oil whirl rotordynamic instability large induction motor

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1 OIL WHIRL ROTORDYNAMIC INSTABILITY PHENOMENON- DIAGNOSIS AND CURE IN LARGE INDUCTION MOTOR Copyright Material IEEE Paper No. PCIC-2009-35 Sumit Singhal Rajendra Mistry, P.E. Member, IEEE Member, IEEE Siemens Energy & Automation Siemens Energy & Automation 4620 Forest Avenue 4620 Forest Avenue Norwood, OH ,USA Norwood, OH , USA [email protected] [email protected] Abstract - Reliable design and operation of rotating machinery is critical to achieve bottom line production goals and financial results in petrochemical, transportation and chemical products processing industries. Of the many factors which may cause failure and or reliability issues in these industries, one of the most frustrating and involving relates to bearing failures or high vibration problems. One of the choices for rotating equipments is oil lubricated journal bearings. Oil film journal bearings are susceptible to large- amplitude, lateral vibrations due to a “self-excited rotordynamic instability” known as oil whirl. During oil whirl the rotor orbits in its bearings at a frequency approximately half the rotor speed. Oil whirl is usually caused by lightly loaded bearings, excessive bearing clearances, weak foundation and/or improper oil viscosity. At the inception of non-synchronous oil whirl the amplitude of rotor motion progressively builds at a frequency of approximately less than half of that of the rotor speed and does not damped out. If left uncontrolled oil whirl may lead to catastrophic bearing failures and equipment damage due to high vibration values. This paper will discuss the cause of oil whirl, its effect, how to detect it and its solution. Index Terms — Oil whirl, sleeve bearing, induction motor, lateral vibrations I. INTRODUCTION Rotating machinery such as compressors, pumps, blowers, motors etc. form the backbone of the petroleum and chemical industry processing and transportation application. Due to their high reliability, good efficiency and simplicity, induction motors are the most popular type of prime mover for rotating equipments. There are many factors which may cause motor failure. One of the most frustrating and involving is bearing failure or high vibration. Although the cost of a bearing is a small fraction of the induction motor cost, their failure may cause additional costly equipment damage and expensive downtime. Most of the large induction motors which drive high speed compressors (1800 or 3600 RPM) or any other rotating equipment run on an oil lubricated sleeve bearings. Bearings can be either ring lubricated or force lubricated. Lateral vibrations of a motor rotor operating at high speeds may occur for different reasons as discussed in [1]. One of the most common reasons of high lateral vibration is due to a presence of unbalance in the rotor. The frequency of this kind of lateral vibration is always the same as the speed of the motor. This type of vibration does not depend on the bearing and can be cured by balancing of the rotating system. Rarely but some times induction motor sleeve- bearing configurations are susceptible to large-amplitude lateral vibrations due to a “self-excited instability” also known as oil whirl. Oil whirl is independent of rotor unbalance or misalignment and is a self-excited instability caused by the forces generated in the lubricating oil film due to hydrodynamic action. During oil whirl the rotor orbits in its bearing clearance at a frequency approximately less than half the rotor angular speed in the same direction as the rotor. If not controlled, this non-synchronous self-excited orbiting motion of the rotor will grow without bound inside the bearing clearance circle which may lead to catastrophic bearing failure and equipment damage. During oil whirl, rotor behavior is unlike the critical speed resonance where the amplitude of motion builds up as the rotor reaches its critical speed and then decreases once it has passed through the critical speed. At the inception of non-synchronous whirl the amplitude of rotor motion progressively builds up at a frequency less than half that of the rotor speed and stays high. This paper discusses the basic theory of oil lubricated sleeve bearings, causes and prediction of oil whirl instability. Also it presents the case study on oil whirl instability diagnosis and control. II. THEORY A. Operation of journal bearing A journal bearing is used to support the rotor in horizontal machines and to restrict radial movement in vertical machines. Most oil lubricated journal bearing consist of a stationary cylindrical body (sleeve) separated from the rotating shaft by a layer of lubricant. In a journal bearing such as plain cylindrical, lobe or tilting pad bearings pressure or hydrodynamic lift is generated within the thin lubricant oil film which separates the shaft and the bearing, thus preventing metal-to-metal contact. The nomenclature of oil lubricated sleeve bearing is shown in Fig. 1. In oil lubricated sleeve bearings oil may be supplied to the bearings by gravity feed, external lubrication system or oil rings. Oil adheres to the journal due to its viscous properties and pumping action as journal rotates. Fluid pressure is generated due to the hydrodynamic action of the fluid film. This pressure counteracts the weight of the rotor and lifts the journal from 978-1-4244-3800-6/09/$25.00 ©2009 IEEE

