NASA/TM—2010-216094

Advancements Toward Oil-Free

Rotorcraft Propulsion

Samuel A. Howard and Robert J. Bruckner

Glenn Research Center, Cleveland, Ohio

Kevin C. Radil

U.S. Army Research Laboratory, Glenn Research Center, Cleveland, Ohio

March 2010
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NASA/TM—2010-216094

Advancements Toward Oil-Free

Rotorcraft Propulsion

Samuel A. Howard and Robert J. Bruckner

Glenn Research Center, Cleveland, Ohio

Kevin C. Radil

U.S. Army Research Laboratory, Glenn Research Center, Cleveland, Ohio

Prepared for the

65th Annual Forum and Technology Display (AHS Forum 65)

sponsored by the American Helicopter Society

Grapevine, Texas, May 27–29, 2009

National Aeronautics and

Space Administration

Glenn Research Center

Cleveland, Ohio 44135

March 2010
Acknowledgments

The authors would like to acknowledge the fi scal support of NASA’s Subsonic Rotary Wing Program,

part of the Aeronautics Research Mission Directorate.

This report contains preliminary fi ndings,

subject to revision as analysis proceeds.

Trade names and trademarks are used in this report for identifi cation

only. Their usage does not constitute an offi cial endorsement,

either expressed or implied, by the National Aeronautics and

Space Administration.

Level of Review: This material has been technically reviewed by technical management.

Available from

NASA Center for Aerospace Information National Technical Information Service

7115 Standard Drive 5285 Port Royal Road

Hanover, MD 21076–1320 Springfi eld, VA 22161

Available electronically at http://gltrs.grc.nasa.gov
Advancements Toward Oil-Free Rotorcraft Propulsion

Samuel A. Howard and Robert J. Bruckner

National Aeronautics and Space Administration

Glenn Research Center

Cleveland, Ohio 44135

Kevin C. Radil

U.S. Army Research Laboratory

Glenn Research Center

Cleveland, Ohio 44135

Abstract

NASA and the Army have been working for over a decade

to advance the state-of-the-art (SOA) in Oil-Free

Turbomachinery with an eye toward reduced emissions and

maintenance, and increased performance and efficiency

among other benefits. Oil-Free Turbomachinery is enabled by

oil-free gas foil bearing technology and relatively new hightemperature tribological coatings. Rotorcraft propulsion is a

likely candidate to apply oil-free bearing technology because

the engine size class matches current SOA for foil bearings

and because foil bearings offer the opportunity for higher

speeds and temperatures and lower weight, all critical issues

for rotorcraft engines. This paper describes an effort to

demonstrate gas foil journal bearing use in the hot section of a

full-scale helicopter engine core. A production engine hot-core

location is selected as the candidate foil bearing application.

Rotordynamic feasibility, bearing sizing, and load capability

are assessed. The results of the program will help guide future

analysis and design in this area by documenting the steps

required and the process utilized for successful application of

oil-free technology to a full-scale engine.

Introduction

Previously, NASA conducted a feasibility study to assess

the potential for an Oil-Free rotorcraft core (Ref. 1). The

analysis, based upon a rotordynamic study of a T700 engine

core, concluded that gas foil bearing technology was advanced

to the point that an Oil-Free core was technologically possible.

A General Electric T700 engine was used as the model

because it is common in the rotorcraft community, and it is a

reasonable size class for the current SOA of foil bearings.

Following from this analysis, an experimental effort has begun

to demonstrate a gas foil bearing at the hot-end of a T700 core

rotor. NASA and the Army have acquired used engine flight

hardware from a T700, seen in Figure 1, to conduct the

demonstration. Using a phased approach, the engine core will

eventually be rotated on gas foil bearings to demonstrate

successful full-scale Oil-Free engine operation.

NASA, in conjunction with industry, has developed a fourstep process for integration of gas foil bearings into

turbomachinery (Ref. 2). The four steps are: 1) assess the

rotordynamic feasibility of the application, 2) component

testing of candidate bearing designs, 3) rotordynamic system

testing with a simulated rotor and the selected bearing design

from step 2, and 4) full-scale demonstration after successful

rotordynamic simulation. In this paper, the first two steps are

described for the T700 hot-section application. A more

relevant rotordynamic feasibility study is done, and a foil

journal bearing is tested for load capacity and power loss. The

bearing size and design are based upon the conclusions of the

feasibility study, and analytical bearing design tools. The

results indicate that a T700 core rotor is a viable candidate for

a hot-section gas foil bearing, and a 76.2- by 63.5-mm bearing

is chosen as the initial size for simulation testing.

