5 Successful Predictive Vibrations_March_2012

Rotordynamics of PumpsPart I: Single-StageOverhung PumpsBY MALCOLM LEADER, P.E.

VibrationsMarch 2012 | Volume 29, Number 1

For the Vibration Analyst Community

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146273 Vibrations r7.indd 1 3/20/12 11:20 AM

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Thank youIn the 25 years that Judy Eshleman has served as the Editor of Vibrations, it has evolved from a newsletter into the flagship publication for the Vibration Institute and the community it serves. Under Judy’s direction and guidance, and the work of Ron Eshleman as Technical Editor, Vibrations has become the standard for Vibration Analysts throughout the world. Judy and Ron were well suited for this endeavor –Judy had been senior editor for Encyclopaedia Britannica and The Shock and Vibration Digest and Ron was a science advisor for IIT before he became Director of the Vibration Institute

We would like to acknowledge and thank both Judy and Ron for all of their work over the years to make Vibrations the publication it is today.

The Officers, Board of Directors and Staff of the Vibration Institute


VibrationsMarch 2012Volume 29, Number 1ISSN 1066-8268

letter from the President

Successful Predictive Maintenance Programs Require a Skilled AnalystWhen I started in the vibration analysis business in 1974, most people were still using swept sine analyzers for obtaining frequency data. This was also about the time that the first digital Fast Fourier Transform (FFT) analyzers, like the HP 5450 and 5451, were beginning to penetrate the market. The HP 5451A Fourier Analyzer that I used at the Air Force Avionics Laboratory at Wright Patterson Air Force Base had optional, external anti-aliasing filters and consisted of a full rack of equipment, certainly not portable. Analysis of remote data meant collecting data using a tape recorder, identifying data for playback using either a sound track or a time-code signal such as IRIG, and then sitting at the system playing back data segments one at a time for analysis. Thus, widespread data collection and analysis of industrial machinery were not practical, if even possible.

In the early to mid-1980s, with the advent of the personal computer (PC) and very-large-scale integration (VLSI) technology, the birth of portable data collectors and predictive maintenance (PM) as we know them today emerged. VLSI technology provided both the processing speed and downsizing of electronics required to manufacture truly portable data collection devices that made it practical for large- scale monitoring of industrial machinery. Additionally, PCs provided the power to set up machine databases, store large amounts of data, and analyze the data at our desks. Thus, wide-scale condition monitoring as we know it today was born.

It the 1980s and ’90s, several companies manufactured powerful, portable data collectors for collecting route-based data on industrial machinery. Most of these units had off-route or analysis capabilities so that additional readings could be taken and analyzed in the field or stored and downloaded to the PC for further analysis. The interest from Industry in these new systems was high and competition for sales was fierce. I know this from personal experience: I sold these systems during that time. Sales took off because of the potential to reduce unexpected failures, downtime and lost revenue. But, while some companies had success with the new technology, many programs floundered or never got off the ground. Why was such great technology failing in many cases? A key ingredient was missing.

Over the years, technology and industry have done an incredible job of giving us powerful tools to collect and analyze vibration data. Unfortunately, what production

david a. corelli

Vibration Institute President

4 | MARCH 2012 Vibrations

VibrationsVibrations is published quarterly in March, June, September and December by the Vibration Institute. Statements of fact and opinion are the responsibility of the authors alone and do not imply an opinion on the part of the officers or members of the Vibration Institute. Acceptance of advertising does not imply an endorsement by the Vibration Institute.

© Copyright 2012 by the Vibration Institute.All rights reserved. Materials may not be reproduced or translated without the express written permission of the Vibration Institute.

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contents4 letter from the President

– March 2012 – Successful Predictive Maintenance Programs Require a Skilled Analyst By Dave Corelli

6 feature article Rotordynamics of Pumps Part I: Single-Stage Overhung Pumps By Malcolm Leader, P.E.

case histories

14 Bearing Clearance Issue in Blower By Greg Henson

16 Coupling Related Vibration in MG Set By Scott Tilley

20 Resonance: Cutting It Down to Size By Ashok Bhogle

22 newly certified Individuals 2011-2012

23 certification corner Vibration Institute Vibration Analyst Certification Renewal Requirements Explained By Brian Biby and Nancy Denton, P.E.

managers want and need is information, not merely data. They basically want to know two things. First, can they continue to run their machines successfully until the next scheduled outage? Second, what needs to be fixed when the outage occurs so they can line up the parts and people necessary to perform the maintenance in a cost-effective, efficient manner?

The universal question: How do you take vibration data and turn it into useful information? The answer: That requires a skilled vibration analyst. The trained vibration analyst traditionally has been and, in many cases, still is, the missing ingredient in many monitoring programs. Analysts must be trained for this task since it is the only way to make a predictive maintenance program successful, short of hiring a competent vibration consultant.

The stated mission of the Vibration Institute is to “disseminate practical information on evaluating machinery behavior and condition” — in essence, to help people and companies achieve the missing ingredient in condition monitoring programs. The Vibration Institute does this through training, certification and the other products and services the organization provides. Our charter during the 3 years of my term is to improve on this mission.

During my tenure, we will focus on making the Vibration Institute more accessible and helpful to its membership and the entire vibration community. We will strive to appeal to a larger professional base through effective use of state-of-the art electronic media including a new website that was launched in January. We will work to update and improve training, educational materials and our ANSI accredited vibration certification program. We will also work to improve peer-to-peer networking of our members and the vibration community, allowing easy access to the experience of other professionals.

On our new website, you will soon find a searchable database of current and past technical papers available for download free of charge to members in the fall of 2012. We hope you are pleasantly surprised by this issue of the revamped Vibrations which is available as a PDF on our website.

We are in the process of upgrading our training materials with more up-to-date content and state-of-the-art graphics, and that includes our correspondence courses. We will be encouraging all of our chapters to create links from their websites to ours so you can easily find out what is going on in your area.

You will see some surveys coming to your in-box throughout 2012. We want to hear what you want and need from the Vibration Institute. We want to tailor services and the annual training conference to provide the information you seek. Help us put together a bigger and better training conference in 2013 by taking the time to provide us with feedback. And don’t forget to sign up for the 2012 Training Conference in Williamsburg, VA on June 19-22.

The Vibration Institute can help your professional development and performance as a vibration analyst. If you are already a member, I hope you will become more involved; if you are not, please consider joining to enjoy the benefits of membership. Set a goal to become certified during the next year or, if you are already certified, take it up a level to the next category. If you’re thinking, I’ll get started tomorrow, it might never happen. Begin today developing into the missing ingredient for your successful predictive maintenance program.

| 5 MARCH 2012 Vibrations



By Malcolm Leader, P.E.

t he rotordynamics of pumps covers a wide range of machinery types. Pumps are designed to move liquids from point to point and come in hundreds of different

varieties and configurations. This paper will primarily focus on centrifugal pumps, which make up 80 percent of the pumps in service today. Part I covers single-stage overhung pumps. Part II will cover between-the-bearing pumps.

