Luis San Andrés Mast-Childs Professor Fellow ASME

1. ASME Turbo Expo 2009: Power for Land, Sea, and Air June, 2009 Measurements of Structural Stiffness and Damping Coefficients in a Metal Mesh Foil Bearing Luis San Andrés Mast-Childs Professor Fellow ASME Thomas Abraham Chirathadam Research Assistant Tae-Ho Kim Research Associate Texas A&M University ASME GT2009-59315 accepted for journal publication Supported by TAMU Turbomachinery Research Consortium

2. Metal mesh foil bearings Metal mesh ring and top foil assembled inside a bearing cartridge. Hydrodynamic air film will develop between rotating shaft and top foil. • Metal mesh resilient to temperature variations • Damping from material hysteresis • Stiffness and viscous damping coefficients controlled by metal mesh material, size (thickness, L, D), and material compactness (density) ratio. Potential applications:ACMs, micro gas turbines, turbo expanders, turbo compressors, turbo blowers, automotive turbochargers, APU

3. MMFB Assembly Simple construction and assembly procedure METAL MESH RING BEARING CARTRIDGE TOP FOIL

4. TAMU past work on Metal Mesh Dampers METAL MESH DAMPERS proven to provide large amounts of damping. Inexpensive. Oil-free Zarzour and Vance (2000)J. Eng. Gas Turb. & Power, Vol. 122 Advantages of Metal Mesh Dampers over SFDs Capable of operating at low and high temperatures No changes in performance if soaked in oil Al-Khateeb and Vance (2001)GT-2001-0247 Test metal mesh donut and squirrel cage( in parallel) MM damping not affected by modifying squirrel cage stiffness Choudhry and Vance (2005)Proc. GT2005 Develop design equations, empirically based, to predict structural stiffness and viscous damping coefficient

5. Recent Patents: gas bearings & systems Thrust foil Bearing Foil Journal Bearings Turbocharger with hydrodynamic foil bearings Ref. Patent No. US7108488 B2 ‘Air foil bearing having a porous foil’ Ref. Patent No. WO 2006/043736 A1 A metal mesh ring is a cheap replacement for a “porous foil”

6. Metal Mesh Dampers for Hybrid Bearings Recent work by OEM with MM dampers to maximize load capacity and to add damping in gas bearings Ertas &Luo (2008)ASME J. Gas Turbines Power., 130, pp. 032503-(1-8) MM damper force coefficients not affected by shaft eccentricity ( or applied static load) Ertas (2009) ASME J. Gas Turbines Power, 131 (2), pp. 022503-(1-11) Two metal mesh rings installed in a multiple pad gas bearing with flexural supports to maximize load capacity and damping. Bearing stiffness decreases with frequency & w/o external pressurization; and increases gradually with supply pressure Ertas et al. (2009)AIAA 2009-2521 Shape memory alloy (NiTi) shows increasing damping with motion amplitudes. Damping from NiTi higher than for Cu mesh (density – 30%) : large motion amplitudes (>10 um)

7. Metal Mesh Foil Bearings (+/-) No lubrication (oil-free). NO High or Low temperature limits. Resilient structure with lots of material damping. Simple construction ( in comparison with other foil bearings) Cost effective • Metal mesh tends to sag or creep over time • Damping NOT viscous. Modeling difficulties

8. MMFB dimensions and specifications PICTURE

9. Static load test setup Lathe chuck holds shaft & bearing during loading/unloading cycles. Load cell Eddy Current sensor Stationary shaft Lathe tool holder Test MMFB Lathe tool holder moves forward and backward : push and pull forces on MMFB

10. Static Load vs bearing displacement Start 3 Cycles: loading & unloading Nonlinear F(X) Large hysteresis loop : Mechanical energy dissipation Displacement: [-0.12,0.12] mm Load: [-120, 150 ]N MMFB wire density ~ 20%

11. Derived MMFB structural stiffness MMFB wire density ~ 20% During Load reversal : jump in structural stiffness Lower stiffness values for small displacement amplitudes Max. Stiffness ~ 2.5 MN/m

12. Dynamic load tests Motion amplitude controlled mode • 12.7, 25.4 &38.1 μm Accelerometer Force transducer MMFB • Frequency of excitation : • 25 – 400 Hz (25 Hz interval) • Waterfall of displacement Electrodynamic shaker Test shaft Test shaft Fixture Eddy Current sensors MMFB motion amplitude (1X) is dominant

13. Dynamic load vs excitation frequency 38.1 μm Dynamic load decreases around bearing natural frequency, but increases with further increase in excitation frequency. Dynamic load decreases with increasing motion amplitudes Motion amplitude decreases 25.4 μm 12.7 μm Around bearing natural frequency, less force needed to maintain same motion amplitude

14. X(t) F(t) Parameter identification model 1-DOF equivalent mechanical system Equivalent Test System

15. Parameter identification (no shaft rotation) Harmonic force & displacements Material LOSS FACTOR Viscous Dissipation or Hysteresis Energy Impedance Function

16. Model of metal mesh damping material Stick-slip model (Al-Khateeb & Vance, 2002) Stick-slip model arranges wires in series connected by dampers and springs. As force increases, more stick-slip joints among wires are freed, thus resulting in a greater number of spring-damper systems in series.

