan Mass Unbalance with Induced Blade Pass Frequency

Figure 13.28. Fan Operation Off System Characteristic

13.31

Figure 13.29. Support Looseness

Cooling Unit Resonance

The cooling unit shown schematically in Figure 13.30 was experiencing numerous struc-

tural failures. The unit was well built from stainless steel and adequately balanced. However,

resonances were not considered in its design. The following forcing functions were present in

the system.

mass unbalance and misalignment — 1x

misalignment — 2x

impeller passing — 3x

propeller passing — 4x

Thus forcing frequencies of approximately 20, 40, 60, and 80 Hz were available. The

structural resonance in the unit were eliminated by detuning – using resonance testing for identi-

fication and stiffness for correction.

} Speed of rotating unit — 1200 rpm

13.32

The critical speeds of the slender shaft/propeller unit involved more than simple reso-

nance testing due to overhung propeller. Figure 13.31 shows an interference diagram for this

unit.

It shows the stiffening effect of the gyroscopic moments caused by the whirling propeller.

Bump tests showed the propeller/drive shaft natural frequency at 22 Hz – very near the mass

unbalance frequency, 19.33 Hz. Note that the first mode is actually near 35 Hz due to the large

stiffening effect from the gyroscopic moments. This means that the first mode could be excited

by a 2x vibration excitation rather than the 1x of mass unbalance. For this reason it was im-

portant in this system to avoid any 2x vibration – due to asymmetric support stiffness, heavy

mass unbalance induced nonlinearities, or excessive misalignment. This case shows the ex-

treme dependence of the lateral natural frequencies on the rotor whirling – spin effect (gyro-

scopic moments) when overhung wheels, propeller, etc. are present.

Fig-

ure 13.30.

Port-

able Cool-

ing Tow-

er

13.33

Figure 13.31. Overhung Fan Interference Diagram

Overhung Blower Balance Sensitivity and Resonance

A group of 200 hp, 3600 rpm motor blower units shown in Figure 13.32, were experienc-

ing a balance sensitivity problem; i.e., they required endless small balance corrections to keep

them operating. The subject units consisted of a large overhung fan supported on two external

pillow block bearings which were mounted on a fabricated steel base. The blower shaft was

connected to the motor with a steelflex coupling. These units experienced continued bearing

fail-

ures and

out- ag-

es.

13.34

Figure 13.32. Schematic Diagram of an Overhung Blower

In a typical situation, one unit was balanced to a maximum of .05 in./sec. (vertical).

Several days later the unit had excessive radial vibration. It took only 3.75 oz.-in. to rebalance

the fan. These units required this type attention until the matter was given engineering atten-

tion.

Experimental analysis using ringing and coast down tests established the nonrotating

and rotating first natural frequency frequencies at 1750 cpm and 2200 cpm respectively. The

stiffening of the unit results from the rotation of the large overhung fan. It was obvious from

these tests that the unit was operating slightly under the second natural frequency. Figure

13.33 shows the interference diagram. Operation near the natural frequency causes a balance

sensitivity problem — a large vibration is obtained for a small additional amount of mass unbal-

ance (Figure 13.34). In this case slight flexing of the fan may have been adjusting the mass un-

bal- ance

enough to in-

crease the vi-

bration.

13.35

Figure 13.33. Blower Interference Diagram

Figure 13.34. Blower Bodé Plot

The solution to the problem was to raise the second natural frequency enough to elimi-

nate the balance sensitivity problem but not raise the first natural frequency appreciably. In-

spection of the unit revealed flexibility of the unit in the shaft, bearings, and pedestal. It was de-

cided that the fabricated steel base would be the easiest element to change. To raise the natu-

ral frequency it had to be stiffened. The fabricated base was turned upside down — steel re-

bars were welded in the base and it was filled with concrete. This action raised the natural fre-

quencies enough to eliminate the balance sensitivity problem (Figure 13.34).

Critical Speeds and Balance Sensitivity of Overhung Fans

Many severe vibration problems with overhung fans are attributable to the fact that criti-

cal speeds or support resonances occur close to the desired operating speed of the fan. This

condition can be a result of improper mounting of either the fan on flexible structural members

13.36

or vibration isolators. In some cases the fan itself operates close to a critical speed due to bear-

ing stiffness or the shaft span.