A follow-on activity, based upon successful completion of

this work, will focus on spinning the core rotor with the foil

journal bearing tested here installed in the hot-end of the rotor

(the location between the compressor and the turbine,

hereafter called the rear bearing). That work will represent

step 3 of the four-step process. This combined effort

compliments work being done by the engine manufacturer.

Bearing and Rotor Design Considerations

In general, Oil-Free Turbomachinery rotors are designed to

be rigid with respect to the operating speed range. Shafting is

often large in diameter and thin-walled in order to keep the

weight low and the stiffness high. Increased rotor diameter

also results in larger load capacities for foil bearings. This is in

direct opposition to typical rolling element bearing shaft

designs where rotor diameters at bearing locations are often

kept small to reduce the bearing DN (diameter times speed)

rating for longer useful life. Thus, in retrofitting a rotor

bearing system designed for rolling element bearings, one

must address this contradiction.

The T700 rotor of interest (compressor rotor) in this effort

is not extremely small in diameter because it is hollow with

the power turbine rotor running inside. However, it is still

fairly flexible, and in fact, has a bending critical speed below

the idle speed. Ideally, if this system were a complete

redesign, an attempt would be made to stiffen the rotor to

move the bending critical speed above the maximum operating

speed. However, the size and layout of the rotor system is

constrained to fit inside the existing engine envelope

NASA/TM—2010-216094
Figure 1.—Cut-away view of GE T700 engine.

n ,

s

Figure 2.—Rotordynamic model of GE T700 core rotor with ball

bearing in front, foil bearing in rear.

precluding a major redesign. Thus, successful Oil-Free

conversion of the hot-end of the T700 rotor depends upon the

ability to accelerate the rotor through the bending critical

speed. In order to traverse a critical speed, particularly a

critical speed with a bending mode shape, there must be

adequate system damping to limit the amplitude of vibration at

the critical locations of the rotor, such as the compressor and

turbine blade tips. Therefore, in this case, there must be

enough damping in the front rolling element bearing and the

rear foil bearing to sufficiently suppress the vibration levels at

the critical speed.

In the previous feasibility study (Ref. 1), it was determined

that foil bearings in both the front and rear bearing locations of

the core rotor would likely offer sufficient damping to

successfully navigate the original T700 operating envelope.

Two rotor designs were considered, one slightly modified

from the original geometry to accept foil bearings, but not

change the structural characteristics, and one modified to

accept foil bearings while stiffening the rotor to force the

bending critical speed above the maximum operating speed.

The stiffer rotor design results in a more robust design by

moving the bending critical speed above the maximum

operating speed. Also, by increasing the bearing span, the

rotor weight is better distributed, allowing a more heavily

loaded front bearing which results in increased system

stability. Unfortunately, the previous study was conducted

under the auspices of having the freedom to change the rotor

layout and use foil bearings in both the front and rear

locations. Therefore, it has little applicability to this effort in

which the front rolling element bearing is retained and the

geometry of the rotor is constrained to fit in the existing

architecture. In other words, the previous analysis assumed a

complete redesign of the engine, where the current effort is a

less drastic retrofit. Therefore, another rotordynamic analysis

was conducted with the specific bearing configuration planned

for the T700 retrofit to determine the feasibility of success.

Rotordynamics

The previous feasibility analysis indicated that a 76.2-mm

diameter by 50.8-mm length foil bearing would yield desirable

rotordynamic performance. That analysis was aimed at

determining the feasibility of a foil bearing supported engine

rotor, and sought to find a bearing size with potential for

success. The analysis does not suggest that the bearing size

found is the only possible size. Therefore, in choosing a

bearing size to analyze and test in this work, a bearing with a

76.2-mm diameter and 63.4-mm length is used because that

size had been manufactured before, and therefore was faster

and less costly to produce. The added length increases the load

carrying capability of the bearing, while potentially decreasing

the stability due to a lighter unit loading. As long as the

rotordynamics of the system with the larger bearing are

acceptable, the added length is viewed as a benefit in the

context of added load capacity.