IntROductIOnFor the purposes of this paper, a pump is defined as a device that transports liquid from one location to another. The earliest methods of transporting liquids simply involved scooping up a small volume of liquid in a container and dumping it out somewhere else. Today, continuous liquid transport is done by taking suction from a region of lower pressure, increasing the pressure in the liquid through the addition of mechanical energy, and discharging the liquid, usually through a piping system to the desired location. There are many methods of adding the energy to the liquid to raise its pressure. There are positive displacement types like screw pumps or reciprocating pumps and various other specialty pumps. This paper will discuss centrifugal pumps in single-stage configurations.

The rotordynamics of pumps is similar to other rotating machinery in many ways. They consist of rotors supported by bearings. References [1-8] are provided that contain a significant amount of basic information on rotordynamics

in general. The two primary factors that differentiate the rotordynamics of pumps from other types of machinery are the handling of an incompressible fluid and liquid seals. A seal is any device that restricts liquid flow from an area of high pressure to an area of lower pressure. In pumps, the primary seal that protects leakage to outside the pump casing are mechanical face seals. While vital to safe pump operation, mechanical face seals have very little effect on pump rotordynamics. In some mechanical seals there may be some lateral forces generated (e.g., weight) but compared to impeller eye seals and bushings, these can largely be ignored.

Centrifugal pumps rarely encounter traditional rotor critical speeds unless they are operated in a dry condition. This is because of the liquid annular seals that are used to prevent significant leakage from high pressure regions to lower pressure regions. This pressure differential causes axial flow across the seal. When the rotor deflects, one side of the seal has a greater clearance than the opposite side. The side with the smaller clearance develops a higher local pressure. This creates a restoring force, a stiffness, commonly called the Lomakin effect.

The strength of the Lomakin effect depends on the seal diameter, length, and clearance and the pressure differential across the seal. Other contributing factors are the liquid density, viscosity and inlet swirl ratio. In pumps with

feature article

Rotordynamics of Pumps

Part I: Single-StageOverhung Pumps

Malcolm leader, P.e.Owner of Applied Machinery Dynamics CompanyDurango, CO



many seals, the pump rotordynamics are often much more significantly affected by the seals than by the normal bearings. The bearings themselves can be affected by hydraulic loads, particularly when the pump is operated at an off-design point.

Pumps are also susceptible to structural concerns. Weak or misplaced supports can significantly affect the dynamics of a pump system. Vane-pass pulsations in pumps can also affect the rotordynamics. These higher-order pulsations can cause resonance of the pump rotor and/or the pump structural elements.

Figure 1 illustrates a simple single-stage centrifugal pump. Liquid enters axially into the impeller which is spinning counterclockwise here from this view. The higher pressure liquid exits radially. Figure 2 shows the housing towards the rear that contains the mechanical seal and bearings supporting the shaft connected to the impeller. Note the location of the impeller eye wear ring.

SMall SIngle-Stage OVeRhung PuMPSSingle-stage overhung pumps (SSOH) are ubiquitous. A typical chemical plant or refinery might have more than a thousand such pumps. These can range from small to very large. For the sake of limiting the discussion, a “typical” 100 hp motor-driven single-stage pump with rolling element bearings and a speed of 3,580 rpm was selected. The rotating element weighs 73 pounds, including the half-coupling as shown in Figure 3, which is a finite element model of the

pump rotor.

This generic pump rotor example is completely fictitious but illustrates the dynamics of such machines. The shaft and the impeller has both an inlet-eye wear ring seal and a hub ring seal on the back side. Generally, only high-performance pumps have a hub seal like this. Bearings are located at the springs in the finite element model. Because of the overhung load of the impeller, the coupling end bearing is loaded up 62 pounds and the bearing closest to the impeller is loaded down 135 pounds. One common misconception is that rolling element bearings are extremely stiff. The equation [9] for the direct stiffness of an angular-contact deep groove ball bearing is:

32,500(DFZ2cos5α)1/3

Where:D = Ball Diameter (inches) F = Radial Force (pounds) Z = Number of Balls α = Contact Angle (degrees)

For the bearings assumed for this example, the coupling end bearing stiffness is 480,000 lb/in and the coupling end bearing has a calculated stiffness of 630,000 lb/in. The damping in rolling element bearings is not zero, but it is very small, on the same order as structural damping, which is generally assumed to be 1-2 percent of critical damping.

figure 1. cutaway View of Simple Single-Stage centrifugal Pump

figure 2. drive end View of a Small Pump

figure 3. example Single-Stage Overhung Pump Rotor

ImpellerEye WearRing



The actual bearing loads can be affected by hydraulic loads from the impeller/fluid interaction. The hydraulic load depends on both the pump design and how it is operated. Pump impellers operate inside a volute, which can be a single cutwater or a double volute or a diffuser. The pump is designed to have low hydraulic reaction force at the best efficiency point (BEP). When operated at any flow or pressure condition away from the BEP, side loads are generated on the impeller that, in turn, load the bearings. Depending on the hydraulic load and the direction of the load, this can add to or subtract from the gravity load on the bearings. The Hydraulic Institute publishes a methodology for calculating the impeller loads. Since this is a very complex subject, for the purposes of this example we are going to ignore the hydraulic load effect.

The first critical speed of an overhung rotor will be pivotal, as illustrated in Figure 4. The maximum radial motion is at the impeller eye with a nodal point between the two bearings. Because there is large amplitude at the wear ring, any stiffness and damping effect generated by this seal will have an effect on this resonance. Note that, in this relatively rigid body mode, there is very little bending in the shaft at resonance. On most single-stage overhung pumps, the first critical speed is above operating speed. This does not mean that this resonance cannot affect the vibration of the pump.

Figure 5 shows the mode shape of the second critical speed. Here the impeller has almost no motion. This means that stiffness and damping from the eye-seal would have minimal effect on this mode. This critical speed is usually many times the operating speed of the pump. However, if this resonance were to coincide with a system frequency like vane-pass frequency, it could cause high vibration at that frequency. This phenomenon has been observed in pumps in the field, especially when they are operated off-BEP, which tends to magnify the vane-pass pulsations. If the internal construction in a pump is not optimized, vane-pass frequency can be magnified.

Modeling a small SSOH pump without the eye-seal effects and without a casing substructure will not give realistic results. When the unbalance response is calculated, the result will be similar to Figure 6. This gives the impression the critical speed is lurking just above operating speed when in fact it is not. If just a flexible pedestal is added, the predicted critical speed could drop down to operating speed, which does not happen in practice unless the pump is run dry. Thus the need to understand seals.

SealS and the lOMaKIn effectLiquid annular seals are designed to restrain liquid flow from the discharge area of the pump to the suction area (eye seals) and the area behind the impeller (hub seals). These seals are almost always plain flat rings, as this is a simple and efficient design. Lomakin [10] explained that these seals work because of the pressure drop across the seal creates an apparent stiffness and damping restraining force. This occurs when the rotating part of the seal moves eccentrically in the annulus. On the side with the greater radial clearance, the pressure will be reduced and an apparent restoring force will be generated on the side with the smaller clearance.

