A Novel Thermal Management Approach for Radial 

Foil Air Bearings 

by Kevin Radil and CDT Zach Batcho 

ARL-MR-0749 July 2010 

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Army Research Laboratory 

NASA Glenn, Cleveland, OH 44135-3191 

ARL-MR-0749 July 2010 

A Novel Thermal Management Approach for Radial 

Foil Air Bearings 

Kevin Radil 

NASA Glenn Research Center 

Cleveland, OH 

and 

CDT Zach Batcho 

U.S. Military Academy 

West Point, NY 

Approved for public release; distribution unlimited. 
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A Novel Thermal Management Approach for Radial Foil Air Bearings 

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Kevin Radil and CDT Zach Batcho 

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U.S. Army Research Laboratory 

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14. ABSTRACT 

We experimentally evaluated a novel thermal management technique for radial foil air bearings. The technique is based on 

injecting air directly into the internal circulating fluid-film to reduce bulk temperatures and axial thermal gradients. The tests 

were performed on a single top foil, Generation III, radial foil bearing instrumented with three thermocouples to monitor 

internal temperatures. A thru hole in the bearing shell coincident with the gap between the top foil’s fixed and free ends 

provided entry for the injection air. The tests were conducted at room temperature with the bearing operating at speeds from 

20 to 50 krpm while supporting 222N. Two different mass flow rates of injection air were evaluated for this method, 0.017 

and 0.051 kg/min. Test results suggest that the air injection approach is a viable thermal management technique capable of 

controlling bulk temperatures and axial thermal gradients in radial foil air bearings. 

15. SUBJECT TERMS

Foil air bearing, gas bearing, turbomachinery 

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Kevin Radil 

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3 

Contents 

List of Figures 4

List of Tables 4

1. Introduction 5

2. Test Apparatus 9

3. Procedure 11

4. Test Results and Discussion 11

5. Conclusion 16

6. References 17 

Distribution List 19 
4 

List of Figures 

Figure 1. Schematic of the test bearing showing locations of thermocouples and inlet hole 

for air injection. ..........................................................................................................................9

Figure 2. Hardware layout for injecting air into the bearing. .......................................................10

Figure 3. High-speed, high-temperature, journal foil air bearing test rig. ....................................11

Figure 4. Time history plot of bearing temperatures and maximum axial thermal gradient 

showing the effects of applying air injection flows of 0.017 and 0.051 kg/min. Speed = 

30 krpm, load = 222N. .............................................................................................................13

Figure 5. Bearing temperatures demonstrating an excessive thermal gradient situation. 

Injection air at 0.051 kg/min was delivered at the 3.5 min mark followed by 0.068 kg/min 

at the 4.5 min mark. Test was discontinued once the thermal gradient surpassed 25 

°C/cm. Speed = 50 krpm. Load = 89N. ..................................................................................14

Figure 6. Repeat of operating conditions that led to failure (figure 5) except a constant air 

injection of 0.068 kg/min was applied at the beginning of the test. Speed = 50 krpm. 

Load = 89N. .............................................................................................................................14

List of Tables 

Table 1. Bearing temperature results for the two injection air flows at three different 

operating speeds while supporting 222N. Injection air temperature = 23 °C. ........................12

Table 2. Effect of air injection location with respect to the load support region on bearing 

temperatures and thermal gradient. Speed = 20 krpm. Load = 89N. Injection air 

temperature = 22 °C. ................................................................................................................15
5 

1. Introduction 

Foil bearings are hydrodynamic bearings that generate a thin gas film (~10–5m) from the natural 

pumping action of a rotating shaft that drags air into a convergent wedge-shaped gap to produce 

pressure. The basic elements of a foil air bearing are the shell, top foil, and bump foils. The top 

foil is a thin flexible membrane that retains the generated fluid pressure and is supported by 

elastic support structures called bump foils. Bump foils are shaped similar to the corrugations in 

a cardboard box with each bump acting as a structural spring that allows the top foil to move in 

response to hydrodynamic pressure changes due to static and dynamic loads, shaft centrifugal 

and thermal growth, and misalignment. Bump foils are also responsible for the bearing’s 

stiffness and damping properties and can be adjusted for specific applications through their 

judicious design. The material of choice for foil bearings is typically a nickel based superalloy 

because of its high temperature stiffness properties but other materials, chosen with respect to 

operating requirements and conditions, may also be used. 

