A Three-Dimensional Foil Bearing Performance Map 

Applied to Oil-Free Turbomachinery 

by Kevin C. Radil and Christopher DellaCorte 

ARL-TR-4473 April 2009 

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Army Research Laboratory 

NASA-Glenn Research Center, Cleveland, OH 44135-3191 

ARL-TR-4473 April 2009 

A Three-Dimensional Foil Bearing Performance Map 

Applied to Oil-Free Turbomachinery 

Kevin C. Radil 

Vehicle Technology Directorate, NASA-Glenn Research Center, ARL 

and 

Christopher DellaCorte

NASA-Glenn Research Center, Cleveland, OH

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A Three-Dimensional Foil Bearing Performance Map Applied to Oil-Free 

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Kevin C. Radil (ARL) and Christopher DellaCorte (NASA-Glenn) 

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14. ABSTRACT 

To effectively apply compliant foil gas bearings to increasingly larger and more challenging turbomachinery, a comprehensive 

method that compares a foil bearing’s capabilities with the application’s operating requirements is needed. Extensive 

laboratory and field experience suggests foil bearing failure is generally due to thermal stress brought on by excessive viscous 

power loss; therefore a map that graphically relates component and system-level parameters (bearing size, applied loads, and 

shaft rotational speeds) directly to bearing power loss is more elucidating than a map based on a lumped speed/load parameter 

like the Sommerfeld number. In this report we describe a performance map featuring a three-dimensional contour plot that 

illustrates the expected power loss in a foil bearing as a function of applied load and shaft speed. Using this performance map, 

bearing capabilities can be examined at the anticipated system operating conditions and safety margins between an operating 

point and incipient bearing failure can be ascertained. To demonstrate the concept’s features and usefulness, we present a 

performance map generated from foil bearing power loss test data. We expect these maps, combined with other predictive 

tools, will help evaluate a foil bearing’s general suitability for a candidate rotor system and will lead to more robust and 

successful Oil-Free turbomachinery designs. 

15. SUBJECT TERMS

Foil bearing, gas bearing, turbomachinery 

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Kevin Radil 

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iii 

Contents 

List of Figures iv

List of Tables iv

1. Introduction 1

2. Methods, Assumptions, and Procedures 4

2.1 Foil Bearing and Journal .................................................................................................4 

2.2 Results and Discussion....................................................................................................7

3. Conclusions 12

4. References 14

Distribution List 16 
iv 

List of Figures 

Figure 1. Cross-sectional view of a radial foil air bearing..............................................................2 

Figure 2. Foil bearing performance map (12) illustrating its general shape and the two 

operating regions........................................................................................................................3 

Figure 3. Foil bearing high-speed test rig with a stack of three Anviloy rings in place.................5 

Figure 4. Exploded view of Anviloy rings capable of producing bearing loads from 

approximately 44 to 176 N. (Steel housing for 35 mm bearing not shown)..............................6 

Figure 5. A 3-D performance map for the 35 mm diameter foil bearing generated from three 

load curves. ................................................................................................................................7 

Figure 6. A 2-D performance map for the 35 mm diameter foil bearing using the three load 

curves. ........................................................................................................................................8 

Figure 7. Representative performance map showing hypothetical operating points. Operating 

points A and C are in the low load region and point B is in the high load region.....................9 

Figure 8. Drawing of the 51-mm diameter, 51-mm long foil bearing. The type K 

thermocouple is located in the middle of the bearing and on the spacer block near the 

leading edge. ............................................................................................................................10 

Figure 9. A 2-D performance map for the 51-mm diameter, 51-mm long foil bearing................11 

List of Tables 

Table 1. Test speed increments for data collection.........................................................................7 

Table 2. Sample power loss data including air film temperature at the midpoint near the 

leading edge for the 51-mm diameter bearing. ........................................................................11 
1 

1. Introduction 

Foil bearings are self-acting, hydrodynamic gas bearings that use air as their working fluid. 

