Luis San AndrésMast-Childs Professor

Fellow ASME

ASME GT2011-45257ASME 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 http://rotorlab.tamu.edu

Metal Mesh Foil BearingEffect of Motion Amplitude, Rotor Speed, Static Load, and Excitation Frequency on

Force Coefficients

Oil-Free Bearings for Turbomachinery

JustificationCurrent 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

 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

Gas Foil Bearings

Used in many oil-free rotating machinery: high load capacity (>20 psig), rotordynamically stable, tolerance of misalignment and shocks…….

: Bump spot weld X

Rotor spinning

Housing

Bump strip layer

Top foil

Gas film

Y

Top foil spot weld

Ω

Ө

…… but expensive with intellectual property restrictions. A low cost proven alternative needed.

Metal Mesh Foil Bearing (MMFB)

MMFB COMPONENTS: Bearing cartridge, metal mesh ring and top FoilHydrodynamic 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 bearing5 cm

Metal Mesh Ring

Top Foil Coated with MoS2

Bearing Cartridge

Rotor spinning

Slot

MMFB components

Top Foil

0.12 mm top foilChrome-Nickel alloyRockwell 40/45

Heat treated at ~ 450 ºC for 4 hours and allowed to cool. Foil retains arc shape after heat treatment

Sprayed with MoS2

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

Zarzour and Vance (2000) J. Eng. Gas Turb. & Power, Vol. 122Advantages of Metal Mesh Dampers over SFDsCapable of operating at low and high temperaturesNo changes in performance if soaked in oil

Al-Khateeb and Vance (2001) ASME GT-2001-0247Test metal mesh donut and squirrel cage( in parallel)Metal Mesh damping not affected by modifying squirrel cage stiffness

Choudhry and Vance (2005) ASME GT-2005-68641Develop 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

Ertas &Luo (2008) ASME J. Gas Turbines Power, 130MM damper force coefficients not affected by shaft eccentricity (or applied static load)

Ertas (2009) ASME J. Gas Turbines Power, 131Two 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 2009-2521Shape 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

Past work in MMFBsSan 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 GT2009-59920Demonstrated operation to 45 krpm with early rotor lift off. Educated undergraduate students.

San Andrés et al. (2010) J. Eng. Gas Turb. & Power, 132Start 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, 133Rotordynamic coefficients from unidirectional impact loads. Estimated stiffness and damping force coefficients at 50 krpm.

EXPERIMENTS with a PRIOR MMFB (larger mesh thickness)

1. Structural stiffness and damping

2. Friction factor with airborne operation

0

0.5

1

1.5

2

2.5

0 100 200 300 400Frequency [Hz]

Str

uctu

ral s

tiffn

ess

[MN

/m]

12.7 um25.4 um38.1 um12.7 um Prediction25.4 um Prediction38.1 um Prediction

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 dependent12.7 m

25.4 m38.1 m

Motion amplitude increases

San Andres et al., 2010, ASME J. Eng. Gas Turbines Power, 132 (3)

10

100

1000

10000

100000

0 100 200 300 400Frequency [Hz]

Equi

vale

nt v

isco

us d

ampi

ng [N

s/m

]

12.7 um25.4 um38.1 um12.7 um Prediction25.4 um Prediction38.1 um Prediction

MMFB eq. damping vs. frequency

Amplitude increases

12.7 μm25.4 μm38.1 μm

MMFB equiv. viscous damping

decreases as the excitation

frequency increases and

as motion amplitude increases

Al-Khateeb & Vance modelSan Andres et al., 2010, ASME J. Eng. Gas Turbines Power, 132 (3)

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 speedDry sliding Airborne (hydrodynamic)

Dead weight

(WD= 3.6 N)

Increasing static load (Ws) to 35.6 N (8 lb)

f = (Torque/Radius)/(Net static load)

