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CHAPTER 7

FAULT DIAGNOSIS OF CENTRIFUGAL PUMP AND

IMPLEMENTATION OF ACTIVELY TUNED DYNAMIC

VIBRATION ABSORBER IN PIPING APPLICATION

7.1 INTRODUCTION

Vibration due to defective parts in a pump can be an annoying

problem resulting in unnecessary maintenance and can affect the pumping

system performance and endurance. This chapter focuses on diagnosing of

faults of centrifugal pump by vibration analysis. Also, to control the amplitude

of vibration in a pipe line due to hydraulic pulsation frequency, this is the speed

result of operating condition of the pump. The developed SMA based actively

tuned dynamic vibration absorber was used to control the amplitude of

vibration for varying excitation frequency.

7.2 CENTRIFUGAL PUMP

Centrifugal pumps are one of the most important elements in almost

all industries. The pumps are the key elements in food industry, waste water

treatment plants, agriculture, oil and gas industry, paper and pulp industry, etc.

Its purpose is to convert energy of a prime mover (an electric motor or turbine)

first into velocity or kinetic energy and then into pressure energy of a fluid that

is being pumped. The energy changes occur by virtue of two main parts of the

pump, the impeller and the volute or diffuser. The impeller is the rotating part

that converts driver energy into the kinetic energy. The volute or diffuser is the

stationary part that converts the kinetic energy into pressure energy. The cut

section model and fluid path through centrifugal pump is shown in Figure 7.1.

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Figure 7.1 Fluid path through the centrifugal pump (Suhane A, 2012)

7.3 SIGNIFICANCE OF FAULT DIAGNOSIS USING

VIBRATION ANALYSIS

The most revealing information on the condition of rotating

machinery is a vibration signature. Vibration parameters provide the needed

frequencies due to the flow and recirculation. When analyzing the vibration

data, an FFT vibration spectrum may be broken down into several frequency

ranges to help to determine the machine problem. Vibrations externally

measured on a pump have been used to monitor the operating condition of the

pump and diagnose the fault, if there is any, without interfering with the normal

operation. The most common method employed for examining mechanical

vibration is spectral analysis. Condition monitoring and fault diagnostics are

useful to ensure the safe running of machines. Vibration signals are often used

for fault diagnosis in mechanical systems because they carry dynamic

information from mechanical elements. These mechanical signals normally

consist of a combination of the fundamental frequency with a narrowband

frequency component and the harmonics. Most of these are related to the

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revolutions of the rotating system since the energy of vibration is increased

when a mechanical element is damaged or worn. Some of the conventional

techniques used for fault signal diagnosis include power spectra in time domain

or frequency domain. These can provide an effective technique for machinery

diagnosis provided the signals are stationary.

7.4 CAUSES OF VIBRATION IN CENTRIFUGAL PUMP

Vibration due to unbalance

Dynamic imbalance in centrifugal impeller or shaft can cause heavy

vibration and transmits it to piping which can be cured by balancing the fan

with shaft on balancing machine. Simple unbalance, uncomplicated by other

problems can be identified by the following characteristics (Brain PG 2011):

a) Amplitude occurs at 1 X RPM of the shaft

b) The radial vibration is reasonably uniform and not highly directional

c) If the specific component such as impeller or fan is the source of

unbalance, it will have high amplitude at 1 X RPM frequency

Vibration due to misalignment

Misalignment of direct coupled machines is the most common cause

of machinery vibration. In spite of self-aligning bearings and flexible

couplings, it is difficult to align two shafts and their bearings, which will cause

vibration.

Although machines may be well aligned initially, several factors can

affect alignment, namely, operating temperature, setting up of the base or

foundation and deterioration or shrinkage of grounding. Misalignment can be

clearly identified from the following characteristics:

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a) Predominantly occurs at 2 X RPM

b) Amplitude is high in axial direction when compared to horizontal

direction

Vibration due to hydraulic pulsation

The problems caused by hydraulic pulsation in pumps are easy to

recognize because the resultant vibration will occur at frequency which is the

product of number of impeller vanes and machine speed (rpm). The amplitude

of vibration due to hydraulic pulsation in a pipe line is inevitable because of the

working of the pump. It is not unusual to detect some vibration at the vane or

blade passing frequency on nearly every pump. It would be impossible to build

a machine where no hydrodynamic forces are present. However, when the

amplitude of hydraulic pulsation is excessive, a problem is indicated.