Their use in rotating systems eliminates the weight, maintenance requirements, speed, and 

temperature limitations associated with conventional oil-lubricated rotor supports (i.e., bearings, 

dampers, seals and oil-system) (1–3). Gas foil bearing technology has been deployed in many 

high-speed rotating shaft systems such as air cycle machines, turboexpanders, and compressors 

and in small micro-turbine systems (4–6). Additionally, foil bearings have been demonstrated as 

“proof-of-concepts” in diesel engine turbochargers, auxiliary power units (APUs), and selected 

hot section bearings in gas turbines (7–9). While these Oil-Free turbomachines represent the 

successful final outcomes of foil bearing integration efforts, most were achieved through arduous 

“make and break” hardware-intensive development programs. This inefficient approach is 

often employed because, despite foil bearings’ long use, little effort has been undertaken to 

develop a systematic approach to streamline their integration into new, high-speed 

turbomachinery systems. 

Figure 1 shows a cross section of a radial foil air bearing. This type of fluid-film gas bearing 

employs a series of thin sheet metal foils, from which its name is derived. The foils are typically 

made from a nickel-based superalloy, because this material retains a majority of its elastic spring 

properties at high temperatures. Bearing construction is based on three main components: the 

top foil, the bump foil, and the shaft. The top foil acts as a compliant surface that traps and 

supports the hydrodynamic gas film against the rotating shaft. It rests on a corrugated foil layer, 

known as the bump foil, which serves as an elastic spring foundation that provides the bearing its 

compliancy. As the shaft rotates, it pumps air into the hydrodynamic gas film via a viscous 

dragging mechanism. Consequently, the generated gas film pressure, and therefore resulting 

bearing performance, is directly proportional to speed. During start-up, shut-down, and periods 

of very low shaft speed, there is an insufficient lubricating gas film to maintain separation and 

sliding contact occurs between the top foil and the shaft. To address these contact episodes, solid 

lubricants are applied to the top foil and/or shaft to reduce friction and minimize wear. 
2 

Figure 1. Cross-sectional view of a radial foil air bearing. 

Recently, we have been advocating a four-step process designed to guide the development of foil 

air bearing supported turbomachinery and reduce risk. The four steps include (1) a rotordynamic 

feasibility and layout trade study, (2) bearing sizing and testing, (3) experimental rotordynamic 

simulator validation tests, and (4) a system-level demonstration. The first two steps contain 

component and system-level experimental work supplemented by various analytical and 

empirical modeling tools. For instance, the rotordynamic feasibility step combines computerbased, finite element rotordynamic modeling with empirical equations regarding load capacity 

and stiffness and damping characteristics to predict the stability of a proposed shaft design in the 

form of a critical speed map. The bearing sizing effort in step two encompasses experimental 

verification of the bearing’s load capacity to ensure that its size is adequate for the application. 

The last two steps focus on the overall system by demonstrating operation first on a rotor 

simulator to uncover any unforeseen integration problems requiring attention followed by the 

system-level demonstration. Even though the four-step process is a viable developmental 

approach, there are still missing elements that challenge our ability to predict system-level foil 

bearing performance. 

For example, one such poorly understood aspect of foil bearing performance recently identified 

is thermal management and stability (10, 11). Depending on the speed, foil bearings generate a 

moderate amount of heat due to viscous shearing of the thin gas film. If left unchecked, build-up 

of this heat can cause localized hotspots, resulting in excessive thermal gradients and possible 

bearing seizure. The most common approach to removing this heat is to flow an adequate 

amount of air axially behind the top foil and through the bearing’s bump foils. Another effective 

method (10) is to impinge a jet of air directly on the inner surface of a hollow rotating shaft 

coincident with the bearing, thereby turning the shaft into a rotating heat exchanger for the 
3 

gas film. Regardless of the strategy, research suggests proper thermal management is a key 

fundamental necessity for the successful deployment and operation of foil bearings in 

high-speed turbomachinery systems. 