The rotor model used previously is resurrected and used for

the current rotordynamic analysis. Figure 2 shows the model

with the bearing locations indicated by zig-zag lines and the

added mass for blades and other non-structural components

represented by circles.

A baseline model was run in the original analysis to verify

that the model accurately represents the actual engine. To that

end, the first three critical speeds of the rotor were calculated

and compared to the actual critical speeds of the engine as

measured by the manufacturer. Table 1 lists the calculated

critical speeds and the error with respect to the measured

critical speeds as reported in (Ref. 1). Since the errors are

quite small, the model is assumed to represent the engine well.

For the current effort, as stated earlier, there is little freedom

to modify the rotor significantly in the area where the foil

bearing will be located. Therefore, the foil bearing is modeled

in the same axial location, and it is assumed the geometry of

the rotor will not be changed enough to significantly alter the

structural characteristics of the rotor. The bearing properties

are changed to reflect those of the chosen foil bearing size.

Table 2 list the pertinent bearing parameters used in the

model. The bearing force coefficients are calculated using the

most recent version of the bearing analysis tool used in the

previous feasibility analysis (Refs. 3 and 4).

TABLE 1.—CRITICAL SPEED AND MASS COMPARISONS OF

BASELINE MODEL TO EMPIRICAL ENGINE DATA

Characteristic Baseline model

predictions

Error compared

to engine

1

st critical speed 6,260 rpm 0.6 %

2nd critical speed 12,700 rpm 3.5 %

3

rd critical speed 26,600 rpm 1.2 %

4

th critical speed 76,600 rpm not measured

Mass 21.9 kg –1.7 %

NASA/TM—2010-216094 2
TABLE 2.—BEARING DYNAMIC FORCE COEFFICIENTS

USED IN ROTORDYNAMIC MODEL FOR

76.2 mm BY 63.5 mm BEARING

Speed, Kxx(N/m) Kxy Kyx Kyy

rpm Cxx(Ns/m) Cxy Cyx Cyy

10,000 1.42e7 9.75e5 1.55e6 1.07e7

3.89e3 –2.30e3 1.75e3 4.38e3

30,000 1.39e7 –4.16e5 1.71e6 1.36e7

1.30e3 –4.10e2 6.22e2 1.22e3

50,000 1.48e7 –3.14e5 2.02e6 1.42e7

7.44e2 –1.46e2 2.37e2 5.60e2

Since the current analysis is based upon spinning the rotor

with a foil bearing at the rear, and a ball bearing at the front,

the front bearing (outboard of the compressor) remains the

same as in the baseline rotor configuration. Therefore, the

rotordynamic coefficients used in the baseline analysis are

used again here. The configuration analyzed and proposed

here is the most likely to be implemented first in a production

engine for at least two reasons. First, a majority of the benefits

of oil-free bearing technology can be realized by eliminating

the oil-lubricated bearing from the hot-section. Second,

utilizing a ball bearing in the front section eliminates the need

for an oil-free thrust bearing, the technology of which is not as

advanced as oil-free journal bearings. Thus, the configuration

offers a way to relatively quickly reap many of the oil-free

benefits, with less development effort and risk. It is a

configuration being considered by several engine

manufacturers in various applications.

The results of the rotordynamic analysis indicate that the

combined ball bearing/foil bearing configuration is feasible.

The critical speeds and mode shapes of the resulting system

are only slightly changed from the original engine. Table 3

lists the critical speeds of the baseline model compared to the

current model. One can see that the foil bearing at the rear has

very little effect on the natural frequencies of the rotor system.

The foil bearing has similar stiffness properties to the rolling

element bearing of the baseline configuration resulting in the

small changes in critical speeds. However, the stability of the

system is just as important as the frequencies of the critical

speeds.