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Page 1: Oil Whirl Rotordynamic Instability Large Induction Motor

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OIL WHIRL ROTORDYNAMIC INSTABILITY PHENOMENON- DIAGNOSIS AND CURE IN LARGE INDUCTION MOTOR

Copyright Material IEEE

Paper No. PCIC-2009-35

Sumit Singhal Rajendra Mistry, P.E. Member, IEEE Member, IEEE Siemens Energy & Automation Siemens Energy & Automation 4620 Forest Avenue 4620 Forest Avenue Norwood, OH ,USA Norwood, OH , USA [email protected] [email protected]

Abstract - Reliable design and operation of rotating machinery is critical to achieve bottom line production goals and financial results in petrochemical, transportation and chemical products processing industries. Of the many factors which may cause failure and or reliability issues in these industries, one of the most frustrating and involving relates to bearing failures or high vibration problems. One of the choices for rotating equipments is oil lubricated journal bearings. Oil film journal bearings are susceptible to large-amplitude, lateral vibrations due to a “self-excited rotordynamic instability” known as oil whirl. During oil whirl the rotor orbits in its bearings at a frequency approximately half the rotor speed. Oil whirl is usually caused by lightly loaded bearings, excessive bearing clearances, weak foundation and/or improper oil viscosity. At the inception of non-synchronous oil whirl the amplitude of rotor motion progressively builds at a frequency of approximately less than half of that of the rotor speed and does not damped out. If left uncontrolled oil whirl may lead to catastrophic bearing failures and equipment damage due to high vibration values. This paper will discuss the cause of oil whirl, its effect, how to detect it and its solution.

Index Terms — Oil whirl, sleeve bearing, induction motor,

lateral vibrations

I. INTRODUCTION

Rotating machinery such as compressors, pumps, blowers, motors etc. form the backbone of the petroleum and chemical industry processing and transportation application. Due to their high reliability, good efficiency and simplicity, induction motors are the most popular type of prime mover for rotating equipments. There are many factors which may cause motor failure. One of the most frustrating and involving is bearing failure or high vibration. Although the cost of a bearing is a small fraction of the induction motor cost, their failure may cause additional costly equipment damage and expensive downtime. Most of the large induction motors which drive high speed compressors (1800 or 3600 RPM) or any other rotating equipment run on an oil lubricated sleeve bearings. Bearings can be either ring lubricated or force lubricated. Lateral vibrations of a motor rotor operating at high speeds may occur for different reasons as discussed in [1]. One of the most common reasons of high lateral vibration is due to a presence of unbalance in the rotor. The frequency of this kind of lateral vibration is always the same as the speed of

the motor. This type of vibration does not depend on the bearing and can be cured by balancing of the rotating system. Rarely but some times induction motor sleeve-bearing configurations are susceptible to large-amplitude lateral vibrations due to a “self-excited instability” also known as oil whirl. Oil whirl is independent of rotor unbalance or misalignment and is a self-excited instability caused by the forces generated in the lubricating oil film due to hydrodynamic action. During oil whirl the rotor orbits in its bearing clearance at a frequency approximately less than half the rotor angular speed in the same direction as the rotor. If not controlled, this non-synchronous self-excited orbiting motion of the rotor will grow without bound inside the bearing clearance circle which may lead to catastrophic bearing failure and equipment damage. During oil whirl, rotor behavior is unlike the critical speed resonance where the amplitude of motion builds up as the rotor reaches its critical speed and then decreases once it has passed through the critical speed. At the inception of non-synchronous whirl the amplitude of rotor motion progressively builds up at a frequency less than half that of the rotor speed and stays high.