CURRENT MMFB & ROTORDYNAMIC TEST RIG

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, t 2.667 mm

mesh inner diameter 36.74 mm

Copper mesh density 20 %

Wire diameter 0.30 mm

Steel top foil thickness, tF 0.12 mm

Bearing diametral clearance

ID-2(t+tF)

~ 0.0 mm

5 cm

Top foil: Chrome nickel alloy

Metal mesh: copper

Bearing Cartridge:

Stainless steel

2.7mm

MMFB rotordynamic test rig

Max. operating speed: 75 krpmTurbocharger driven rotorRegulated air supply: 9.30bar (120 psig)

Test Journal: length 55 mm, 36.5 mm diameter

Journal press fitted on Shaft Stub

TC cross-sectional viewRef. Honeywell drawing # 448655

Twin ball bearing turbocharger, Model T25, donated by Honeywell Turbo Technologies

Bearing

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

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 temperature5 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

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

R2 = 0.95

KSY = 21.2 kN/m

MSY = 0.88 kg

CSY = 6.6 Ns/m

Curve fit for the transfer Function

0 25 50 75 100 125 1500

5

10

15

20

25

ExperimentalCurve Fit

Frequency [Hz]

Acc

eler

ance

[(m

/s^2

) /N

]

2

1/ 22 22Y Y Y

Y

YS S S

a

FK M C

Accelerance function = physical

model equation

Estimated test system mass =0.88 kg

Y

Impact load

Bearing overhang on squirrel cage

Bearing

Sq. cage

Identification model

KS,CS: soft SQ stiffness

and damping

MS : effective mass

X

YKYY, CYY

KXY, CXY

Shaker force, FY

Bearing

Journal

KYX, CYX

KXX, CXX

Ω

X X X

Y Y Y

S X S X S XX XY XX XY X

YX YY YX YY YS Y S Y S

M a C v K X C C K K Fx x

C C K K FM a C v K Y y y

EOM:

Shaker force, FX

KSX, CSX

KSY, CSY

SoftSupport structure

Kij ,Cij: test bearing stiffness

and damping

Identification model

Forces: Sine sweep excitations (200-400 Hz), amplitude controlled

Responses: measure bearing accelerations and displacements relative to journal

Kij ,Cij : bearing stiffness and damping vs.

frequencyX

YKYY, CYY

KXY, CXY

Shaker force, FY

Bearing

Journal

KYX, CYX

KXX, CXX

Ω

Shaker force, FX

KSX, CSX

KSY, CSY

SoftSupport structure

X X

X

Y Y

Y

S S2SX XXX XX XY XY

YX YX YY YY S SY Y2S

C KMF AxK j C K j C j

yK j C K j C C KF AM j

( ) ( )

( ) ( )

( )

( )

Process data in frequency domain to obtain:

Y

W

X

Dynamic load and displacements

Static load resolved along X and Y = 15.5 N

Shaft speed=50 krpm (833 Hz)

15.5 N 15.5 N

Fixed end Force along X

Displacement along X

Displacement along Y

Dynamic load 25-100 NExcitation 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

Forces & disps. vs. frequency

Static loads along X and Y =15.5 N

Shaft speed=50 krpm (833 Hz)

15.5 N 15.5 N

Fixed end

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

0 100 200 300 400 5000

5

10

15

20

XXXYYXYY

Frequency [Hz]

Fo

rce

[N]

Fo

rce

[N]

Frequency [Hz]

0 100 200 300 400 5000

2

4

6

XXXYYXYY

Frequency [Hz]

Disp

lacem

ents

[um

]

Dis

pla

cem

en

t [μ

m]

Frequency [Hz]

0 100 200 300 400 5000

2

4

6

XXXYYXYY

Frequency [Hz]

Dis

pla

cem

ents

[u

m]

0 100 200 300 400 5000

2

4

6

XXXYYXYY

Frequency [Hz]

Dis

plac

emen

ts [u

m]