TABLE 3.—CRITICAL SPEED COMPARISONS OF BASELINE

MODEL TO CURRENT MODEL WITH REAR FOIL BEARING

Characteristic Baseline model Current model

predictions predictions

1

st critical speed 6,260 rpm 7,300 rpm

2nd critical speed 12,700 rpm 12,900 rpm

3rd critical speed 26,600 rpm 27,200 rpm

4

th critical speed 76,600 rpm 77,000 rpm

A stability analysis indicates that all three modes below the

maximum speed of the rotor are stable up to and including

50,000 rpm which represents a roughly 10% over-speed

condition. Figure 3 is a plot of the log decrement of the lowest

three forward mode shapes as a function of speed. The system

1E AWMW561"S

iGn ti

-a& .

la: '25000 ;50000!

i. :ZLv 5 ;5peedl(r,prri,)

Figure 3.—Logarithmic decrement versus speed for

the first three mode shapes.

is considered stable as long as the log decrement is positive.

Based upon the critical speed analysis and the stability

analysis, a 76.2-mm diameter by 63.5-mm long foil bearing

appears to be a possible bearing size for the rear of the T700

core rotor.

Preliminary Bearing Tests

Given the promising results of the rotordynamic analysis, a

76.2- by 63.5-mm foil bearing was chosen for preliminary

testing. The following tests can be thought of as preliminary

screening to determine if the bearing size is appropriate for the

T700 application. The bearing is of the generation III type as

defined in DellaCorte and Valco (Ref. 5). A short description

of a generation III foil bearing is one in which the compliant

support structure is designed with variable stiffness in at least

two directions (i.e., axial and radial, axial and circumferential,

etc.). By comparison, generation I and II bearings have

structural properties that are constant everywhere and vary in

only one direction, respectively. In general, bearing

performance improves from generation I to II and from II to

III, such that generation III bearings are considered to be the

SOA in foil bearing technology.

The foil bearing tested was of the design described in

Heshmat (Ref. 6). Typical bearing qualification tests include

load capacity, and torque vs. speed vs. load (torque data is

used to generate a power loss map described in (Ref. 7)).

Unfortunately, with the bearing size of interest here, a load

capacity test was not possible because the maximum load that

can be applied by the test rig is approximately 222 N. The

anticipated load capacity of the 76.2- by 63.5-mm bearing,

using a previously developed empirical rule-of-thumb (Ref. 5),

is around 1,800 N at 18,000 rpm (the test speed). However, the

static load (weight of the rotor) the bearing will be required to

support in the T700 application, 170 N, is within the

capabilities of the test rig. Similarly, a complete power loss

map was not possible due to limited speed and load

capabilities of the existing test rig relative to a bearing of this

size.

NASA/TM—2010-216094
The test rig, shown in Figure 4, was designed to do

start/stop life testing of smaller bearings. As such, it has a

maximum speed of 21,000 rpm and maximum load of 222 N.

Therefore, only preliminary bearing testing is conducted to

determine how well the bearing performs subjected to the

steady state rotor weights of 170 and 145 N as described

below.

The test rig allows for two different methods of loading a

test bearing. In the first method, a pneumatic actuator is used

to pull downward on the test bearing by way of a cable and

yoke mechanism. Unfortunately, for the large loads of interest

here, the cable mechanism also constrains the bearing in the

rotational direction, making it impossible to accurately

measure torque in this configuration. In the second method of

applying load to the bearing, concentric rings of a very dense

material (Anvaloy) are mounted on the outside diameter of the

bearing, as seen in Figure 5. The rings are balanced in an

attempt to minimize any rotational force due to nonconcentricity that would affect the torque measurement. The

advantage of the first loading method is the ability to load the

bearing between zero and the upper limits of the test rig,

approximately 222 N. The obvious disadvantage is the lack of

capability to measure torque. The second method allows for

measurement of torque, but limits the applied load capability

to several discrete loads up to a maximum of 145 N. From

here forward, the concentric ring loading mechanism is

referred to as a dead-weight loading since the load is passive.

In order to test the capability of the bearing to support a

steady-state load of 170 N, the cable loading mechanism was

used. To gauge the power loss, though, a second test was

conducted using the dead-weight loads. Typically, a power

loss map is generated to determine if a bearing is operating in

a light, moderate, or heavy loading regime, as described in

(Ref. 7). However, since this was a preliminary screening test,

no attempt was made to generate a full power loss map, rather

the power loss with the maximum possible dead load, 145 N,

was measured. The value of this test was simply to determine

if the power loss at a load near that of the expected load is

consistent with a healthy load condition. If the power loss

were excessive, it would suggest the bearing size may not be

appropriate for the application.

