modal analysis on probe bracket of hydro turbine.pdf

Modal Analysis with SCOUT/vb8 — Vibration Probe Mounting used for Condition Monitoring of a Hydro Turbine 1

GE Measurement & Control

Modal Analysis with SCOUT/vb8 — Vibration Probe Mounting used for Condition Monitoring of a Hydro Turbine

Roengchai Chumai | Technical Leader | Bently Nevada Machinery Diagnostics Services | GE Oil & Gas | [email protected]

Abstract

The paper discusses a classic resonance problem of a vibration probe mounting installed on a hydro turbine during retrofit project commissioning. Fluctuating vibration was detected within the online machine condition monitoring system. Experimental Modal Analysis (EMA) was performed to confirm the presence of a resonance issue at each probe bracket mounting at the Turbine Guide Bearing (TGB). Actual natural frequency was identified with associated mode shape and damping. Finite Element Analysis (FEA) was then used to perform modal/frequency analysis of the probe bracket with calibrating to match the actual modal impact test result. Consequently, the model was used for structure modification and redesign of a new probe bracket with higher natural frequency to prevent natural frequency excitation. Multiple disciplines and tools are required to analyze and solve this problem.

Introduction

Modal analysis has been used in many industries and applications to find natural frequencies along with associated mode shapes within relevant frequency ranges in both engineering design and field troubleshooting work [1, 2]. The main objective is to avoid excessive vibration amplitude due to resonance excitation which can shorten the operating life of the test object or machine part due to fatigue, or even prevent the machine from operating reliably in some cases. Experimental Modal Analysis (EMA) results are also used for correlation with Analytical Modal Analysis (AMA) models which are constructed using Finite Element Analysis (FEA) based on precise dimensions, geometry, material properties, and boundary condition (e.g. restrains/constrains, pre-stretch, etc.). Once the FEA model is calibrated to match actual modal test data from EMA with acceptable tolerance, the model can be used to find the optimum design for future fabricating or manufacturing without a trial and error approach (saving cost and time). The machine and associated parts are free from natural frequency excitation throughout operating speed range. Consequently, vibration amplitude is acceptable for long-term continuous operation.

Application Information

This project was performed to provide an online machine condition monitoring and diagnostics system for a hydro power generation plant covering four Francis hydro turbine-generator sets (4 x 250 MW) and two Pelton hydro turbine-generator sets (2 x 37.5 MW). Examples of hydro turbine types installed in this plant are shown in Figure 1 below. All units have casing vibration transducers installed at all bearings, and a Keyphasor® transducer on each rotor. Vibration signals are then connected to the existing vibration monitor rack, serving as a simple machinery protection system. However, the existing machinery protection system is not sufficient for machinery diagnostics and long-term machine condition monitoring. Hence, the new project work scope includes designing and installing a shaft vibration transducer at all bearings; proper monitors; software for an online Condition Monitoring System (CMS); and system integration between the machine control and online machine condition monitoring system. Selected process variables are integrated into the CMS via OPC (OLE for Process Control) protocol. The Francis hydro turbine-generator set has three radial journal bearings installed: Generator upper and lower guide bearings and a turbine guide bearing. The Pelton hydro turbine-generator set has two bearings: Generator upper and lower guide bearings. The Francis turbine discussed in this paper incorporates a 15 blade runner, 26 wicket gates, and stay vanes. The tilting pad thrust bearing has six pads.



Project delivery depended on machine overhaul schedule so it was started with unit #3. The first set of vibration probe brackets was designed based on engineering judgment and experience from past projects without calculations. As commonly known, a shaft vibration probe measures relative vibration between the rotor and the bearing housing where the probe is mounted. The probe bracket and its mounting must be rigid with natural frequencies well above machine operating speed and possible exciting frequency range depending on machine configuration and operation such as blade passing frequency, fluid induced vibration, etc.