The fact that a fan operates close to a critical speed or a structural resonance creates a

balance sensitivity problem. Any degradation of balance due to wear, corrosion, or uneven

product buildup causes a severe vibration problem. Figure 13.35 shows the amplification of the

vibration response of a system if the operating speed is close to a natural frequency. Amplifica-

tion factors of five or more are not uncommon if isolators or support structures are improperly

mounted.

A small change in condition can

therefore result in large changes in vibration

level. The problem can be solved by moving

a natural frequency of the system up or

down, thereby tuning it. Depending on the

structure and the nature of the fan, move-

ment in either direction will be advantageous.

A support-oriented natural frequency

can be determined by bump or resonance

tests (Chapter 6). A natural frequency that

is shaft or bearing oriented, however, must

be determined by a coast-down test (Chap-

ter 6). In this case the natural frequency

usually varies with speed because of the

gyroscopic moments exerted by the over-

hung fan. The gyroscopic effect tends to ef-

fectively stiffen the shaft and raises the natu-

ral frequency.

Errors in natural frequency as high as 15 percent can occur by bumping a stationary fan.

Figure 13.36 shows an interference chart of a fan operating close to its second natural frequen-

cy. Note that a bump test on the stationary fan would have identified 50 Hz as the natural fre-

quency rather than 60 Hz. The true critical speed is 3600 rpm.

Figure 13.35. Vibration Response to Mass Unbalance

13.37

Figure 13.36. Fan Interference Diagram

Fan Mounting Resonance

The fan-motor unit shown schematically in Figure 13.37 was mounted on isolators

through a flexible steel skid. The 50 HP, two pole motor (3,594 RPM) drove a centrifugal fan

(3,180 RPM) with belts. The excessive vibration (Figure 13.38) of the skid and ducting varied

from .7 ips to 2.0 ips causing structural and bearing failures at regular intervals.

An analysis of the vibration indicated beating between the excessive fan and motor vi-

bration (Figure 13.38) – resonance was suspected. Impact tests were conducted to (a.) deter-

mine if resonance is the problem and (b.) determine mode shapes if resonance is the problem.

Figure 13.39 shows the measurement points on the skid. Figure 13.40 shows sample two

channel data taken at point 4 in the vertical direction. Figure 13.40a shows coherence (valida-

tion of the test) and frequency response – mobility (ips/lb) while Figure 13.40b shows phase and

mobility at the same point in the vertical direction. Natural frequencies were obtained at 48 Hz

and 59.75 Hz. Thus motor speed is directly on a rocking natural frequency and the fan speed is

13.38

10% off a torsional natural frequency (Table 13.3). At other points, the torsional mode was

measured at 52 Hz, which is right on the fan speed. So resonance is the principal problem.

Figure 13.37. Schematic of Fan-Motor, Skid Isolator Mounting

13.39

Figure 13.38. Fan Base Resonance (Improper Isolators)

Figure 13.39. Measurement

Points on Skid

13.40

13.41

Figure 13.40a. Impact Test Measurements – Coherence Top and Mobility Bottom – IPS/lb

Figure 13.40b. Impact Test Measurement – Phase (Top) and Mobility (Bottom)

13.42

Table 13.3. Summary of Natural Frequencies

Measurement Location

Natural Frequencies, Hz

Hammer Vibration 1 2 3 4 5 6 7 8

4V 4V 14 19.5 33.7 48 59.75 90

4V 3V 19.5 40 59.75 90

4V 2V 19.5 39 47 55

4V 1V 19.5 39 55 90

4H 3V 15 42 52 70 90

4H 1H 15 40 51.75 72 90

4H 2V 51

4H 2H 16 40 47 60 72

1A 2A 25 50.5 55 60 72 88

1A 3A 59.5 88

Mode Type Hori- zontal

Hori- zontal

Vert- ical

Tor- sional

Rock Rock Struc- tural

Struc- tural

A skid mounted on isolators will have six (6) rigid body (frame does not deflect) modes –

vertical, horizontal (2), rocking (2), and torsional (twisting). In this case, the skid had higher fre-

quency deflection modes – 72 Hz and 90 Hz. Some frame operating deflection shapes are

found in Figure 13.41.

