GAS FOIL BEARINGS FOR OILFREE ROTATING MACHINERY

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Escuela Politecnica Nacional, Quito, Ecuador
Metal Mesh Foil Bearings for Oil-Free
Turbomachinery
Luis San Andrés Mast-Childs Professor Fellow ASME
STLE
Texas AM University
Julio 9, 2009
Supported by TAMU Turbomachinery Research
Consortium Honeywell Turbocharging Technologies
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Micro Turbomachinery (lt 0.5 MW)
ADVANTAGES

• High energy density
• Compact and fewer parts
• Portable and easily sized
• Lower pollutant emissions
• Low operation cost

http//www.grc.nasa.gov/WWW/Oilfree/turbocharger.h
tm
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Gas Foil Bearings Bump type

• Series of corrugated foil structures (bumps)
  assembled within a bearing sleeve.
• Integrate a hydrodynamic gas film in series with
  one or more structural layers.

PROVEN TECHNOLOGY!! Applications Aircraft ACMs,
micro gas turbines, turbo expanders, turbo
compressors,

• Damping from dry-friction and operation with
  limit cycles
• Tolerant to misalignment and debris, also high
  temperature
• Need coatings to reduce friction at start-up
  shutdown
• Often need cooling flow for thermal management of
  rotor-GFB system

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Foil Bearing Research at TAMU
2003-2009 Funded by NSF, Capstone Turbines, NASA
GRC, Turbomachinery Research Consortium
Reference DellaCorte (2000) Rule of Thumb

• Test Gas Foil Bearing (Bump-Type)
• Generation II. Diameter 38.1 mm
• 25 corrugated bumps (0.38 mm of height)

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Gas foil bearings for oil free machinery

• Eliminate complex bulky oil lubrication system

• Operate at elevated temperatures

• Eliminate sealing

• Reduce system overall weight ( High power density)

• Enhanced reliability at high rotating speeds
• Large inherent damping prevents potentially
  harmful rotor excursions

• Extended maintenance intervals

• Low power loss

• Simple assembly procedure using cheap,
  commercially available materials

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Part I
Measurements of Structural Stiffness and Damping
Coefficients in a Metal Mesh Foil Bearing
Thomas Abraham Chirathadam Research Assistant
Tae-Ho Kim Research Associate
Luis San Andrés Mast-Childs Professor Fellow ASME
STLE
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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, APU

• Large damping (material hysteresis) offered by
  metal mesh
• Tolerant to misalignment, and applicable to a
  wide temperature range
• Suitable tribological coatings needed to reduce
  friction at start-up shutdown

Cartridge
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MMFB Assembly
Simple construction and assembly procedure
METAL MESH RING
BEARING CARTRIDGE
TOP FOIL
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Recent Patents gas bearings systems
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
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Metal Mesh Dampers for Hybrid Bearings
METAL MESH DAMPERS proven to provide large
amounts of damping. Inexpensive. Oil-free
Recent work by OEM with MM dampers to maximize
load capacity and to add damping in gas bearings
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
(gt10 um)
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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

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MMFB dimensions and specifications
PICTURE
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Static load test setup


Lathe chuck holds shaft bearing during
loading/unloading cycles.


Eddy Current sensor


Stationary shaft
Lathe tool holder
Test MMFB
Lathe tool holder moves forward and backward
push and pull forces on MMFB
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Static Load vs bearing displacement

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
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MMFB Research at TAMU
MMFB wire density 20