Some damping is also generated by this type of seal. The pressure drop across the seal from one end to the other is the same at the narrow gap as at the large gap. Thus, a higher flow velocity occurs in the higher gap than in the lower gap. Bernoulli’s equation tells us that this gives higher pressure applied at the smaller gap than at the larger gap. The net force

figure 4. first critical Speed Mode Shape of SSOh Pump

figure 5. Second critical Speed Mode Shape of SSOh Pump

figure 6. Predicted unbalance Response of SSOh Pump without Pedestal and Seal effects



from these two pressures causes a centering direction force. The restoring forces are proportional to the displacement, velocity and acceleration. In this respect, the seals resemble fluid-film bearings. However, the governing equations for an annular seal are quite different from the “normal” turbulent Reynolds equations used in a bearing analysis. Quite a few researchers have refined the work of Lomakin, most notably Black and Jenssen [11], who applied bulk flow analysis, and later by Childs [12], who applied Hirs’ lubrication equation, which includes the influence of fluid inertia terms and inlet swirl. The mathematics involved in these analyses is fairly complex. The details are readily available from many Internet sources.

Per Lomakin, if the impeller perturbations are small, the radial stiffness (KR) of a seal can be estimated if the pressure drop (ΔP), radial clearance (C), length (L) and diameter (D) are known. The more refined methods give better results and should be used. Thus the longer the seal, the bigger the diameter of the seal, the larger the ΔP and the smaller the clearance, the stiffer it will be.

KR = 0.2(ΔPDL/C)

In practice, the stiffness and damping generated by a liquid annular seal are a function of the geometry (length, diameter), fluid viscosity and density, and pressure drop. The inlet-swirl ratio is usually assumed to be 0.5 unless swirl brakes are used. Reducing the inlet swirl through the seal will reduce the cross-coupling generated and increase stability, although pump instabilities are relatively rare compared to machines handling compressible fluids like compressors.

Taking an example eye-seal for the pump in Figures 7 and 8, assume a diameter of 12 inches, an axial length of 1 inch, hot water with a density of 8.821 X 10-5 lb/in3 and a viscosity of 3.307 X 10-8 Reyns, an inlet swirl ratio of 0.5 and a radial clearance of 0.01 inches. The pressure drop across the seal is assumed to be a function of speed squared. Assuming the full-speed BEP pressure drop is 200 psi, and using the method of Black and Jenssen, Figure 7 is the calculated stiffness and Figure 8 is the calculated damping. Using Childs’ method, the direct stiffness is 30 to 40 percent higher and the cross-coupled stiffness and direct damping are about 8 percent higher.

SSOh PuMP wIth Seal and PedeStal effectSWhen the seal and pedestal effects are included, the model changes to look like Figure 9. Here the numbered black boxes represent the pedestal mass, stiffness and damping. One percent of critical damping (CC) was used where CC = 2(KM)½

This model will show that the critical speed is eliminated, as shown in Figure 10. However, this is for synchronous unbalance as the forcing function. This curve is really valid only up to the operating speed of the pump (3,580 rpm).

figure 7. typical Seal direct Stiffness

figure 10. Predicted unbalance Response of SSOh Pump with Seal and Pedestal effects

figure 9. complete Model with flexible Pedestals and Seals

figure 8. typical Seal direct damping



ReSPOnSe tO Vane-PaSS fRequencyTo consider the effect of vane-pass frequency on the example pump rotordynamics, a 5X tracking force was applied at the impeller location and the speed was varied from 2,000 to 4,000 rpm meaning that the forcing function varied from 10,000 to 20,000 cpm. Since any system like this does have some resonances, any natural frequency in this frequency range will be excited, as Figure 11 shows. It doesn’t take too much imagination to realize that, with the right pedestal stiffness characteristics, vane-pass could easily “tuneup” a structural resonance. Indeed, this author has seen this phenomenon more than a few times.

RefeRenceS lISt1. Nicholas, J.C., and Barrett, L.E., “The Effects of Bearing Support Flexibility on Critical Speed Prediction,” ASLE Transactions, 29 (3), July 1986

2. Leader, M.E., “Practical Rotor Dynamics,” Proceedings of the Vibration Institute 26th Annual Meeting, June 2002

3. Leader, M.E., “Rotor Dynamics as a Tool for Solving Vibration Problems,“ Proceedings of the 27th Vibration Institute Annual Meeting, July 2003

4. Leader, M.E., “Rotordynamics of Semi-Rigid and Overhung Turbomachinery”, Proceedings of the 28th Vibration Institute Annual Meeting, July 2004

5. Kirk, R.G. and Gunter, E.J., “The Effect of Support Flexibility and Damping on the Synchronous Response of a Single Mass Flexible Rotor,” ASME Journal of Engineering for Industry, 94(1), February 1972

6. Nicholas, J.C., Whalen, J.K., and Franklin, S.D., “Improving Critical Speed Calculations Using Flexible Bearing Support FRF Compliance Data,” Proceedings of the 15th Turbomachinery Symposium, Texas A&M University, pp. 69-80, 1986

7. Leader, M. E., “A Solution for Variable Speed Vertical Pump Vibration Problems, Proceedings of the 2nd International Pump Symposium, Texas A&M University, 1985

8. “Rotordynamics Tutorial: Lateral Critical Speeds, Unbalance Response, Stability, Train Torsionals, and Rotor Balancing”, API RP684, American Petroleum Institute, August 2005

9. Chen, W. J., and Gunter, E. J., “Introduction to Dynamics of Rotor Bearing Systems”, Trafford Publishing, 2005

10. Lomakin, A.A., Calculation of the critical speed and the conditions to ensure dynamic stability of the rotors in high pressure hydraulic machines, taking account of the forces in the seals (in Russian). Energomashinostroenie, 14, No.4, pp. 1-5, 1958

11. Black, H. F. and Jenssen, D. N., “Dynamic Hybrid Bearing Characteristics of Annular Controlled Leakage Seals,” Proc Instn Mech Engrs, Vol. 184, pp. 92-100, 1970

12. Childs, D. W., “Dynamic Analysis of Turbulent Annular Seals Based On Hirs’ Lubrication Equation,” ASME Journal of Lubrication Technology, Vol. 105, pp. 429-436, 1983.

13. Corbo, M.A., Malanoski, S. B., “Pump Rotordynamics Made Simple,” Proceedings of the 15th International Pump Symposium, Texas A&M University, 1998

14. Leader, M. E., “Introduction to Rotordynamics of Pumps without Fluid Forces,” Proceedings of the 1st International Pump Symposium, Texas A&M University, 1984

Malcolm Leader, P.E. owner of Applied Machinery Dynamics Company in Durango, Colo., has been a turbomachinery consultant for 25 years. Focusing on providing practical solutions, he specializes in lateral rotordynamics including bearing and seal optimizations and

steady-state and transient torsional analyses.

Leader holds BSME and MSME degrees from the University of Virginia. He is a Certified Vibration Analyst Category IV by the Vibration Institute and is a member of the Institute’s board of directors. He has more than 28 publications in the machinery and rotordynamics fields and holds one patent. He is an ASME Fellow and a registered Professional Engineer in Texas. He can be reached at [email protected]

figure 11. SSOh Pump Vibration due to 5X Vane-Pass frequency excitation

certification Surveillance

In order to protect the Institute members who have justly earned certification as a Vibration Analyst, the Vibration Institute wants to pursue individuals who falsify Institute certification in any manner.

If you are aware of any instance in which you believe an individual is falsifying their certification status, please call or contact the Vibration Institute immediately.