This paper discusses the basic theory of oil lubricated sleeve bearings, causes and prediction of oil whirl instability. Also it presents the case study on oil whirl instability diagnosis and control.

II. THEORY

A. Operation of journal bearing A journal bearing is used to support the rotor in horizontal

machines and to restrict radial movement in vertical machines. Most oil lubricated journal bearing consist of a stationary cylindrical body (sleeve) separated from the rotating shaft by a layer of lubricant. In a journal bearing such as plain cylindrical, lobe or tilting pad bearings pressure or hydrodynamic lift is generated within the thin lubricant oil film which separates the shaft and the bearing, thus preventing metal-to-metal contact. The nomenclature of oil lubricated sleeve bearing is shown in Fig. 1. In oil lubricated sleeve bearings oil may be supplied to the bearings by gravity feed, external lubrication system or oil rings. Oil adheres to the journal due to its viscous properties and pumping action as journal rotates. Fluid pressure is generated due to the hydrodynamic action of the fluid film. This pressure counteracts the weight of the rotor and lifts the journal from

978-1-4244-3800-6/09/$25.00 ©2009 IEEE

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c. Unstable Orbit

b. Steady State orbit with unbalance

the bearing surface preventing metal to metal contact. There are several key operating and geometric parameters which influences the generation of reliable oil film such as operating speed, load, clearance between journal and bearing, oil viscosity, surface finish as well as the radius and length of the bearing.

Some of the important parameters involved in the design of oil lubricated sleeve bearings systems are summarized below and have been divided into various sub-categories:

A. Operating Parameter a) Bearing Load (W) b) Operating speed (N) c) Oil viscosity (μ) B. Geometric Parameter

a) Length of bearing (L) b) Diameter of bearing (D) c) Radial clearance between journal and bearing

(C) C. Performance Parameters

a) Eccentricity (e) b) Eccentricity ratio (ε) = e/C c) Minimum film thickness ( minh )

Usually operating parameters are determined by the rotor load and motor speed, which are known to the mechanical designer. In order to obtain optimized performance parameter for a given set of operating parameters, the designer can control items such as geometric and operating parameters such as oil viscosity to obtain the optimal and most reliable bearing system possible. B. Motion of the Rotor Inside the bearing clearance Rotor center can exhibit following type of trajectory inside the bearing clearance

a) Steady state orbit without unbalance b) Steady state orbit with unbalance c) Unstable orbit

Steady state rotor orbit without unbalance is shown in Fig. 2

(a). During this kind of rotor motion rotor center will rotate and if measured with non contacting proximity probes the orbit will be a point in the clearance circle and will remain there indefinitely. During this kind of rotor motion minimum film thickness will remain at fixed location. However this kind of rotor motion will not happen in real world situation as their is always some degree of rotating unbalance present in the rotor. Steady state rotor orbit with rotating unbalance is shown in Fig. 2 (b) during this kind of rotor motion rotor center will rotate around a point in the clearance circle forming a fixed repeatable stable orbit. During this kind of rotor motion minimum film thickness will not be at fixed location but makes rotation along with rotor center. Non contacting proximity probes mounted on the bearing housing measures the amplitude and phase of this shaft orbit. The amplitude of the rotor orbit grows with degree of the unbalance forces. In order to have reliable bearings vibration amplitude, this orbit should be kept as low as possible. API 541 limits residual unbalance to 4W/N at each end of the motor rotor. During unstable rotor orbit as shown in Fig. 2 (c) unlike steady state orbit with unbalance as shown in Fig. 2(b), rotor never settles down to a fixed point and never

hmin

Load

Clearance Circle

Shaft

Bearing

Lubricant

Eccentricity

Fig. 1 Schematic of Journal Bearing

Sensor 1 Sensor 2

a. Steady State orbit with no unbalance

Sensor Sensor

Fig. 2 Types of Rotor Motion inside bearing clearance

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500 1000 1500 2000 2500 3000

