A good example of this situation occurred on a refrigeration compression train. This two-case train consisted of a horizontally split compressor containing seven impellers, with a rotor weight of 750 pounds. The steam turbine driver contained a three-stage rotor that weighed 800 pounds. This turbine was rated for 5,000 HP at 9,000 RPM. The machines were connected with a lubricated gear type coupling. Both machines contained five pad tilting pad bearings, with nominal diametrical clearances between 5 Mils and 7 Mils. The refrigeration train had a good operating history, with only occasional problems at the turbine exhaust end bearing. In order to ensure continued reliability of this train, a turbine bearing inspection was scheduled for an upcoming turnaround.

Bearing inspection of the turbine exhaust end bearing revealed the pads and the shaft journal in excellent condition. The measured diametrical clearance of 6 Mils was within specifications. No other maintenance work was performed and the housing was reassembled using the same bearing assembly. Maintenance personnel were confident that the train was in excellent condition and the previous issues on the turbine exhaust bearing were nonexistent.

Successful startup by operations occurred late one evening. The ramp up was smooth and the machine was lined out at a constant speed of 8,955 RPM. Overall shaft vibration amplitudes at each of the four radial bearings were acceptable and thrust positions for both rotors had returned to their previous operating positions. The only abnormality was a high temperature of 210 F at the turbine exhaust end bearing. This bearing historically operated between 160 F and 170 F as measured by a thermocouple embedded in the bottom bearing pad.

Bearing temperature increased to 215 F during the warmth of the next day. At this point, concern began to develop about the longevity of this bearing. Various inspections and examinations were performed with inconclusive results. In an effort to understand the shaft vibration, plant personnel employed a spectrum analyzer to produce FFT plots of the proximity probes. A typical set of spectrum plots across the coupling are shown in Figure 1.

FFT data from the Y-axis proximity probe at the turbine exhaust and the coupling end compressor bearing are displayed in Figure 1. Note that both spectrum plots reveal acceptable amplitudes at 1x rotational speed, plus a string of running speed harmonics (i.e., 2x, 3x, 4x, etc.). On the turbine exhaust, the amplitude of the second harmonic was twice as large as the fundamental 1x component. This vibration data was of concern to the plant personnel and two different opinions were soon openly debated.

Some people were convinced that the series of multiple harmonics were due to mechanical looseness of the turbine bearing. Others had the opinion that the vibratory behavior was due to misalignment across the gear coupling. One proposed solution was to monitor vibration as the compressor's hold down bolts were loosened and the compressor was allowed to move into a state of satisfactory hot alignment. For the uninitiated, it must be recognized that this type of action is extraordinarily dangerous. There is massive energy contained in an 800-pound rotor rotating at 8,955 RPM. The danger of releasing that energy by unloosening the hold down bolts is obvious and is not justifiable. Fortunately, this activity was not implemented and it was agreed to just continue monitoring the machinery.

Both theories were downgraded following several days of operation. Each day, shaft vibration amplitudes remained constant and the turbine exhaust bearing temperature cycled between 210 F and 215 F. Clearly, if the turbine bearing was loose, the behavior would tend to degenerate with time and that did not occur. The misalignment theory was also discounted. Specifically, if heat generation in a bearing was due to misalignment, it is logical to believe that the bearing would either fail or relieve some clearance due to the applied preloads. It is hard to believe that any significant misalignment would appear as constant vibration and elevated temperature without any change. Furthermore, the maintenance work performed during the previous turnaround did not disturb the coupling alignment. At this time, it was necessary to examine the machinery behavior in more detail to resolve this problem.

Time domain presentation of the vibration signals are shown in Figure 2. At both measurement planes, the time base signals are corrupted by a series of repetitive spikes. These spikes are generally indicative of shaft surface imperfections. There is a distinctively different pattern between the turbine signal and the compressor probe signal. However, both cases are representative of rough shaft surfaces below the respective probes. Comparison of the Y-axis data with the associated X-axis probes at each measurement location (not shown) reinforces the fact that the spikes are shaft surface scratches. Hence, the majority of the harmonic activity shown in Figure 1 is simply due to shaft surface imperfections.

Continued hot operation of the turbine bearing was an indisputable fact. Bearing housing measurements with a surface pyrometer confirmed that the housing was hotter than normal. Unfortunately, the approaching hot summer days would only aggravate this situation. It was hard for the plant personnel to justify an entire plant shutdown to investigate this hot bearing without any real inspection objectives. At this point, a set of casing velocity measurements was acquired across the coupling. The resultant FFT plots are exhibited in Figure 3. This data shows that higher order harmonics do not exist on the bearing housings. As with many turbines of this general size, the exhaust end bearing is a fairly simple assembly. There is normally a close relationship between the frequencies of vibration components measured on the shaft versus the casing. More specifically, if the turbine shaft was really subjected to a severe misalignment condition, a strong twice rotational speed component should be visible.

No real evidence appeared to support the previous theories of loose bearing or misalignment. Thus, the source of the hot turbine bearing remained unanswered. Like other machinery problems, it is mandatory to thoroughly examine the operating equipment. This approach includes "go out and look, touch, feel, smell and listen to the machinery." In many cases, you do not know what you are looking for so the best advice is to look for any peculiarities. On this train, it was finally observed that the oil flowing through the bearing drain sight glass was minimal. This simple observation was previously overlooked by everyone. There is an old adage that states: "For every 10 drops of oil, only one drop is for lubrication and the remaining nine drops are for cooling." On this bearing, there was evidently enough oil flow for lubrication, but not much left over for cooling.

Testing this hypothesis of restricted oil flow was accomplished by raising lube oil supply pressure from 20 psig to 25 psig. This change was carefully monitored to ensure there was no detrimental effects to the other machine train bearings. As the oil supply pressure was gradually increased, the bearing temperature dropped. At an oil pressure of 25 psig, the 215 F bearing temperature was reduced to 203 F. In addition, the oil flowing through the drain sight glass did perceivably increase.

In conclusion, the refrigeration train operated successfully through the hot summer in this manner and ran for another six months. At that time, the bearing was opened during a short plant outage. It was discovered that Permatex® was blocking the oil inlet to the turbine exhaust bearing. After this blockage was removed, subsequent turbine operating bearing temperatures returned back to normal levels between 160 F and 170 F.

Some Lessons To Be Learned:

  1. Don't make quick judgments based on poor, inconclusive, or conflicting data.
  2. On complex dynamic vibration data, don't get locked into a single form of data evaluation and presentation.
  3. Don't get fooled by the data processing characteristics of any analytical instrument.
  4. Always consider overall vibration characteristics of the machinery, including relative shaft vibration and position, plus absolute casing motion.
  5. Always exam traditional machinery operational parameters, such as oil flow, valve positions, pressures, temperatures, seal parameters, etc.
  6. Always look for a confluence of information and be cautious of diverging indicators.

This article was extracted from Robert Eisenmann's book: Machinery Malfunction Diagnosis and Correction, pages 343 to 347. Original Publication Date: November 26, 1997. Author currently owns copyright.

Robert C. Eisenmann, Sr., is President of Wilpat, Inc., providing technical training in machinery malfunctions and vibration analysis. He has dealt with process machinery throughout his 45-year career beginning with operations and maintenance experience at Shell Oil and Northern Petrochemical. He is a retired registered Professional Engineer in Illinois, Nevada and Texas. His initial undergraduate work was at the University of Illinois, and he is a 1965 Mechanical Engineering graduate of the Illinois Institute of Technology. www.wilpat.biz

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