Figure 1: Cutaway of Francis (left) and Pelton (right) hydro turbines

A machine train diagram after all transducers were installed is shown in Figure 2 below. There are two casing vibration transducers installed and two X-Y shaft vibration probes installed at each bearing. A Keyphasor® transducer is also installed to observe a notch on the shaft surface so as to provide the vibration phase angle and supplementary machine speed. The online machine condition monitoring and diagnostic software platform retrieves all vibration data from the vibration monitor rack and processes the variables data from the machine control system. The data is processed, stored in a database, and presented as various kinds of plots — Such as trend, time base, orbit, spectrum, bode, polar, average shaft centerline plots, and others. Plant personnel and machinery specialists can then utilize those plots to view vibration data in correlation with applicable process or environmental variables to monitor and diagnose machine problems.



Figure 2: Machine train and transducer layout of unit #1 to #4 of Francis hydro turbine-generator set



Vibration Analysis after Commissioning

As part of a Site Acceptance Test (SAT), a system functions check was carried out including a vibration data review at all measurement points. Vibration readings were found to be normal except at the turbine guide bearing shown in Figure 3. Vibration amplitude at one probe appeared to be fluctuating over a period of steady state operation. The trend plot in Figure 4 shows fluctuating overall vibration amplitude at the Y probe while the X probe appeared to be stable at a normal level at low load operating condition. The vibration waveform data presented in an orbit plot (Figure 5) shows high frequency vibration induced at the Y probe but normal frequency at the X probe. XY probes are installed at the same bearing. Consequently, measured vibration frequency is normally similar, but naturally with some difference in amplitude. The observed vibration characteristic was concluded to have been most likely not caused by real mechanical vibration produced from the machine.

Figure 3: Shaft vibration measurement at turbine guide bearing with original probe bracket design

Y Probe X Probe



Recognized best practices suggest normally having seismic transducers installed adjacent to rotor vibration proximity transducers. However, this could not be done in this case due to limitations of probe mounting and fixture — The reason being that the seismic transducers were installed by the Original Equipment Manufacturer (OEM). When XY measurements were added during the provided retrofit, the existing bracket accommodation did not provide adequate angular shift between the X-probe and the visible seismic transducer. The situation was the same between the Y-probe and the next visible seismic transducer.

Figure 4: Trend plot of overall shaft vibration amplitudes (X and Y probes) measured at turbine guide bearing after commissioning

Figure 5: Orbit plot of shaft vibration measured at turbine guide bearing after commissioning

Y Probe

X Probe

Y Probe

X Probe

Less than 10% load condition



Vibration frequency at the guide bearing measurement point was reviewed and found predominantly at 95 Hz at the Y probe (Figure 6). There was some low amplitude vibration presented at 120 Hz at the X probe (Figure 7). There was no high frequency vibration excited at the generator upper and lower bearings (Figure 8).

The waterfall plots show clear evidence of excitation at the turbine guide bearing, only with higher amplitude at the Y probe. Nevertheless, those high vibration frequency components are not coincident with any vibration frequency generated from the machine and associated components. Consequently, it was suspected to be caused by a resonance problem, which natural frequency of the probe bracket and mounting that could be excited, resulting in fluctuating vibration amplitude predominantly at the Y probe. The difference of vibration amplitude and frequency between the X and Y probes could be due to varying stiffness of the probe mounting, as the physical installation of these two probes is different. We are most likely seeing guide bearing angular stiffness anisotropy (at the bearing shown in Figure 3). For example, the X probe mounting may be influenced by the adjacent oil cooler structure, which can increase stiffness. Somehow, higher frequency excitation in this case has lower energy as lower excited vibration amplitude. The source of excitation was unknown at this stage. A modal impact test was required to identify the natural frequency of the fixture’s mechanical structure and associated mode shapes.

Figure 6: Waterfall plots of spectra versus time of the Y probe measured at the turbine guide bearing after commissioning

1X vibration component and its harmonics

Vibration component at 95 Hz



Figure 7: Waterfall plots of spectra versus time of X probe measured at turbine guide bearing after commissioning

Figure 8: Waterfall plots of spectra versus time of X probe measured at generator lower guide bearing after commissioning

1X vibration


Vibration component at 120 Hz



Experimental Modal Analysis

In order to confirm the suspected resonance problem, a modal impact test was carried out on each probe bracket installed at the turbine guide bearing. A Bently Nevada SCOUT140-Ex portable data collector (equivalent to Commtest vb8) was used together with an accelerometer and impact force hammer. The test setup is shown in Figure 9. To get the best response the accelerometer was installed close to the end of the probe bracket where the shaft vibration probe is installed.