To address the lack of system-level design tools, we introduced the concept of a performance 

map for foil air bearings. This map is a plot of bearing power loss versus speed covering start-up 

to high-speed operation and describes bearing operating regimes, performance safety margins, 

and the effects of load. The map, as seen in figure 2 from reference 12, is for a constant applied 

load and resembles a Stribeck curve for bearing friction. It is divided into two distinct operating 

regimes, designated as the highly loaded, thin-film regime and the lightly loaded, viscositydominated regime. The minimum in the curve represents the instance where a sufficiently thick 

gas film develops to support the load, but the speed is still low to prevent excessive viscous 

losses. The area to the left of this minimum point is the high load region and is where the 

bearing’s power loss is dominated by the thin-film shear occurring in the high-pressure, load 

support zone. A foil bearing operating in this region will experience large fluctuations in its 

power loss in response to very small changes to speed and/or load and is susceptible to a failure 

mechanism known as thermal runaway. 

Figure 2. Foil bearing performance map (12) illustrating its general shape 

and the two operating regions. 

Thermal runaway is a consequence of the thermal growth mismatch between the shaft and 

bearing and manifests as a continual increase in preload. During operation, most of the viscous 

heat generated in the gas film is conducted into the shaft and bearing (13). However, the small 

cross section of the bump foils limits the rate of heat conduction to the surrounding structure, 

effectively isolating the thin top foil. This causes the top foil to quickly reach the temperature of 

the adjacent gas film leaving the shaft to absorb most of the continuous supply of viscous energy. 

As the shaft expands, it squeezes the air film uniformly around the bearing’s circumference, 

which mimics an increase in preload and expends available load capacity. This extra load on the 

bearing produces additional heating and further growth of the shaft, thereby beginning a thermal 
4 

repetitive cycle that will eventually overload the bearing and cause seizure. Understandably, 

bearings designed with an initially high preload for proper stiffness and damping properties are 

more susceptible to this type of failure mechanism. 

Alternately, the area to the right of the load curve’s minimum has been termed as the low load 

region and is characterized by a more thermally stable operating point for the bearing. In this 

region the entire circumference of the bearing contributes to the power loss via a windage 

mechanism instead of from localized viscous shear in the high pressure zone. However, the high 

speeds in this region can cause the thermal gradients to grow to dangerous levels that will result 

in bearing failure unless an adequate thermal management strategy is employed. 

Despite its importance to Oil-Free systems, little appears to be publicly known regarding 

assessing the thermal robustness of a foil bearing application. Put another way, there is no clear 

and accepted method to judge how much performance margin exists for a foil bearing with 

respect to system-level parameters such as shaft speed, load, and thermal stability. This report 

addresses this shortfall by proposing a three-dimensional (3-D) map that parametrically and 

experimentally characterizes foil bearing power loss as a function of operating speed and load 

and provides a measurable margin of safety for an anticipated application. 

2. Methods, Assumptions, and Procedures

2.1 Foil Bearing and Journal 

The test bearing used to demonstrate the concept of a 3-D performance map is a Generation III 

version of the bearing shown in figure 1. Generation III is a category label for foil bearings with 

improved performance in both load capacity and stiffness and damping properties due to a 

spatially varying support structure in the circumferential, axial, and/or radial directions (14). The 

test bearing was nominally 35 mm in diameter, 27 mm long, and weighed 1.42 N. 

The mating test journal was prepared by first plasma spraying PS304, a high temperature solid 

lubricant coating, onto the surface then placing it into a 540 °C oven for 100 h (15). After the 

heat treatment process, the journal was affixed to the test rig and machined with an in-place 

grinding system to the nominal final diameter of 35 mm, resulting in a bearing preload of 

7.1 kPa. Finally, a computer program, based on the two-plane influence coefficient method, was 

used to balance the rig to 55 krpm. 

