A steam turbine-generator set was overhauled for about two (2) months before it was attempted to put back to services again. However, excessive vibration amplitudes were noticed predominantly at generator drive end (DE) bearing causing the unit to trip several times before reaching full speed. The generator rotor was cleaned and inspected with its bearing reconditioned. Shaft alignment was performed after unit assembly. Excessive vibration amplitudes occurred during startup when turbine speed was around 5500 to 5600 rpm according to plant personnel.
Machine description
The machine train consists of condensing steam turbine with an extraction, speed reduction gearbox, and generator. Steam turbine operating speed is 5816 rpm with overspeed trip at 6398 rpm. Inlet steam pressure and temperature are 5.0 MPa(a) and 410 °C. Extraction steam condition pressure is 1.8 MPa(a) at 8.3 t/h and exhaust steam pressure is 0.008 MPa(a). Gearbox sped ratio is 1.9388. Generator is 2 poles with operating speed at 3000 rpm and output 9900 kW.
Displacement vibration sensors, XY probes are installed at steam turbine drive end (DE) and non-drive end (NDE) bearing, gearbox high-speed (HS) DE bearing, gearbox low-speed (LS) DE bearing, generator DE and NDE bearings. A permanent/online keyphaser is installed at turbine NDE bearing to measure vibration phase angle and speed. All vibration signals are connected to vibration monitoring system (VMS) installed and located inside local control room. A laser keyphaser is temporarily installed at generator DE bearing to get phase angle data of all vibration measurement points of LS shaft. XY sensors are installed as 45°right and left from vertical axis at all bearings. Multi channels vibration analyzer (AS-1250FE with 18+2 channels) and its companion software AS-410 Machine Analyzer was used for data collection, processing, storing, and preparing report.
Vibration data discussion
Low amplitude instability vibration was noticed predominantly at generator DE bearing around 2251 rpm as subsynchronous component 0.5X tracking with machine speed as shown in cascade plots. The amplitudes were then rapidly increasing at speed around 2814 rpm which is about two (2) times of first critical speed of generator at 1425 rpm. The dominant frequency still maintained at around 0.5X with machine speed independent and overall amplitudes about 200 microns, pp.
With vibration monitoring in real time, machine speed was slowly increased with manual speed control mode to observe vibration characteristics and prevent the unit to trip. The speed was then held or reduced when vibration amplitudes increased. Then, machine speed was increased again when amplitudes are stabilized. Finally, it could reach full speed no load condition with acceptable vibration amplitudes. However, this is not a normal startup curve and operator could not accept this new procedure for long-term operation. Cold startup seems to be problematic compared to hot startup as the turbine unit has to be heat soaked at 5400 rpm for about one (1) hour before ramping up to full speed without tripping. This is to improve hot alignment condition as well as clearance between rotor and casing due to thermal growth, hence, increased stability margin to prevent excessive vibration amplitude.
Figure 1: Cascade plots of vibrations spectrum versus machine speed during startup
Conclusions
Observed vibration characteristics were due to fluid induced instability phenomena with threshold speed around 2251 rpm and machine speed dependent at 0.5X vibration so called “Oil Whirl” which is originated at generator DE bearing. It became more severe problem with excessive vibration amplitudes when oil whirl frequency excited first balance critical speed of generator (1425 rpm) and locked to this frequency when machine speed increased e.g., speed independent vibration so called “Oil Whip” phenomena. The corresponding orbit shape is not as circular as expected [1] which could be due to elliptical bearing shape, hence, preload from the bearing. Vibration response was sensitive to rotor eccentricity as vibration amplitudes increased with instability phenomena when the rotor position is approaching bearing center (lower eccentricity) and vice versa. Change in rotor eccentricity can change rotor-bearing stiffness, damping, and average oil velocity within the bearing. This influences change in stability threshold speed. There is higher stiffness and damping but lower average oil velocity at higher rotor eccentricity (close to bearing wall), hence, higher stability threshold speed. The problem disappeared when the stability threshold speed is higher than running speed. The formula of this fundamental relationship can be found in [2]. A similar case study with analytical model can be also found in [3] when hot alignment condition plays an important role to control rotor-bearing stability.
Figure 2: Average shaft centerline plots with overlay orbit shape when instability vibration occurred
Recommendations
The unit revealed excessive vibration amplitudes after major overhaul with some work activities highlighted above, therefore, there is suspected improper shaft alignment condition and/or excessive bearing clearances causing instability vibration problem. It is recommended to check and adjust alignment condition in the way to increase rotor eccentricity. Oil and air seal rings should be also checked for evidence of rubbing as a consequence of excessive vibration amplitude due to instability. It shows no excessive vibration amplitude throughout speed range during startup as well as load variation until baseload after realignment. The problem was solved.
References
[1] D. E. Bently, Fundamentals of Rotating Machinery Diagnostics, NV, United States: Bently Pressurized Bearing Co, 2003.
[2] D. E. Bently and A. Musynzynska, "Fluid-Generated Instabilities of Rotors," Orbit Magazine, April 1989.
[3] A. Vania and P. Pennacchi, "Effects of the Hot Alignment of a Power Unit on Oil-Whip Instability Phenomena," International Journal of Rotating Machinery, 2010.