Vibration Certification Case Studies Vertical Pump ...rushengineering.com/VFD-Pump-MultiChannelCaseV.pdfFirst, consider the Kingsbury Thrust Bearing in Figure E1. The complete upper

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. Page 2 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 47 of 101 CASE V: Vertical Can Pump Oil Whip Excitation Vibration Certification of Vertical Pump Machinery Controlled with Variable Frequency Drives

CASE V: Vertical Can Pump Oil Whip Excitation This case involves 1000 HP vertical turbine short set can pumps. Initially, REAI became involved on the pump installation due to the inability to operate the pumps without vibration trips. A dual channel analyzer was employed to evaluate the vibration and measurements were made at several speeds that indicated that the vibration was well below the trip level. Later presentation of the plant control system vibration records indicated high vibration within a relatively small speed range (<5% delta RPM). To establish the cause of the vibration trip a multi-channel digital signal recorder/analyzer was employed to capture the trip event for evaluation. From this evaluation it was established that the motor, which employed Kingsbury Thrust Bearings, had an oil whip critical speed due to motor reed mode resonances. The initial design evaluation using finite element analysis had predicted the pump resonances with reasonable accuracy. However, the analysis did not include an evaluation of the oil whip critical speed. Prediction of oil whirl excitation is not an established science and the finite element analysts are unlikely to consider it. By employing the multi-channel digital signal recorder/analyzer it was determined that an increase in lubrication grade could remedy the oil whip condition. Since Kingsbury Thrust Bearings provide very long life, their use in high horsepower variable speed machines can be expected to rise. This case provides guidance in how to deal with oil whip conditions. There were fifteen identical pumps and the use of the multi-channel digital signal recorder/analyzer was mandatory because the machines had to be operated under production load conditions. This meant very limited test time and the multi-channel digital signal recorder/analyzer could capture all the necessary vibration data with a single start and run up to maximum speed. This procedure acquired over a gigabyte of vibration data on each machine and that is well beyond the capabilities of predictive maintenance machinery vibration analyzers (MVA). DISCUSSION First, consider the Kingsbury Thrust Bearing in Figure E1. The complete upper bearing assembly is contained within a lubricant bath with a water cooled jacket of coils. The journal bearing is bolted to the shaft with the thrust load transferred to the thrust bearing plate. Around the journal bearing is the upper sleeve guide bearing. With a complete oil bath the potential for wear is greatly reduced and long maintenance free life can be expected. Hydrodynamic oil film journal bearings have an oil wedge that is pushed by the journal as shown in Figure E2. Viscosity of the oil pulls the wedge between the journal and the sleeve where pressure is built up to carry a radial load. This load in horizontal bearings is essentially the weight of the rotor and any specific lateral loads required for operation, such as belts, gears, etc. As a result the load vector is stable at whatever angle the journal lifting load is required. If the journal is lightly loaded the forces in the oil wedge will exceed the applied journal loads and the wedge will whirl around the journal. This whirling wedge of oil produces radial forces that are reacted by the journal and by the sleeve. The frequency of the whirl is speed dependent and it is a characteristic fraction of the journal rotational speed, generaly between 40% and 49%, that is affected by the residual mechanical radial load between the journal and the sleeve and the structural sensitivity to speed. When the oil film lubricated journal is in the vertical position as in Figure E1 the mechanical radial load is essentially zero. This means that the location of the oil wedge is not defined and it is free to move in accordance with the viscosity and frictional characteristics of the lubricant and the journal and sleeve elements. As a result the oil wedge hydrodynamic pressure is a revolving force between the journal and the



