INVESTIGATION OF HIGH VIBRATION ON LOW PRESSURE STEAM TURBINES

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1 INVESTIGATION OF HIGH VIBRATION ON LOW PRESSURE STEAM TURBINES Ray Beebe Senior Lecturer Monash University School of Applied Sciences and Engineering Summary: During operation of a large steam turbine generator at lower than full load, the vibration levels measured at the Low Pressure bearings increased to high levels. Despite changes to operation, the levels continued to increase and the machine was tripped. During coastdown to rest, the vibration levels exceeded the maximum scale on the recording charts. The machine was later returned to service without incident. Condition monitoring techniques were summoned to investigate this disturbing problem. This paper describes the investigation and identification of the cause as operating at a very low exhaust pressure. Apparent high noise levels were also a concern when operating at high exhaust pressure. The outcome of both these investigations was to derive or confirm allowable operating limits of exhaust pressure. A third case study emphasises the importance of knowing plant detail when interpreting vibration levels. Keywords: Low Pressure steam turbine, turbine operation, vibration monitoring, vibration severity, noise, condition monitoring 1. INTRODUCTION Vibration monitoring and analysis is well established for detecting and monitoring changes in the internal condition of machinery to assist in or guide the need for maintenance and its extent. In power generation, the prime movers and often also major auxiliary machines have permanently installed vibration monitoring equipment. The purpose of this is basically for operator guidance and surveillance, but the recorded data can also be useful for condition monitoring, and can also be used as screening to initiate more complex test measurements. Transducer arrangements are summarised in Table I: Type Transducer and position Quantity measured Comments 1 Velocity transducers or accelerometers mounted on the bearing caps Vibration of the bearing cap 2 Velocity transducers mounted on shaft-riding probes 3 Displacement transducers mounted on the bearing to measure shaft vibration relative to the bearing, often with two transducers at right angles. 4 Combinations of these eg displacement transducer measuring relative shaft vibration and bearing cap transducer. Vibration of the shaft relative to space (ie absolute). Vibration of the shaft relative to the bearing All of the above. Presented at ICOMS-2002, International Conference of Maintenance Societies Brisbane Australia Output commonly converted to displacement, sometimes assuming that all the vibration is at 50Hz The amplitude of shaft vibration in most such machines is much greater than that of the bearing cap. The twin probes enable the shaft centreline position to be monitored, both during start-up and its orbit at service speed and load conditions. Gives absolute shaft vibration, relative shaft vibration, and bearing cap vibration. Less used than above systems. Table I: Vibration monitoring transducer arrangements for major turbomachinery As for many other key instruments, the output of these systems is shown on any, some or all of: panel meters, usually in the control room, paper chart recorders, paperless chart recorders, and plant computer system displays. A displacement probe is also often fitted to give a shaft angular position reference, to allow the phase angle of the rotating speed

