Paper Ref: S1147_P0508 3rd International Conference on Integrity, Reliability and Failure, Porto/Portugal, 20-24 July 2009

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FREQUENT AND PREMATURE FAILURE OF ANTI-FRICTION BEARINGS: DIAGNOSIS APPROACH?

Jyoti K. Sinha School of Mechanical, Aerospace, and Civil Engineering (MACE) B3, Pariser Building, Sackville Street The University of Manchester P.O. Box.88 Manchester M60 1QD Email: Jyoti.Sinha@manchester.ac.uk

ABSTRACT

Several vibration based methods are used in practice to identify the fault in the anti-friction bearings (ball bearings and roller bearings), however experience suggests that these methods are just indicative to the bearing health and not sufficient to identify the root cause if the failure is premature and frequent.

INTRODUCTION

Many rotating machines have rotors that are supported through anti-friction ball or roller bearings. Faults in such bearings always develop and detection/ appearance of these faults well in advance is important so that the remedial action to be taken before any catastrophic failure. A photograph of a typical premature failed roller bearing is shown in Figure 1 (Sinha and Rao, 2004).

Figure 1 Photograph of a premature failed bearing (Sinha and Rao, 2004)

The most common vibration based techniques for detection of faults in anti-friction bearings are crest factor and kurtosis measurements (Bendat and Piersol, 1985, Rao, 1999, Barkov et. al., 1995a). The envelope, or more precisely the amplitude demodulation at the carrier frequency (usually the rotating speed of a machine) is often used to locate the exact nature of fault by identifying the bearing characteristic natural frequencies (Barkov et. al., 1995a &b, Randall and Gao, 1996, Bosmans, 1982). McFadden and Smith (1984) gave the review of the techniques available till 1984 for the anti-friction bearing diagnosis. All these approaches give the health of

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the bearing only without looking into the real root cause problem. If the failure is purely due to the bearing ageing related problem, then these methods are fine, but if the failure is premature and frequent, whether the approaches used in practice are enough to downsize the maintenance overhead and increasing machine availability. The answer is negative. The experience shows that the monitoring of the bearing health alone is not sufficient if failure is frequent. For such cases, it has been observed that the dynamics of the complete machine unit comprises of the rotor, bearings and foundation is responsible for such premature and frequent failure (Sinha and Rao, 2006, Sinha and Balla, 2006). The paper discusses few such case studies.

CASE STUDIES Here two typical case studies (Sinha and Rao, 2006, Sinha and Balla, 2006) have been discussed where machines were operated for long periods (years together) without any significant damage then it has been observed that the frequent failure (2 or more times in a year) in the anti-friction bearings in either machines. The regular kurtosis and the envelope measurements were always used to identify the faults in the bearings and replaced them in time to avoid any catastrophic failure. However these bearing diagnosis methods have not been able to provide any root cause for such frequent failures of the bearings which is expected the way the anti-friction diagnosis methods have been developed. The modal testing (Ewins, 2000) and/or details vibration diagnosis on the complete machines during the normal machine operation (rather than vibration measurements on the bearing pedestals) that have helped in solving the problems of the bearing frequent failures which have been discussed here.

CASE#1 (Sinha and Rao 2006)

It is a case of a centrifugal pump commissioned in 1985 (Sinha and Rao, 2006). The pump had no failure till 1990, and then it has a long history of anti-friction bearing failure. The vibration based condition monitoring on the pump (Prasad et al., 2002) had usually shown high 2X component (two times the pump RPM) in the vibration spectrum, and faults in the bearings was detected by the kurtosis and the envelope analysis of the measured responses on the bearing pedestals. This information has always been utilized to replace the faulty bearings in well planned shutdown before any major failure occurred. The photograph of a typically failed bearing of this pump is already shown in Figure 1 (Sinha and Rao, 2004). However the condition monitoring could not identify the root cause for frequent failure of bearings. Hence the modal tests were performed on the pump assembly to understand the dynamic behaviour.

Figure 2 shows schematic of the pump assembly. It is a horizontally mounted centrifugal pump with the axial inlet, and the radial outlet. The pump and the motor shafts are rigidly coupled to a shaft carrying a flywheel (FW). The FW is supported by a grease lubricated radial bearings on pump side and oil lubricated taper roller thrust bearing on the other side. The pump is driven by a 540kW electric motor operating at 1492RPM with a discharge rate of 21 klpm at 11kg/sq. cm (Sinha and Rao, 2006). The pump is mounted directly on the base plates embedded to the rigid concrete floor.

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Figure 2 Schematic of a Pump assembly and the measurement (dot) locations (Sinha and Rao, 2006)

Modal Testing and Diagnosis Sinha and Rao (2006) have conducted modal tests on the pump assembly using an instrumented hammer to give impulse excitation to the pump assembly in a frequency band up to around 500Hz, and the responses of the assembly from different locations and in different directions (the vertical, the axial along shaft axis, and the lateral to shaft axis) were collected from number of accelerometers. The measurement locations are also shown in Figure 2.