Vibration due to cavitation

Pumps are designed to operate at certain flow conditions including

suction and discharge pressures, flow rates, head pressures, product density or

specific gravity, etc. If operated beyond or outside these designed parameters, a

high amplitude of vibration generally results. Pumps that are forced to operate

amount of fluid enters the pump is insufficient. This creates vacuum pockets in

the fluid that are unstable and can even collapse or explode.

Cavitation can be identified by the following characteristics:

a) Cavitation occurs between 20000 CPM (333.33 Hz) and 150000 CPM

(2500 Hz)

b) Vibration can be detected at any location in the pump

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c) A hydraulic problem and it may be due to the design of vanes

d) Causes haystack of vibration

Vibration due to bearing defects

When a rolling-element bearing develops flaws on the raceways and/

or on rolling elements, there are actually a number of vibration frequency

characteristics that can result, depending on the extent of deterioration.

Thus, identifying these characteristic frequencies can not only help to verify that a

bearing is definitely failing, but it can also give some indication on the extent of

deterioration. Bearing defects can be identified by the following characteristics:

a) Defects on bearing occurs between 20000 CPM (333.33 Hz) to 150000

CPM (2500 Hz)

b) The vibration can only be detected at the place of bearing

c) The haystack of vibration will increase and spread out as the day progresses

Apart from the above stated reasons, piping system might

experience vibrations due to improper supports, fittings and water

hammer. In case of gas pipe lines, the vibration can also come from

the pulsations generated by reciprocating compressors.

7.5 INFLUENCING PARTS OF PIPING SYSTEM

The following parts are subjected to vibration and defect due to the

working of the pump.

Pipe

All piping supports

Hangers

Snubbers

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Pipe to pipe interfaces

Machinery or devices attached to the pipe

All these items can influence the pipe vibration patterns. The vibrations

produced in the pipelines contain various risks concerned with the industry as well

as the domestic applications. A pipe will not vibrate if it is prevented from moving.

However, this does not necessarily help the piping system design

from the standpoint of its ability to absorb differential thermal expansion.

Therefore, when addressing a vibration problem, the flexibility design of the

piping system must also be considered. Restraints that are added to reduce

vibration must not increase the pipe thermal expansion stresses or end-point

reaction loads to unacceptable levels.

7.6 SPECIFICATION OF PUMP

It is necessary to obtain amplitude vs frequency spectrums or FFTs

absolutely no value to the vibration analyst unless some specific details about

the machine are known. Specific machinery problems are identified by relating

their vibration frequencies to the rotating speed (RPM) of the machine

components, along with other machine features such as the number of teeth on

gears, the number of blades on a fan etc., are shown in Table 7.1.

Table 7.1 Specification of centrifugal pump

1. Rotating speed of shaft - 2880 RPM 2. Type of bearings - Roller bearing 3. No of rolling element in each bearing - 7 4. No of impeller vanes - 6 5. Head range - 20-100 feet 6. Output range - 500- 1500 lpm

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7.7 EXPERIMENTAL PROCEDURE FOR FAULT DIAGNOSIS Experimental setup used in this study for fault diagnosis of a

centrifugal pump is shown in Figure 7.2 (a). Centrifugal pump is rigidly fixed

to the foundation, so that the vibration caused due to the looseness with the

foundation could be avoided in experimental results. Accelerometers are

mounted at the inboard bearing of centrifugal pump as shown in Figure 7.2 (b).

(a) Major parts (b) Location of accelerometer

Figure 7.2 Experimental setup for fault diagnosis of centrifugal pump

Experimental setup is made to run as per the testing condition

standards. By changing the accelerometer orientation, data was recorded in

other two axes at the same bearing location.