If you are aware of any individual whom you believe is violating basic ethics, please contact the Institute as soon as possible. Failure to do so degrades the reputation of your certification and the Vibration Institute.

(630) [email protected]

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figure 1. Photo of blower Showing Motor and gearbox


R ecently the utility section of a manufacturing plant site chose to upgrade the vibration supervisory systems on some large blowers installed in 1976. The project was

very straightforward: replace the existing proximity probes and replace the old supervisory system with a new modern system that allows live data to be taken directly from the panel. This would create a reliable system, with no power board or miscellaneous board failures to worry about causing false trips on the blowers. Seems simple, right?

Figure 1 is a picture of the blower units. The blower units provide a high-volume, low-pressure air supply for the site’s fermentation operation. The machines installed in 1976 are single-stage blowers with a 24,000 ACFM rating. The inlet pressure is ambient air and the discharge pressure is 27.2 psia. The blowers are driven with a 1750 hp, 4160V, 1800 rpm motor with a gearbox that increases the speed of the blower to 11000 rpm.

The site’s vibration monitoring system was old enough that, if a component failed, replacement parts might not be available, and the manufacturer no longer had trained or qualified people to work on the systems. A new system was purchased and installed. When qualifying the system on start-up, alarms went

off on several of the units. The probes were placed in the same location as the previous probes. So why were the readings so much higher than before? Were the signature patterns different? What was causing the change?

Bearing Clearance Issue in Blower

caSe Study

by greg hensonElanco/Eli Lilly


The blower uses a pressure-feed oil system. The blower impeller is attached directly to the output shaft of the gearbox. The bearings in the gear case are tilting pad bearings. The machine also has set points that do not allow for start up until certain parameters are met, one of these being oil temperature. In an effort to complete the project, the machine was started as soon as the parameter was met.

Figure 2 illustrates a vibration spectrum that shows the result of this start up. The peak at approximately 50 percent of operating speed was suspected to be a subsynchronous resonance, possibly caused by the low oil temperature. Even before the project, this machine had suspect indications of bearing wear. It was in probable need of inspection soon.

With the signature pattern indicating possible subsynchronous resonance due to excessive clearance, it was decided to warm the oil 5-10 degrees above the initial start up temperature and try to restart the machine. Figure 3 shows the vibration spectrum for that start up. The subsynchronous peak and other associated amplitudes were reduced or gone.

The bearings were inspected and replaced, and at the time of replacement, the bearing nearest the blower was found to have more than a 0.010-inch clearance. This is at least 0.004 inches more than recommended. After bearing replacement, the machine has not experienced any of the previous issues.


Greg Henson received his degree in mechanical engineering from Rose-Hulman Institute of Technology, he spent almost 20 years with PSI and Cinergy, helping to establish their predictive maintenance practices. After leaving the power industry, Henson has been with Elanco/Eli Lilly and Company for the past 13 years, improving

their reliability and predictive maintenance practices across the manufacturing organization. He has had multiple papers published at both local and national conferences. Henson also holds certifications from the Vibration Institute as a Vibration Analyst Category III and SMRP Certified Maintenance Reliability Professional. He can be reached at [email protected] 2. frequency Spectrum of Vibration

mils

(pea

k-pe

ak)

Frequency (cpm)

figure 3. frequency Spectrum of Vibration after Increasing Oil temperature

mils

(pea

k-pe

ak)

Frequency (cpm)


figure 1. Photo of the Mg Set


M y company, Burford Electric Service, was asked to troubleshoot a vibration problem in a customer’s DC generator driven by an AC motor, commonly

referred to as an MG set. The generator provides DC power to a carriage drive at a hardwood sawmill that, in turn, runs logs back and forth across a band saw. The AC motor is rated at 350 hp and 880 rpm and is coupled to the generator with a rigid-style coupling. The motor rotor is supported on both ends with rolling element bearings, and the generator rotor is supported on the outboard end with a rolling element bearing. The drive end of the generator shaft has no bearing support. It is supported by the coupling and the inboard motor bearing. Figure 1 shows the MG set in question. The 350 hp AC motor is located on the left-hand side of the picture, and the DC generator on the right-hand side.

The customer had noticed a lot of vibration in the unit during a coast down after the power was shut off. The MG set is located on the second floor of the mill, and the customer informed us that the whole floor vibrated during the coast down. The customer requested that we check the vibration during a coast down. Since we arrived on-site early, it was decided to acquire vibration readings on the MG set while it was still running under normal conditions.

Table1 lists the overall vibrations levels acquired from the unit while it was operating. These data clearly indicated that the MG set not only had high vibration when it was coasting down but also had high vibration while it was running. Overall vibration levels exceeded 1.2 ips peak velocity at the outboard horizontal measurement point on the AC motor. Over 90 percent of the vibratory energy was occurring at 1x turning speed, which was recorded at 892 rpm using a strobe light.

Coupling Related Vibration in MG Set

caSe Study

by Scott M. tilleyBurford Electric Service



Elevated vibration was seen throughout the motor and the generator, with the highest readings recorded in the horizontal direction

table 1. Mg Set Overall Vibration amplitudesVibration Amplitude

Measurement Point In/sec peakMotor Outboard Horizontal 1.28Motor Outboard Vertical 1.06Motor Outboard Axial 0.27Motor Inboard Horizontal 0.87Motor Inboard Vertical 0.89Motor Inboard Axial 0.22Generator Outboard Horizontal 0.56Generator Outboard Vertical 0.56Generator Outboard Axial 0.15

Table 1 shows the overall vibration levels were much higher than recommended for reliable operation. This supported an initial diagnosis that damage or excessive material build-up in the rotor body might be causing imbalance and amplifying the vibration. To investigate the validity of this diagnosis, cross- channel phase readings were obtained at different locations on the motor and generator. The phase readings began telling an interesting story.

The phase across the coupling in the radial direction was between 160 and 170 degrees. The radial vibration was out-of-phase, prompting a closer investigation of the coupling. By means of a strobe light, the display of coupling rotation was slowed down without completely freezing the shaft. Allowing the display of the shaft to continue turning while viewing it with a strobe light can help reveal excessive amounts of run-out or eccentricity. As the rotation of the shaft was slowed down with the strobe light, it was obvious the generator shaft and coupling were not running true with the motor shaft and coupling.

The unit was shut down. A dial indicator placed on the couplings to measure the amount of radial run-out revealed a large amount of coupling run-out (0.075 inch) on the generator as well as run-out on the motor coupling (0.007 inch).

Figure 2 is a picture of the coupling. It is a rigid-style coupling that bolts up tight. Apparently, over time, wear on the inside of the coupling had allowed the generator coupling and shaft to move, resulting in the excessive vibration amplitudes.

Based on this study, we recommended that the coupling be repaired or replaced. We advised that the coupling could probably be “reset” into its proper position, which would likely lower the vibration amplitudes. The customer agreed to reset the coupling and asked us to perform the work. To reset the generator coupling and shaft, we measured the coupling run-out on the generator and put the high point at top dead center. The next step was to loosen all of the coupling bolts until

they were completely backed off the lockwasher, revealing that all the bolts were already slightly loose, which may have contributed to the onset of the vibration problem. Once all the coupling bolts were completely loosened, the generator coupling and shaft dropped into its proper position and the coupling bolts were retightened. With the coupling bolts tight, the radial run-out was remeasured with a dial indicator. The motor and generator couplings run-outs were both recorded at 0.002 inch. We were confident this adjustment would lower the vibration amplitudes toward acceptable levels. After the work was completed, the MG set was restarted. And, in fact, the overall vibration amplitudes were much lower.

