Vibration Certification Case Studies Vertical Pump · PDF fileRush Equipment Analysis, Inc....

RUSH EQUIPMENT ANALYSIS, INC. [email protected] 714-998-7433 2401 EAST SEVENTEENTH ST., #181 SANTA ANA, CALIFORNIA 92705 FAX 714-998-0121

Vibration Certification Case Studies Vertical Pump Machinery

Controlled with Variable Frequency Drives

Copyright REAI 2008 http://www.rushengineering.com/VFD-Pump-MultiChannelIntroduction.pdf 09/07/2008

Rush Allen

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RUSH EQUIPMENT ANALYSIS, INC. [email protected] 951-927-1316 43430 EAST FLORIDA AVE., #F331 HEMET, CA 92544-7210 FAX 951-927-0512

Rush Allen

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Rush Equipment Analysis, Inc. of 101 Introduction Vibration Certification of Vertical Pump Machinery Controlled with Variable Frequency Drives

Executive Summary

The use of variable frequency drives, or variable frequency controllers, in the pumping industry has become very common place over the last five years. Pump manufacturers and pump machinery operators have been plagued with the operation of these machines due to vibration issues made very complex by variable speed operation. This document presents eight case histories of vertical pump machinery that have had difficult certification processes due to the variable speed operation. The objective of the document is to alert operators, machinery suppliers, contractors, and pump station designers of the potential complications and some remedies that REAI or REAI clients have implemented. Additionally, the vibration analysis technologies and the vibration test technologies have evolved such that the operation of variable speed machinery can be successfully implemented at the certification phase when proper steps are implemented early in the design phase of the pump stations and the pumping machinery and followed up at the equipment startup certification phase. Design analysis of machinery using finite element analysis has been implemented for over 30 years and has been integrated into the personal computer work station for over 15 years. By appropriate finite element models the machinery can be constructed to minimize dynamic problems associated with variable speed operation. Even so, there are situations where the technology hand off from pump station system designers to pump machinery manufacturers and pump machinery operators has resulted is severe start up vibration problems. By the examples presented herein the reader can become aware of the principle issues and some remedial actions. Machinery vibration analysis has witnessed an explosion of electronic devices designed to be used in predictive maintenance of machinery over the last two decades. These devices started out as single channel data collectors and then dual channel data collectors. These machinery vibration analyzers (MVAs) can gather tremendous volumes of data and they can manipulate the data through semi-automated reporting systems for predicative maintenance vibration surveys. As such, these instruments are especially useful for simplifying the startup vibration certification process of new machinery. For well over forty years startups of large machinery have included the use of tape recorders and signal analyzers. The instruments were quite large and expensive and were not of use in the startup of vertical pump machinery due to cost ratio of the machinery and the instrumentation. Over the last five years or so the microelectronic technology industry has provided multi-channel digital signal recorder/analyzers (DSRAs) for use in startup vibration certification. A modern 24 channel digital signal recorder/analyzer will easily fit in a shoe box. Some of these devices can be daisy chained to well over one hundred channels for aerospace and turbo machinery startup tests. No special evaluation of these analyzers is provided herein, although they are employed in all the examples presented. For vertical pump machinery over 500 HP the use of a multi-channel digital signal recorder/analyzer can dramatically reduce the data acquisition time. On very large machines in excess of 2000 HP these multi-channel signal analyzers can act like medical CAT Scan devices for a complete evaluation of the machinery health at relatively low cost and time expended compared to the cost of operation of the machinery. The multi-channel digital signal analyzer can be used to perform resonance evaluations through the operating range without performing complex bump test procedures. Essentially the variable speed operation and pump system transients excite the resonances and the digital signal analyzer captures the complete event. The effectiveness of multi-channel digital signal recorder/analyzers for vertical pump machinery vibration certification is well established by the case studies presented herein.


Rush Equipment Analysis, Inc. Page 25 of 101 CASE II: Vertical Non-Clog Pump Lateral Resonance Vibration Certification of Vertical Pump Machinery Controlled with Variable Frequency Drives