2 vibration component to be found. Dedicated computer monitoring is also used, where the vibration signals can be analysed into their frequency content and continually sampled, analysed, and compared with historical data. Data can also be taken during transients, such as start up and coast down, to provide further information on condition, and relative phase is then particularly important. Depending on the extent of permanent systems fitted, vibration data is also taken with manual collection and remote processing, as is used in many industries where permanent or continuous monitoring is not considered justified. With the simpler systems, analysis of incidents after the event is more difficult, and relies on manual correlation of the available data. This paper describes such an incident, and also further related investigations into machine vibration. 2. CASE STUDY 1 : THE INCIDENT The large steam turbine generator in this case study is of the general layout shown in Figure I. It is a 3-casing machine, with combined High Pressure/Intermediate Pressure casing, and two Low Pressure casings, all tandem-compound the so-called TC4DF layout. The condensers are mounted rigidly on their foundations, with flexible connecting joints to the turbine Low Pressure casings. The machine had permanent vibration monitoring with velocity transducer mounted on shaft riders at each bearing, as in Type 2 of Table I. Vibration trends were logged as displacement on a paper chart recorder in the control room, along with selected temperatures and other data. Since commissioning, the normal amplitudes of vibration were very low, indicating a good state of balance and alignment. HP P HP-IP LP A LP B Generator Couplings Journal bearings Figure I: schematic layout of turbine generator On a weekend day when less production was required, the unit output was reduced to below half rated output. Inside an hour, a Vacuum Warning alarm showed. Suspecting increased air leakage into the condensers, the operator started the standby air ejector pump. Inside another hour, vibration increased at No 4 bearing (ie the generator end bearing of LP casing A). Vibration at other bearings then showed a more rapid increase to levels not seen before, and reached the Vibrations high alarm, set at about 100μm peak to peak. Output was reduced in a successful attempt to reduce the vibration. However, on then increasing load the vibrations also increased beyond alarm levels, so the unit was tripped manually. During the coastdown, vibration levels on LP A and LP B exceeded the chart range of 350μm peak to peak. After some hours on turning gear (low speed barring) the unit was started up and reloaded without incident. During this time, the vacuum warning transmitter was found to be faulty. All operating indications were normal at and after the return to service. 3. CASE STUDY 1: THE INVESTIGATION Investigation of this disturbing incident was initiated the next day, and data was correlated from several chart recorders and data logging output. Two recordings of the LP casing exhaust temperature were available on one chart, output on another, and vibrations on another. The turbine exhaust pressure was not recorded, but the saturation temperature is directly related to it. Combining this data on the one plot gave the trends shown in Figure II. The effect on exhaust pressure (ie temperature) of removing more of the air leakage is evident. The vibration transducers were checked and found to be in calibration, so the recorded readings were confirmed as correct. 2

3 Figure II: trend plots of data The lowest exhaust pressure during this incident was below 4 kpa, which is much lower than normally met in operation. This pressure is related directly to the condenser pressure and varies in service. It is reduced : with lower cooling water inlet temperature, which is a function of cooling tower or pond performance, which in turn varies with the ambient weather. This remained unchanged during this incident. if air leakage is reduced or removed more effectively. output reduces, as shown in Figure III. higher cooling water flow, as shown in Figure III (but the flow is normally not changed). Condenser Pressure Low flow eg 1 pump only (kpa) High flow 2 pumps Unit output (MW) Figure III Schematic behaviour of turbine exhaust (ie condenser) pressure with cooling water flow and output 3

4 As the bearings are supported on the exhaust hood structure, it was deduced that the whole structure had been distorted due to the very low internal pressure, sufficiently to affect the alignment of the rotor line. The result was the very high vibrations. This hypothesis matched the known facts. If it had been possible to analyse the vibration during or after the event, bearing instability due to unloading of a bearing may have been indicated. As these casings look strong and stiff, with thick steel reinforcing ribs, etc. this was perhaps a surprising conclusion. It was therefore recommended that the exhaust hood pressure not be allowed to get too low, by setting a lower operating temperature of 35 C. An alarm was arranged accordingly. As the return to service was without incident, it was deduced that no permanent damage had occurred. Bearing and support structure Concrete foundation Flexible connection from turbine casing to condenser Figure IV: Low Pressure casing cross-section showing bearing support It was later found that the operating instructions require one of the cooling water pumps be taken from service at outputs below half load. No reason for this stipulation was given. The turbine manufacturer later confirmed that the reason for this was to prevent very low exhaust pressures, and prevent such occurrences as this incident, which had been experienced on other similar machines. Changes in alignment would be expected to result in changes in loading on the bearings. Each of these bearings has a thermocouple set through its steel shell to near to the outer surface of the whitemetal (ie Babbitt) lining. A change in load would increase bearing metal temperature. During this incident, the thrust and both bearings of the HP-IP rotor were found to have increased by about 0.5C. Thermal inertia in this case would delay response, and the thermocouples would probably need to be into the whitemetal to be indicative enough. The effect observed in this case would not be expected to occur in machines with a different Low Pressure exhaust design, where the condenser neck is welded to the turbine exhaust casings, and the condenser supported on springs to allow for thermal expansion. 4. CASE STUDY 2: INVESTIGATION OF HIGH NOISE LEVELS Higher noise levels had been reported in summer conditions when the exhaust pressure was higher than usual. Given that the experience described above had revealed the need to decide an allowable lower limit on casing temperature 4