Rotor Speed (Rpm)

0.0

2.0

4.0

6.0

8.0

10.0

Vib

ratio

n A

mpl

itude

Oil Whirl

forms a fixed repeatable stable orbit. During this kind of rotor motion, rotor rotates around the bearing center with orbits which may grow unbounded inside the bearing clearance circle ultimately reaching the limit of the bearing clearance circle causing bearing wipe out. This kind of rotor orbit occurs during self excited lateral vibration instability such as oil whirl. III. OIL WHIRL PHENOMENON, DIAGNOSIS & CONTROL

There are many electrical and mechanical forces present in induction motors that can cause vibrations. Some of mechanical vibrations are listed below.

a) Resonance vibration at the critical speed of the rotor b) Self excited vibration in oil lubricated journal

bearings due to oil whirl. A. Critical Speed Resonance Resonance vibration at the critical speed of the rotor is the forced vibration due to an imbalance present in the rotor. When the rotating speed coincides with the natural frequency of the rotor system, high vibration occurs as a result of the resonance. Oil lubricated bearings reduce high vibrations while rotor is passing through the critical speed. Fig. 3 shows the typical signature plot of rotor speed v/s vibration amplitude of an induction motor passing through the critical speed and can be observed that the amplitude of motion builds up as the rotor reaches its critical speed and then decreases once rotor had passed through the critical speed. The amplitude of the orbit of rotor center grows as the speed ramps up to the critical speed and then shrinks as it moves away from the critical speed. Since critical speed resonance is not self excited vibration the rotor orbit is stable and does not grow without bounds inside the bearing clearance circle. Amplitude of vibration during critical speed and operating speed can be reduced by bearing oil film damping and precision balancing of the rotor respectively.

B. Self excited vibration: Oil Whirl Phenomenon Newkirk and Taylor [2] in 1925, based on their experimental findings, reported a new kind of self-excited rotordynamic instability associated with hydrodynamic type journal bearings. They observed that during this self excited instability the rotor orbits in its bearing clearance at a

frequency less than half of the rotor speed. They found that this lateral vibration can be controlled by shutting off the oil supply to the bearings. From these observations they concluded that these types of lateral vibrations of the rotor are due to the action of lubricating oil film and referred to this self-excited rotordynamic instability as oil whirl. From their experiments they also concluded that these sub-synchronous vibrations may be prevented by misaligning the bearings slightly, by the use of friction damped bearings or by avoiding the lightly loaded shafts. Since the discovery of this phenomenon by Newkirk many experimental and theoretical investigations have been conducted to understand and predict the onset of this non-synchronous whirling. Hagg [3] provided some theoretical insights into the phenomenon of oil whirl. He stated that during the stable motion or steady state condition of the rotor in a bearing, the hydrodynamic fluid forces developed by the oil film is equal to the external load. However during the whirling motion of the shaft, the hydrodynamic forces overcome the external load and act as an “energy source” which accelerates the circular orbit of the shaft in the direction of rotation of rotor. Oil whirl is independent of rotor unbalance and cannot be controlled by balancing [4]. Fig. 4 shows the typical signature plot of rotor speed v/s vibration amplitude of an induction motor showing oil whirl vibration. During oil whirl the rotor behavior is unlike critical speed resonance as discussed in previous section. During oil whirl the amplitude vibration increases with speed and never decreases or damped out. Also the rotor orbits around bearing center. Orbit of the rotor is unstable and grows without bound inside the bearing clearance circle. If go undetected or not diagnosed and controlled oil whirl can cause catastrophic bearing and equipment damage.