The vibratory forces generated by the rotating components of a

machine are passed through the bearings. Vibration readings for both detection

and analysis are taken on the bearings whenever possible. Ideally, vibration

readings taken in horizontal and vertical directions are taken directly on or as

close as possible to the bearings with the accelerometer pointing towards the

centerline of the shaft. Axial vibration readings are taken on the bearing as

close to the shaft as possible. Adhesive mounting is used to mount the

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accelerometer over the bearing which provides frequency response up to

540000 CPM (9000 Hz) that is sufficient for this experimentation. The things

that must be taken care while adhesive mounting is machine surface must be

flat, smooth and clean to ensure secure bonding.

The layer of adhesive should be kept as thin as possible to provide

maximum frequency response. Figure 7.3 shows the Lab VIEW block diagram for

this experimentation. NI USB DAQ 9461 with 4 input channels and delta type

DAQ is used to connect the accelerometer with Lab VIEW software. Initially, the

measurement was taken to check for background noise, which clearly indicates

that there is no significant level of surrounding as shown in Figure 7.4.

Figure 7.3 Lab VIEW block diagram

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Figure 7.4 Lab VIEW measurement for surrounding noise level

Also, the frequency spectrum of the healthier pump was recorded, so this

Figure 7.5.

Then, the frequency spectrum of defected pumps was recorded. From the data

collected, the defects are identified through vibration analysis by finding the

difference in amplitudes. Also, according to pump standards the amplitude of

vibration is within 0.15 inches which is acceptable due to any cause of vibration

(Brain PG 2011).

Figure 7.5 Vibration level of a good pump

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Few pumps were taken from pump industry which has customer

complaints for vibration analysis in order to find out the causes for complaints.

From Figure 7.6 (a) & (b), it can be clearly noticed that the peak amplitude of

vibration occurs at 50 Hz, which is clearly a characteristic of an unbalanced

impeller.

The magnitude of unbalance is little higher in case of second pump

when compared to first one. But in both cases, at frequency of 50 Hz,

amplitude of vibration is higher than the permissible value which has to be

addressed. Vibration was measured both in axial and radial direction.

(a) Radially

(b) Axially

Figure 7.6 Vibration measured for an unbalanced pump

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The vibration levels measured in the radial and axial axes of the

pump are shown in Figure 7.7 (a) & (b). From the Figure 7.7 (b), it can be

clearly seen that the axially measured amplitude vibration of a pump is 700%

greater than that of radially measured vibration. Both are located at 100 Hz.

Also, since it has different magnitudes in different axes, it is found to be

directional. It shows that the amplitude at 100 Hz is clearly the characteristic of

misaligned shaft pump. Further, it can be understood that the vibration is at risk

because of the misalignment which has to be addressed.

(a) Radially

(b) Axially

Figure 7.7 Vibration measured for misalignment in a pump

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Similarly a number of pumps were taken for experimentation to

study the effect of cavitation and hydraulic pulsation in various pumps.

Common pump defects and its corresponding frequency is shown in Figure 7.8

which will be helpful for further fault detection in defected or normal pumps.

Figure 7.8 Vibration characteristics of centrifugal pump with defects

7.8 VIBRATION CONTROL IN A PIPE LINE USING SMA

BASED ATDVA Vibration in the pipelines may be caused by various factors as

discussed above. However, the vibration due to the hydraulic pulsation is

caused when the impeller in a pump continuously transmits tiny pockets of

water into the pipeline. The hydraulic pulsation is calculated as given in

Equation 7.1.

Hydraulic pulsation frequency = NZ (7.1)

where N is the speed of the shaft and Z is the number of vanes on the impeller.

An attempt is made to develop an absorber system to control the amplitude of

vibration caused by the frequency which is the result of impeller rotation.

Hydraulic pulsation frequency caused by the speed is a result of operating

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condition of the pump. The variation in the speed due to voltage fluctuation,

slippage in rotor and variation in load conditions that leads to increase in

vibration in the pipe line. This may increase the chances of causing damages to

pipeline, the seals and joints which results in leaks. The leakage of the toxic

chemical or gases may pollute the atmosphere or may even affect the

surroundings. Further, the pipe line produces sound and causes loss of the fluid.