The concept of a foil bearing was first described in a report over 50 years ago by Blok and Van 

Rossum (1). Since their discovery, foil bearings have been commonly used as rotor supports in 

air cycle machines (ACM’s) that are part of an aircraft’s cabin environmental system (2). This 

application is relatively benign and requires only simple bump foil designs because speeds are 

high and temperatures and loads are low. Transitioning these simple designed foil bearings into 

more complex systems was subsequently attempted but progressed no further than the research 

stage (3, 4). However, during the last 15 years, more advanced, higher load capacity bearings 

and high temperature coatings for wear protection have been developed, driving the technology 

into larger, more demanding systems. Current Oil-Free commercial products using advanced 

bearings include 60 and 200 kW microturbines for power generation and air compressors (5). 

Recent industrial research activities are suggesting that future Oil-Free turbomachinery will also 

include turbochargers, larger microturbines, auxiliary power units (APU’s) and gas turbine 

engines (6, 7). 

With air as the lubricant, foil bearings can successfully support high speed rotors in extreme 

temperature environments without the threat of fluid degradation. The ability to operate at high 

temperature, however, does not preclude them from having their own unique thermal 

management issues that require attention during the design of an Oil-Free system. The thermal 

profile in a foil bearing results from a complex interaction of both internal and external factors. 

In a gas turbine engine environment, for example, a foil bearing’s temperature would be 

influenced by its proximity to high temperature components, such as the combustor, compressor 

or turbine and the heat that it produces internally from the viscous shear occurring in the thin air 

film and any compression work. Even though the viscosity of air is low, the high surface speed 

required by the bearings for adequate load support coupled with a thin film results in a large 
6 

velocity gradient that generates a moderate amount of heat. The low thermal capacity of air 

makes it difficult for any appreciable amount of heat to be removed with side leakage leaving 

most to dissipate into the shaft and bearing. In fact, estimates place the heat absorbed by these 

two components at about 80% with the remaining 20% convected by side leakage (8). The end 

result is a shaft-bearing interface operating hotter than its immediate surroundings. 

Breaking the energy distribution down further, it’s expected that the shaft accepts a majority of 

this 80% because of its large exposed surface area and extension into cooler parts of the engine. 

The bearing’s inability to dissipate its share of heat stems from its construction; specifically, the 

bump foils that enable its compliant nature. From a thermal viewpoint the bump foils isolate the 

top foil from the rest of the bearing because the support points are in line contact. This limits the 

conduction of energy which causes the thin top foil to quickly approach thermal equilibrium with 

the adjacent air film leaving most of the heat to flow into the shaft. Any heat that does pass 

through the bearing is controlled by thermal emission to the surrounding hardware and any 

energy transfer that does occur at the top foil/bump foil contact points. 

If the localized temperatures grow unabated, bearing performance can be detrimentally affected 

in a number of ways. One concern is overheating of the bump foil material. Foil bearings are 

typically made from Inconel, a nickel based superalloy, because it retains an elastic modulus 

(spring rate) at high temperature better than most materials. However, as with most materials, it 

does experience a drop in its modulus with an ever increasing temperature. This softening effect 

increases the bearing’s compliance, which reduces the bearing’s maximum load capacity as 

demonstrated analytically and experimentally (9, 10). In addition, research suggests that changes 

to the bearing’s stiffness (and possibly damping) properties will occur which can impact a 

system’s rotordynamic behavior (11). 

Foil bearings can also suffer from thermal runaway, a condition that usually leads to failure and 

is caused by a continuous increase in the bearing’s preload. Preload in a foil bearing is obtained 

by designing the bearing’s ID to be slightly smaller than the OD of the shaft, similar to an 

interference fit. When the top foil is manually retracted to allow its insertion onto the shaft, the 

bump foils elastically deform and produce reaction forces (preload) that push back against the 

shaft. This preload force acts over the entire top foil area. For manageable start-up torque, liftoff speed, and load capacity considerations, preload is typically around 6.89 to 10.34 kPa. 

Lighter preloads can adversely affect a bearing’s stiffness and damping properties and, if low 

enough, can lead to rotor instability. On the other hand, a heavier preload will create higher 

start-up torque and lift-off speed requirements and a reduction in load carrying capacity. 