Figure 3 shows the testing we performed on the high-speed rig. The rig consists of a drive shaft 

in an overhung configuration supported on ceramic, hybrid ball bearings that are lubricated by a 

closed-loop oil system. Rig bulk temperatures are controlled by circulating chilled water through 

the housing. An air impulse turbine attached to the main shaft is capable of driving the journal to 

a maximum rotational speed up to 60 krpm. The rod extending from the ring stack acts as a 
5 

moment arm to relay the bearing’s operating torque to a load cell for power loss calculations. To 

effectively unload the bearing during startup and shutdown, a pulley-cable system using a 

pneumatic cylinder was installed to suspend the stack. A scatter shield placed around the bearing 

provided a safe operating environment (16). 

Figure 3. Foil bearing high-speed test rig with a stack of three Anviloy rings in place. 

Since the PS304 coating in an as-ground condition will undergo break-in during initial operation, 

the test bearing and journal were together subjected to approximately 1500 start-stop cycles at 

540 °C to eliminate this effect from influencing the data. The start-stop cycles act as a polishing 

wear mechanism between the asperities of the top foil and coating, resulting in a smooth, 

glossy, oxide surface on the PS304 coating and the development of a solid lubricant layer on 

the foil surface (17). After the break-in cycles, the final coating surface roughness was 

approximately 0.2 μm. 

To obtain accurate power loss measurements, four dead weights in the form of rings made from 

Anviloy, a heavy tungsten alloy, were fabricated to apply radial loads to the bearing. Each ring 

weighed about 44 N and was sized so that they could be stacked to simulate loads of 44, 88, 132, 

and 176 N (figure 4). Holes in the rings provided access for the torque rod to be screwed into the 

innermost ring to hold the stack in place. Metal shims placed in the small space between the 

rings centered the weight assembly with respect to the bearing to facilitate single plane static 

balancing of the stack. Because of size differences, the test bearing had to be first placed in a 

steel ring that weighed 21.4 N and secured before inserting into the innermost Anviloy ring when 

higher radial loads were required. 
6 

Figure 4. Exploded view of Anviloy rings capable of producing 

bearing loads from approximately 44 to 176 N. (Steel 

housing for 35 mm bearing not shown). 

The power loss tests were conducted at room temperature to provide a known constant ambient 

environment. The first test began by inserting the bearing into the steel ring and securing it with 

a set screw. The bearing assembly was then placed on the journal, the torque cable attached to 

the load cell and the load support cable hooked to the torque rod. The test load, which included 

the bearing weight and steel ring, was 23 N. To provide a more thermally stable system, a guide 

tube was positioned to inject 0.20 m3/min cooling air into the journal after placement of the 

scatter shield. This is known as the indirect journal cooling method and is further described in 

reference 10. Testing began by supporting the stack assembly with the pneumatic loader and 

accelerating the journal to about 25 krpm to guarantee lift-off. The speed was then reduced and 

held at a constant speed of approximately 20 krpm while the load cable was manually removed 

from the torque rod to allow the bearing to freely support the weight. With the bearing floating 

on the journal, the speed was dropped to the lowest attainable speed for the given stack weight 

(determined through earlier tests) that would put operation in the high load region. Collection of 

torque data began at this point and continued for each increment in speed according to the test 

plan shown in table 1 until reaching 55 krpm. After recording the last data point, the speed was 

dropped back to 20 krpm and the load cable reattached to the torque rod in preparation for rig 

shutdown. The bearing-stack assembly was then suspended with the pneumatic loader while the 

journal came to a complete stop. Once cool the stack was removed from the journal, a ring was 

added, and the stack was balanced and then reinstalled on the rig. This test preparation and 

operation was repeated for loads of 66 N and 109 N. Because of high ball bearing defect 

frequencies in the rig, radial loads greater than 109 N were not attempted. 
7 

Table 1. Test speed increments for data collection. 

