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Fig. 7. The damaged 3™ bearing rotor orbit/time base plot at 2758 rpm during shut down

Figure 7 The damaged 3™ bearing rotor orbit/time base plot at 2758 rpm during shut down

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Fig. 1. Turbounit scheme: high DSR, middle VSR and low MSR pressure steam turbine rotors, generator and exciter rotors (GR, ZR); 1, 3-10 journal bearings, 2-thrust tilting pad bearing; 1XY,...,8X Y-non- contacting sensors; EC1, SP1,..., SP3, ARSP1,2, KAP1 - displacement transducers; RSG - rotation speed sensors; FK - Keyphasor; 1GV, 2GA...,10GH - seismic transducers (V vertical, H horizontal and A axial) ee The rotating system of 110 MW power turbounit and 3000 rpm rotation speed with three steam turbines, electric generator, and exciter and equipped with new stationary condition monitoring and diagnostic system BNC 3500 (USA) illustrated in Fig. 1. The modern MDS was installed after 25 years of the continuous operation of the machine. Primary stationary condition monitoring system was based only on bearings housings absolute vibration velocities measurements V;ms values without diagnostics. Vibration displacement measurements of the rotors relative to the bearings had not been performed earlier.
Fig. 1. Turbounit scheme: high DSR, middle VSR and low MSR pressure steam turbine rotors, generator and exciter rotors (GR, ZR); 1, 3-10 journal bearings, 2-thrust tilting pad bearing; 1XY,...,8X Y-non- contacting sensors; EC1, SP1,..., SP3, ARSP1,2, KAP1 - displacement transducers; RSG - rotation speed sensors; FK - Keyphasor; 1GV, 2GA...,10GH - seismic transducers (V vertical, H horizontal and A axial) ee The rotating system of 110 MW power turbounit and 3000 rpm rotation speed with three steam turbines, electric generator, and exciter and equipped with new stationary condition monitoring and diagnostic system BNC 3500 (USA) illustrated in Fig. 1. The modern MDS was installed after 25 years of the continuous operation of the machine. Primary stationary condition monitoring system was based only on bearings housings absolute vibration velocities measurements V;ms values without diagnostics. Vibration displacement measurements of the rotors relative to the bearings had not been performed earlier.
The condition monitoring and diagnostics of turbounit was performed continuously between two overhauls and temporal variation of relative vibration displacements of eight bearing shafts during operation at maximum loading is presented in Fig.2. Vibration displacements monitoring results indicate that low pressure rotor MSR 5" bearing had severe vibration displacement values since the first overhaul up to the second. Fig. 2. Maximum values of relative vibration displacements smax of the rotor bearing shafts under maximum loading at continuous operation: 1 — 2005 (after first overhaul), 98 MW; 2 — 2007, 99 MW; 3 — 2008; 106 MW and 4 — 2009 (before second overhaul), 102 MW
The condition monitoring and diagnostics of turbounit was performed continuously between two overhauls and temporal variation of relative vibration displacements of eight bearing shafts during operation at maximum loading is presented in Fig.2. Vibration displacements monitoring results indicate that low pressure rotor MSR 5" bearing had severe vibration displacement values since the first overhaul up to the second. Fig. 2. Maximum values of relative vibration displacements smax of the rotor bearing shafts under maximum loading at continuous operation: 1 — 2005 (after first overhaul), 98 MW; 2 — 2007, 99 MW; 3 — 2008; 106 MW and 4 — 2009 (before second overhaul), 102 MW
Fig. 3 reveals that the most sensitive to the loading from the vibration intensity point of view are DSR 2™ and VSR 3™ bearings when loading varies from maximum value 102 MW up to free run (FR) when tested before scheduled second overhaul (2009). The 2™ and 3" bearing vibration values indicate significant changes in vibration displacement amplitudes and in phases. Fig. 3. Maximum values of relative vibration displacement Smax (2 Smax ~ Sp-pmax) Of the eight bearing shafts measured before 2009 scheduled major overhaul at varying loading from 102 MW up to FR: 1 — 2009 03 20, 102 MW; 2 — 2009 04 15, 86 MW; 3- 29 MW and 4 - 2009 04 16, FR
Fig. 3 reveals that the most sensitive to the loading from the vibration intensity point of view are DSR 2™ and VSR 3™ bearings when loading varies from maximum value 102 MW up to free run (FR) when tested before scheduled second overhaul (2009). The 2™ and 3" bearing vibration values indicate significant changes in vibration displacement amplitudes and in phases. Fig. 3. Maximum values of relative vibration displacement Smax (2 Smax ~ Sp-pmax) Of the eight bearing shafts measured before 2009 scheduled major overhaul at varying loading from 102 MW up to FR: 1 — 2009 03 20, 102 MW; 2 — 2009 04 15, 86 MW; 3- 29 MW and 4 - 2009 04 16, FR
Fig. 4. 3 bearing’s shaft relative vibration displacement Sp-p and phase trend plots at shutdowns (rubbings at ~2789 rpm direct S))=185-250 um; 1X=46,48 Hz frequency vibration amplitude s),=71um; at resonance 1398 rpm Sp.pres=1 40-151 um
Fig. 4. 3 bearing’s shaft relative vibration displacement Sp-p and phase trend plots at shutdowns (rubbings at ~2789 rpm direct S))=185-250 um; 1X=46,48 Hz frequency vibration amplitude s),=71um; at resonance 1398 rpm Sp.pres=1 40-151 um
© VIBROENGINEERING. JOURNAL OF VIBROENGINEERING. DECEMBER 2010. VOLUME 12, ISSUE 4. ISSN 1392-8716
© VIBROENGINEERING. JOURNAL OF VIBROENGINEERING. DECEMBER 2010. VOLUME 12, ISSUE 4. ISSN 1392-8716
Fig. 5. Abnormal ASCP of the damaged 3" bearing measured at shutdown and run-up before overhaul (see Fig. 4), 2009
Fig. 5. Abnormal ASCP of the damaged 3" bearing measured at shutdown and run-up before overhaul (see Fig. 4), 2009
Fig. 8. Finite elements: a — rotor FEM; b — journal bearing FEM rotating system. The rotor dynamics is simulated by FEM when finite element consists of two nodes and five degrees of freedom (DOF) at each node (Fig. 8 a). The first and second DOF are displacements along y and z axes and the last three DOF are angles about X, Y and Z axes. Journal bearing hydrodynamics is simulated by FEM when the isoparametric nine-node finite elements are used (Fig. 8 b). Solution of dynamic equilibrium equations The system of equations of motion of the rotating system can be written in the form:
Fig. 8. Finite elements: a — rotor FEM; b — journal bearing FEM rotating system. The rotor dynamics is simulated by FEM when finite element consists of two nodes and five degrees of freedom (DOF) at each node (Fig. 8 a). The first and second DOF are displacements along y and z axes and the last three DOF are angles about X, Y and Z axes. Journal bearing hydrodynamics is simulated by FEM when the isoparametric nine-node finite elements are used (Fig. 8 b). Solution of dynamic equilibrium equations The system of equations of motion of the rotating system can be written in the form:
Fig. 9. Geometry of the journal bearing
Fig. 9. Geometry of the journal bearing
Fig. 10. Geometry of the sliding surface of the journal bearing
Fig. 10. Geometry of the sliding surface of the journal bearing
Fig. 12. Pressure (Pa) distribution on the 3™ journal bearing surface: a- w=O degrees;b- w = 200 degrees
Fig. 12. Pressure (Pa) distribution on the 3™ journal bearing surface: a- w=O degrees;b- w = 200 degrees
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