C. Why the frequency of oil whirl is 0.45X - 0.5 X

A cross-section of the journal bearing is shown in Fig. 5. The center of the bearing of radius R is labeled J . The shaft has radius r with center S . For simplicity if pressure variation in the bearing is ignored then the velocity profile of the lubricant is approximately linear, and varies from 0=v

Critical Speed (w)

0.0

2.0

4.0

6.0

8.0

10.0

Vib

ratio

n A

mpl

itude

500 1000 1500 2000 2500 3000 3500

Rotor Speed (Rpm)

Fig. 3 Vibration Data of Motor passing through critical speed Fig.4 Vibration Data of Motor with oil whirl

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on the bearing internal face to rv Ω≅ on the outside surface of the journal. Hence, incompressible flow continuity in the control volume 1-2-3-4-5-6-1 implies

( ) ( ) rveCreCr 221

21 =−Ω−+Ω (1)

where v is the speed of J . The constant whirl speed implies circular motion with rv ω= thus follows from (1) that

erre ω2=Ω , (2)

or

2Ω=ω . (3)

Hence the whirling frequency of the shaft equals half of the operation speed of the journal when pressure variation is ignored. However, bearings support of shaft load is by virtue of pressure variations. When the pressure varies the whirling speed is less than half of the shaft speed.

D. Prevention and Control of Oil whirl Lightly loaded bearing or very large bearing clearances are

most common reasons for oil whirl issues in induction motor. Various design charts have been developed to predict the onset of oil whirl based on theoretical and experimental study which are based on operating, geometrical and design parameters for various rotor stiffness values [5]. These design charts can be used to check bearing stability once geometrical parameters are calculated for optimized performance parameters. If speed of operation is higher than the threshold speed of instability then geometrical parameters such as length, diameter or clearance can be changed to increase the stability of the bearing. In some cases due to the combination of bearing design parameters and operating requirements such as for high speed lightly loaded bearing operation it may be difficult to design a plain cylindrical bearing which is free from oil whirl instability. To overcome this issue, bearings such as offset 2 lobes, 3 lobe, four lobe, and tilting pads as shown in Fig. 6 can be used to

suppress oil whirl and improve bearing stability. A few of the disadvantages of using offset or lobe bearings are

a) These are more expensive than cylindrical bearings b) Due to non uniform cross section oil rings may not

be used with 3 lobe or 4 lobe bearings hence it requires external lubrication system. This bearing system requires redundant or backup oil lubrication system in case of emergency coast down due to power failure.

c) Since it is special bearing design, spare parts may have long lead time.

The onset of oil whirl instability is sensitive to bearing clearance and overall system stiffness values. A situation may arise that the motor is running free from oil whirl for many years after commissioning, and it may suddenly run into oil whirl issue after several years of operations. This may happen due to increase in bearing clearances which may be caused due to wear of the babbit or the journal. Oil whirl instability may also be caused by weak motor foundation which reduces overall stiffness of the system. Sometime it is possible to control oil whirl on fields by changing the viscosity of lubricant. Lubricant viscosity plays a major role in oil whirl instability. Under normal operating conditions, the lubricant undergoes a significant change in temperature which causes a change in viscosity and this affects other bearing performance parameters such as minimum film thickness and load carrying capacity. Singhal et al [7] presented the comprehensive bearing design charts to control oil whirl by increasing or decreasing the oil supply temperature.

