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Luis San Andrés Mast-Childs Professor Fellow ASME ASME GT2011-45257 ASME J. Eng Gas Turbines & Power (in print) Thomas Abraham Chirathadam Research Assistant.

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Presentation on theme: "Luis San Andrés Mast-Childs Professor Fellow ASME ASME GT2011-45257 ASME J. Eng Gas Turbines & Power (in print) Thomas Abraham Chirathadam Research Assistant."— Presentation transcript:

1 Luis San Andrés Mast-Childs Professor Fellow ASME ASME GT2011-45257 ASME J. Eng Gas Turbines & Power (in print) Thomas Abraham Chirathadam Research Assistant Texas A&M University ASME TURBO EXPO 2011, Vancouver, Canada (June 2011) Presentation available at Metal Mesh Foil Bearing Effect of Motion Amplitude, Rotor Speed, Static Load, and Excitation Frequency on Force Coefficients

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. Gas Bearings allow weight reduction, energy and complexity savings higher temperatures, without needs for cooling air improved overall engine efficiency

3 Ideal gas bearings 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) available

4 Gas Foil Bearings Used in many oil-free rotating machinery: high load capacity (>20 psig), rotordynamically stable, tolerance of misalignment and shocks……. …… but expensive with intellectual property restrictions. A low cost proven alternative needed.

5 Metal Mesh Foil Bearing (MMFB) MMFB COMPONENTS: Bearing cartridge, metal mesh ring and top Foil Hydrodynamic air film develops between rotating shaft and top foil. Potential applications: ACMs, micro gas turbines, turbo expanders, turbo compressors, turbo blowers, automotive turbochargers, APUs Large damping (material hysteresis) offered by metal mesh Tolerant to misalignment, and applicable to a wide temperature range Coatings needed to reduce friction at start- up & shutdown Metal mesh foil bearing 5 cm

6 MMFB components Top Foil 0.12 mm top foil Chrome-Nickel alloy Rockwell 40/45 Heat treated at ~ 450 ºC for 4 hours and allowed to cool. Foil retains arc shape after heat treatment Sprayed with MoS 2 sacrificial coating Metal mesh pad Compressed weave of copper wires Compactness (density)=20% Stiffness and damping of MMFB depend on metal mesh compactness Bearing cartridge ( +top foil+ metal mesh ) Metal mesh pad and top foil inserted in steel bearing cartridge. Top foil firmly affixed in a thin slot made with wire-EDM machining Simple to manufacture and assemble

7 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) ASME GT-2001-0247 Test metal mesh donut and squirrel cage( in parallel) Metal Mesh damping not affected by modifying squirrel cage stiffness Choudhry and Vance (2005) ASME GT-2005-68641 Develop design equations, empirically based, to predict structural stiffness and viscous damping coefficient METAL MESH DAMPERS provide large amounts of damping. Inexpensive component. Past work in Metal Mesh Dampers

8 Ertas &Luo (2008) ASME J. Gas Turbines Power, 130 MM damper force coefficients not affected by shaft eccentricity (or applied static load) Ertas (2009) ASME J. Gas Turbines Power, 131 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 larger than for Cu mesh (density – 30%) : large motion amplitudes (>10 um) Recent work by OEM with MM dampers to maximize load capacity and to add damping in gas bearings Metal Mesh Dampers in Hybrid bearings

9 Past work in MMFBs San Andrés et al. (2010) J. Eng. Gas Turb. & Power, 132(3) Assembled the first prototype MMFB (L=D=28 mm). Load vs Deflection with hysteresis shows large structural damping   0.7). Frequency dependent stiffness agree with predictions. San Andrés et al. (2009) ASME GT2009-59920 Demonstrated operation to 45 krpm with early rotor lift off. Educated undergraduate students. San Andrés et al. (2010) J. Eng. Gas Turb. & Power, 132 Start and shut down to measure torque and lift-off speed. Low friction factor ~ 0.01 at high speed 60 krpm. San Andrés and Chirathadam (2011) J. Eng. Gas Turb. & Power, 133 Rotordynamic coefficients from unidirectional impact loads. Estimated stiffness and damping force coefficients at 50 krpm.

10 EXPERIMENTS with a PRIOR MMFB (larger mesh thickness) 1. Structural stiffness and damping 2. Friction factor with airborne operation

11 Al-Khateeb & Vance model MMFB structural stiffness vs. freq. At low frequencies (25-100 Hz), stiffness decreases At higher frequencies, stiffness gradually increases Bearing stiffness is frequency and motion amplitude dependent 12.7  m 25.4  m 38.1  m Motion amplitude increases San Andres et al., 2010, ASME J. Eng. Gas Turbines Power, 132 (3)

12 MMFB eq. damping vs. frequency Amplitude increases 12.7 μm 25.4 μm 38.1 μm MMFB equiv. viscous damping decreases as the excitation frequency increases and as motion amplitude increases Al-Khateeb & Vance model San Andres et al., 2010, ASME J. Eng. Gas Turbines Power, 132 (3)