CASE II: Vertical Non-Clog Pump Lateral Resonance SUMMARY This case involved a unique installation of a booster pump in a sewage pipe line. The long pipe line required a means to avoid water hammer damages during transient power conditions. The pump manufacturer designed a flywheel arrangement between the non-clog pump assembly and the 80 HP driver. In the process of designing the driver installation it was determined that a standard pump stand would be insufficient because of the heavy flywheel assembly. The designers employed a double weir arrangement with a spanner sole plate to support the non-clog pump. Unfortunately, the spanner plate caused rocking mode resonance of the pump to be within the 2xRPM speed range. Since the non-clog pump had a two vane impeller the resonance was at the frequency of the maximum forcing function. In this case finite element analysis and detailed testing were employed to establish a remedy that brought the pump to be within the tolerance of the HI Standards. DISCUSSION The appropriate vibration standard is provided in Figure B1 as ANSI/HI Figure 9.6.4.10. The driver support structure for the actual installation is also shown in Figure B1. At the time of the photograph the contractor had installed temporary braces between the sole plate and the pump bearing housing. The figure also shows the locations used for a multi-channel digital data acquisition setup during trial certification and troubleshooting tasks. Figure B2 illustrates the vibration of greatest concern in the orbital movement at the top of the pump bearing housing. This orbit has an rms amplitude of 0.6089 in/sec perpendicular to the discharge nozzle and 0.2905 in/sec rms parallel to the discharge nozzle. The allowable overall vibration is indicated in Figure B1 as 0.30 in/sec rms at 80 BHP. The vibration orbit provides evidence of excess vibration but not the frequency of the vibration, which will be shown below to be 3578 CPM, or twice the shaft speed of 1789 RPM. The narrow shape of the orbit indicates that it is a resonance response with the greatest deflection in the direction of maximum compliance for the vibration inducing forces. Pure misalignment forces or pressure pulsation forces at 2xRPM would not generally produce a narrow vibration orbit because the excitation is not unidirectional. Figure B3 presents an impulse response compliance function (microinch / lb). The classic single degree of freedom resonance characteristic has resonance compliance ten times the static compliance with a peak frequency of 3345 CPM. The 2xRPM critical speed for this resonance is 1672.5 RPM. An operating deflection shape (ODS) study was performed with the results presented in Figure B4. The primary static flexure is the diaphragm deflection of the sole plate which produces 43% of the deflection at the top of the bearing housing. An additional 33% of the ODS is due to flexure of the pump casing. The ODS also shows a theoretical rotor deflection for the measured casing deflection at 24% of the ODS. Both rigid rotor and a flexible rotor deflection shapes are shown in Figure B4, with the flexible rotor deflection projected to the top of the casing for referencing the relative amplitude. Although there were other issues with regard to the machinery vibration, the ODS identifies the essential concern. Subsequently the pump manufacturer had an FEA model prepared and the two pump reed modes are illustrated for the FEA model in Figure B5. The FEA modes are of sufficient accuracy to reveal the vibration problem. The mode frequencies increase due to gyroscopic stiffening and bearing effects.



Figure B6 presents an interference diagram showing two orthogonal lateral modes of the pump casing within the operating speed range on the 2xRPM order line. The model also provided three other resonance conditions that do not have critical speeds for the 1xRPM and 2xRPM orders of rotation. This information should have been sufficient for the manufacturer to redesign the pump mounting configuration prior to installation. However, since the interference is with the 2xRPM order of rotation the possibility that excitation will not occur does exist. The focus of the design engineers was on ensuring that there was no critical speed for the 1xRPM order of rotation. A two vane non-clog end suction pump normally produces 2xRPM pulsations. The strength of the pulsations is dependent upon liquid flow condition and pump operating conditions. Thus, since the pump met the ANSI/HI 9.6.4.10 factory test criterion applied by the manufacturer, it was assumed that it would meet the criteria in the field installation. The logic of this conclusion is faulty because the factory test criterion is not defined by ANSI/HI 9.6.4.10. The standard has the following wording.

The values in Figures 9.6.4.4 to 9.6.4.14 are not applicable to factory or laboratory acceptance tests. Experience has shown that vibration levels measured on temporary factory setups may be as much as two times higher than those obtained in the field.