Before any testing was conducted, the bearing was brokenin, or conditioned, by completing 5,800 start-up and shutdown cycles at 538 °C. This conditioning routine is typical to

obtain a conforming surface between the journal and the

bearing, and to improve the surface finish for better

performance (Ref. 8). A photograph of the conditioned

bearing shows normal regions of wear from the break-in

process in Figure 6.

The bearing had three thermocouples (the leads can also be

seen in Fig. 6) installed at the fixed end of the top foil. The

thermocouples were attached to the spacer block in the gap

between the fixed and free ends of the foil as shown in

Figure 7. They were located outboard, in the middle, and

inboard of the bearing in the axial direction relative to the

motor side of the journal. The thermocouple data was used in

conjunction with the torque data to monitor the health of the

bearing, and to compare to typical foil bearing behavior for

screening purposes during the test with the dead-weight load.

Figure 4.—Photograph of the test rig used for GE T700 bearing

tests.

Figure 5.—Photograph of the test bearing with concentric dead

weight loading rings and torque arm.

NASA/TM—2010-216094 4
Figure 6.—Photograph of the test bearing after

conditioning showing normal wear.

C605-82 4

Figure 7.—Schematic showing the thermocouple

location inside the test bearing.

Preliminary Bearing Test Results

As described above, the first test was to apply a 170 N load

via a cable and pneumatic loader. With 170 N, the bearing ran

for about an hour with no problems observed. The purpose of

this test was to see if the bearing would lift-off under an

applied load consistent with the T700 application. The

determination of lift-off is based mostly on experience, but the

bearing was clearly floating and was free to drift back and

forth on the journal. This first test served to verify that the

bearing size was likely sufficient.

Next, the concentric load rings were installed on the

bearing, and the more quantitative, yet still preliminary, deadweight test was conducted. Figure 5 shows the bearing with

the three concentric dead-weight loads mounted. One can also

see the torque arm that attaches by cable to a load cell for

measuring the running torque. During the dead weight test,

three objectives were met. First, the temperature was

measured from start-up until a steady state thermal condition

was reached. Second, running torque with a 145 N deadweight load (as close to 170 N as possible using existing deadweights) was measured for a power loss assessment. Lastly,

speed versus time data for a coast-down event was recorded to

observe the speed at which the bearing started to rub (touchdown or lift-off speed).

Figure 8 shows the temperature recorded by the three

thermocouples as the bearing reached steady-state at about

45 min. into the test. Two observations are immediately

obvious from the figure: the steady state-operating

temperature at 18,000 rpm is around 60 to 70 °C depending on

axial location. Interestingly, the outboard temperature was

very similar to the middle temperature, while the inboard

temperature was about 10 °C lower during most of the test.

This was most likely due to the positioning of the

thermocouples, wherein the outboard thermocouple was

located about 6 mm in from the edge of the foil, while the

inboard was located very near the edge of the foil. The reason

for the discrepancy was simply a matter of installation

logistics. The thermocouples enter the gap from the outer

diameter through holes, and then are bent to align with the

spacer block. In this case, the technician bent all the

thermocouples the same direction such that the outboard

thermocouple was bent toward the middle of the bearing upon

exiting the hole, and the inboard thermocouple was bent away

from the middle as it exited the hole. This cause may not be

the sole cause, as there could be other contributing factors.

One such factor is the relative looseness of the inboard and

outboard edges. If the bearing were tighter, due to

manufacturing variation, on the outboard edge, the above

observation could result. The variation in temperature is not a

reason for concern, and the main conclusion one can make

from this data is that the bearing performed well, based upon

temperature, for a load of 145 N at 18,000 rpm. Temperatures

of 60 to 70 °C are relatively low for an uncooled bearing

(Ref. 9), indicating that 145 N is likely a light-load condition

for the bearing at this speed.

NASA/TM—2010-216094
Time (Minutes)'

Figure 8.—Plot of temperature at the fixed end of the foil

Figure 9.—Typical power loss gas foil bearing performance

map.

A. 1115 0

Time (seconds )

Figure 10.—Coast-down speed plot.

It has been observed that one can use power loss as a

measure of performance in gas foil journal bearings (Ref. 7).