During Load reversal jump in structural
stiffness
Max. Stiffness 4 MN/m
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Dynamic load tests
Motion amplitude controlled mode
12.7, 25.4 38.1 µm
Frequency of excitation 25 400 Hz (25 Hz
interval)
Waterfall of displacement
MMFB motion amplitude (1X) is dominant
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Dynamic load vs excitation frequency
Dynamic load decreases around bearing natural
frequency, but increases with further increase in
excitation frequency. Dynamic load increases with
increasing motion amplitudes
Motion amplitude decreases
Around bearing natural frequency, less force
needed to maintain same motion amplitude
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Parameter identification model
1-DOF equivalent mechanical system
Equivalent Test System
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Parameter identification (no shaft rotation)
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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.
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Design equation MMB stiffness/damping
Empirical design equation for stiffness and
equivalent viscous damping coefficients
(Al-Khateeb Vance, 2002)
Functions of equivalent 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 (?)
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Real part of (F/X) vs excitation frequency
Frequency of excitation 25 400 Hz ( 25 Hz
step)
Motion amplitude increases
Real part of (F/X) decreases with increasing
motion amplitude
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MMFB structural stiffness vs frequency
Frequency of excitation 25 400 Hz (25 Hz
step)
At low frequencies (25-100 Hz), stiffness
decreases At higher frequencies, stiffness
gradually increases
Motion amplitude increases
MMFB stiffness is frequency and motion
amplitude dependent
Al-Khateeb Vance model reduction of stiffness
with force magnitude (amplitude dependent)
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Imaginary impedance (F/X) vs frequency
Frequency of excitation 25 400 Hz ( at 25 Hz
interval)
Motion amplitude increases
Im(F/X) decreases with motion amplitude
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Predictions vs. test data Viscous Damping
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 coefficients
in good agreement with measurements
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Loss factor vs excitation frequency
Frequency of excitation 25 400 Hz ( at 25 Hz
step)
Structural damping or loss factor is the
largest around the MMFB natural frequency
Loss factor nearly similar for all motion
amplitudes
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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
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Part II
Measurements of Drag Torque, Lift-off Speed and
Temperature in a Metal Mesh Foil Bearing
Thomas Abraham Chirathadam Research Assistant
Tae-Ho Kim Research Associate
Luis San Andrés Mast-Childs Professor Fellow ASME
STLE
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MMFB rotordynamic test rig
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 ,
Weight0.22 kg
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Test Rig Torque Lift-Off measurements
Thermocouple
Force gauge
String to pull bearing
Shaft (F 28 mm)
Static load
MMFB
Top foil fixed end
Torque arm
Positioning (movable) table
Preloading using a rubber band
Eddy current sensor
Calibrated spring
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Test procedure

• Sacrificial layer of MoS2 applied on top foil
  surface
• Mount MMFB on shaft of TC rig. Apply static
  horizontal load
• High Pressure cold air drives the ball bearing
  supported Turbo Charger. Oil cooled TC casing
• Air inlet gradually opened to raise the turbine
  shaft speed. Valve closing to decelerate rotor to
  rest
• Torque and shaft speed measured during the entire
  experiment. All experiments repeated thrice.

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Journal 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
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Varying steady state speed torque
Manual speed up to 65 krpm, steady state
operation, and deceleration to rest
61 krpm
50 krpm
37 krpm
24 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
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Startup torque vs applied static load
Top foil with worn MoS2 layer shows higher starup
torques
Worn MoS2 layer
Fresh coating of MoS2
Larger difference in startup torques at higher
static loads
Startup Torque Peak torque measured during
startup
Dry sliding operation
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DRY friction coeff. vs static load
Friction coefficient f (Torque/Radius)/(Static
load)
With increasing operation cycles, the MoS2 layer
wears away, increasing the contact or
dry-friction coefficient.
Worn MoS2 layer
Enduring coating on top foil required for
efficient MMFB operation!
Fresh MoS2 layer
Dry sliding operation
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Data derived from bearing torque and rotor speed
vs time data
Bearing drag torque vs rotor speed
Side load increases
WD 3.6 N
Steady state bearing drag torque increases with
static load and rotor speed
4.5
35.6 N (8 lb)
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Rotor not lifted off
26.7 N (6 lb)
3.5
3
17.8 N (4 lb)
2.5
Bearing torque N-mm
8.9 N (2 lb)
2
1.5
Increasing static load (Ws) to 35.6 N (8 lb)
1
Dead weight (WD 3.6 N)
0.5
0
20
30
40
50
60
70
80
Rotor speed krpm
airborne operation
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Friction coefficient vs rotor speed
Friction coefficient f (Torque/Radius)/(Static
load)
Friction coefficient f increases with rotor
speed almost linearly
Increasing static load (Ws) to 35.6 N (8 lb)
Dead weight (WD 3.6 N)
f decreases with increasing static load
airborne operation
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Bearing drag torque vs rotor speed
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
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Friction coefficient vs rotor speed
Friction coefficient ( f ) decreases with
increasing static load
f 0.01
f rapidly decreases initially, and then
gradually raises with increasing rotor speed
Rotor accelerates
Dry sliding
Airborne (hydrodynamic)
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Lift-Off speed vs applied static load
Side load increases
WD 3.6 N
Lift-Off Speed Rotor speed beyond which drag
torque is significantly small, compared to
Startup Torque
Lift-Off Speed increases linearly with static
load
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Top foil temperature (bearing outboard)
Room Temperature 21C
Top foil temperature measured at MMFB outboard end
Side load increases
Top Foil Temperature increases with Static Load
and Rotor Speed
Only small increase in temperature for the
range of applied loads and rotor speeds
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Conclusions