sleeve. This revolving radial force acts like a mass imbalance force with the rotational frequency of the oil wedge. This is called the oil whirl frequency. The important point to understand is that there is a rotating radial force produced by the hydrodynamic journal at a frequency below the shaft speed. The typical range of the oil whirl frequency identified in the published literature is 45% to 47% of the shaft speed. However, as this study demonstrates that mechanical dynamics of the bearing support can widen the oil whirl frequency range. The lateral stiffness is primarily due to the structure with the oil film stiffness providing additional stiffness or compliance that modifies the critical speed by a few percentage points. Additionally, the gyroscopic stiffening effects of the rotor also modify the critical speed. As a result, precise numerical values for the fractional speeds of oil whirl and oil whip conditions are not within simple theoretical description. The potential for oil whip can be identified within a 40-49% boundary of rotational speed and the installed condition must be evaluated accordingly. A numerical analysis prior to fabrication can guide the manufacture of the support structure to keep the oil whip condition, which is the critical resonance speed for the oil whirl loads, outside the operating speed range. This was not done in the present case. Figure E3 presents the machinery installation for this case study. The 1000 HP motor sits upon a motor support integral to the top of the discharge head. The discharge head has orthogonally spaced nozzles at the same elevation, the suction and discharge nozzles. The flow enters the suction nozzle above the discharge head base and flows down around the pump column and enters the pump bowl assembly where it is drawn up by the pump as depicted in the inset schematic. The motor contained hydrodynamic oil film journals at the top at the Kingsbury thrust bearing and at the bottom guide bearing. The nominal side load on the motor shaft is approximately zero with a small but negligible magnetic centering force. Impact tests were performed on the motor frame as depicted in Figure E4 to establish the reed modes of the installation, and any other modes of significance. The millwright is impacting the motor frame at the level of the upper journal bearing with the motor shut down. Figure E5 and Figure E6 illustrate the mobility response functions (in/sec / lb) acquired by the test. The data indicates Motor 1st Reed Modes at 570 CPM and 600 CPM. The data also indicate Motor 2nd Reed Modes at 2542.5 CPM and 2550 CPM for the Crossline and Inline impacts. These directions are shown in Figure E3 to be parallel to the suction and discharge nozzles, respectively. It is common practice to use the discharge nozzle orientation as the inline or parallel designation. In both impact response functions there are blips near 950 CPM. These blips represent an internal mode of the motor that is not significantly excited by impact forces on the motor casing. Figure E7 indicates a trip that was captured by the digital signal recorder/analyzer between 11:58 and 12:12 on 9/27/07. The black dashed line indicates the shaft speed with the scale to the right of Figure E7. Between 12:04 and 12:10 the motor speed was increased in small steps from 58% Load (1044 RPM) to 64% Load (1152 RPM). As the speed increased in this range the vibration values increased in various proportions depending upon the location of the measurement. The maximum vibration was at the Motor Top Suction MTC location with a value of approximately 0.224 in/sec peak at the 64% Load condition. Then at 12:10:27 the motor speed was increased to 68% Load (1222 RPM) and the machine tripped when the MTC vibration exceeded 0.5 in/sec peak. This trip value was set above the normal 0.3 in/sec peak trip value so that vibration readings at the maximum vibration condition could be acquired. The trip demonstrates the initial problem with the pumps. The plant operators could not bring them up to system speed without tripping. Of the fifteen pumps only about one third of them experienced the trip conditions. This was the result of variances in damping and bearing whirl sensitivities for all pumps had essentially the same reed mode frequencies.



Between 12:22 and 12:34 on Figure E7 tests were performed to determine if flow and pressure operating points had an effect on the trip condition. This test involved adjusting the discharge valve in Figure E8 between 100% Open and 13% Open. The data between 12:22 and 12:34 in Figure E7 indicate that the vibration was essentially independent of the pump operating point under constant speed. Figure E9 illustrates the power of the digital signal recorder/analyzer with Digital Analog Tape (DAT) recording capabilities. The graph is called a sonagram and it shows the variation in time, frequency, and amplitude of the vibration signal acquired in the crossline direction at the top of the motor during a trip condition similar to that of Figure E7. The three parameter graph in the sonagram at the upper right quadrant of Figure E9 provides a wealth of information. Three orders of rotation can be identified, two resonances can be identified, and the low frequency hydraulics and electrical noise are identifiable. The flat line labeled Unidentified Motor Dynamic was subsequently associated with the “blip” in the impact functions described earlier, and its amplitude was not significantly affected by operation once the motor came up to speed. This behavior indicates that the dynamic is the analytically determined 15-16 Hz (900- 960 RPM) shaft resonance, as opposed to a motor frame resonance. The Motor 1st Reed Modes were excited by ambient conditions at the pumping facility because the suction and discharge line were tied into the operating system with several other pumps in operation. When the motor was started up the reed mode frequency did not change significantly, and this is evidence that the mode is a structural mode of the motor, specifically the 1st Reed Mode representing cantilevered motion from the discharge head base flange. The upper left quadrant of Figure E9 presents a vibration velocity spectrum with a dominant peak at 9.38 Hz (562.8 CPM) with an amplitude of 0.3231 in/sec rms at 22m35s500 (22 minutes 36.5 seconds) into the data recording. The lower right quadrant in Figure E9 presents a Cut across the sonagram at 9.375 Hz (562.5 CPM). The spectrum graph and the cut graph are indicating that the vibration is dominated by the 562.5 CPM energy. This compares very favorably with the non-operating impact response resonance in Figure E5 at 570 CPM. The Cut Graph also duplicates the behavior of the MTC data in Figure E7 when the machine tripped on the first run up. This data is sufficient to conclude that resonant excitation of the 1st Reed Mode of the Motor in the crossline direction was the cause of the motor trip in Figure E7. The trip in the Cut Graph of Figure E9 was caused by over pressure considerations when the motor was taken rapidly from approximately 1200 RPM to 1781 RPM at 22m55.500 as indicated by the c2 cursor on the sonagram and in the table at the lower left quadrant in Figure E9. A finite element analysis (FEA) of the motor had been performed for the manufacturer by their analysis engineers and the results of the structural resonance analysis are presented in Figure E10 for the dry pump condition. The impact response test resulted in 570 CPM and 600 CPM for the installed pump while the FEA indicated 787 CPM for the 1st Reed Mode. The interference diagram provided by the FEA and operating analysis performed by the manufacturer and presented in Figure E11 shows acceptable agreement with the impact test results for the Motor 1st Reed Mode (Motor Housing Pitch Mode). The Oil Whirl Excitation Frequency at 46% of the shaft speed has been overlaid on the manufacturer’s analysis. It shows that interference occurs at the minimum operating speed for the pump. This interference was not identified by the manufacturer during the design analysis because the analyst(s) did not consider the potential for an oil whirl excitation. LUBRICANT GRADE EFFECT Figure E12 demonstrates variation of the oil whirl excitation. When interference exists for an operating speed and the oil whirl at a lateral resonance of the bearing support, the oil whirl condition will result in an