5 (albeit as an indicator of pressure), it was decided to use vibration analysis to investigate the effect of operation at higher exhaust pressures. Operating instruction limits were an alarm setting at an exhaust pressure of 16.6kPa (absolute), corresponding to a temperature of 56 C; a unit trip at a pressure of 24.8kPa (absolute) corresponding to 66 C; and also a unit trip at an LP hood temperature of 107 C. Windage overheating could cause this latter temperature, even if the pressure was low, and cooling sprays were provided to automatically turn on before this temperature was reached. There had been some design concerns expressed as to increased blade loading when the same mass flow of steam resulted in increased velocity at the lower density. It was felt that vibration measurements might help quantify the effects of operating limits and perhaps minimise any output reductions made to comply with them. A sound level meter was set up at a constant position. A vibration transducer was mounted at an arbitrary point on the casing. Vibration levels and also spectra up to 1000Hz were obtained at three increasing casing temperatures. These were arranged by reducing the cooling water flow. The results of the noise measurements and vibration analysis at significant frequencies are shown in Table II: Casing temperature C Noise level dba Overall vibration mm/s Vibration components from spectra 50Hz 100Hz 200Hz Datum level % increase 5% increase 220% increase % increase 30% increase 320% increase Table II: Effect on LP casing vibration of operation at higher casing temperatures There was no significant change in overall rotor vibrations during these tests, nor in the bearing metal temperatures. It was therefore concluded that increased noise was normal at higher casing pressures, and probably also occurred on may not have been noticed on older machines where the noise from the generator may have been much greater and have made it not evident It should be noted that there are no accepted standards for severity of vibration as measured on a casing only those measured on bearings or on shafts (Ref 2-7). As can be seen in Table II, the vibration in general increased as casing temperature was raised. Presumably this occurs from increased buffeting noise from changes in the distribution of steam flow inside the exhaust casing as steam velocities move away from the design values. Some noise tests run by Bigret in France (1), but at varying loads and constant exhaust pressure, gave somewhat similar results. In this case, no reason emerged as to why the manufacturer s operating limit for casing temperature (at least up to that tested here) should be changed. 5. OTHER ALIGNMENT INVESTIGATIONS These machines are aligned cold with a large vertical offset at the generator coupling to allow for thermal growth, and the investigation described above led to study of alignment changes as the machine reached its normal operating temperatures. The bearings are equipped with high pressure jacking oil tappings at bottom dead centre to reduce bearing friction at starting up. On some machines, once the jacking oil is shut down, the local pressure gauges then indicate the wedge pressure in the bearing, which is directly related to the loading exerted by the shaft on the bearing. This is a rule of thumb used by commissioning engineers in checking alignment. This machine required minor piping modifications to enable the pressures to be read using test transducers. The alignment was expected to show some change with the machine at normal service speed and load from the cold to hot condition. No changes were observed in the wedge pressures. The bearings are of the tilting pad type, and may show different behaviour from the elliptical clearance types where the initial experience was gained. Given the smooth running of these machines, the investigation was not taken further. 6. CASE STUDY 3: THE MISNAMED BEARING VIBRATION Another design of large steam turbine generator unit had an overhaul where an LP rotor had been removed. When half full output was reached during the return to service, the vibration on the bearing between the two LP rotors reached the alarm level. Placing a test vibration instrument adjacent to it checked the service transducer. The permanent system was 5