These undesirable effects are produced due to the variation in hydraulic

pulsation frequency is addressed by the development of SMA based actively

tuned vibration absorber. Figures 7.9 & 7.10 show the centrifugal pump

apparatus and the experimental setup including the interfacing the system with

LabVIEW.

Figure 7.9 Experimental setup

Figure 7.10 Experimental setup with Lab VIEW

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Based on the centrifugal pump selected for vibration analysis, the

hydraulic pulsation frequency was calculated by using Equation 7.1.

Speed of the motor (N) = 2880 rpm

Total no of vanes (Z) =7

Excitation frequency of the pump = 2880 x 7

= 20160 rpm (hydraulic pulsation frequency)

=20160/60

Excitation frequency of the pump = 336 Hz

When the vibration is measured on the pipe line, it is evident from

the Figure 7.11 that the peak occurs at nearly 336Hz and it is the frequency

which has the highest amplitude. This amplitude of vibration is due to

hydraulic pulsation (Excitation Frequency) which was considered for active

vibration control. But in the experimental results, it was found that the peak

frequency is varying between 336 Hz and 339 Hz. This may be due to the

voltage fluctuations, varying load conditions and slippage in rotor.

For the varying frequencies from 336 Hz to 339 Hz, the

conventional absorber has to be redesigned which is cumbersome. Also, the

SMA spring available may not offer the stiffness for the frequency around

336 Hz, Hence, it is decided to develop a combined absorber system with

conventional spring and SMA spring so the equivalent stiffness was considered

for calculation of excitation frequency.

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Figure 7.11 Amplitude of vibration due to hydraulic pulsation

7.8.1 Development of conventional absorber

In this study, the conventional absorber is developed for the

excitation frequency of 336 Hz. The variation in excitation frequency due to

the impeller rotation is taken care by the SMA spring. Approximately a change

speed of 180 rpm (3 Hz) is accounted for reduction in amplitude of vibration.

Angular velocity x 3.14 x f

where f = 336 Hz

= 336 x 2 x 3.14

= 2110.08 rad/s

modulus E = 2.1x105 N/mm2 (spring steel)

=0.3

Wire diameter d= 1.8 mm

Mean diameter D=3.72mm

No of turns N=9

Stiffness of the conventional spring K =228757.1363 N/mm

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7.8.2 Development of an absorber with conventional SMA spring

The stiffness offered by the SMA spring is not sufficient for

controlling this very high excitation frequency. Hence, to improve the stiffness,

a parallel system of springs has been used which consists of both the

conventional and SMA springs as shown in Figure 7.12.

Figure 7.12 DVA with conventional spring and SMA spring

K1 = stiffness of the conventional spring.

K2 = stiffness of the SMA spring.

Equivalent stiffness (Ke) = K1 + K2

In martensite state,

Equivalent stiffness of the parallel system = K1 + K2

K1=221533.7667 N/m

K2=1660.67 N/m in martensite state of SMA spring

Ke= 223194.436N/m and the corresponding frequency is calculated as

2 x x

where stiffness Ke = 223194.436N/m and suspended mass m=50g

rad/s.

Frequency f = 336.4307 Hz.

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In austenite state,

K1 =221533.7667 N/m

K2=4863.281 N/m in austenite state for SMA spring

Ke = 226397.046 N/m and the corresponding controlling frequency is

calculated where

Ke=226397.046N/m and for the same suspended mass of 50g

2 = 226397.046/0.050

= 4527940.92

rad/s

= 2 x 3.14 x f

f= 339 Hz

Hence, the absorber was designed for controlling the frequency in the range of

336 to 339 Hz.

7.9 EXPERIMENTATION USING SMA BASED ATDVA

The excitation frequency of the pump is calculated as 336 Hz. But in

the experimental results, it was found to be varying between 336 and 339 Hz.

For the varying frequencies of 336 to 339 Hz, an SMA based actively tuned

absorber system has been developed to take care of this 3Hz change in

excitation frequency due to hydraulic pulsation. Figure 7.13 shows that the

peak occurs at 337 Hz with amplitude of 2.3mm. But the theoretical calculation

shows that the excitation frequency of the pump occurs at 336 Hz according to

the Equation 7.1. This variation in frequency may be due to voltage fluctuation,

varying load conditions and slippage in rotor.