oil whip of the revolving oil wedge. The forces in the oil whirl do not necessarily increase as much as the relative compliance of the motor bearing support increases. The result is elevated vibration and in this case the motor trips. The B03 MTC Suction Original Oil data curve in Figure E12 indicates a severe oil whip condition occurred between 64% Load and 73% Load. Assuming that the 1sr Reed Mode frequency at 9.375 Hz (562.5 CPM, 31.25% Load) in Figure E9 remains constant the oil whirl ratio varies between 48.8% RPM and 42.8% RPM for the 64% Load to 73% Load conditions of Figure E12, respectively. Several corrections to the motor installation were attempted to remedy the oil whip condition including balance correction, misalignment correction, flange tightness variation, and viscosity grade change of the lubricant. Only the change in lubricant had significant affect on the oil whip condition. The original lubricant was ISO Grade 68. When the lubricate was changed to ISO Grade 100 the speed range of elevated vibration was dramatically reduced with a narrow peak near 67% Load for vibration at 46.6% RPM that was very near the vibration trip level for the machinery. Further increase in lubricant viscosity from ISO Grade 100 to ISO Grade 150 nearly eliminated the oil whip condition. The ISO Grade 150 data in Figure E12 indicates oil whip excitation between 65% Load and 70% Load (48.1% RPM and 44.6% RPM) without a large excursion in amplitude. This condition was considered to be an effective remedy for the vibration trip condition of the machinery. All fifteen pumps were modified to specify ISO Grade 150 lubricant. The next higher grade lubricant, ISO Grade 220, was considered but the motor manufacturer objected on the grounds that the motor bearing temperature increase would not be kept within acceptable limits for proper motor operation. During the tests with the ISO Grade 150 lubricant the bearing temperatures were monitored but no significant increase in bearing temperature was noted. This may be the result of the water cooling coils and the location of the temperature sensor not resulting in a proper measure of the local bearing temperature. The vibration was well controlled with the ISO Grade 150 lubricant and there was no need for the higher grade lubricant. VIBRATION CERTIFICATION OPERATING TESTS After all pumps were retrofitted with the ISO 150 Grade lubricant and the plant was brought on line the vibration certification testing was performed. Knowing that the oil whip condition existed and that these pumps were required to operate under plant system controls the procedure for performing the vibration certification test required a continuous data acquisition system with simultaneous measurement at all locations for extended lengths of time. The test procedure included mounting thirteen vibration probes on the motor, pump, and piping as well as a tachometer and two sound level meters. The digital signal analyzer was operated in the DAT mode and the pump operating point was modified by adjusting the discharge valve shown in Figure E3 and in Figure E8. Figure E13 illustrates the vibration spectra cascade time map for a complete 35 minute vibration certification run for a pump. Excitation of the reed modes and the pump shaft mode are quite evident during the startup sequence. After reaching minimum operating speed the vibration was very well behaved with no elevated vibration levels at any of the three modes for oil whip frequency forces or for shaft speed frequency forces. The first seven minutes of the startup transient is illustrated on another pump in Figure E14 with the various components labeled. Since the machines were under system flow the control strategy of throttling the discharge valve increased the discharge pressure and the motor increased in speed to maintain the required system flow. Figure E15 illustrates the operating points for various speeds evaluated on one pump between 1280 RPM and



