Failure of a Test Rig Operating with Pressurized Gas Bearings: a Lesson on Humility

Transcription

1 Proceedings of ASME Turbo Expo 2015: Turbine Technical Conference and Exposition, June 15-19, 2015, Montreal, Canada GT Failure of a Test Rig Operating with Pressurized Gas Bearings: a Lesson on Humility Luis San Andrés Mast-Childs Chair Professor Fellow ASME Michael Rohmer Graduate Research Assistant Texas A&M University Sangshin Park Professor Yeungnam University 1

2 Oil-Free Bearings for Turbomachinery Justification Current advancements in vehicle turbochargers and midsize gas turbines need of proven gas bearing technology to procure compact units with improved efficiency in an oil-free environment. DOE, DARPA, NASA interests range from applications as portable fuel cells (< 60 kw) in microengines to midsize gas turbines (< 250 kw) for distributed power and hybrid vehicles mandate on + efficiency for IC engines: materials and oil-free bearing systems will enable 55 mpg (23 km/h). 2

3 Gas Bearings Ideal gas bearings for micro turbomachinery (< 0.5 MW ) must be: Simple low cost, small geometry, low part count, constructed from common materials, manufactured with elementary methods. Load Tolerant capable of handling both normal and extreme bearing loads without compromising the integrity of the rotor system. High Rotor Speeds no specific speed limit (such as DN) restricting shaft sizes. Small Power losses. Good Dynamic Properties predictable and repeatable stiffness and damping over a wide temperature range. Reliable capable of operation without significant wear or required maintenance, able to tolerate extended storage and handling without performance degradation. +++ Modeling/Analysis (anchored to test data) readily available 3

4 Objective Evaluate the performance of gas lubricated hybrid radial and thrust bearings for high speed rotating machinery: Externally pressurized bearings allow for rub-free operation at start up & shut down. Major issues with gas bearings: Little damping & Instability (whirl & hammer) 4

5 Gas bearings at TAMU Develop experimentally validated computational tools for predicting the performance of radial and thrust gas bearings (GFB, MMFB etc). Performance Characteristics: - Structural mechanics - Drag torque, power loss - Dynamic force coefficients - High temperature performance Bump type foil bearing Aid to system development: - Oil-free turbocharger - Water aeration systems - CO2 turbo expanders Tilting pad bearing Metal mesh foil bearing Work concentrated on radial gas bearings. 5

6 Brief Literature Review ½ frequency whirl is well known in (lightly loaded) hydrodynamic bearings. The whirl is usually benign until locking to a natural frequency to produce an instability (large amplitude motion). ASME J. Eng. Gas Turbines Power, 128 Osborne and San Andrés (2006) On a rigid rotor on hydrostatic gas bearings: As supply pressure into bearings increases, the system critical speed increases and the damping ratio decreases the threshold speed of instability increases. Int. Conf. on Noise and Vibration Engineering Int. Conf. on Sustainable Construction and Design Waumans et al. (2006, 2011) Micro-turbine rotor on gas bearings: Sub synchronous whirl is more apparent at low supply pressure and high rotor speed. Methods to improve the stability of gas bearings decrease film clearance and modify bearing geometry. 6

7 A thrust bearing test rig Funded by USET (AF) program ( ). 7

8 Program Objectives Hydraulic shaker load shaft test rotor coupling drive motor Validate hybrid thrust bearings predictive tools for application to cryogenic turbo pumps. Measurements of forced performance of water lubricated hybrid thrust bearings for operation with high speed (25 krpm) and high supply pressure (70 bar). Measurements show remarkable correlation with predictions from XLHYDROTHRUST model. 8

9 Test Rig Description Water lubricated bearings (a) Motor drives rotor through coupling. (b) Two radial bearings support rotor. (c) Load shaft applies load to test thrust bearing and pushes on thrust collar in the shaft. (d) Rotor displaces and slave thrust bearing reacts load. 9

10 Test Rig Description Water lubricated bearings Aerostatic Bearings Test Thrust Bearing Journal Bearings Slave Thrust Bearing Radial Bearing Thrust Bearing Load Shaft cm Rotor Housing 10

11 Coupling Description a) Quill shaft threads into rotor. b) Hub clamp attaches to motor shaft. c) Diaphragm isolates motor from axial loads. d) Diaphragm and quill shaft are flexible to allow for misalignment. 11

12 2013 Modified thrust bearing test rig Redesign and manufacture radial bearings for operation with air. Major objective is was to evaluate the dynamic forced performance of hybrid thrust bearings and thrust foil bearings (gas lubricated). 12

13 Gas Bearing Description Flexure, Pivot Tilting Pad Hydrostatic Gas Bearing 330 Bearing Bronze Pivot Ø0.25 mm Ø1.4 mm Main Pad (Flexible) Supplementary Arc (Fixed) Inner Diameter Outer Diameter Length 3.81 cm 7.62 cm 3.81 cm Journal Rotation # Pads 4 Arc Length 72 Pivot Offset 0.6 Radial Clearance 76 μm Air Injection Orifice Diameter 1.4 mm Arc Length 18 Radial Clearance 76 μm Orifice Diameter 0.25 mm 13