Power loss is calculated from the torque measurement, and

represents an indication of how much heat is generated in the

bearing. There is a significant speed effect and to a lesser

extent a load effect on power loss. At very low speed, the

power loss is high due either to extreme sheer rates of the gas

film or rubbing if the speed is below the lift-off speed. As

speed increases, power loss drops to a minimum, then begins

to rise again. Figure 9 shows a sample power loss map (not for

this bearing). For various reasons, the desirable operating

regime is at speed and load combinations that keep the bearing

on the high side (with respect to speed) of the minimum power

loss. A full power loss map was not created in this work

because of the limited test rig capabilities. However, a torque

measurement of 0.094 N-m at 18,000 rpm and a dead-weight

load of 145 N was measured when the temperature reached

steady-state during the test. Torque of 0.094 N-m corresponds

to a power loss of 37 kW/m2

, a relatively low number based

upon experimental results where typical power loss levels at

failure are approximately 155 kW/m2

. While a full power loss

map would need to be constructed to fully characterize the

bearing, this one data point suggests that at 18,000 rpm under

a load of 145 N, the bearing tested is near the minimum power

loss level. An increase in speed, up to the T700 full speed of

roughly 45,000 rpm, would likely put the power loss in the

desired zone above the minimum power loss level.

The final screening test to characterize the bearing was a

coast-down test. The bearing was run at 18,000 rpm and

allowed to warm-up to a steady-state temperature condition as

in the first test. The motor power was shut-off, and the speed

was recorded as the rotor coasted to a stop. Plotting speed

versus time, as in Figure 10, one can see the approximate

speed at which the bearing begins to contact the rotor, i.e., the

touch-down, or lift-off speed (called the lift-off speed from

here on). Looking at Figure 10, the speed of the rotor

decreases with positive curvature from 18,000 rpm to

approximately 8,000 rpm (the curvature of the speed curve is

the third derivative of position with respect to time, called

jerk). After 8,000 rpm, the sign of jerk changes from positive

to negative, and the speed trace curves downward. It is

believed the transition from positive to negative jerk marks the

transition from a fully hydrodynamic fluid film to a boundary

type lubrication regime. That transition also translates roughly

to the speed on the performance map where the power loss

minimum occurs. In this case, that speed is approximately

10,000 rpm, as seen in Figure 9. At some speed, the fluid film

completely breaks down, and the slope of the curve becomes

nearly vertical, indicating full sliding contact, which can be

seen around 2,000 rpm on the coast-down plot. Typically, the

minimum power loss is 2 to 6 times the lift-off speed (Refs. 7

and 10), which is consistent with the coast-down plot here

where lift-off speed is near 2,000, and 10,000 rpm is near the

transition to a full film.

All three of the test results point to the same conclusion,

that a 145 to 170 N load on a 76.2- by 63.4-mm foil journal

NASA/TM—2010-216094 6
bearing is a reasonable load at 18,000 rpm. The T700

application has a speed range requirement of about 20,000 to

45,000 rpm. Though the three screening tests were limited in

scope, due to test rig constraints, they all indicate that the

bearing size is viable for further consideration in the

application.

Conclusions

The first two steps of a four-step design process have been

completed for assessing the performance of a gas foil journal

bearing in the hot-section of a rotorcraft propulsion system.

The rotordynamic feasibility study concluded that a bearing

with a 76.2-mm diameter and 63.5-mm length would be a

proper size for the applied load of 170 N, and would offer

sufficient rotordynamic performance based upon predicted

bearing force coefficients.

The series of tests performed here indicate: 1) The bearing

could lift off with a 170 N load and maintain operation for a

prolonged period of time, 2) The bearing running with a 145 N

dead-weight load reached steady state operating temperatures

of 60 to 70 °C, which were relatively low, indicating that the

bearing was not heavily stressed, 3) The power loss of

37 kW/m2 at 18,000 rpm with a 145 N load was low compared

to typical values observed at failure of 155 kW/m2, and 4) The

lift-off speed was found to be around 2,000 rpm and the

transition to full film occurs at roughly 5 times that speed at

10,000 rpm, thus, the T700 speed range of 20,000 to

45,000 rpm will be slightly above where the transition to a full

film occurs.