Table 2 compares vibration levels before and after the work. The highest vibration amplitudes dropped from over 1.2 ips to less than 0.1 ips after the excessive run-out was removed. The overall vibration amplitudes were much lower at all data locations once the excessive coupling/shaft run-out was removed from the generator.

table 2. Mg Set Vibration amplitudes before and after corrections Vibration Amplitude In/sec – peakMeasurement Point BEFORE AFTERMotor Outboard Horizontal 1.28 0.08Motor Outboard Vertical 1.06 0.09Motor Outboard Axial 0.27 0.06Motor Inboard Horizontal 0.87 0.07Motor Inboard Vertical 0.89 0.05Motor Inboard Axial 0.22 0.05Generator Inboard Horizontal 0.56 0.07Generator Inboard Vertical 0.56 0.05Generator Inboard Axial 0.15 0.05

Figures 3 and 4 show the frequency spectra and time waveforms of data acquired from the outboard horizontal location on the motor before and after the work. Note that the vibration scale in Figure 3 is different from the one in Figure 4.

figure 2. Photo of the Mg Set Rigid coupling


This case history demonstrates that, along with analysis of the vibration signature, utilizing some fairly simple diagnostic tools such as phase angle, a strobe light and a dial indicator will help find and correct many vibration problems.

Scott M. Tilley has performed field service work in vibration analysis, balancing and laser alignment for Burford Electric Service, an electric motor repair shop in Columbus, MS, for 14 years. A graduate of Mississippi State University, Tilley earned his initial Vibration Analyst Certification with the Vibration Institute in 2001, and

Category III certification in 2002. Mr. Tilley can be reached at [email protected].

figure 3. ac Motor Vibration before corrections upper Plot: frequency Spectrum lower Plot: time waveform

figure 4. ac Motor Vibration after corrections upper Plot: frequency Spectrum lower Plot: time waveform

Route Waveform 07-Oct-11 09:03:15 RMS = .2355 PK(+/-) = .4796/.4239 CRESTF= 2.04

0 0.3 0.6 0.9 1.2 1.5

-0.6

-0.4

-0.2

0.0

0.2

0.4

0.6

Time in Seconds

Acc

eler

atio

n in

G-s

SMT - MG SET01 -MOH MOTOR OUTBOARD BRG HORZ

Route Spectrum 07-Oct-11 09:03:15 OVERALL= 1.28 V-DG PK = 1.27 LOAD = 100.0 RPM = 892. (14.87 Hz)

0 20000 40000 60000 80000

0

0.4

0.8

1.2

1.6

Frequency in CPM

PK

Vel

oci

ty in

In/S

ec

Freq: Ordr: Spec:

892.24 1.000 1.259

Route Waveform 28-Oct-11 11:01:35 RMS = .0399 PK(+/-) = .1223/.1102 CRESTF= 3.06

0 0.3 0.6 0.9 1.2 1.5

-0.18

-0.12

-0.06

0.00

0.06

0.12

0.18

Time in Seconds

Acc

eler

atio

n in

G-s

SMT - MG SET01 -MOH MOTOR OUTBOARD BRG HORZ

Route Spectrum 28-Oct-11 11:01:35 OVERALL= .0865 V-DG PK = .0862 LOAD = 100.0 RPM = 899. (14.99 Hz)

0 20000 40000 60000 80000

0

0.02

0.04

0.06

0.08

0.10

Frequency in CPM

PK

Vel

oci

ty in

In/S

ec

Freq: Ordr: Spec:

898.43 .999 .06559


Recertification Requirements

Certification as a Vibration Analyst is valid for five years from the date of current certification level. After five years, and in compliance with ISO 18436: Part I, certified Vibration Analysts are required to recertify. Re-certification at the current level of certification can be achieved in one of two ways:

Renewal. You may provide evidence of continuing education experience, training and/or technical activity. Points for renewal can be earned for vibration-related activities including work experience, professional development, attending industry, association or chapter meetings, and vibration-related presentations and published articles.

Re-examination. You may take the certification exam at the level you are currently certified. This requires scheduling an examination and securing a proctor per established Vibration Institute protocol.

Vibration Analysts are certified on the basis of ability to function at a specified level. The motivation for re-certification is to ensure that the Vibration Analyst maintains the capability to function at the level certified.

Points toward recertification can be earned in various ways. The Vibration Institute Certification Committee has approved renewal requirements as follows:

Category I: 24 points (beginning January 2011)Category II: 28 points (beginning January 2011)Category III: 32 points (beginning January 2013)Category IV: 36 points (beginning January 2014)

Visit www.vi-institute.org and click on Certification to learn more about earning points for re-certification!


Tel: (610) 415-0412Fax: (610) 415-0413

[email protected]

Rotor Bearing Technology & Software1041 West Bridge StreetPhoenixville, PA 19460, USA

RBTS, Inc.

TM

ARMD

COMPREHENSIVE ENGINEERING SERVICESRBTS offers professional engineering services in rotating machinery dynamics, bearings, bearing systems and seals. Rotating machinery services offered by RBTS include:

Field Vibration Measurement and AnalysisTorsional Critical Speeds, Mode Shapes and Response of Drive Trains Including Reciprocating Machinery, Synchronous and Variable Speed Drive Motors

Rotor Dynamics (Natural Frequencies, Stability, Critical Speeds & Response)Bearings, Bearing Systems, and SealsFailure AnalysisDesign Audit

THE MOSTCOMPLETE WAY TO

EVALUATE ROTATINGMACHINERY DYNAMICS

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h igh vibrations due to resonance in variable speed machines are not uncommon in industry. Simple solutions are available to correct these issues. A resonance problem in

the motor support of variable speed pumps was identified by the Condition Monitoring Team at the Minara Resources Limited Murrin Murrin site in Western Australia, whose primary business is to produce nickel, cobalt and ammonium sulphate from nickel laterite ore. The six first-stage pumps in the Ore Leach Plant were all displaying similar vibration characteristics.

Figure 1 illustrates one of the pumps. The pumps are pulley- driven with a variable speed motor, with a pulley ratio of 1:1. The motors are supported by four threaded rods extending from a fabricated steel base. (Note that the picture of the pump was taken at the conclusion of this study, after modifications had been made. The threaded fasteners supporting the motor were longer before modification.)

Figure 2 shows a survey of vibration levels versus speed that indicates the motors were probably resonant between speed ranges of 1024 rpm and 1278 rpm, with vibrations reaching 63 mm/sec (2.48 ips rms).

Figure 3 is a frequency spectrum of vibration at the most sensitive speed. It is dominated by response at the rotation speed of the pump and motor.