The impact force hammer was then used to hit the bracket in the vertical direction, which is the selected measurement plane in this case. Based on bracket confirmation and geometry, it is likely to deflect and vibrate in the vertical direction. Impulse force input from the impact force hammer generates broadband exciting frequencies. Frequencies that coincide with the probe bracket natural frequency show up as peak vibration amplitude with a 90 degree phase shifted which can be picked up by the installed accelerometer. There are generally an infinite number of natural frequencies of the test object but the frequency range of interest that was selected for this particular test was 0 to 1000 Hz. Coherence values are used to check and confirm data quality based on the relationship of input force and output vibration. It is expected to be more than 90% for good and acceptable results.

It was identified that the first natural frequency of the Y probe mounting structure was 91.25 Hz with 99.83% coherence (Figure 10). For the X probe bracket it was 116.30 Hz with 100% coherence (Figure 11). This confirms the high vibration frequency presented in vibration readings of the X and Y probes was due to probe mounting structure resonance excitation. However, there is some difference in the observed vibration frequency during turbine operation and during the test. This could be due to two different reasons: First, a “mass loading” effect, since a miniature accelerometer was not available at the site. Hence, we used a normal, large one with a heavy weight for the test instead. Accelerometer mass with a magnetic base can alter the test object (probe bracket) mass, as happened in this case. This can affect the test result accuracy, as the identified natural frequency might be lower than the actual value. However, it is acceptable for a troubleshooting job to produce an approximate result. Second, there is different behavior of the TGB shell (different modal mass, different modal stiffness) during turbine operation and during the test. According to ISO 7919-5 standard[3], the lowest natural frequency of the transducer support structure should be more than seven times that of the machine operating speed (333 rpm or 5.55 Hz) which is about 38.85 Hz in this case. The observed resonance frequency is much higher than this guideline. However, field test data showed that it is not safe from excitation. Consequently, probe bracket and mounting modification is required.



Figure 9: Photo of modal impact test performed at site with heavy accelerometer

Figure 10: Modal impact test result of Y probe with original fixing

NT2_Francis_U3 - 2 - +Y - Modal Accelerance 1000 Hz - Coherence

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Heavy accelerometer

with magnetic base

Seismic vibration

transducer



Figure 11: Modal impact test result of X probe with original fixing

Figure 12: Frequency Response Function (FRF) of original fixing

NT2_Francis_U3 - 1 - +X - Modal Accelerance 1000 Hz - Coherence

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Mode#1 Mode#2

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After natural frequency was obtained as a preliminary check, it was confirmed to be a resonance problem of the vibration probe installation. The test was continued to identify mode shape at each particular frequency. An impact force hammer was used as a roving point and the accelerometer was fixed where it was expected to sense high vibration amplitude near the end of the bracket. A grid of the measurement points was drawn on the bracket, so that data collection could be carried out accordingly. Frequency Response Function (FRF) was then calculated for each point as a ratio of acceleration/force. Curve fitting might be implemented to find the damping value at each mode [4] and select the number of interest modes — E.g. the first three modes are identified in Figure 12. The first mode shape is shown in Figure 13 with maximum deflection at the bracket end where the vibration probe is installed. This can partially explain how a resonance problem influenced the excessive vibration amplitude observed at the Y probe. Similarly, the second mode shape is identified in Figure 14 with only minimal deflection. Therefore, we did not consider it at later stages of analysis. The first mode is the main contributor to the observed problem.

Figure 13: Experimental modal analysis result of first mode shape with original bracket design



Figure 14: Experimental modal analysis result of second mode shape with original bracket design

The example in Figure 15 shows proper instrument setup to avoid a “mass loading” problem which can affect the accuracy of the test result. It also presents an alternative transducer attachment for a more complete modal test — The transducer axis has the same orientation as the axis of the non-contacting probe. However, we are still not able to improve our test due to change of modal characteristics (modal mass and stiffness) for running versus not running unit conditions of the TGB shell that supports the bracket, creating the mechanical system that resonates.