200-400 Hz Force FYY

FXX

FXY

FYX

Displacement YY

XxXY

YX

N

m

X Y

-1

-0.5

0

0.5

1

200 250 300 350 400Frequency [Hz]

Sti

ffn

es

s [

MN

/m]

-1

-0.5

0

0.5

1

200 250 300 350 400Frequency [Hz]

Sti

ffn

es

s [

MN

/m]

-1

-0.5

0

0.5

1

200 250 300 350 400Frequency [Hz]

Sti

ffn

es

s [

MN

/m]

-1

-0.5

0

0.5

1

200 250 300 350 400Frequency [Hz]

Sti

ffn

es

s [

MN

/m]

K KK K

XX XYYYYX

MMFB stiffnesses: varying rotor speeds

22 N static load

50 krpm

40 krpm0 krpm

45 krpm

KxxKYY

KYX

KXY

KXXKYY

KYX

KXY

KXXKYY

KYX

KXY

Kxx

KYX

KXY

KYY

At rest (0 rpm) direct stiffness is

structural only. Direct K

decreases with rotor speed.

With rotation, KYX changes sign.

Small cross-stiffneses.

Direct stiffnesses gradually

increase with frequency

-500

-250

0

250

500

200 250 300 350 400

Frequency [Hz]

-500

-250

0

250

500

200 250 300 350 400

Frequency [Hz]

Eq

. v

isc

ou

s d

am

pin

g [

Ns

/m]

-500

-250

0

250

500

200 250 300 350 400

Frequency [Hz]E

q.

vis

co

us

da

mp

ing

[N

s/m

]

-500

-250

0

250

500

200 250 300 350 400

Frequency [Hz]

Eq

. v

isc

ou

s d

am

pin

g [

Ns

/m]

C C

C CXX XY

YX YY

MMFB damping: varying rotor speeds

40 krpm0 krpm

50 krpm45 krpm

CYX

CYY CXX

CXY

CXXCYY

CXY

CYX

CYY

CXYCYX

CXX CXXCYY

CYX CXY

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.

-1

-0.5

0

0.5

1

200 250 300 350 400Frequency [Hz]

Sti

ffn

ess

[MN

/m]

-1

-0.5

0

0.5

1

200 250 300 350 400Frequency [Hz]

Sti

ffn

ess

[MN

/m]

K KK K

XX XY

YX YY

-1

-0.5

0

0.5

1

200 250 300 350 400Frequency [Hz]

Sti

ffn

es

s [

MN

/m]

MMFB stiffnesses: varying motion amplitudes

20 m 25 m

30 m

KXX

KYX

KXY

KYY KXX

KYX

KXY

KYY

KXX

KYX

KXY

KYY 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)

-500

-250

0

250

500

200 250 300 350 400

Frequency [Hz]

Eq

. vis

cou

s d

amp

ing

[N

s/m

]

-500

-250

0

250

500

200 250 300 350 400

Frequency [Hz]

-500

-250

0

250

500

200 250 300 350 400

Eq

. vis

cou

s d

amp

ing

[N

s/m

]

XYYX YY

-500

-250

0

250

500

200 250 300 350 400

Frequency [Hz]

Eq

. vis

cou

s d

amp

ing

[N

s/m

]

C CC C

XX XYYX YY

MMFB damping: varying motion amplitudes

30 m

25 m20 m

CYX

CXXCYY

CXY

CXX

CYY

CYX CXY

CXY

CYYCXX

CYX

Direct damping decreases slightly with increasing motion amplitude.

Direct C increases with frequency.

With increasing motion amplitude, cross-damping CYX decreases

22 N Static

load

50 krpm Rotor speed

(833 Hz)

MMFB K & C: varying applied static load

22 N

36 N

36 N

22 N

For increasing static load: force

coefficients are similar in

magnitude and show same trend

in frequency.