In this situation the factory test engineers used the wording “may be as much as two times higher” as a factory test standard. The factory test engineers misinterpreted the statement that the standard is “not applicable to factory or laboratory acceptance tests.” The consequence of the subjective interpretation was a very drawn out struggle between the factory, the installation contractors, the plant engineers, and the owner, on the proper certification of the pump as installed. During the attempts to resolve the vibration issues resonance testing was performed and the factory performed the optional lateral dynamics analysis under duress. The purchase specification identified the ANSI/HI 9.6.4 standard but only specified lateral analysis of the rotor shaft and not the pump assembly. After completion of the pump assembly lateral analysis when the machinery did not meet the vibration limit criterion the analysis results were compared to field tests. Figure B7 illustrates the comparisons. The model includes the complete driver assembly with the flywheel and the pump assembly but without any piping. The Original Model showed significant improvement by the addition of factory designed Pump Bearing Braces for the Mode 3 and Mode 4 Pump Assy Rocking or 1st Reed Modes. The Mode 8 Pump Assy Sole Plate Bounce mode was also moved beyond the influence of the 2xRPM energy. In Figure B7, Mode 2 Driver Assy ZY Rocking, or 1st Driver Reed, placed a restriction on the minimum operating speed due to interference at 1xRPM. This critical speed is sensitive to unbalance and cannot be ignored. However, the bump test results indicated that the mode was lower than the analysis with an install critical speed at 1368 RPM. This placed the mode 2.3% below the minimum practical operating speed of the pump at 1400 RPM. Although the specified range was 1350-1800 RPM, the pressure required to move the water in the pipe line could not be obtained below 1400 RPM. As such, Mode 2 was not a critical speed for actual operating conditions, and a compromise could be made for the 1st ZY Driver Reed Mode. Mode 5 and Mode 6 in Figure B7 were difficult to excite as 2xRPM critical speeds and test data indicated that these resonances did not represent critical speeds. However, Mode 7 of Figure B7 was a different matter. This mode represented structural torsion of the driver support frame. Since the pump was a two vane pump it would produce torque pulses at 2xRPM. The bump test data placed the critical speed at 3571 CPM, which was at the very top end of the potential 2xRPM range. The test data indicated that this mode was critical and it limited the upper operating speed of the installation to 1750 RPM with a 2% margin.



Figure B8 illustrates the overall parallel plane vibration at the top of the motor for a rundown of the machinery from 1781 RPM. The rundown data indicates that the gyroscopic stiffening increased the critical speed for the Mode 1 Driver XY Rocking Mode from 1288 RPM for the non-operating bump test data to 1365 RPM with an overall vibration amplitude of 0.65 in/sec rms. At 1400 RPM the vibration was just below the ANSI/HI 9.6.4 allowable value of 0.3 in/sec rms. This set the minimum operating speed and it was acceptable for the pump installation due to pressure limitations that required greater than 1400 RPM to move the water through the pipeline. Good design practice should have reduced in frequency to 1260 CPM or lower. Figure B9 presents the overall pump bearing parallel plane vibration during a run up of the modified machinery. The ANSI/HI 9.6.4 allowable vibration limit was exceeded at 1660 RPM. This severely limited the pump operation and additional fine tuning of the installation had to be performed. Subsequently, the vibration was made acceptable through 1700 RPM. This case has a great number of dynamic considerations that resulted because of the unique requirement of the pipeline for surge protection in the event of power transients. The pump was in a retrofitted booster station that was moving the water through mountainous terrain. Figure B10 illustrates the pipeline operating region between the C=160 and the C=130 curves. These curves indicate the new and old flow restriction curves for the pipeline theoretical head loss. These requirements were fully identified during the purchase phase of the retrofit of the pump station. However, when we compare these requirements to the Preferred Operating Region of the installed pump we can see an immediate conflict. The Preferred Operating Region, or POR, represents the region within which the pump produces minimum turbulence and vibration. Since the pipeline operating region was acknowledged by the pump manufacturer it identifies the manufacturer’s Allowable Operating Region, or AOR. Some standards, such as ANSI/API 610, allow a 25% increase in vibration for operation in the AOR. However, ANSI/HI 9.6.4 does not state allowance increases for operation outside the POR. ANSI/HI 9.6.4 makes the following statement.

If the rate of flow were below an appropriate value, the vibration acceptance standards would not be applicable because the pump was operating below the minimum of preferred operating range. These vibration values are to be used as a general acceptance guide with the understanding that vibration levels in excess of these values may be acceptable by mutual agreement if they show no continued increase with time and there is no indication of damage, such as an increase in bearing clearance or noise level.

Figure B11 presents the two operating region boundaries with actual vibration readings acquired during acceptance testing in the tables. Operation near 1493 RPM resulted in overall vibration at the pump bearings between 0.26 and 0.30 in/sec rms. These levels of overall vibration are marginal for POR operation in accordance with the ANSI/HI 9.6.4 standard in Figure B1. For operation near 1640 RPM the overall vibration at the pump bearings was between 0.28 and 0.59 in/sec rms. As the data (blue squares) in the speed range indicate the vibration was acceptable within the POR and unacceptable outside the POR. Clearly, the pump manufacturer had an argument for acceptance using the POR values. However, the acceptable vibration was only achievable in a closed short line test at the pump station and the C=160 and C=130 curves show that that POR flow conditions could never occur during normal operation of the pump station. If the pump manufacturer had not warranted operation within the pipeline curves there would not have been a dispute. But, that was not the case, for the pump manufacturer was fully aware of the intended head and flow conditions.