Speed Range 

(rpm) 

Speed Increments 

(rpm) 

Up to 20,000 1000 

20–30,000 2000 

30–55,000 5000 

2.2 Results and Discussion 

Figure 5 shows the concept of a 3-D performance map for the test bearing, which was 

constructed with the power loss data collected at three different radial loads. The map has speed 

on the x-axis, power loss on the y-axis, and load on the z-axis. Presenting the data in a contour 

plot format not only quantifies the bearing’s resulting power loss under steady-state conditions 

but also demonstrates its response to changes in both speed and load. As shown in the graph, as 

the bearing’s radial load increased, the curves moved up and to the right. It is expected that 

additional load curves representing heavier loads will continue to shift up and to the right 

following the current trend. Each curve has a minimum that separates the low load and high load 

regimes and is the reference point for the concept of a margin of safety. The definition of a 

margin of safety is only relevant for a bearing operating in the low load region and is a measure 

of how much of a speed reduction can be tolerated before moving into the high load region. At 

this time, there is no set rule-of-thumb for how large the margin should be. It depends on the 

bearing’s application, variations in the operating conditions, and the experience of the designer. 

Figure 5. A 3-D performance map for the 35 mm diameter foil bearing generated from three load curves. 
8 

The usefulness and simplicity of the performance maps as a design tool can also be elucidated by 

presenting the test data in figure 5 on a two-dimensional (2-D) graph, as shown in figure 6, with 

power loss as a function of speed for different radial loads. This more user-friendly format 

enables the designer to directly predict the bearing’s behavior throughout its expected operating 

envelope and also the margins of safety with respect to the high load region and then ascertain if 

the bearing’s abilities are adequate to manage the system’s speed and load requirements. To 

demonstrate, refer to the representative performance map in figure 7. If a system requires the 

bearing to idle at point A and withstand the maximum maneuver load conditions at B at full 

speed, the performance map indicates the bearing is thermally stable at idle since point A lies in 

the bearing’s low load region and has a modest margin of safety. However, as the conditions 

change from A to B, the increase in speed and load causes the bearing’s operating point to travel 

vertically up and to the right from one load curve to another until finally stopping at point B, 

which is an unstable condition that puts the bearing at risk of failure due to thermal runaway. If, 

for example, the bearing was allowed to operate at B and thermal runaway did manifest, the 

increase in preload force would most likely force a decrease in speed. This reaction would be 

represented on the map as point B traveling up and to the left from one load curve to another, 

moving deeper into the high load region, eventually ending in failure. A much more stable 

application for this size bearing would be in a system with higher operating speeds (or lower 

loads), where the idle condition was again at A but the maximum load condition was at point C, 

since both A and C remain in the low load region and have acceptable margins of safety to 

accommodate unforeseen fluctuations in system operating speed. 

0

10

20

30

40

50

60

70

0 10 20 30 40 50 60

Speed, krpm

Power Loss, Watts

23 N 66 N 109 N

Figure 6. A 2-D performance map for the 35 mm diameter foil bearing using the three load curves. 
9 

Figure 7. Representative performance map showing hypothetical operating points. 

Operating points A and C are in the low load region and point B is in the high 

load region. 

Performance maps for other foil bearings types (overlapping leaf, multiple top foils), support 

complexity (backing springs, multiple bump layers), and size are expected to exhibit the same 

general behavior since they all rely on the principle of hydrodynamics but will vary according to 

their own design characteristics. For example, the effect of bearing size on power loss was 

studied by testing the bearing in figure 8, which was of similar design to the 35 mm bearing but 

larger. This bearing’s inner diameter and length were nominally 51 mm and had a type K 

thermocouple tack welded to the middle of the spacer block and positioned just below the plane 

of the top foil near the leading edge to monitor the bearing’s internal temperature. Bearing 

preload was not measured but is estimated to fall in the range of 6.9 to 10.3 kPa. The bearing 

weighed 4.4 N and did not require a steel spacer to properly fit in the smallest Anviloy ring. To 

address the ball defect frequencies the rig was rebuilt with new ball bearings. We only tested 

radial loads of 48 N and 179 N, and followed the same test set-up and procedure, except thermal 

management was administered by directly cooling the journal (10). As can be seen in figure 9, 

the resulting 2-D plot does not show a huge difference between load curves, indicating that 

heavier weights are needed for this size bearing to generate a more informative performance 

map. Unfortunately, high synchronous vibrations prohibited testing loads greater than 178 N. 