An absorber designed with the help of SMA spring along with

conventional spring was used to control the amplitude of vibration due to the

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change of 3 Hz in excitation frequency. The amplitude of reduction in vibration

at 337 Hz is shown in Figure 7.14. From this, it is evident that the amplitude of

vibration reduces from 2.3 mm to 0.8 mm which accounts for around 63.04 %

reduction of amplitude due to the SMAs stiffness changing ability. Also, the

experiments were carried out by intentionally varying the excitation frequency

within the permitted level of change in stiffness of SMA spring.

Figure 7.13 Amplitude of vibration without absorber at 337 Hz

Figure 7.14 Amplitude of vibration with absorber at 337 Hz

When the excitation frequency is varied to 338 Hz, the amplitude of

vibration is measured on the pipeline. Figure 7.15 shows that the peak

amplitude occurs at 338 Hz, owing to the variation in the excitation frequency.

The SMA spring will be supplied with required current by the control system,

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in order to produce the desired stiffness to reduce the amplitude of vibration,

which is the result of change in frequency from 337 to 338 Hz. The amplitude

of vibration is reduced from 2.1 mm to 0.8 mm which is shown in Figure 7.16.

This results in 58.09% reduction in vibration of the pipe line.

Figure 7.15 Amplitude of vibration without absorber at 338 Hz

Figure 7.16 Amplitude of vibration with absorber at 338 Hz

When the excitation frequency is increased further to 339 Hz, it

tends to develop peak amplitude in pipe line at 339 Hz. The amplitude is found

to be 2.25 mm at 339 Hz as shown in Figure 7.17.

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Figure 7.17 Amplitude of vibration without absorber at 339 Hz

This leads to 62.2% reduction in amplitude of vibration at the

frequency of 339 Hz which is depicted in Figure 7.18. If the excitation

frequency is reduced from 336 to 334 Hz. It is not possible to use the combined

DVA with conventional and SMA, because the conventional DVA was

designed to take care of the frequency of 336 Hz.

Figure 7.18 Amplitude of vibration with absorber at 339 Hz

The SMA spring is attached with conventional spring will increase

the stiffness leads to account only increase in excitation frequency for vibration

control. This problem can be addressed when only SMA springs of high

stiffness ranges can be combined parallel nature or co-axial in nature results in

more variation in excitation frequency range both below and above the

frequency of 336 Hz.

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Figure 7.19 shows the reduction in amplitude of vibration in the pipe

line for the frequencies of 336 to 339 Hz. Table 7.2 shows the reduction in

amplitude of vibration with percentage.

Figure 7.19 Reduction in amplitude of vibration for the frequency range of

336 to 339Hz

Table 7.2 Comparison of percentage reduction in amplitude of vibration

for 336 to 339 Hz of excitation

Frequency (Hz )

Without absorber (mm )

With SMA spring (mm)

Percentage of reduction (% )

337 2.3 0.85 63.04

338 2.1 0.88 58.09

339 2.25 0.85 62.20

7.10 CONCLUDING REMARKS

Thus, the vibration due to different defects in centrifugal pumps has

been analyzed and the frequency at which the defects are happening has been

found out by the experimentation procedure. These frequencies would be

useful for the fault diagnosis of the centrifugal pumps without dismantling the

assembly. Experiments were further carried out with the help of SMA springs

0

0.5

1

1.5

2

2.5

337 338 339

Ampl

itude

(mm

)

Frequency (Hz)

Without absorber

With SMA spring(mm)

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on centrifugal pump apparatus. In order to demonstrate the concept of actively

tuned dynamic absorber for varying excitation frequencies due to hydraulic

pulsation results in more amplitude of vibration in the pipe lines connected

with centrifugal pump. The change in excitation frequency in the inclining

trend was addressed with the help of ATDVA. The reduction in amplitude of

vibration around 50-65% was attained with the help of developed SMA based

actively tuned dynamic vibration absorber. The variation in excitation in

decreasing trend, which can be addressed only with SMA springs of high

stiffness ranges can be combined parallel nature or co-axial in nature resulting

in more variation in excitation frequency range in the decreasing trend.