6 found to be reading correctly. Frequency analysis showed that the largest vibration component was at 300Hz, or 6 rotor service speed. On some types of turbines, the bearing shell has a retaining keep with metal-to-metal contact to the outside, which is the part seen and accessible for taking vibration measurement. On some other designs, such as both types featured in this paper, there is a gap of some 50mm between the cover and a direct metal-metal path to the bearing. To the observer from outside the machine, both of these arrangements look similar. Figure V shows how the permanent velocity transducer was mounted, and clearly labelled Bearing 5 Vibration, with this designation carried through to the control room monitoring system (Figure V). Installed adjacent was one proximity transducer arranged to indicate the absolute shaft vibration. Figure V: Vibration transducer on turbine: labelled and appearing to be Bearing Vibration Investigation of this incident was done under some pressure, with management, maintenance and contractor staff standing by anxiously. Knowing of the alternative designs, a cross-section drawing was obtained (Figure VI), and it was established that the cover did not communicate directly with the bearing itself. The permanent Bearing Vibration transducer was sensing vibration of the cover, not of the bearing. It was verified that a plugged hole on the top of the cover could be opened without danger to machine or people. A rod was inserted through it to press firmly on to the bearing top. An accelerometer was held firmly on top of this rod and the vibration of the bearing measured using the test instrument. The level so found was 5.1mm/s rms ( Hz), which was in the satisfactory range of severity according to available standards from experience for bearing cap vibration (Ref 2), [which is consistent with the more recent ISO Standards Ref (3) and (4)]. Frequency analysis revealed that the major component was at 50Hz, with no 300Hz component present. It was therefore recommended that the machine loading up proceed as normal, and further service proceeded with out incident. It was deduced that the cover had a resonance at 300Hz, and similar behaviour was eventually confirmed on other machines of this design. It was unclear why this cover did not show this behaviour before. It was also found that the permanent monitoring system converted the velocity signal from the transducer into displacement, assuming that all vibration was at 50Hz. Such an arrangement may be satisfactory for routine monitoring, but of little use in comparing levels with those contained in accepted Standards. 6

7 Air gap between cover and bearing Cover with transducer mounted on it Top of bearing Figure VI: Cross-section of bearing showing air gap between cover and bearing In this assessment, the shaft absolute vibration measurement was available, but it was not possible to check this locally. Experience with assessing shaft vibration severity is less well established, and standards such as Ref (5), (6) were not available on site at the time of this investigation (which was approaching midnight on a weekend). Ref (7) is the start of a series of ISO Standards, and further developments may result from the current ISO/TC108/SC5 activities. This example shows the importance of being sure of the location and mounting arrangement of permanent transducers, or for that matter, of vibration transducers used for routine manual monitoring. There are no established standards for bearing cover vibration! 7. CONCLUSIONS The results of investigations into some disturbing vibration and noise incidents on large steam turbines showed that the techniques used for condition monitoring are useful also for investigation, the full details of machine construction are needed before established severity standards can be used, overall vibration levels are useful, but more refined diagnosis would be useful. The results should be of interest to those who perform condition monitoring using vibration analysis on such turbomachines. 8. REFERENCES (1) Bigret, R Vibrations and noises of low pressure casing turbines Transactions of the American Nuclear Society Vol 46 Supplement #1, pp (1984) (2) AS2625, Part Rotating and reciprocating machinery mechanical vibration Measurement and evaluation of vibration severity of large machines in situ (3) ISO :1995 Evaluation of machine vibration by measurements on non-rotating parts Part 1: General guidelines (4) ISO :2001 Evaluation of machine vibration by measurements on non-rotating parts Part 2: Landbased steam turbines and generators in excess of 50MW with normal operating speeds of 1500 r/min, 1800 r/min, 3000 r/min, and 3600 r/min. (5) VDI 2059 Part 2 June 1990 Shaft vibrations for steam turbosets for power stations Measurement and evaluation 7

8 (6) ISO7919-1:1996 Mechanical vibration of non-reciprocating machines measurement on rotating shafts and evaluation criteria Part 1 General guidelines (7) ISO :1998 Rotating shaft vibration measuring systems Part 1: Relative and absolute sensing of radial vibration Presented at the International Conference of Maintenance Societies, (ICOMS), Brisbane