Recommendations for a fix included

1. New isolators do tune the two resonances.

2. Replace isolators with steel blocks to hard mount the unit.

3. Replace isolators with steel spacers and a layer of damping material.

The owner elected to use number three (3) recommendation.

13.43

Figure 13.41. Frame Operating Deflection Shapes

13.44

Fan Induced Floor Resonance

A complaint was received from a high rise building that floor vibration was causing em-

ployee discomfort. It was noted that the air handling systems on the floor above the affected

area had 50 fans operating at different times. Floor impact tests (Figure 13.42) indicated a nat-

ural frequency of 7.125 Hz. After measuring many fans, the exciter fan (Figure 13.43) was

found to be operating at 427 RPM. Therefore, this fan was exciting the floor natural frequency

below. It was found that one of the original sheaves had been replaced with one of a differing

size – thus changing the fan speed to be tuned to the floor natural frequencies. A new sheave

of original diameter was installed to correct the problem.

Fan Duct Acoustical Vibration

A large ID fan operating at 930 RPM was encountering high 1x vibration in the axial (dis-

courage) direction – Figure 13.44. The vibration (about 1 ips) was causing metal fatigue and

cracks in the discharge ducting (Figure 13.45). Structural impact tests shows no natural fre-

quency at 930 cpm. It was noted on the coast down test that the vibration would disappear al-

most immediately after the power was out. Calculations on the discharge duct length indicated

a possible acoustical natural frequency at fan speed. The fan speed was reduced slightly by

changing the drive sheave thus eliminating the resonance.

Fan — Loose Asymmetric Supports

The FD fan shown schematically in Figure 13.46 was experiencing high vibration at op-

erating speed in the horizontal direction. Figures 13.47 and 13.48 show data that are RMS and

synchronous time averaged respectively. The data show no difference between RMS and syn-

chronous time averaging. This means that all vibration is mechanical related to operating

speed.

Figures 13.47 and 13.48 shows 1x vibration with a long series of orders – indicating

pedestal looseness or excessive bearing clearance. Detailed analysis of the unit’s support

structure confirmed excessive bearing clearance. Figure 13.49 shows 1x filtered data. The ver-

tical vibration (top) is one-third the amplitude of the horizontal vibration (bottom). The 90° phase

difference between horizontal and vertical confirms that this portion (.2 ips) of the 1x (total .21

ips) horizontal vibration was mass unbalance excited. Analysis of the pedestals shows that the

fan was not resonant in the horizontal direction but was structurally flexible. The easiest solu-

tion is to eliminate the looseness and balance the fan to the horizontal plane.

13.45

Fig- ure 13.42 Floor Impact Test – Tape Speed 10x

Fig- ure 13.43. Fan Vibration – Tape Speed 10x

13.46

Figure 13.44. Axial Vibration on the Fan

Figure 13.45. Original Duct Configuration

13.47

Figure 13.46. Fan Configuration

Figure 13.47. RMS Averaged Data

13.48

Figure 13.48. Synchronous Time Averaged Data

Figure 13.49. Fan Inboard-Filtered, Horizontal and Vertical

13.49

DYNAMIC COMPRESSORS

Due to pressures involved, most centrifugal compressors have massive casings and

small lightweight rotors that make seismic measurements difficult. Compressor faults are simi-

lar in nature to those encountered in steam turbines and pumps occurring subsynchronous to

operating speed, at operating speed, or as multiples of operating speed. For fault analysis, see

Chapter 12 on rotor and bearing faults. Compressors have a minimum flow point called the

surge limit. The operation of the machine is unstable below the surge limit. The surge limit is a

function of compressor type, gas properties, inlet temperature, blade angle and speed.

INTRODUCTION

Dynamic compressors develop a pressure differential by the action of rotating blading

that imparts velocity and pressure to the flowing medium.

The compressor is used to increase the energy of a fluid in the form of pressure. Rotat-

ing units are divided into radial flow (centrifugal) and axial flow types depending on the flow path

and the design of the impeller wheels — centrifugal and blading (axial). Compressors normally

can be direct driven by steam and gas turbines. Motors use a gearbox to attain efficient com-

pressor speeds. The larger the volume of air to be handled, the larger must be the diameter of

the impeller. On account of centrifugal stresses, top speeds are limited to available materials.