A two channel portable data collector with a modal test feature (such as SCOUT140-Ex or vb8) is required. A miniature accelerometer is used together with a small magnetic base or wax for mounting to suit the frequency range of interest. An appropriate impact force hammer should be used with consideration of tip material and size to ensure sufficient frequency response range as well as enough force to strike the test object in order to get a good response. The harder the tip is, the higher the frequency response range will be. The product datasheet should be referred to for more detail and precise data.



Figure 15: Example of typical setup for modal impact test with SCOUT140-Ex and miniature accelerometer

Minimum 2 channels analyzer

Miniature accelerometer

Impact Force Hammer



Analytical Modal Analysis

Because we don’t have the option of changing the modal characteristics of the TGB shell, we can only try to improve the real situation by changing the construction of the bracket to improve the resonance characteristics of the mechanical system formed by the TGB shell and transducer bracket. Before bracket modification/redesign we used Finite Element Analysis (FEA) to model the original bracket and then performed modal/frequency analysis to find its natural frequency and associate mode shape. The result was calibrated with EMA result which was considered actual data. This approach can eliminate “trial and error”, saving cost and fabrication/manufacturing time. The model was built with inputs of precise dimensions and geometry, material properties, and boundary condition[5]. Bracket material was carbon steel code JIS G3101 SS400 with elasticity 210000 N/mm2, Poisson’s ratio 0.26, and mass density 7860 kg/m3. The bracket was fixed at two spots where it was bolted to the bearing cover. This prevented the bracket from moving in translation and rotationally. The contact surface between the bracket and bearing housing was modeled as an elastic support which can control bracket movement in the vertical direction. Support stiffness was interpolated to match modal analysis results with actual data from the EMA test — Hence the model was calibrated. Model assumption includes homogenous material properties, uniform support stiffness throughout the contact face, properly tied mounting bolts, and the actual material used being the same as the specification. The first mode of the original bracket was 92.7 Hz (Figure 13) which closely matches the actual data in both frequency and mode shape.

Figure 16: First mode shape of original bracket design at 92.65 Hz



A new bracket was redesigned by changing the dimensions and geometry to increase thickness, add more ribs, and then re-run modal analysis. Such bracket construction will result (through the bracket fixing) in increasing of TGB shell stiffness in the horizontal direction as well. First mode of the new bracket was increased to 179 Hz (Figure 17) which is about 32 times the machine running speed at 333.33 RPM (5.56 Hz). This is considered safe from excitation. The new bracket was then fabricated and installed in the field.

Figure 17: First mode shape of new bracket design at 179.05 Hz



Vibration Results of New Probe Bracket

New probe brackets were redesigned and installed onsite as shown in Figure 18. FEA simulation showed that the first mode was increased from 92 to 179 Hz which is approximately 32 times the machine running speed. Vibration data appeared to show no sign of a probe bracket resonance problem as shown in orbit plot (Figure 19) and spectrum plot (Figure 20). However, this data was collected at a machine operating condition of 235 MW with 75% inlet guide vane (IGV) opened. The waterfall plot in Figure 21 was then used to view the vibration spectrum of the Y probe at various machine operating conditions. It was found that low amplitudes of resonance vibration at around 183 Hz were revealed at some operating conditions. This implies that the resonance vibration problem relates to the machine operating condition such as inlet flow, IGV position, generator load, etc.