Fig. 5 Journal bearing undergoing oil

J

S

1

2

3 45

6

er ω+Ω

er ω−Ω

Ω

r

R

Fig. 6 Different kind of bearing configuration

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IV. CASE STUDY

Before we discuss a Case Study we will discuss about the practical aspect of finding oil whirl vibration using orbit, and orbit time base, and also spectrum plots. Oil whirl is self excited mechanism; motion is forward, is greater than 50% of rotational speed. Fig. 8 shows the orbit and time base of the oil whirl vibration and Fig. 9 shows the spectrum clearly indicating the sub synchronous vibration of the oil whirl. In Fig. 8 the upper plot shows the high unfiltered vibration while the lower plot shows low 1 time rotational vibration. [10] circular, and typically appears between 35% and 49% of shaft rotational frequency. The direct orbit is predominantly forward and circular, while filtered to the instability frequency, will always be circular and forward. Fig. 7 shows the behavior of the Keyphasor dots. (Blank bright sequence) Number of Keyphasor dots per rotational speed represents the relationship between sub synchronous frequency to the running speed. The direction of the dots with respect to the rotational speed tells us the value of the of sub synchronous frequency. When dots appear to move

in a counter rotation direction, the sub synchronous component occurs at less than 50% of running speed. If the dots appear to move in the same direction as shaft rotation, the sub synchronous component occurs at a frequency that Fig. 8 shows the orbit and time base of the oil whirl vibration and Fig. 9 shows the spectrum clearly indicating the sub synchronous vibration of the oil whirl. In Fig. 8 the upper plot shows the high unfiltered vibration while the lower plot shows low 1 time rotational vibration. [10] A. Case Study: Oil Whirl of a Vertical Motor

The following is a case study where a vertical motor exhibits oil whirl at shop testing. The motor was 3500 HP, 1200 RPM for a pump application.

Fig. 7 Keyphasor Dot for Sub-Synchronous Vibration [9]

Fig. 8 Orbit Plot – Direct & One time rotational

Fig. 9 Spectrum Plot – One time rotational

Fig. 10 Upper bearing Housing of Vertical Motor

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Fig.10 shows the upper bearing housing and bearing arrangement. The upper bearing housing consists of a thrust bearing; guide bearing and thrust block which is tightly assembled on the rotor shaft. The thrust bearing has six pivoted thrust pads with top surface of pads lined with the Babbitt material. Plain cylindrical guide bearing in upper bearing assembly keeps the thrust block in radial position. The inner surface of this guide bearing is also lined with Babbitt. The reservoir filled with suitable viscosity oil serves the lubrication of both thrust and guide bearings. Water cooler keeps the oil temperature controlled.

Fig. 11 shows the lower bearing housing and the bearing arrangement. The guide block is tightly fitted on to the shaft and runs within the guide bearing and is lubricated with the suitable oil. Both upper and lower guide bearing has a light radial load from the rotor. Any load exerted on these bearings is due to misalignment uneven magnetic force or tilt of the rotor. The motor was tested uncoupled and unloaded. The shaft vibration at upper and lower bearing was monitored using non contact probes position 90 o apart. Motor showed a high bearing vibration during the shop test.

Fig. 12 is the orbit plot for the sub synchronous and 1 time rotational of the upper bearing. Sub synchronous shows 2 keyphasor dots per revolution and it is less than 180o

indicating oil whirl sub synchronous frequency less than 50% of rotational frequency. Also sub synchronous vibration amplitude is much higher than the synchronous vibration amplitude. The spectrum plot in Figure 13 shows the sub synchronous frequency of 9.25 Hz. The rotational frequency is 20 Hz. 9.25 Hz is 46.25% of rotational speed.

Fig. 14 is the orbit plot for the sub synchronous and 1 time rotational of the lower bearing. Similar information can be achieved from these plots as from the upper bearing plots. Sub synchronous shows 2 keyphasor dots per revolution and it is less than 180 o indicating oil whirl sub synchronous frequency less than 50% of rotational frequency. Comparing upper and lower sub synchronous frequencies are very close. Also sub synchronous vibration amplitude is much higher than the synchronous vibration amplitude.