The performance of a centrifugal compressor can be stated in terms of the volume of

gas, density relative to the air at the same temperature and pressure, and the ratio of the dis-

charge pressure to the inlet pressure. To compare compressors operating under different con-

ditions, the desired performance is converted to an equivalent performance under standard op-

erating conditions – inlet conditions 60° F and 14.4 PSIA. The discharge pressure under this

condition is called equivalent air pressure (EAP).

The flow through a centrifugal compressor is directly proportional to speed; the head is

proportional to speed squared; and horsepower, to speed cubed.

The specific speed is a parameter used to classify compressor impellers on the basis of

their performance and proportions. Specific speed is the speed in RPM at which the impeller

would rotate if reduced proportionately in size to deliver one cubic foot of gas per minute against

a total head of one foot.

13.50

Specific Speed = 4/3s

H

QNN

N = operating speed, RPM

Q = design flow, CPM

H = design head, FT

Flow coefficient and head coefficient are commonly used nondimensional parameters used for

rating an impeller’s performance.

For a multi-stage compressor (compressed about 3-1 per stage), the gas passes

through a crossover and enters the vanes that guide the flow uniformly into the next impeller

inlet. Machines such as the cross section shown in Figure 13.50 are split horizontally at the

centerline for reasons of maintenance. For higher pressures, the casings (Figure 13.51) are

built as complete cylinders (barrel compressor) with the rotor removed axially. The flow in axial

compressors is axial and the compression cycle involves passing gas through alternating rows

of rotating and stationary blading (Figure 13.52).

PERFORMANCE CHARACTERISTICS

The relationship between the inlet volume, flow head, speed, efficiency, and power of a

dynamic compressor are characteristic curves – Figures 13.53 and 13.54 for centrifugal and ax-

ial compressors respectively. These characteristics are developed from the laws of thermody-

namics and properties of gases.

From Figure 13.53 it can be seen that at any constant speed when the system re-

sistance increases or the flow is throttled, flow through the compressor is decreased. By follow-

ing the constant speed curve with lower flows, a peak head is reached that is the surge point

(surge line for multi speeds). Flow instability takes place at this point on the curve causing the

flow to pulsate – operation must be maintained with flow volumes above this level. A point of

maximum flow and minimum head occurs at the other end of the constant speed curve. This is

the choked flow where the compressor impeller cannot accept any more flow volume.

13.51

Figure 13.50. Cast Case Centrifugal Compressor

13.52

Figure 13.51. Cross Section of Barrel Compressor

Figure 13.52. Axial Compressor

13.53

Figure 13.53. Typical Multistage-Centrifugal-Compressor Characteristic

Figure 13.54. Typical Axial-Compressor Characteristic

13.54

Figure 13.55. Centrifugal Compressor Rotor

DESIGN

Since dynamic compressors rotate at high speeds (5000 RPM to 12,000 RPM) to devel-

op pressure by rotating blading that imparts energy to the flowing medium, rotor design is a sig-

nificant aspect of the compressor. The rotor consists of a shaft, shrunk on impellers, a balance

drum, and a thrust collar (Figure 13.55). The rotor for an axial compressor can be a forging or

built-up assembly with individual blades mounted (Figure 13.56). Centrifugal compressor impel-

lers (Figure 13.57) are usually shrink-fitted to the

shaft and stack balanced in the process. Sometimes these

rotors are balanced in a high speed balance facility at operat-

ing speed.

Cen-

trifugal com-

pressors

normally use

tilt pad bear-

ings (Figure

13.58) be-

cause of their

high speed and relative light rotors. Axial

compressors can have any bearing from

plain to cylindrical with preload to pressure dam. The seals on compressors vary according to

design and purpose including labyrinths, carbon rings, brushings, and contact seals (Figure

13.59). The relative leakage for dry sealing is provided in Table 13.4.