• Metal mesh foil bearing assembled using cheap,
  commercially available materials.
• Bearing break away torque, during start up,
  increases with applied static loads. A
  sacrificial coating of MoS2 reduces start up
  torque
• Bearing drag torque, while bearing is airborne,
  increases with static load and rotor speed
• Top foil steady state temperature increases
  with static load and rotor speed

Metal mesh foil bearing Promising candidate for
use in high speed oil-free rotorcraft applications
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Acknowledgments

• Thanks to
• TAMU Turbomachinery Research Consortium
• Honeywell Turbocharging Technologies

Learn more at http//phn.tamu.edu/TRIBGroup
Preguntas ?
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Tribology Group
Mission To advance state of the art in fluid
film bearings for high performance turbomachinery

• Types of funded research projects
• Engineering science fundamentals development of
  computational engineering tools,
• Experimental Bearing Seal parameter
  identification
• Test-product verification design validation
• Education

Luis San Andres
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Engineering Fundamentals Tools

2009 Fortran codes GUIs integrated into XLTRC2
tool Fast programs Physics based engineering
driven
Gas foil bearings for high temperatures NASA GRC
(Sept 07- August 09) 284 k Low cost license to
TRC members Summer 2009
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Industry Grants
HYBRID BRUSH SEALS TO IMPROVE GAS TURBINE
EFFICIENCYSiemens (06-June 09) 182,943
Hybrid Brush Seal
High temperature seal test rig (300 C572 F)
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Industry Grants - Siemens
Compare leakage four seal types
Flow factor F for four seal types versus
inlet/exhaust pressure ratio Ps/Pe for
increasing gas inlet temperatures (300 C).
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Industry Grants
SFD EXPERIMENTAL TESTING ANALYTICAL METHODS
DEVELOPMENT Large load (600 lbf) SFD test
rigPratt Whitney (08/08-12/10) 350k
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08-09 TRC funding
Metal Mesh Foil Gas Bearings for Oil-Free
Turbomachinery TRC (08-09) 40 k
Gas Bearings for Oil-Free Turbomachinery TRC
(08-09) 40 k
Foil Bearings for Oil-Free Turbomachinery TRC
(08-09) 40 k
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Virtual Tool for Turbocharger rotordynamics
Honeywell Turbocharging Technologies (2001-Jan
2009) 438 k
Predicted shaft motion
Measured shaft motion
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http//reumicro.tamu.edu
To conduct hands-on training and research in
mechanical, manufacturing, industrial, or
materials engineering topics related to
technological advances in microturbomachinery.
To develop microturbines to enhance defense,
homeland security, transportation, and aerospace
applications.
ASME Paper GT2009-59920
L. San Andres (MEEN) W. Hung (ENT)
13 students in Summer 2009
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Tribology Group Performance
Statistics 08/09 210k 120k (TRC)new
research funding 549.9 k research
expenditures (FY 08)
Projects on Gas Foil Bearings (NASA GRC, TRC),
Hybrid Brush Seals (Siemens), Turbochargers
(HTT), SFDs (Pratt Whitney)
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A pain in the neck!
Lubrication path
Wear !
Bionic 6c man
PEEK disks
After Oct 08
BEFORE Oct 08
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Off to Singapore