1650 RPM with the flow control steady near 3800 gpm. The pumps could not be taken to full speed because the system pressure would cause a trip and the whole plant would be shut down. With fifteen 1000 HP pump operating near 3800 gpm the trip would represent a loss of 15,000 HP (11 megawatts) and 57,000 gpm (82 MG/D). The electrical costs for such transients did not warrant a vibration certification test with such tight operating parameters that inadvertent trips would have been inevitable. As it turned out there were two trips caused by inappropriate change in the discharge valve throttling during the initial tests. Figure E16 presents 63 minutes of vibration data during the discharge valve controlling sequence. The jumps and drops in the 1xRPM trace were a direct result of the manual operation of the discharge valve. Essentially a change in the discharge valve position caused a change in the speed as the motor control responded to the change in flow. These changes were counter intuitive since a closing of the valve caused an increase in motor speed, and vice versa. After several tries the valve operator(s) would realize that opening the valve would actually reduce the motor speed because the flow stayed constant. On several occasions a quick small adjustment of the discharge valve resulted in a dramatic change in motor speed and the valve operator had to react quickly to keep the motor from an over speed setting. Once the desired speed change was accomplished the pump was allowed to settle and the flow, pressure, and HP, readings were acquired from local instrumentation at the pump. CASE V CONCLUSION The data presented in this case demonstrate the versatility of a digital signal recorder/analyzer in pump machinery vibration certification. As a trouble shooting instrument the “big picture” could be acquired in minimum time. Several trial fixes could be applied and the global effect measured with a single run of the machinery. During the certification process, which was extremely difficult to control, the DAT analyzer allowed the operator full leeway to maintain operation of the pump without losing any data. As such, REAI highly recommends the implementation of the DAT Analyzer for vibration certification of large pumps with complex control schemes. The additional information gathered from this case study includes the advisory that any finite element analysis should be aware of the potential for excitation by hydrodynamic oil film forces. If feasible, the design should be modified so that no encroachment of the oil whirl frequency leads to an oil whip critical speed. Finally, as a practical guide to the solution of oil whip conditions, increasing the viscosity of the lubricant can dramatically reduce the oil whip vibration amplitude. There is no guarantee that lubricant viscosity change will always result in the elimination of oil whip. However, the cost of changing the lubricant viscosity is not significant and changing the viscosity is highly advisable as a first step solution. The final alternative would be to modify the pump support structure to prevent interference of the oil whirl frequency with operating speed conditions. The manufacturing costs and construction delay costs to prevent oil whirl encroachment would have been enormous considering that all fifteen of the 1000 horsepower machines would need to be modified. The change in lubricant was accomplished with less than a week delay in the construction schedule at the cost of the lubricant and the labor to change the lubricant in all fifteen motor. There should be no doubt that the use of the multi-channel digital recorder/analyzer for this certification project was a practical requirement.



Figure E1: Top Motor Bearing Elements

Figure E2: Fluid Film Journal Bearing Static and Dynamic Load Components



Figure E3: Feed Pump B03 Flow and Measurement Directions

Figure E4: Feed Pump B03 Crossline Impact Vibration Measurement



Figure E5: Motor Top Crossline Impact Mobility Response Function Perpendicular to Discharge

Figure E6: Motor Top Inline Impact Mobility Response Function Parallel to Discharge



Figure E7: Feed Pump B03 551 CPM Vibration on September 27, 2007 between 11:58 and 12:34

Figure E8: Feed Pump B03 Discharge Manual Valve Settings



Figure E9: Feed Pump B03 MTC Vibration Sonagram, 9/27/07 12:12-12:36

Figure E10: Feed Pump Finite Element Analysis Model Results



Figure E11: FEA Interference Diagram

Figure E12: Feed Pump MTC Perpendicular Vibration Summary Comparison



Figure E13: Feed Pump A01 Cascade Plot, Motor Top Inline, In/Sec rms

Figure E14: Feed Pump C01 Startup Cascade Plot, Motor Top Inline, Mils p-p rms, 58% NOR



Figure E15: Feed Pump A01 Operating Points for Running Tests

Figure E16: Feed Pump C01 Cascade Plot, Motor Top Inline, Mils p-p rms, Running Tests


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Vibration Certification Case Studies Vertical Pump ...rushengineering.com/VFD-Pump-MultiChannelCaseV.pdfFirst, consider the Kingsbury Thrust Bearing in Figure E1. The complete upper