Figure 13.56. Typical Axial-Compressor Rotor

Figure 13.57. Impeller

13.55

Figure 13.58. Typical Tilting-Pad Bearing

Table 13.4. Seal Leakage

Seal Leakage Index

Straight Pass labyrinth

Staggered Labyrinth

Segmented Carbon Rings

Dry contact

100

56

20

2

13.56

a. Axial Staggered Labyrinths b. Segmental Carbon Ring

c. Bushing Seal d. Carbon Face-Contact Seal

Figure 13.59. Seals

13.57

MEASUREMENT AND ANALYSIS

Due to the pressure involved, most centrifugal compressors have massive casings and

small lightweight rotors that make seismic measurements difficult. For this reason, it is almost

mandatory that x-y noncontacting displacment probes be mounted on the compressor near the

bearings. Common faults encountered in dynamic compressors are listed in Table 13.5. See

Chapter 12 for additional information.

Table 13.5. Centrifugal Compressor Faults

FAULT SYMPTOM

Excessive Bearing Clearance

High 1x vibration with orders – cannot be balanced – may go unstable and/or change critical speed

Wiped Bearings Same symptoms as above if severity generates excessive

clearance

Rough Journals Excessive noise on probe signal

Bent or Bowed Rotor High 1x vibration – vibration dropout at discrete speed if phased against mass unbalance

Build-up of Product Deposits on Rotor or Casing

High 1x and/or rubs

Unbalanced Rotor High 1x, fixed phase angle

Shaft Misalignment High 2x, in one direction with 1x axial

Dry Gear Coupling Same as misalignment

Worn or Damaged Coupling High 1x

Liquid “Slugging” Random vibration including natural frequencies

Operating in Surge Range: Insufficient Flow

Pressure impact on rotor rings at natural frequencies

Oil Whirl/Whip Excessive vibration at subsynchronous frequencies

Subharmonic Resonance High vibration at exact subharmonic

Gas Flow Buffeting in Shrouded Impellers

Random noise

Piping Pulsation Check piping for resonance between compressor vane pass fre-quency and piping natural frequency

The procedures used to measure the shaft vibration of a centrifugal compressor are rela-

tively similar to those used with a steam turbine. Active thrust is toward the active thrust bear-

ings; some compressor manufacturers counterbalance the aerodynamic thrust to place the rotor

initially on the inactive side of the float zone. As labyrinths and other seals wear, the rotor even-

tually wanders across to the active position.

13.58

The balance piston pressure should be measured either statically in the chamber or ac-

tively in the balance line. Additions to API 617 (1979) allow a flow measurement in the balance

line to permit better assessment of deterioration of the thrust balance. Such deterioration oc-

curs when the balance piston labyrinths wear and more flow is required in the balancing line.

Surge detection devices determine reverse flow and surge and are used to control aero-

dynamic performance. They affect mechanical performance and can cause large increases in

vibration levels. The response to surge on an axial compressor causes much more damage

than that which occurs on a centrifugal compressor.

High Speed Centrifugal Compressors

High speed centrifugal compressors provide special challenges in measrement and

analysis. These units (Figure 13.60) are usually motor driven through a speed increaser with

speeds as high as 70,000 RPM. With light speed rotors and heavy casings this yields special

measurement and analysis problems. Units could have rolling element and/or fluid film bear-

ings. Many of these machines have permanently mounted proximity probes. The noncontact-

ing displacement probes can be used to evaluate rotor-bearing faults and condition: however,

accelerometers must be used to evaluate gear mesh frequencies which could be as high as 25K

Hz.

The accelerometer should be located as close to each stage bearing as possible in the

axial direction. Since the acceleration range could approach 50 kHz, care must be taken to se-

lect appropriate accelerometers, which have a calibrated frequency range that will cover two to

three times gear meshing frequency and ten (10) times the ball pass frequency of the inner

race.

Table 13.6. High Speed Centrifugal Compressor Faults

Mass unbalance

Gear meshing

Gear breakage (cracked, broken, chipped)

Distortion

Rolling element faults

Fluid film bearing faults

Blading/diffusers*

Alignment

Oil pump hydraulics

13.59

* Blade Pass Frequency = RPMxNb

Blade Rate Frequency (BRF) = RPMxK

NxN db

where

Nb = number of vanes on impeller

Nd = number of diffuser vanes

K = highest common factor of Nb and Nd

BRF quantifies pulse rate due to compression of air between rotating and stationary

vanes.