The conclusion was that the length of the threaded fasteners caused the stiffness of the structure supporting the motor to be low. The natural frequency (fn) of the motor support equals:


Resonance: Cutting It Down to Size

caSe Study

by ashok bhoglePredictive Maintenance Engineer

figure 1. Photo of first Stage Pump

Motor Lowered by 175 mmbelts changed from SPC3150 to SPC 2800



The two options considered for correcting the resonance problem concentrated on increasing the natural frequency by increasing the stiffness of the motor support. These options were:

1. Stiffening the motor support by welding cross braces between the base to which the motor is attached and the fabricated steel base that supports the threaded fasteners.

2. Lowering the motor by 175 mm (6.89 inches) and change the belts. A sketch of this option is included in Figure 4.

The second option was selected for the trial since it did not require major mechanical work. It required only modification of the guard and replacement with a shorter belt when the motor was lowered. Since the type of belt was not modified, the load factors for the belt did not change.

Follow-up vibration surveys confirmed that the equipment was no longer resonant at operating speeds. Figure 5 shows a survey of vibration levels versus speed after modification. Vibration reduction from levels of 63 mm/sec (2.48 ips) to 5 mm/sec (0.20 ips) was a success for the Condition Monitoring Team. The solution was implemented for the rest of the pumps of the same design with resonance issues.

Ashok Bhogle is a predictive maintenance engineer for the Murrin Murrin operations of Minara Resources Limited in Western Australia. He has been actively involved in the field of condition monitoring for over 21 years in the petrochemical and mining industries. He is certified by the Vibration Institute as a Vibration Analyst: Category

III. He can be reached at [email protected].

figure 2. Motor Vibration versus Pump Speed (before Modification)

note: Vertical axis: Vibration (mm/sec)horizontal axis: Pump Speed (rpm)

figure 3. frequency Spectrum of Motor Vibration before Modification

figure 4. Sketch of Modification to Motor Support

figure 5. Motor Vibration versus Pump Speed (after Modification) note: Vertical axis: Vibration (mm/sec) horizontal axis: Pump Speed (rpm)

Motor Lowered by 175 mm, belts changed from SPC3150 to SPC 2800



Newly Certified Individuals

newly ceRtIfIed IndIVIdualS December 1, 2011 – February 29, 2012

categORy I eXaMJuan Carlos Albavera Hernandez, Juan Manuel Almeida Lugo, Jose De Jesus Avina Cueto, J. Jesus Ayala Medina, Noel Belmonte Gonzalez, Sergio GPE Bio Felix, Ismael Damian Calderillo, Sergio Gabriel Calvillo Rodriguez, Keith Camara Loria, Luis Antonio Camerana De La Fuente, Ing. David Campos Villagomez, Delfino Carmona Ramirez, Jose Manual Augusto Carreto Fernandez, Ricardo Everado Carrillo Napolez, Alejandro Cerda Lopez, Eric Cervantes Cruz, Angel Enrique Chavez Valdes, Erick Paul Contreras Rivera, Ivan Cortes Pelaez, Jorge Jesus Damian Loyola, Ricardo Delgado Arellano, Jose Luis Dominguez Mendez, Naveen Avil Oswald Dsouza, Hector Duran Novelo, Ricardo Elias Espino Osuna, Keith Falcao, Rafael Guadalupe Flores Revoreda, Eloy Alejandro Flores Ruiz, Eloy Alberto Garcia Villanueva, Juan Carlos Gutierrez Bautista, Jorge Antonion Herrera Cahuich, Eugene Lai, Roberto Christian Lara Roldan, Ing. Alejandro Lopez Ramirez, Hemanath Manohar, Jesus Alberto Mariscal Garcia, Victor Hugo Mata Ibarra, Armando Mendoza Nava, Jorge Armando Michel Diaz, Juan Manuel Monroy Violante, Jose Oscar Moreno Cortes, Juan Pablo Naarro Orozco, Antonio Esteban Neria Merino, Juan Manuel Olvera Olvera, Juan Osorio Ontanon, Alfredo Ortiz Oviedo, Ruben Pecero Sobrevilla, Jazz Keith Peters, Hector Efrain Prado Sanchez, Alfredo Rivera Feregrino, Jesus Enrique Rodriguez Campos, Carlos Roberto Rodriguez Cardenas, Javier Alejandro Rodriguez Govea, Angel Antonio Salas Perez, Ing. Eleuterio Santiago Ramirez, Mike Skvarka, Oscar Solar Campa, Brian Tafolla, Gustavo Adolfo Valenzuela

Cueva, Abraham Vidal Peralta, Christian Henry Villadonga Guzman, Pedro Zapata Menchaca, Jose Arturo Zea Monhera, Jose Arturo

categORy II eXaMDan Addington, Michael Akin-femi Akinuli, Nasser Hamed Al-Hinai, Amr Hassan Ali, Yousuf Mohammed Al-Jabri, Nasser Sulaiman Nasser Alkhanjari, Tariq Khalifa Ahmed Al-Sabahi, Bader Monsour Salim Al-Salmi, Chinniah Amburose, Doug Ayers, Ghouse Shahinshah Basha, Gautam Nikhilchandra Bhattacharya, Andrew Boggs, David Bray, Neil Brooks, Andres Caceres Lara, Justo Canales Ferre, Brian Cary, Carolina Izquierdo Castilla, Henry Cruz Huiman, Justin Davis, Steve Deases, Serief Atef El Beshlawy, Osama Elbshir Elnahrawy, Beau Fulford, Barry Gallant, Juan Jose Gambos, Suryakant Vithal Gawde, Charles Lee Henderson II, Mohamed Hussein Abdel Rahman Hussein, Kuddush Syed Ibrahim, Jobin P. Jacob, Dinesh Bhawarlal Jain, Balasubramanian Krishnakumar, Sujit Kshatri, Pathakota Sudheer Kumar Reddy, Luis Torres Lagos, Arshad Mansoor, Hecner Merino, Youssef Ibrahim Mikhail, M. Meeran Mohideen, Abel Ortega Mollar, Muhamad Saiin Mustofa, Thangavelu Muthukumaran, Abdulrahman Noordeen Shamsudeen, Christopher Olsen, Joseu Rosemberg Coutino Ozuna, Alex Rogelio Pachas Sulca, Pratap Ramrao Patil, Richard Earl Pratt, Alex Quispe Quispe, Sudhar Rajagopalan, Palanisamy Ravi, Imanel Kostany Rebello, Purushotham Reddy, Wilmers Cruz Rodriguez, Rafael Rozo, Mohamed Salahudeen K.H., Ghilmar Jhonatan Santos Chauca, Manish Kumar Saxena, Lawrence V. Seger, Scott Self, Murugesan Sellapillai, Ramasany Senthil Kumar, Rajeev Rajkumar Singhai, Phil Slifkin, Kevin