Fig. 11 Lower bearing Housing of Vertical Motor

Fig. 13 Spectrum of upper bearing

Fig. 12 Orbit plots at Lower bearing

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Fig. 15 also reveals the same information for the sub synchronous frequency. It is 9.25 Hz. and it is 46.25% of rotational speed.

Various methods were tried to solve the problem like reducing oil level, changing oil temperature by cooling the oil. None of the methods gave an acceptable result. Finally reduced bearing to journal clearance applied, which resulted in a lower vibration level. In the field one can deliberately slightly misalign the driver and driven equipment within the acceptable limit. This can produce a radial force on the guide bearing and either minimize or eliminate the vibration due to oil whirl.

V. CONCLUSION

Oil whirl is a self-excited vibration and can affect the rotor

stability. It is uncontrolled and changes to whip at the critical speed. Various factors, such as lightly loaded bearing, excessive clearance, and change in oil properties can produce oil whirl. Proper attention at design, manufacturing and operating stage will prevent oil whirl.

In the field by changing oil properties or adding bearing load will reduce the oil whirl vibration.

VI. REFERENCES

[1] Finley, William R., Hodowanec, Mark M., and Holter,

Warren G., “An Analytical Approach to Solving Motor Vibration Problems,” IEEE Trans. Ind. App., Vol. 36, No. 5, Sept./Oct. 2000, pp. 1467-1480.

[2] Newkirk, B. L., and Lewis, J. F., ”Shaft Whipping due to Oil Action in Journal Bearings,” General Electric Review, 1925, 559-568.

[3] Hagg, A. C. “Oil Whip.” Westinghouse Research Laboratories.

[4] Pinkus, O., “Experimental Investigation of Resonant Whip.” Trans. ASME (1957): 975-983.

[5] Chauvin, D “An Experimental Investigation of Whirl Instability Including Effects of Lubrication Temperature in Plain Circular Journal Bearings”, Master thesis, LSU, 2004.

[6] Nicholas, J “Hydrodynamic Journal Bearings- Types, Characteristics and Applications”, Vibration Institute 20th Annual Meeting, 1996.

[7] Singhal, S; Khonsari, M. “A Simplified Thermo hydrodynamic Stability Analysis of Journal Bearings”, Proceedings of the Institution of Mechanical Engineers—Part J – Journal of Engineering Tribology, 2005, Vol 219 Issue 3, p225

[8] Fredric F. Ehrich; Handbook of Rotordynamics 1992. [9] Machinery Malfunction Diagnosis and Correction by

Robert C Eisenmann, Sr. and Robert C Eisenmann, Jr [10] Fundamentals of Rotating Machinery Diagnostics by

Donald E Bently, Charles T. Hatch, and Bob Grissom

Fig. 14 Orbit plots at Lower bearing

Fig. 15 Spectrum of Lower bearing

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VII. VITAE Sumit Singhal graduated with a BSME degree from Bhilai Institute of Technology, India in 2000 and received a Master of Science degree in Mechanical Engineering from Louisiana State University in 2004. Sumit worked for Center for Rotating Machinery (CEROM) as Research Assistant where he conducted research in an area of Rotordynamic Instability problems. He has been a Mechanical Engineer in the Above NEMA motor development engineering group at Siemens Energy and Automation since 2004. He is a member of IEEE and ASME. He is a reviewer in Journal of Tribology, IMECH: Tribology Transactions and has published several papers related to bearing design and induction motor.

Rajendra Mistry, PE received his B.E. degree in Mechanical Engineering in India and a Bachelor of Technology in Electrical Engineering in the U.K. He is currently a consulting product engineer at Siemens Energy & Automation, Inc. (Norwood) in the engineering development department responsible for developing above NEMA induction motors. In addition to his industry role, he has attended several courses in vibrations, design for manufacturing, concurrent engineering, and digital signal processing. He is a certified vibration analyst; category II and III per ISO/FDIS 18436-2. He is a member of ASME, ASM International, and Vibration Institute. He holds two patents for components in hydraulic elevators.