14 Gas lubricated thrust bearing test rig Modifications took several months to complete. After rotor-bearing-coupling system assembly and alignment, verification of operation without pneumatic hammer, test system was ready to operate with shaft rotational speed (thrust bearings NOT active)..... and the following happened on day one of test rig operation. 14

15 Maiden Operation No active thrust bearings Amplitude [μm, p-p] rpm 5.14 bar (a) air into bearings FAILURE c) Contact between shaft and bearings at 5.14 bar(a) and 28 krpm. d) Emergency stop initiated Rotor Speed [rpm] a) Operating conditions a) Air supply pressure to bearings = 7.89, 6.52, 5.14 bar(a) b) Rotor speed = 0 25 krpm c) Supply pressure = 5.14 bar(a), rotor speed > 25 krpm b) Natural frequency ~ 6 krpm (100 Hz), damping ratio ~ 6% 15

16 Troubleshooting Operation a) Large 1X amplitude rotor motion recorded while crossing lowest critical speed (6 krpm). b) Above 12 krpm, operator taps rotor to excite it. Rotor motion appears at first natural frequency (~ 100 Hz) and decays stable. c) At high speed (~23 krpm), decay time increases to ~7 s while rotor vibrates at first natural frequency. 80 Findings: Rotor-bearing system has little damping (6%) at its first natural 60 frequency. When operating above krpm, rotor is sensitive to show 20 subsynchronous whirl motions at its 0 first natural frequency

17 Maiden Operation Decrease in supply pressure Rotor displacement recorded at collar on rotor free end: 134 Hz 142 Hz Speed increases 7.89 bar(a) air into bearings 126 Hz 6.52 bar(a) air into bearings 17

18 Maiden Operation Decrease in supply pressure 116 Hz Findings: Subsynchronous rotor motion appear. As supply pressure increases, amplitude and frequency of subsynchronous vibration (SSV) increases. Speed increases 5.14 bar(a) air into bearings Whirl frequency ratio (WFR)=0.50 Bearing Supply Pressure 5.14 bar 6.52 bar 7.89 bar 7.89 bar Rotor Speed 233 Hz (14 krpm) 250 Hz (15 krpm) 267 Hz (16 krpm) 283 Hz (17 krpm) Frequency of SSV Amplitude of SSV 116 ± 2 Hz 11 μm 126 ± 2 Hz 13 μm 134 ± 2 Hz 18 μm 142 ± 2 Hz 31 μm 18

19 Post-Mortem

20 Damaged Components Tilting Pad Welded Rotor contacted radial bearings and generated profuse heat that welded the pads to the bearing. Rotor severely worn & twisted. Damage at Bearing Support Locations Damage at Eddy Current Sensor Locations 20

21 Damaged Components Ruptured Diaphragm Emergency stop placed large torque that ruptured coupling diaphragm while rotor kept spinning. Sensors damaged, seals damaged, pins and connecting bolts sheared off. 21

22 What Caused the Failure?

23 What do we need to know to explain/predict the failure? a) Structural analysis of rotor and coupling. b) Force coefficients of gas bearings. c) Accurate model of rotor-bearing system. d) System natural frequencies and damping ratios. e) System imbalance response. 23

24 Rotor and Coupling Natural Frequencies Measurement [Hz] Prediction [Hz] Mode Shape Rotor 1,760 ± 8 1,817 Coupling 496 ± ,504 ± 8 1,535 Finding: Structural models of rotor and coupling predict well their free-free mode natural frequencies. 24

25 Gas Bearing Stiffness vs. Speed Supply pressure: 5.14 bar(a) W = 19 N T = 24 C Bearing Stiffness [MN/m] Kxx Kyy Kxy -0.2 Kyx Rotor Speed [rpm] Findings: Bearing direct (hydrostatic) stiffnesses grow slowly with shaft speed. Small cross-coupled stiffnesses due to flexure webs. ASME J. Tribol., 128 San Andrés (2006) Hybrid Flexure Pivot-Tilting Pad Gas Bearings: Analysis & Experimental Validation 25

26 Gas Bearing Damping vs. Speed Supply pressure: 5.14 bar(a) W = 19 N T = 24 C Bearing Damping [N-s/m] Cyy Cxx Cyx Cxy Rotor Speed [rpm] Findings: Direct damping coefficients decrease slowly as shaft speed increases. Cross-coupled damping coefficients are negative and a fraction of the direct damping. ASME J. Tribol., 128 San Andrés (2006) Hybrid Flexure Pivot-Tilting Pad Gas Bearings: Analysis & Experimental Validation 26

27 Gas Bearing Stiffness vs. Pressure Shaft speed: 28 krpm W = 19 N T = 24 C Bearing Stiffness [MN/m] Kxx Kyy Kxy Kyx Air Pressure Supply [bar(a)] Findings: Bearing direct (hydrostatic) stiffnesses increase little with supply pressure. Small cross-coupled stiffnesses due to flexure webs. ASME J. Tribol., 128 San Andrés (2006) Hybrid Flexure Pivot-Tilting Pad Gas Bearings: Analysis & Experimental Validation 27