KXX

KYX

KXY

KYY

KXX

KYX

KXY

KYY

CXY

CYYCXX

CYX CXY

CYYCXX

CYX

50 krpm Rotor speed

(833 Hz)

Net static load = 22 N ( W/LD=0.16

bar) & 36 N ( W/LD=0.26 bar)

W

-1

-0.5

0

0.5

1

200 250 300 350 400Frequency [Hz]

Sti

ffn

es

s [

MN

/m]

-1

-0.5

0

0.5

1

200 250 300 350 400

K KK K

XX XYYX YX -1

-0.5

0

0.5

1

200 250 300 350 400Frequency [Hz]

Sti

ffn

ess

[MN

/m]

-1

-0.5

0

0.5

1

200 250 300 350 400

K KK K

XX XYYX YX

-500

-250

0

250

500

200 250 300 350 400

Frequency [Hz]

-500

-250

0

250

500

200 250 300 350 400

Eq

. vis

cou

s d

amp

ing

[N

s/m

]

XYYX YY

-1

-0.5

0

0.5

1

200 250 300 350 400

C CC C

XX XYYX YX

-1

-0.5

0

0.5

1

200 250 300 350 400Frequency [Hz]

Sti

ffn

ess

[MN

/m]

-1

-0.5

0

0.5

1

200 250 300 350 400

C CC C

XX XYYX YX

MMFB: estimation of loss factor

Proportional structural damping C K

T T

t T t T

V M

t t

E dt E dt

z Cz z K zEnergy (material damping) = Energy (viscous damping)

2 2

2 2

XX X YY Y

XX X YY Y

C V C V

K V K V

For elliptical orbits:

For circular orbits, XX YY

XX YY

C C

K K

Energy dissipation in MMFB largely due to mechanical hysteresis. A

loss factor (γ) best represents the material damping

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 FBsTypical BFB loss factor ~ 0.1-

0.4Kim et al. (2008)

Waterfalls of force and displacement

Dominant displacement amplitude corresponds to excitation frequency

No sub-synchronous whirl detected.

020406080

100120140160

0 200 400 600 800 1000

Frequency[Hz]

Fo

rce

[N

]

0

10

20

30

40

50

60

0 200 400 600 800 1000

Frequency [Hz]

Dis

pla

ce

me

nt

[um

]

400 Hz

200 Hz

200 Hz

400 HzSynchronous ( 833 Hz)

15.5 N 15.5 N

Fixed end 22 N static

load

50 krpm Rotor speed

(833 Hz)

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

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/in3

damping coeff. C~ 0.1-1.0 (L x D) lbf-s/in3

MMFB stiffness coeff. K~ 1,330 (Lx D) lbf/in3 [360 MN/m3] damping coeff. C~ 0.93 (L x D) lbf-s/in3 [252 MN/m3]

Test MMFB is structurally soft with large damping: 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)

Acknowledgments/ Thanks to

http://rotorlab.tamu.eduLearn more at:

Honeywell Turbocharging Technologies

Turbomachinery Research Consortium

Questions (?)

Extra slides - >

Rotor acceleratesRotor accelerates

Comparison: MMFB&BFB Friction factor vs rotor speed

f = (Torque/Radius)/(Static load)

f ~ 0.03

f ~ 0.03

Friction coefficient decreases with increasing applied static loads and rotor speed (due to lift-off)

MMFB BFB

Static load

0.01

0.1

1

0 10 20 30 40 50 60 70

Rotor speed [krpm]

Fri

ctio

n f

acto

r [-

]

35.6N

26.7N

17.8N

0.01

0.1

1

0 10 20 30 40 50 60 70

Rotor speed [krpm]

Fri

cti

on

facto

r [-

]

35.6N

26.7N

17.8N

Future work: MMFB force coefficient prediction

Θ

Θl

Θt

Θp

X

Y

eY

eX

h

r

r

p

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

Km 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

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

krpmc) Compare thermal performance of MMFBs with Gen. I bump-foil bearings

Future work: High temperature operation

Metal Mesh Foil Bearing