In summary, all indications are that the tested bearing is

expected to perform well in the hot-section of a T700

rotorcraft propulsion engine. With a viable bearing design, the

future plan is to move ahead with the third step of the design

process, rotordynamic simulator testing.

References

1. Howard, S.A., Bruckner, R.J., DellaCorte, C., and Radil, K.C.,

“Preliminary Analysis for an Optimized Oil-Free Rotorcraft

Engine Concept,” NASA/TM—2008-215064, 2008.

2. Valco, M.J., and DellaCorte, C., “Emerging Oil-Free

Turbomachinery for Military Propulsion and Power

Applications,” Proceedings of the Army Sciences Conference,

Ft. Lauderdale, FL, 2003.

3. Carpino, M., and Talmadge, G., “A Fully Coupled Finite

Element Formulation for Elastically Supported Foil Journal

Bearings,” Tribology Transactions, Vol. 46, pp. 560–565, 2003.

4. Carpino, M., and Talmadge, G., “Prediction of Rotordynamic

Coefficients in Gas Lubricated Foil Journal Bearings with

Corrugated Sub-Foils,” Tribology Transactions, Vol. 49,

pp. 400–409, 2006.

5. DellaCorte, C., and Valco, M.J., “Load Capacity Estimation of

Foil Air Journal Bearings for Oil-Free Turbomachinery

Applications,” Tribology Transactions, Vol.43, No. 4, pp.795–

801, 2000.

6. Heshmat, H., “High Load Capacity Compliant Foil

Hydrodynamic Journal Bearing,” U.S. Patent No. 5,988,885,

Nov. 1999.

7. DellaCorte, C., Radil, K.J., Bruckner, R.J., and Howard, S.A.,

“A Preliminary Foil Gas Bearing Performance Map,”

NASA/TM—2006-214343, 2006.

8. Radil, K.J., and DellaCorte, C., “The Effect of Journal

Roughness and Foil Coatings on the Performance of Heavily

Loaded Foil Air Bearings,” NASA/TM—2001-210941, 2001.

9. Radil, K.J., DellaCorte, C., and Zeszotek, M., “Thermal

Management Techniques for Oil-Free Turbomachinery

Systems,” NASA/TM—2006-214358, 2006.

10. DellaCorte, C., Radil, K.J., Bruckner, R.J., and Howard, S.A.,

“Design, Fabrication and Performance of Open Source

Generation I and II Compliant Hydrodynamic Gas Foil

Bearings,” NASA/TM—2007-214691, 2007.

NASA/TM—2010-216094 7
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1. REPORT DATE (DD-MM-YYYY) 2. REPORT TYPE 3. DATES COVERED (From - To)

01-03-2010 Technical Memorandum

4. TITLE AND SUBTITLE 5a. CONTRACT NUMBER

Advancements Toward Oil-Free Rotorcraft Propulsion

5b. GRANT NUMBER

5c. PROGRAM ELEMENT NUMBER

6. AUTHOR(S) 5d. PROJECT NUMBER

Howard, Samuel, A.; Bruckner, Robert, J.; Radil, Kevin, C.

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WBS 877868.02.07.03.01.01.04

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John H. Glenn Research Center at Lewis Field E-17147

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13. SUPPLEMENTARY NOTES

14. ABSTRACT

NASA and the Army have been working for over a decade to advance the state-of-the-art (SOA) in Oil-Free Turbomachinery with an eye

toward reduced emissions and maintenance, and increased performance and efficiency among other benefits. Oil-Free Turbomachinery is

enabled by oil-free gas foil bearing technology and relatively new high-temperature tribological coatings. Rotorcraft propulsion is a likely

candidate to apply oil-free bearing technology because the engine size class matches current SOA for foil bearings and because foil bearings

offer the opportunity for higher speeds and temperatures and lower weight, all critical issues for rotorcraft engines. This paper describes an

effort to demonstrate gas foil journal bearing use in the hot section of a full-scale helicopter engine core. A production engine hot-core

location is selected as the candidate foil bearing application. Rotordynamic feasibility, bearing sizing, and load capability are assessed. The

results of the program will help guide future analysis and design in this area by documenting the steps required and the process utilized for

successful application of oil-free technology to a full-scale engine.

15. SUBJECT TERMS

Gas bearings; Foil bearings; Rotor dynamics

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