28 Gas Bearing Damping vs. Pressure Shaft speed: 28 krpm W = 19 N T = 24 C Bearing Damping [N-s/m] Cyy Cxx Cyx Cxy Air Pressure Supply [bar(a)] Findings: Bearing direct damping decreases as supply pressure increases. Cross-coupled damping coefficients are negative and a fraction of the direct damping. ASME J. Tribol., 128 San Andrés (2006) Hybrid Flexure Pivot-Tilting Pad Gas Bearings: Analysis & Experimental Validation 28

29 Rotor-Coupling-Bearing System Model Question: Is coupling dynamics decoupled from rotor dynamics? 29

30 Free-free modes of rotor & coupling Measurement [Hz] Prediction [Hz] Free-Free Mode Shape 104 ± ± ,272 ± 8 1,332 1,944 ± 8 2,029 Findings: Modes show quill shaft is too flexible. For operation below 30 krpm, both quill shaft and rotor operate as single unit. 30

31 System natural frequencies & 1 st critical speed Predicted critical speed agrees with measured one at 6 krpm [100 Hz]. Findings: Gas bearing stiffness does not affect natural frequencies and critical speed (6 krpm [100 Hz]) as flexibility of quill shaft determines its placement. Bearing clearance is too large for adequate stiffness. 31

32 System Damping Ratio Predicted instability at 14 krpm with whirl frequency ratio (WFR)= As supply pressure into bearings increases, threshold speed of instability also increases. Findings: Damping ratio is very low because most motion is at quill shaft (no damping). At natural frequency, predicted damping ratio (ζ = 0.001) is lower than estimated damping ratio (ζ = 0.06). 32

33 Nat freq. & damping ratio vs supply pressure Supply pressure into bearings 5.14 bar 6.52 bar 7.89 bar 1 st Natural Frequency 99 Hz 100 Hz 101 Hz Damping Ratio, ζ Threshold Speed of Instability 14 krpm 16 krpm 21 krpm Whirl Frequency Ratio nd Natural Frequency 563 Hz 563 Hz 563 Hz Damping Ratio, ζ Predictions: Quill shaft dominates rotordynamics: first and second natural frequencies remain constant as supply pressure increases. Damping ratio is very low and threshold speed of instability increases with supply pressure. 33

34 Response from Imbalance Measured Response Prediction vs data 5.14 bar(a) air into bearings Amplitude [μm, p-p] 6000 rpm Imbalance failure Rotor Speed [rpm] Measured Findings: Predicted 1X response is similar to measured response. Operation >25 krpm gives increasing amplitude of motion as system approaches its second natural frequency (34 krpm [563 Hz]). 34

35 Failure Description Test rig experienced catastrophic failure. Rotor and bearings suffered extensive surface damage. Shaft rubbed against bearings. Rotor experienced large amplitude motions. Coupling diaphragm ruptured. Coupling twisted while reacting to rotor torque. Motor instantaneously stopped, rotor spinning. Gas bearings provide small damping. Bearings have large clearance. Poorly designed gas bearings. Operator continued to increase rotor speed. Operator disregarded early signs of SSV (potential road to an instability). Operator rushed to make measurements. Operator initiated an emergency stop. Operator did not allow enough time to safely modify test rig and conduct experiments. 35

36 Conclusion (a) Modified test rig to operate with air lubricated hybrid journal bearings and thrust bearings. (b) During maiden operation: contact between the shaft and bearings led to catastrophic failure of the test rig. (c) Large amplitude rotor motions are a result of hydrodynamic instability from gas bearings plus flexibility of quill shaft. (d) A lesson: incident could have been avoided had operators not ignored early signs of an instability (lack of damping). 36

37 Acknowledgments Thanks to TAMU Turbomachinery Laboratory & Turbomachinery Research Consortium. Questions: Learn more at 37

38 Gas Bearing Stiffness vs. Frequency Supply pressure: 5.14 bar(a) Shaft speed: 28 krpm W = 19 N T = 24 C Bearing Stiffness [MN/m] Kxx Kyy Kxy Kyx Excitation Frequency [Hz] 1X Findings: Bearing direct (hydrostatic) stiffnesses grow slowly with frequency. Cross-coupled stiffnesses decrease. ASME J. Tribol., 128 San Andrés (2006) Hybrid Flexure Pivot-Tilting Pad Gas Bearings: Analysis & Experimental Validation 38

39 Gas Bearing Damping vs. Frequency Supply pressure: 5.14 bar(a) Cyy Shaft speed: 28 krpm W = 19 N T = 24 C Damping [N-s/m] Cxx Cyx Cxy Excitation Frequency [Hz] 1X Findings: Direct damping coefficients decrease as excitation frequency increases. Same for magnitude of negative cross-coupled damping. ASME J. Tribol., 128 San Andrés (2006) Hybrid Flexure Pivot-Tilting Pad Gas Bearings: Analysis & Experimental Validation 39