Figure 13.60. Compressor-Gearbox Assembly

13.60

Monitoring and Analysis

Because of the high speeds of these units, monitoring overall levels is not recommend-

ed except for the permanently mounted proximity probes. Common faults are provided in Table

13.6.

It is recommended that spectrum and time waveforms be examined on a regular basis.

For added sensitivity, it may be prudent to use a log display in the spectrum so the emerging

sidebands can be identified early. Figures 13.61 and 13.62 compare linear and log plots from a

relatively new gearbox (gearmesh = 742.5 Hz, input speed = 1,790 RPM). For particular com-

pressors, the dB range between the center frequency (gearmesh or rolling element bearing

fault), should relate to the condition. For example, if the gearmesh is 2 gs and the pinion speed

sideband is .005 gs, the dB range would be

522

005.log20dB

This is well within the capabilities of data collectors and FFT analyzers if the dynamic

range is adjusted properly. For instance, if a 16 bit instrument is being used, the range is 215 or

32,768 to 1 or dB = 20 log 32,768 = 90.

If the range on the analyzer is set a 10 gs for this measurement, then the base value is

gs0003.x

10X

10or

X

10log2090

20

90

13.61

Fig- ure

13.61. Gearbox Vibration Displayed on a Linear Scale

Figure 13.62. Gearbox Vibration Displayed on a Log Scale

13.62

REFERENCES 13.1. Karassik, IlJ., Krutzsch, W.C., Fraser, W.H., Messina, J.P., Pump Handbook, 2nd Ed.,

McGraw Hill, Inc., NY, 1986. 13.2. Baxter, N., Machinery Vibration Analysis III Notes, Volume II, Vibration Institute. 13.3. Bolletar, U., “Blade Passage Tones of Centrifugal Pumps,” Vibrations, Vol. 4, No. 3, Sept

1988, pp 8-13. 13.4. Guy, K., Case Histories: Power Industry, Vibration Institute, Willowbrook, IL, 1993. 13.5. Wachel, B., Szenasi, F. et al, Vibrations in Reciprocating Machinery and Piping Sys-

tems, Engineering Dynamics Inc., San Antonio, TX, 2002. BIBLIOGRAPHY Welch, Harry J., “Transamerica Delaval Engineering Handbook,” 4th Ed., McGraw-Hill Inc., NY, 1983. Block, Heinz P. and Geitner, Fred K., Machinery Failure Analysis and Troubleshooting, 3rd Ed., Gulf Publishing Co., Houston, TX , 1997. Block Heinz P. and Geitner, Fred K., Machinery Component Maintenance and Repair, Gulf Pub-lishing Co., Houston, TX, 1985. acoustic resonance, 25

asymmetric supports, 43

atmospheric pressure, 5

average fluid velocity, 5

axial pump, 2

back pressure, 7

balance sensitivity, 36

best efficiency design point, 6

blade pass, 28

blade rate frequency (BRF), 57

Bolletar, U., 60

bypass orifice, 9

casing, 5

casing structural resonance, 14

cavitation, 5

centrifugal compressors, 48

centrifugal fans, 27

centrifugal machines, 1

choked flow, 49

compressor, 1

damper position, 28

diffuser, 2

diffuser pump, 2

13.63

diffuser vanes, 2

duct acoustical vibration, 43

dynamic head, 6

dynamic vibration absorber, 19, 20

firewater pump, 14

floor resonance, 43

flow instability, 49

foreign objects, 23

gage suction pressure, 5

gaps, 25

head curve, 9

impeller, 5

impeller eccentricity, 28

impeller-diffuser, 25

improper pump assembly, 13

Karassik, IlJ., 60

liquid vapor pressure, 5

model, 18

multistage pumps, 2

operating deflection shapes, 16

overhung blower, 33

poor inlet conditions, 25

pulse width modulation (PWM) motor excitation, 22

pump, 1

pump shaft deflection, 20

pumps, 2

recirculation, 9

shaft resonance, 20

specific speed, 5

static head, 5

stiffeners, 19

torsional vibrations, 22

total pump head, 4, 5

turbine multistage pump, 3

vapor pressure, 5

vertical pumps, 14

volute, 2

volute pump, 2

Wachel, B., 60

wear ring clearance, 24