Even so, comparing the maps in figure 6 and 9 at identical speeds and applied load conditions 

shows that the power loss signature generated by the 35-mm bearing is lower due to the smaller 

top foil surface area. Consequentially, operating efficiency can be maximized by using the 

smallest bearing that both satisfies the system’s performance requirements and also has an 

acceptable margin of safety. This information can be obtained by constructing maps for each foil 
10 

air bearings being considered for a high speed application. Bearing temperatures for selected 

operating points in figure 9 are provided in table 2 as reference data for the validation of future 

predictive computer models for advanced foil air bearings. 

Figure 8. Drawing of the 51-mm diameter, 51-mm long 

foil bearing. The type K thermocouple is located 

in the middle of the bearing and on the spacer 

block near the leading edge. 
11 

0

25

50

75

100

125

150

175

200

225

0 10 20 30 40 50 60

Speed, krpm

Power Loss, Watts

48 N 179 N

Figure 9. A 2-D performance map for the 51-mm diameter, 51-mm long foil bearing. 

Table 2. Sample power loss data including air film temperature at the 

midpoint near the leading edge for the 51-mm diameter bearing. 

Speed 

(krpm) 

Load 

(Newtons) 

Power Loss 

(Watts) 

Temperature 

(°C) 

10 48 20 27 

30 48 86 44 

50 48 181 65 

20 179 54 38 

30 179 97 47 

50 179 221 69 

In order to prevent the onset of thermal runaway, we recommend that the bearing’s operating 

point always remain in its low load region from idle to maximum maneuver loads. The caveat to 

this stipulation, however, is that the designer has to contend with the possible development of 

excessive axial thermal gradients in the bearing. As explained in reference 18, axial thermal 

gradients are postulated to arise due to the entrapment of a slug of air in the middle of the 

bearing that continuously gets heated from the viscous shear, resulting in a temperature gradient 

that develops across the half-width of the bearing, from the middle to the ends. Whereas thermal 

runaway will result in failure, if left unchecked, the development of small thermal gradients is 

not an indication of impending failure as foil bearings can safely operate in their presence. 

Excessive gradients, those around 20 °C/cm or above, are a concern and can manifest during 

high surface speed operation. When present, they can significantly warp the top foil to a point 

that film formation is disrupted, lowering the available fluid-film for load support and leading to 

high-speed rubbing contact and eventual catastrophic failure (11). As with thermal runaway, its 

effect on the bearing is similar to a gradually increasing radial load and its progression can also 

be explained via the performance map. As the gradient grows the bearing’s operating point 
12 

moves vertically up from one load curve to the next and to the left since there is usually a 

decrease in speed due to the increased power loss. Eventually, the operating point will enter into 

the high load region and continue to do so until complete, and most likely catastrophic, bearing 

failure. This can be averted by employing an effective thermal management strategy based on 

maximum operating conditions to control the thermal gradient, whether by removing excess heat 

through indirect cooling of the journal (as was done for this research) or other effective 

techniques (10). Thermal management does not imply cooling the bearing below a certain level 

since high temperature operation is one of their advantages. Instead, its purpose is keep the 

magnitude of the temperature gradient under control to ensure the bearing’s performance is not 

compromised while operating in the low load region of the performance map. 

Thermal management can also be effective at preventing thermal runaway. This is evidenced by 

the existence of the power loss data in the low load region of this report, but only to a limited 

extent in highly preloaded bearings. Because of the steep slope of the curves in this region, offdesign operation, due to a sudden, small increase in load or decrease in speed, can cause a 

significant jump in power loss, which could overwhelm the existing thermal management 

strategy, rendering it ineffective. With a passive thermal management strategy in place, the 

constant mass flow of air may not be sufficient to remove the extra viscous heat generated. One 

solution is a self-compensating system that continuously monitors the bearing’s operating 

conditions and makes adjustments in the cooling air mass flow to counteract the excess heat. 