Small, Mark V. Stokley, Scott Stranford, Edwin Eugene Todd Jr., Jose Torres Mejia, Subramanian K. Vaidyanathan, Vijayakrishnan Venugopal, Mukesh D. Vyas, Manojkumar B. Wagh, Tiffany Ann Ward, Dan West, Prasada Rao Yaramatti, K.S. Mohamed Yasar Ali, K.S. Mohamed

categORy III eXaMAmr Hassan Abayazeed, Mitchell Anderson, Bernard Boueri, R. Dennis Conroy, Shawn Covington, ShawnMichael Feuser, Ahmed Sayed Mohamed Ismail, Roshan Joseph, Jeff Kenney, Steve Kouma, Sudalaimuthu Muthusamy, Brian Pae, Dakshina Murthy Ram Sekar, Guna Sager Malla, Ramesh Seshan, Muthusamy Vaikundam,

categORy I balancIng eXaMFredy Sansom

2011-2012



Certification Corner

by brian bibyArcelorMittal

by nancy denton, P.e.Professor, Mechanical Engineering TechnologyPurdue University

abStRactThe renewal process within the Vibration Institute’s Vibration Analyst certification program is in place to offer a path to recertification of certificants who are competent and current in their knowledge at a specific level without the burden of taking another examination. There have been recent changes to the Vibration Institute’s Vibration Analyst certification renewal form and underlying policy. These changes are designed for continued compliance with the intent of the renewal process, to be more equitable to certificants, and to continue to satisfy the needs of industry and other stakeholders. This article explains the motivation behind the changes to the renewal process and how they will affect certificants.

bacKgROundThe Vibration Institute is accredited by ANSI to operate in accordance with ISO/IEC 17024 [1] as a third-party certification body for Vibration Analyst classification. The certification scheme used is based on ISO 18436 [2, 3], which states that certificates issued may be valid for no more than five years. The Vibration Institute vibration analyst scheme complies with Parts 1 and 2 of the ISO 18436 standard; Part 2 is specific to Vibration Analyst classification, while Part 1 contains additional language regarding the requirements for certification bodies that deal with the ISO 18436 “Condition monitoring and diagnostics of machines” family of standards. These standards come with requirements for clear audit processes, separation of certification and training programs, definition of the relevant body of knowledge and specific certification exam criteria.

In compliance with these standards, the Vibration Institute has long offered a methodology to renew vibration analysts’ certificates without requiring the certificant to recertify by

examination, providing certain criteria are met. The criteria are primarily continued work in the field without significant interruption, ongoing ethical behavior and verifiable evidence that the certificant has engaged in activities that serve as evidence the certificant continues to comply with the current certification requirements. Some of this evidence is obtained through proactive surveillance activities; some evidence is documented and declared by the certificant on the renewal application. To date, no specific training or examination in specific Body of Knowledge (BOK) subjects is required for renewal.

The purpose of any personnel certification scheme is to provide value to industry and stakeholders by certification of individuals to their various classifications based upon a standardized body of knowledge, job description, and task summary for each. Certification to such an established standard or other set of criteria offers confidence that certified individuals can adequately perform all the job functions safely and correctly, and require no further training to do so. Certification must serve both the industry and the stakeholders; therefore, a certification body must be responsive to industry, stakeholders and the standards.

MOtIVatIOn fOR changeFeedback received by the Vibration Institute and its vibration analyst scheme committee revealed a potential disparity in renewal point requirements between the different classification categories of vibration analysts. The same 30-point requirement was being applied to all four classification categories. This was a simple implementation, but based on the experience, training and depth of knowledge a certificant in each category must possess, clearly too much was being asked of entry-level certificants and too little of more advanced certificants. In addition, confusion about the interpretation of some of the point-eligible activity listings on the renewal process form needed to be addressed.

Renewal POlIcy changeIn response to industry and stakeholder needs, the Vibration Institute vibration analyst scheme committee’s diverse, international and seasoned subject matter experts collaborated to resolve this disparity. (Figure 1 offers more information about the Vibration Analyst Certification Scheme Committee). The solution is a tiered renewal-point requirement based on certificant category, where increasing point values apply as certificants

Vibration Institute Vibration Analyst Certification Renewal Requirements Explained


move to more advanced categories. Although certification category may have little effect on vibration-related experience points, category does have a marked effect on the amount of participation in technical activities a vibration analyst must exhibit to demonstrate continued compliance with certification requirements. The revised point requirements of 24, 28, 32 and 36 based on certification category are being rolled out in progressive stages and are being advertised. The category I and II reductions are already in effect; the Category III and IV increases will be effective at the beginning of 2013 and 2014, respectively.

figure 1. Vibration analyst Scheme committee Roles and Responsibilities

Vibration analyst (Va) certification Scheme committee Roles and Responsibilities

The VA Scheme Committee is a group of approximately 25 Category III and IV certificants who are Vibration Institute members. This committee is responsible for the development, ongoing review, and maintenance of the certification scheme. They meet regularly to:

• Develop, review and revise vibration analyst certification exam questions

• Propose certification policies and procedures to the board of directors

• Ensure compliance with applicable standards

• Address any ethics and/or misconduct issues related to exam takers, certificants and proctors/invigilators

Questions about certification and the renewal process should be submitted Karen Bresson, Vibration Institute executive director. Exam applications and renewal forms are available online at www.vi-institute.org or call (630) 654-2254.

ReVISed Renewal fORMA review of the renewal application form layout, instructions and related details was part of the VA Certification Scheme Commitie’s effort to improve the renewal process. Again, the focus of the review is to better serve industry and the certificant while maintaining the integrity of the certification scheme. Table 1 compares key elements of the legacy and current renewal forms. The current renewal form, VI_Form_CF024_R3_2011-07-10 [4], is now available as a fillable PDF document on the Vibration Institute website.

One challenge of the legacy form was its layout. The format of four half-pages has been modified to four full-size pages to accommodate both typed and handwritten entry. Another challenge was the lack of point limits and requirements in each section. The form has been modified to make it clear to the certificant how many points are required and how many points may be earned from each experience and technical activity section. The current renewal form should be quite adaptable

to online application if the Vibration Institute chooses to accommodate online submission.

The term “vibration-related” appears in both Experience and Technical Activities sections of the form. The legacy form’s use of this term without explanation created confusion. On the revised form, it is well-defined for better understanding and usability by the certificant. “Vibration-related” includes all the subjects and topics of the Body of Knowledge and job task analysis. As reminders, some often overlooked BOK-related items such as alignment, mechanical work electrical work, and a host of companion PdM technologies are listed. Each grouping has its own point cap reflecting the job task analysis and the breadth of the Body of Knowledge. The Institute interpretation of “vibration-related” has not been altered, but now the certificant has a much better understanding of its relationship to the Body of Knowledge, companion technologies, specific maintenance and repair corrective activities and particular point limits attributable to these non-core activities.

The distribution of experience points earned was slightly increased to better reflect the value of ongoing vibration-related work experience. Investment of 10 percent of a work year into vibration-related activities is now recognized with 1 point, while an investment of 60 percent of a work year is needed to earn 4 points. A maximum of 20 points may be earned from work experience during the five-year certification period.

The three sections related to Technical Activities are more comprehensible and better spelled-out to the certificant:

• The Documented Training technical activity now contains more descriptive language about the types of training eligible, points earned, frequency and value of points earned for similar activities, and a cap of 16 points earned from Documented Training activities.

• The Meeting and Conference Attendance technical activity now contains more descriptive language about the eligible activities, points earned, frequency and value of points earned for similar activities, annual point caps on the subcategories of meetings and conference attendance, and a cap of 16 points earned from Meeting and Conference Attendance activities.

• The Presentations and Publications technical activity now contains more descriptive language about eligible activities, points earned, frequency and value of points earned for similar activities, and a cap of 16 points earned from “Presentations and Publications” activities.