13 Friction coefficient ( f ) decreases with increasing static load Rotor accelerates 8.9 N (2 lb) 17.8 N (4 lb) 26.7 N (6 lb) 35.6 N (8 lb) Friction coefficient vs rotor speed f ~ 0.01 f rapidly decreases initially, and then gradually raises with increasing rotor speed Dry sliding Airborne (hydrodynamic) Dead weight (W D = 3.6 N) Increasing static load (W s ) to 35.6 N (8 lb) f = (Torque/Radius)/(Net static load)


15 MMFB dimensions & materials Metal mesh foil bearing Bearing axial length, L 38.0 mm Journal diameter, D 36.5 mm Bearing cartridge OD 63.57 mm Bearing cartridge ID 42.07 mm Copper mesh thickness, t2.667 mm mesh inner diameter36.74 mm Copper mesh density20 % Wire diameter0.30 mm Steel top foil thickness, t F 0.12 mm Bearing diametral clearance ID-2(t+t F ) ~ 0.0 mm 5 cm Top foil: Chrome nickel alloy Metal mesh: copper Bearing Cartridge: Stainless steel 2.7mm

16 MMFB rotordynamic test rig Max. operating speed: 75 krpm Turbocharger driven rotor Regulated air supply: 9.30bar (120 psig) Test Journal: length 55 mm, 36.5 mm diameter Journal press fitted on Shaft Stub TC cross-sectional view Ref. Honeywell drawing # 448655 Twin ball bearing turbocharger, Model T25, donated by Honeywell Turbo Technologies Bearing

17 Rotordynamic test rig (X-Y 100 N shakers) Dynamic load : 25-100 N Rotor speed 50 krpm Freq. identification range: 200 to 400 Hz Motion amplitudes :  m, 25  m & 30  m Static loads: 22 N (15.5 N along X & Y) and 36N

18 Test rig schematic diagram Squirrel cage affixed on turn- knob controlled positioning table Continuous supply of oil lubricates ball bearings in turbocharger center housing Thermocouple measures bearing outboard end temperature 5 cm TC center housing Oil inlet Shaft stub Air outlet Oil outlet BEARING Squirrel cage (Soft elastic support) Static load Y X BEARING Stinger connection to shaker Load sensor Accelerometer Static load along X Static load along Y Bearing weight Net static load

19 Impact load tests : system mass & soft structure stiffness Impact load along Y direction Squirrel cage structure stiffness < 10% of bearing stiffness Damping ratio =0.024 Accelerance function = physical model equation Estimated test system mass =0.88 kg Y Impact load Bearing overhang on squirrel cage Bearing Sq. cage

20 Identification model K S,C S : soft SQ stiffness and damping M S : effective mass X Y K YY, C YY K XY, C XY Shaker force, F Y Bearing Journal K YX, C YX K XX, C XX Ω EOM : Shaker force, F X K SX, C SX K SY, C SY Soft Support structure K ij,C ij : test bearing stiffness and damping

21 Identification model Forces: Sine sweep excitations (200-400 Hz), amplitude controlled Responses: measure bearing accelerations and displacements relative to journal K ij,C ij : bearing stiffness and damping vs. frequency Process data in frequency domain to obtain: X

22 Dynamic load and displacements Static load resolved along X and Y = 15.5 N Shaft speed=50 krpm (833 Hz) Force along X Displacement along X Displacement along Y Dynamic load 25-100 N Excitation frequency 200- 400 Hz Displacement along X ~ 30  m Noticeable cross-directional motion Net static load (static load-bearing weight) = 22 N along vertical direction X Y

23 Forces & disps. vs. frequency Static loads along X and Y =15.5 N Shaft speed=50 krpm (833 Hz) DFT amplitude of dynamic loads and bearing displacements relative to rotor Average of ten consecutive excitations In frequency domain, displacement magnitude decreases (force increases) with frequency 200-400 Hz Force F YY F XX F XY F YX Displacement Y Y Xx XYXY YXYX N mm X Y

24 MMFB stiffnesses: varying rotor speeds 22 N static load 50 krpm 40 krpm 0 krpm 45 krpm K xx K YY K YX K XY K XX K YY K YX K XY K XX K YY K YX K XY K xx K YX K XY K YY At rest (0 rpm) direct stiffness is structural only. Direct K decreases with rotor speed. With rotation, K YX changes sign. Small cross- stiffneses. Direct stiffnesses gradually increase with frequency

25 MMFB damping: varying rotor speeds 40 krpm 0 krpm 50 krpm45 krpm C YX C YY C XX C XY C XX C YY C XY C YX C YY C XY C YX C XX C YY C YX C XY 22 N static load Rotor speed does not affect damping. Major effect is from metal mesh hysteresis. Direct C increases with frequency. At rest (0 rpm) direct damping is structural only. Direct C decreases with frequency.