17. Design equation: MMB stiffness/damping Empirical design equation for stiffness and equivalent viscous damping coefficients (Al-Khateeb & Vance, 2002) Functions ofequivalent modulus of elasticity (Eequiv), hysteresis coeff. (Hequiv), axial length (L), inner radius (Ri), outer radius (Ro), axial compression ratio (CA), radial interference (Rp), motion amplitude (A), and excitation frequency (ω)

18. Natural frequency of test system Real part of (F/X) vs excitation frequency 12.7 μm • Frequency of excitation : • 25 – 400 Hz ( 25 Hz step) 25.4 μm 38.1 μm Motion amplitude increases Real part of (F/X) decreases with increasing motion amplitude

19. MMFB structural stiffness vs frequency 12.7 um • Frequency of excitation : • 25 – 400 Hz (25 Hz step) At low frequencies (25-100 Hz), stiffness decreases At higher frequencies, stiffness gradually increases Motion amplitudeincreases 25.4 um 38.1 um MMFB stiffness is frequency and motion amplitude dependent Al-Khateeb & Vance model : reduction of stiffness with force magnitude (amplitude dependent)

20. Imaginary impedance (F/X) vs frequency • Frequency of excitation : • 25 – 400 Hz ( at 25 Hz interval) 12.7 μm Motion amplitude increases Im(F/X) decreases with motion amplitude 25.4 μm 38.1 μm

21. Predictions vs. test data: Damping Amplitude increases MMFB equiv. viscous damping decreases as the excitation frequency increases and as motion amplitude increases 12.7 μm 25.4 μm 38.1 μm Predicted equivalent viscous damping coefficientsin good agreement with measurements

22. Loss factor vs excitation frequency • Frequency of excitation : • 25 – 400 Hz ( at 25 Hz step) 25.4 μm Structural damping or loss factor is the largest around the MMFB natural frequency 38.1 μm 12.7 μm Loss factor nearly similar for all motion amplitudes

23. Conclusions • Static and dynamic load tests on MMFB show large mechanical energy dissipation and (predictable) structural stiffness • MMFB stiffness and damping decreases with amplitude of dynamic motion • MMFB equivalent viscous damping decreases with motion amplitude, and more rapidly with excitation frequency • Large MMFB structural loss factor ( g ~ 0.7) around test system natural frequency Predicted stiffness and equivalent viscous damping coefficients are in agreement with test coefficients: Test data validates design equations

24. Thanks to TAMU Turbomachinery Research Consortium Honeywell Turbocharging Technologies Acknowledgments Learn more at http://phn.tamu.edu/TRIBGroup Questions ?

25. Current work

26. MMFB rotordynamic test rig MMFB Journal press fitted on Shaft Stub cm 15 10 5 0 (a) Static shaft TC cross-sectional view Ref. Honeywell drawing # 448655 Max. operating speed: 75 krpm Turbocharger driven rotor Regulated air supply: 9.30bar (120 psig) Twin ball bearing turbocharger, Model T25, donated by Honeywell Turbo Technologies Test Journal: length 55 mm, 28 mm diameter , Weight=0.22 kg

27. 5 cm Test Rig: Torque and Lift-Off Measurements Thermocouple Force gauge String to pull bearing Shaft (Φ 28 mm) Static load MMFB Top foil fixed end Torque arm Positioning (movable) table Preloading using a rubber band Eddy current sensor Calibrated spring

28. Constant speed ~ 65 krpm Valve open Valve close 3 N-mm Rotor speed and torque vs time • Applied Load: 17.8 N Rotor starts Rotor stops WD= 3.6 N • Manual speed up to 65 krpm, steady state operation, and deceleration to rest Iift off speed • Startup torque ~ 110 Nmm • Shutdown torque ~ 80 Nmm • Once airborne, drag torque is ~ 3 % of startup ‘breakaway’ torque Lift off speed at lowest torque : airborne operation Top shaft speed = 65 krpm

29. Varying steady state speed & torque Rotor starts Rotor stops • Manual speed up to 65 krpm, steady state operation, and deceleration to rest 50 krpm 61 krpm 24 krpm 37 krpm • Drag torque decreases with step wise reduction in rotating speed until the journal starts rubbing the bearing 57 N-mm 45 N-mm 2.5 N-mm 2.4 N-mm 2.0 N-mm 1.7 N-mm Side load = 8.9 N WD= 3.6 N Shaft speed changes every 20 s : 65 – 50 – 37 - 24 krpm

30. Bearing drag torque vs rotor speed Lift-off speed 35.6 N (8 lb) 26.7 N (6 lb) 17.8 N (4 lb) 8.9 N (2 lb) Rotor accelerates Max. Uncertainty ± 0.35 N-mm Bearing drag torque increases with increasing rotor speed and increasing applied static loads. Lift-Off speed increases almost linearly with static load