The update to the renewal application form and its point system has effectively addressed the concerns raised by industry and certificants about the potential disparity and difficulty of renewal for a certificant who in fact had adequate experience and technical activities but was taxed to support his/her efforts appropriately on the form.

The only significant change to the certification renewal policy is the implementation of tiered increasing point requirements,



which better reflect the participation in technical activities a vibration analyst must exhibit at each classification category to demonstrate continued compliance with certification requirements. The interpretation and evaluation of items on

the renewal application by the Vibration Institute remains essentially the same. What has changed is the ability of the certificant to better understand the criteria and be more aware of the eligible activities and their limits.

Legacy Renewal Policy/Form 2011 Revised Renewal Policy/FormForm Layout

Four pages, 5.5” x 8.5” size, application format.Paper fill-in format only; difficult electronic use.

Four pages, letter size, application format.Paper and electronic fill-in format.Easily adaptable for online submission.

Form Summary and AttestationUnclear, easy-to-miss subcategory items.Single, non-dated, certificant signoff.

Clear and concise; easy to understand layout.Certificant verification initials and date each page.

Vibration-Related Experience25% of work/year to earn 1 point/year.50% of work/year to earn 2 points/year.>50% of work/year to earn 4 points/year.20 points maximum over 5 years.Significant interruptions not noted on form.

10% (220 hours) work/year to earn 1 point/year.20% (440 hours) work/year to earn 2 points/year.40% (880 hours) work/year to earn 3 points/year.60% (1,320 hours) work/per year to earn 4 points/year.20 points maximum over 5 years.Disqualification for significant interruption noted.

Technical Activity - Documented TrainingPoints earned stated, but not for which activities.Underlying limit of similar, repeated activities only valid once

Very limited space for documenting activities.No cap on points earned by documented training.

Points earned stated clearly per activity.Clear limits on repeated activities, with some relief for multiple

Ample space for documenting activities.Cap of 16 points by documented training.

Technical Activity - Meeting and Conference AttendanceUnclearly defined point-earning activities.Silent on point limits of similar, repeated activities per time

Vibration-related committee work status unclear.Very limited space for documenting activities.No cap on points earned in each subcategory; No cap on

Clearly defined point-earning activities.Clear direction on points earned from repeated activities per

Vibration-related committee work included in scope.Ample space for documenting activities.Per-year caps on points earned in each subcategory; Cap of

Technical Activity - Presentations and PublicationsPoints earned per original presentation and original article are

Technical publication clearly defined.Silent on points limit of similar, repeated activities and on

Very limited freeform space for documenting activity.No cap on points earned by presentations and publications.

Points earned per original presentation or original article are

Technical publications are clearly defined.Clear stated point limits for similar, repeated activities and

Ample space for documenting activities.Cap of 16 points earned from presentations and publications.

Informative Text and InformationFew instructional reminders to aid for completion.Clear appeals and complaints information.Silent on definition of “vibration-related” activities.Silent on companion PdM activities; no point caps.Silent on corrective activities; no point caps.

Instructional reminders to aid in form completion.Clear appeals and complaints information.Clear instruction on local definition of “vibration-related”

Clear inclusion of companion PdM activities and cap on

Clear inclusion of certain equipment corrective activities and

per 5 year period. activities.

interval.

points earned per year; No cap on combined points for meeting and conference attendance. attendance.

time interval.

16 points combined for meetings and conference

clearly stated.

when points are not earned for each.

clearly stated.

clearly stated that no points earned are earned fromrepeated activities.

activities.

applicable points earned due to these technologies.

cap on points earned due to those activities.

table 1. Renewal Of Vibration analyst certification Policy/form comparison



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Certificants will likely find it easier to qualify for renewal using the revised form and tiered point system. However, this should not be seen as lowering standards or sacrificing of the integrity of the certification scheme. On the contrary, responding to the needs of all stakeholders by modifying the required points and better educating certificants about the renewal process produces a more robust and equitable method to assess and verify continued compliance with current certification requirements.

The Vibration Institute remains responsive to the needs of industry and certificants alike and will continue to improve and adapt its policies and practices to better serve stakeholders, comply with applicable standards such as ANSI/ISO/IEC 17024 and ISO 18436, reinforce the integrity of its certification schemes, and remain relevant to all parties. As times change, both standards and stakeholder needs change. The Vibration Institute will respond to these changes, ensuring the Vibration Institute remains a premier accredited certification body.

RefeRenceS1. ANSI/ISO/IEC 17024 Conformity assessment - General requirements for bodies operating certification of persons

2. ISO 18436-1 Condition monitoring and diagnostics of machines - Requirements for training and certification of personnel - Part 1: Requirements for certifying bodies and the certification process

3. ISO 18436-2 Condition monitoring and diagnostics of machines - Requirements for training and certification of personnel - Part 2: Vibration condition monitoring and diagnostics

4. Vibration Analyst Certification Renewal form CF024. http://www.vi-institute.org/assets/1/7/VI_FORM_CF024_R3_2011-07-10.pdf.

5. Certification and Accreditation for Condition Monitoring and Diagnostics; David A. Corelli and Brian G. Biby, 2 pp, Sound and Vibration, November 2010. http://www.sandv.com/downloads/1011edit.pdf.

Brian Biby holds a B.S. in Electrical Sciences and Systems Engineering from Southern Illinois University. For the last 20 years he has focused on condition diagnostic technologies, vibration analysis, and related supporting technologies and systems. Currently employed by the world’s largest steel producer, he holds a Category IV vibration analyst certification, is a

member of the Vibration Institute board of directors and chair of the Vibration Institute’s vibration certification scheme committee. He also holds a Certified Lubrication Specialist certification, is active on and has chaired the NWIBRT Reliability Subcommittee, participates on the AIST Maintenance and Reliability Technical Committee and is a member of SMRP. He can be reached at [email protected]

Nancy L. Denton, P.E., is professor and associate department head for mechanical engineering technology at Purdue University, where her special teaching and research interests are machinery health-monitoring and data acquisition. Her industrial experience includes design engineering for the Naval Avionics Center and acoustical engineering

for Digital Equipment Corporation. A member of the Vibration Institute’s board of directors, she chairs the academic committee, serves on the vibration certification scheme committee and is a Category III Vibration Analyst . She is a member of ASME and an ASEE fellow. She can be reached at [email protected].

One-day tRaInIng SeSSIOnS:Identification and Correction of Resonance Problems

Evaluation and Reliability of Electric Motors KeynOte addReSS:A Town Hall Meeting with Thomas Jefferson

2012 VIBRATION INSTITUTE TRAINING CONFERENCERegister Today for the …

For more information or to register: (630) 654-2254 [email protected] www.vi-institute.org

June 19-22, 2012Williamsburg Woodlands Hotel & SuitesWilliamsburg, VA

addItIOnal SeSSIOnS Include:Acceptance TestingBasic MonitoringElectro Static TechnologyJournal Bearing Monitoring & AnalysisMachine Condition EvaluationMachine Vibrations Standards & Guidelines

Motor Current AnalysisOrbital AnalysisSafetySensor MountingSignal ProcessingTime Waveform Analysis



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5 Successful Predictive Vibrations_March_2012

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