The AOR and POR regions for the pump were established using the factory flow-head curves to identify a curve fitted equation representing the data for the 1792 RPM factory test. This curve was then modified using the affinity laws for centrifugal pumps to define the curves for other speeds. Some pump manufacturers provide curves at different speeds; usually the minimum and maximum speeds. But, to determine the proper allowable vibration value in accordance with ANSI/HI 9.6.4 it is necessary to calculate the actual brake horse power as identified in the standard. The brake horsepower and pump efficiency curves are provided on the graphic presentations of Figure B10 and Figure B11. During proper certification for vibration of the non-clog pumps the machinery must be run within the design region for the pump station and not in the design region for the pump. The design region is shown on Figure B11 as the “Allowable Operating Region.” Any operation outside this region is inappropriate for certification because the pump station cannot operate outside the flow-head curves of the pipeline. There were two pumps in the pump station and the certification data acquired for each pump is identified by green squares and the smaller of the two tables of vibration data. The AOR points are identical for each pump. Pump 1 exceeded the allowable vibration limit at 1650 RPM with an overall vibration value at the pump bearing of 0.44 in/sec rms. Pump 2 exceeded the allowable vibration limit at 1500 RPM and 1740 RPM with overall vibration values at the pump bearing of 0.55 in/sec rms and 0.34 in/sec rms. CASE II CONCLUSION In summary, the certification of these pumps took excessive effort because the pump manufacturer failed to perform due diligence when the requirement for the driver resulted in a unique installation for the non-clog pump. The pump manufacturer’s product was clearly not at fault. However, the design of the system led to modifications that had unanticipated results. It is not uncommon to find unique pumping systems designed in the field because the actual operating conditions with variable frequency drives and pump station load conditions are difficult to simulate in a factory. That is why the ANSI/HI 9.6.4 says that the “values in Figures 9.6.4.4 to 9.6.4.14 are not applicable to factory or laboratory acceptance tests.” The pump industry is an old industry and the products are essentially sold like commodities with minimal engineering. When special engineering is required beyond the application engineering criteria, the pump company resources can be severely stretched because one of a kind installations do not fit into the minimum price commodities market for pumps. The finite element analysis that was forced by test results should have been performed prior to fabrication. Even if it had been performed prior to fabrication, the pump manufacturer would have very likely ignored the 2xRPM interference and the possible consequences of a critical speed within the pump operating range. The reason is that the pumps are designed to have minimum vibration in a well defined Preferred Operating Region where pulsation is minimal. This case is an excellent example for establishing the actual operating point prior to fabrication and utilizing that knowledge to predict the potential for interference speeds that might otherwise not be critical speeds. A critical speed is not a speed where a resonance exists. A critical speed is a speed where resonant forces occur that result in excessive vibration. All machines have resonance, but not all resonances go critical because they show interference potential due to some order of rotation.



Figure B1: HIS Figure 9.6.4.10, End Suction, Solids Handling, Vertical and Pump Instrumentation



Figure B2: Pump 1 Bearing Housing Top Inline and Crossline Orbit at 1789 RPM with Tight Tie Rods

Figure B3: Pump 3 Upper Bearing Housing Compliance, Parallel Bump Test with Tight Rods, May 9, 2007



Figure B4: Pump 3 Mode Shapes from Bump Test Data on May 9, 2007, Parallel 3345 CPM, Perpendicular 3180 CPM Illustrating Rotor Flexure Behavior

Figure B5: Original Design Mode 3: 43.33 Hz, 2600 CPM, Pump Assy ZY Rocking, 1st Reed Original Design Mode 4: 47.05 Hz, 2823 CPM, Pump Assy XY Rocking, 1st Reed



Figure B6: Interference Diagram for Torsional and Lateral Analyses

Figure B7: FEA Analysis Comparison to Bump Tests



Figure B8: Pump 3 Run Down Vibration Speed Analysis, Ch 1, MTI N, Motor Top C/L Parallel North

Figure B9: Pump 3 1641-1793 RPM Speed Analysis, Ch 9, PTI N, Pump Bearing Top C/L Parallel North



Figure B10: Preferred and Allowable Operating Regions for Non-Clog Pump 3

Figure B11: POR and AOR for Pump 3 with GPM Scaled Test Data


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