The frequencies 44.51Hz, 57.96Hz, 65.89Hz, and 141.77Hz are the natural frequencies identified by the modal tests. The most interesting observation made was the mode at 57.96Hz had significant deflection only at both the bearing pedestals as shown in Figure 3. This frequency in the FRF at the bearing pedestals appeared as a broad banded peak and has almost 14dB amplification at 2X component of the pump vibration. It is typically seen in the frequency response function (FRF) plot in z-direction (lateral to the rotor axis) at the bearing pedestal (near to motor side) in Figure 4. The drop in the natural frequency close to the 2X, and its broad banded nature must be due to the looseness between the base plate and the concrete resulting in the non-linear interaction between the base plate and the concrete surface. The suspicion of such non-linear behaviour has further been confirmed by the higher order spectrum and the wavelet analysis (Sinha and Rao, 2006).

x z

y

Flow

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Figure 3 The mode shape at 57.96Hz of the pump assembly showing deflection at bearing pedestals only (Sinha and Rao, 2006)

Figure 4 A typical FRF plot at bearing pedestal (near to Motor) in z-direction (Sinha and Rao, 2006)

Bearing Pedestal

Pump

Motor

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Hence the experimentally identified broad banded natural frequency at 57.96Hz and its closeness to 2X component has been identified as the main reason for the failure. It is because a small 2X component generated due to even a small shaft misalignment at the coupling/asymmetric shaft must have triggered the resonance at 57.96Hz which in turn leads to increase in the shaft misalignment. Such induced misalignment could cause damage to the bearings prematurely. It can be solved either by stiffening the roots of bearing pedestals or by proper grouting the base plate in concrete. Sinha and Rao (2006) gave the details of vibration measurements, data analysis, diagnosis, and the solution suggested.

CASE#2 (Sinha and Balla, 2006)

It is a case of a blower system shown in Figure 5, where the problem of the frequent bearings failure is identified, and solved by the ODS analysis (McHargue and Richardson, 1993, Richardson, 1997). The blower system consists of a motor of 40HP, 1460rpm drives the blower at 1070rpm through a V-belt arrangement (Sinha and Balla, 2006, Balla and Rao, 2004). The chassis of the blower system is isolated with the concrete floor by isolation springs. The driven shaft is supported by two anti-friction bearings, which are resting on the bearing support block (bearing pedestal). The Blower is 1 m in diameter and has 16 blades.

The blower system was installed almost 2 decades ago, and the machine had no problem for 15 years. In early 2001, the problem in the machine started with high vibration and the frequent failure of both the bearings (Sinha and Balla, 2006).

Vibration Measurement and Diagnosis The vibration measurements in year 2002 on the bearings in the horizontal, vertical, and axial directions by Balla and Rao (2004) indicate the dominance vibrations at the belt speed (12.5Hz) at the bearing 1 and the blower speed (17.5Hz) and the twice of the blade passing frequency (2 x BPF) at the bearing 2 which is close to the blower fan. Further investigation by Balla and Rao (2004) using the ODS analysis during the machine operation confirms that the out of phase movement of the bearing pedestals with respect to the blower casing and the motor pedestal at the blower RPM. The ODS is shown in Figure 6. This motion must be loading the bearings which lead to the frequent failure.

Hence based on the observations, the two suspicions were raised (Balla and Rao, 2004, Sinha and Balla, 2006)- (1) there may be looseness in the blade anchoring causing the 2 x BPF frequency component (Bosmans, 1982, Goldman, 1999), and (2) some angular misalignment between the driven pulley and the blower shaft (Crawford and Crawford, 1992), the location is highlighted in the encircled area in Figure 5. The site has confirmed the suspects, and the corrected the defects in year 2002. Since then no failure is reported. The details of the measurements and analysis are given the paper by Balla and Rao (2004).

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Top View

Front View

Figure 5 Schematic diagram of the Motor- Blower assembly and the measurement locations (Sinha and Balla, 2006)

1 2

3 4 5

6 7

8

9 10

11 12

Bearing Support

Bearing 1

Bearing 2

Motor

V-Belt Drive

Isolation Springs

Exhaust Fan

Angular misalignment

suspected

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Figure 6 The ODS of the assembly at the blower speed 17.5Hz (Sinha and Balla, 2006)

CONCLUDING REMARKS

Vibration based fault diagnosis methods for the anti-friction bearings are good for identifying the presence of faults well in advance, however if the failure of the bearings are frequent, then it is essential to carry out some additional tests like modal testing and/or ODS analysis to understand the dynamics of the complete machine as a unit to identify the root cause for the repeated failure. Once the root cause has been identified, the appropriate remedy can be done to avoid the repeated failure. It will not only increase the availability of machines, but also downsize the maintenance overhead and enhance the plant safety. This has been successfully demonstrated here through two typical case studies.

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Barkov, A, Barkova N, Mitchell JS. Condition Assessment and Life Prediction of Rolling Element Bearings Part 2. Sound and Vibration 29; 9; 1995b, p. 27-31.

Out of Phase movement

1 2

3 4 5

6 7

8

9 10

11 12

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