Figure 18: New bracket design installed at turbine guide bearing

Y Probe

X Probe



Figure 19: Orbit shape of vibration data measured at turbine guide bearing with new probe bracket design

Figure 20: Vibration spectrum of Y probe installed at turbine guide bearing with new bracket design when the unit was operated at

235 MW with 75% IGV opened




Figure 21: Waterfall plot of vibration spectra versus time measured by Y probe at turbine guide bearing with new bracket design

Broad frequency seismic vibration

No broad frequency vibration

Broad frequency seismic

Resonance frequency at 183 Hz



Possible Sources of Excitation

Operating process variables were reviewed as shown in Figure 22. We noticed a significant correlation between the IGV position and the mechanical vibration of the turbine structure. It is typical of Francis turbines, that below a percentage of nominal power, high fluid flow instabilities may be generated. This can result in structural resonances of various stationary components of the hydro turbine — Generator (HTG) structure. In this unit, we discovered that the excitations resulting in resonance vibrations of turbine structure are present when IGV position is less than 10% when you increase the inlet flow. This means every time the unit is experiencing loading or unloading we can expect transient operation that generates higher vibration levels. However, because the unit is generally operated at base load condition, the structural vibrations do not significantly influence its aging — Due to fatigue.

Flow-induced vibration was revealed by casing vibration data measured at the bearing housing — You can recognize two seismic transducers connected to the TGB shell (Figure 18) as broadband frequency components (Figure 23). The shaft vibration probe bracket is mounted on the bearing housing. Some vibration component frequencies could be coincident with the bracket’s own natural frequency. There was no mechanical vibration resonance excitation revealed in shaft vibration data when no broadband vibration frequency components were present in the casing vibration data. However, it appeared to be low severity mechanical excitation at higher vibration frequency components as lower vibration levels. The new probe bracket design is considered acceptable since the unit is not generally operated at low load condition below 10% flow where there are still some low amplitudes of resonance vibrations. However, now the noise-to-signal ratio is very small, and will not destroy diagnostic analysis of the physical condition of the HTG. Although unnecessary, if required, the mechanical structure of probe fixing can be modified further and better simulated using FEA for the TGB shell to get even higher first mode natural frequency. This can minimize structural mechanical vibration excitations at low inlet flow conditions as well as reduce vibration levels if the mechanical vibrations are excited at a higher frequency.

Figure 22: Trend plots of seismic vibrations measured at TGB (top) correlating with process variables (bottom)

Inlet Flow, ml/s

IGV Position, %

Gen Load, MW

Casing vibration, XV

Casing vibration, YV

Turbine speed



Figure 23: Waterfall spectrum of seismic vibrations measured at shell of TGB

Conclusions

Modal analysis is an effective tool for solving vibration problems as it can be used in the field for troubleshooting, design phase, simulation study, as well as structure modification and improvement. This paper discussed the approach of applying modal analysis to successfully solve vibration problems observed from an online machine condition monitoring system installed on a hydro turbine-generator set during project commissioning. EMA was used to identify a resonance problem of vibration probe mounting at TGB, allowing the natural frequency and associated mode shape to be identified. FEA was then used to model probe bracket structure based on precise dimensions and material properties before modal/frequency analysis was carried out to find the calculated natural frequency and associated mode shape. The result was then calibrated with EMA by adjusting the boundary condition before the structure modification was studied and simulated for probe mounting improvement to increase the natural frequency. FEA can save manufacturing costs and time without doing “trial and error” for new product design. The new bracket design resulted in stiffness increase of the probe mounting, which was acceptable and resulted in improving the noise-to-signal ratio.

References

[1] Avitabile, P., 2001, "Experimental Modal Analysis: A Simple Non-Mathematical Presentation," Sound and Vibration.

[2] Vázquez, J. A., Cloud, C. H., and Eizember, R. J., 2012, "Simplified Modal Analysis for the Plant Machinery Engineer," 41st Turbomachinery Symposium, Texas A&M University, Houston, Texas.

[3] International Standard Organization, 2005, "ISO 7919-5 Mechanical Vibration - Evaluation of Machine Vibration by Measurements on Roating Shafts - Part 5 Machine Sets in Hydraulic Power Genrating and Pumping Plants," Internation Standard Organization.

[4] Lee, M., and Richardson, M., 1992, "Determining the Accuracy of Modal Parameter Estimation Methods," IMAC X.

[5] Richardson, M. H., 1978, "Measurement and Analysis of the Dynamics of Mechanical Structures," Hewlett-Packard Conference for Automotive and Related Industries, Hewlett-Packard, Detroit.


No broad frequency vibration


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modal analysis on probe bracket of hydro turbine.pdf