26 MMFB stiffnesses: varying motion amplitudes 20  m 25  m 30  m K XX K YX K XY K YY K XX K YX K XY K YY K XX K YX K XY K YY Direct stiffnesses decrease with increasing motion amplitudes. Similar to structural test results San Andres et al (2009) At highest displacement amplitude (30  m), cross-coupled stiffness magnitude is large at ~-0.4 MN/m 22 N static load 50 krpm Rotor speed (833 Hz)

27 MMFB damping: varying motion amplitudes 30  m 25  m 20  m C YX C XX C YY C XY C XX C YY C YX C XY C YY C XX C YX Direct damping decreases slightly with increasing motion amplitude. Direct C increases with frequency. With increasing motion amplitude, cross-damping C YX decreases 22 N Static load 50 krpm Rotor speed (833 Hz)

28 MMFB K & C: varying applied static load 22 N 36 N 22 N For increasing static load: force coefficients are similar in magnitude and show same trend in frequency. K XX K YX K XY K YY K XX K YX K XY K YY C XY C YY C XX C YX C XY C YY C XX C YX 50 krpm Rotor speed (833 Hz) Net static load = 22 N ( W/LD=0.16 bar ) & 36 N ( W/LD=0.26 bar ) W

29 MMFB: estimation of loss factor Proportional structural damping Energy (material damping) = Energy (viscous damping) For elliptical orbits: For circular orbits, Energy dissipation in MMFB largely due to mechanical hysteresis. Aloss factor ( γ ) best represents the material damping

30 Material loss factor    is frequency dependent.   does not depend greatly on displacement amplitudes, rotor speed or static loads MMFB: Loss factor vs. frequency  > 1.0 for all test cases. MMFB has more damping than other types of FBs Typical BFB loss factor ~ 0.1- 0.4Kim et al. (2008)

31 Waterfalls of force and displacement Dominant displacement amplitude corresponds to excitation frequency No sub-synchronous whirl detected. 400 Hz 200 Hz 400 Hz Synchronous ( 833 Hz) 22 N static load 50 krpm Rotor speed (833 Hz)

32 MMFB CONCLUSIONS MMFB shows large energy dissipation,  >~ 1.0 Rotordynamic force coefficients estimated for various rotor speeds, motion amplitudes and static loads: a)MMFB stiffness and damping decrease with increasing bearing displacements. b)MMFB direct stiffness and damping largest without journal rotation (structural values). With rotation, cross-stiffnesses are small c)MMFB direct stiffness increases with frequency while damping increases when rotor spins. d)Similar force coefficients obtained for two static loads: 22 N and 36 N e)MMFB loss factor  is nearly independent of motion amplitude, rotor speed or applied static load ASME GT2011-45257

33 MMFB CONCLUSIONS ASME GT2011-45257 Rule-of-thumb (ROT) model (Dellacorte, 2010) Typical foil bearing stiffness coeff. K~ 2,500-7,500 (L x D) lbf/in 3 damping coeff. C~ 0.1-1.0 (L x D) lbf-s/in 3 MMFB stiffness coeff. K~ 1,330 (Lx D) lbf/in 3 [360 MN/m 3 ] damping coeff. C~ 0.93 (L x D) lbf-s/in 3 [252 MN/m 3 ] Test MMFB is structurally soft with largedamping: Mid-range of rule of thumb (ROT) Net static load (applied load-bearing weight) 22 N ( W/LD=0.16 bar ) and 36 N ( W/LD=0.26 bar )

34 Acknowledgments/ Thanks to Learn more at:  Honeywell Turbocharging Technologies  Turbomachinery Research Consortium Questions (?)

35 Extra slides - >

36 Rotor accelerates Comparison: MMFB&BFB Friction factor vs rotor speed f = (Torque/Radius)/(Static load) f ~ 0.03 Friction coefficient decreases with increasing applied static loads androtor speed (due to lift-off) MMFB BFB Static load

37 Future work: MMFB force coefficient prediction Θ Θl Θl Θ t Θ p X Y eY eY e X h r rp rp r+c m r+ c Metal mesh Rotor Top foil radius with assembled clearance Top foil Fixed end Rectangular finite element with 4 nodes 1 2 4 3 K m w p x y z Metal mesh Top foil Analysis steps: 1.Obtain stiffness matrix for MMFB structure + top foil using FEM. 2.Assume small amplitude motions about a static position. 3.Solve Reynolds equations for isothermal, isoviscous ideal gas. 4.Predict force coefficients using dynamic (perturbed) pressure fields Unwrapped Metal mesh and top foil

38 Demonstrate high temperature reliable operation of MMFB with adequate thermal management. a)Construct two MMFB fitting existing test rig dimensions. b)Measure rotor response for temperature as high as 200 ºC, rotor speed up to 50 krpm c) Compare thermal performance of MMFBs with Gen. I bump-foil bearings Future work: High temperature operation Metal Mesh Foil Bearing

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