Setting up an active thermal management strategy may not only be difficult to implement but the 

required additional air flow will be at the expense of system efficiency. Therefore, the 

recommended approach is to design the system to operate only in the bearing’s low load region 

with an acceptable margin of safety and also implement an effective thermal management 

strategy to prevent the formation of large axial thermal gradients. 

3. Conclusions 

In this report, we presented the concept of a 3-D power loss performance map as a function of 

speed and load. The power loss map consists of a series of individual load curves that are 

separated into two operating conditions, the low and high load regimes. The high load region 

sits on the left side of the curve’s minimum point and is characterized by large changes in power 

loss for small variations in speed and load. Bearings designed to operate in this region are at risk 

of failure due to thermal runaway. Alternately, bearing’s operating in the low load region on the 

right side of the minimum are more stable but are susceptible to the development of excessive 

axial thermal gradients that can lead to failure. Employing an effective thermal management 

strategy can control these failure modes, especially in the low load region. 
13 

As a design tool, the power loss map will compliment the load capacity and rotordynamic 

models during the feasibility step of the four-step process in determining the size of the bearing 

by forcing concessions between load capacity, stiffness, and damping and power loss. With this 

data, designers can then target foil air bearings that operate in their low load region throughout 

the system’s entire operating envelope. 
14 

4. References 

1. Gross, W. A. Gas Film Lubrication; John Wiley and Sons: New York, 1962, pp 138–114. 

2. Ma, J.T.S. An Investigation of Self-Acting Foil Bearings. ASME J. of Basic Engineering

1965, 837–846. 

3. Barnett, M. A.; Silver, A. Application of Air Bearings to High-Speed Turbomachinery. SAE

Paper 700720, 1970. 

4. Emerson, T. P. The Application of Foil Air Bearing Turbomachinery in Aircraft 

Environmental Control Systems. Proceedings of the ASME Intersociety Conference on 

Environmental Systems, San Diego, CA, Paper No. 780-ENAS-18, 1978. 

5. Agrawal, G. L. Foil Gas Bearings for Turbomachinery. Proceedings of 20th Intersociety 

Conference on Environmental Systems, Williamsburg, VA, SAE Paper No. 901236, 1990. 

6. Lubell, D.; DellaCorte, C.; Stanford, M. K. Test Evolution and Oil-Free Engine Experience 

of a High Temperature Foil Air Bearing Coating. Proceedings of GT2006: ASME Turbo 

Expo 2006, Barcelona, Spain, GT2006-90572, 2006. 

7. Heshmat, H.; Walton, II, J. F.; DellaCorte, C.; Valco, M. Oil-Free Turbocharger 

Demonstration Paves Way to Gas Turbine Engine Applications. International Gas Turbine 

and Aeroengine Congress and Exhibition sponsored by ASME, Munich, Germany, 200-GT620, 2000. 

8. Suriano, F. J. Gas Foil Bearing Development Program; U.S. Air Force No. AFWAL-TR-81-

2095; 1981. 

9. Heshmat, H.; Walton, II, J. F.; Tomaszewski, M. J. Demonstration of a Turbojet Engine 

Using an Air Foil Bearing,” ASME Paper GT-2005-68404. ASME Turbo Expo 2005, Power 

for Land, Sea and Air, Reno-Tahoe, NV, 2005. 

10. Radil, K.; DellaCorte, C.; Zeszotek, M. Thermal Management Techniques for Oil-Free 

Turbomachinery Systems. STLE Tribology Transactions 2007, 50 (3), 319–327. 

11. Dykas, B.; Howard, S. A. Journal Design Considerations for Turbomachine Shafts 

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and Experiments. ASME Journal of Tribology 2001, 123, 566–571. 
15 

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18. Radil, K.; Zeszotek, M. An Experimental Investigation into the Temperature Profile of a 

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