Because the shaft absorbs most of the heat during bearing operation, it expands faster than the 

bearing which increases the preload. This amplifies the stress on the fluid-film, much like an 

increase in the radial load, and leads to additional heat generation and further expansion of the 

shaft resulting in a self-sustaining cycle (thermal runaway) that will continue until the preload 

force overwhelms the fluid-film and high speed rubbing occurs (12). Understandably, bearing 
7 

types requiring an initially high preload for stiffness and damping considerations are more 

susceptible to this type of failure. 

A third thermally induced problem is the formation of excessive axial thermal gradients across 

the bearing’s half-width (middle to the edge). This phenomenon has been experienced during 

operation of most foil bearing types (overlapping leaf, multi-pad, single top foil) in the authors’ 

lab and is especially prevalent in advanced, third generation (Generation III) bearings where the 

bump foils are designed to limit side leakage by stiffening the support at the bearing’s edges. An 

excessive thermal gradient can sufficiently warp the top foil to the point that it interferes with the 

formation of a fully developed fluid-film. The cause for this phenomenon is hypothesized to be 

a circumferentially flowing slug of air that remains trapped in the middle of the bearing that 

undergoes continuous heating due to viscous shear (13). In fact, results from computer-based 

modeling of the fluid-film mass flow do provide credible evidence to support this theory (14). In 

a real application, non-uniform heating due to external sources can exacerbate the presence of 

these gradients and requires a thorough analysis to avoid dangerous levels. Of the three thermal 

related failure mechanisms, axial gradients are the most dangerous because the bearing can 

exhibit stable operation but fail when pushed to higher speed and/or load levels. 

The importance of thermal management can best be elucidated by considering the report by 

Dykas who recounted a catastrophic bearing failure that produced a dime-sized hole in an 

Inconel journal (15). The amount of energy produced during the failure was substantial, 

considering the melting temperature of Inconel is around 1400 °C. Just before the failure, the 

50.8 mm x 50.8 mm bearing was operating at a speed of 60 krpm in a 538 °C environment and 

supporting 222N load without active cooling. At this size and speed the bearing should easily 

support over 1700N based on the empirical rule-of-thumb equation for estimating load capacity 

for radial foil bearings (16). According to the report, computer-based structural modeling 

predicted that the failure was caused by the superposition of two phenomena, the first being point 

load deformation of the journal due to centrifugal forces from balance correction weights and the 

second being the presence of a large axial thermal gradient in the bearing. This combination 

produced a bulge in the journal that grew as it was continuously heated during operation until 

breaking through the air film, leading to a high-speed rub and bearing failure. As stated above, 

axial thermal gradients are a common occurrence in foil bearings and are an inherent trait 

stemming from their design and operating principle. Lab tests suggest that there is a risk of 

failure when they exceed approximately 22 °C/cm. This can happen in any ambient environment 

once the journal surface speed surpasses around 133 m/s but it depends upon other factors as 

well, such as type, preload, design, and length-to-diameter (L/D) ratio. For example, lab testing 

an L/D = 1, Generation III bearing in a room temperature environment with moderate preload 

(~6.89 kPa) and under low load experienced a destructive axial gradient when operating at 

speeds above 50 krpm and required active cooling to continue operation. 

To address the need for thermal management in Oil-Free systems, the author previously tested 

the effectiveness of three techniques to control the viscous heat in radial foil bearings; direct 
8 

journal cooling, indirect journal cooling, and axial bearing cooling (17). The direct and indirect 

methods were designed to draw heat from the bearing by increasing the rate of thermal 

conduction through the journal. One advantage to this approach is that it addresses the journal 

growth due to thermal expansion and the resulting change in the bearing’s preload force. The 

direct method impinged a jet of air on the inner surface of a hollow journal coincident with the 

bearing to maximize the convection coefficient and heat transfer. The indirect method of 

running air axially along the journal’s inner diameter surface to increase the convection 

coefficient was less effective at bearing cooling. The test results proved that both techniques 

were successful at reducing bulk temperatures and axial thermal gradients but the direct method 

was more effective. The axial cooling technique, on the other hand, simulated the industry 

standard approach to thermal management by forcing air through the bearing’s support structure 

(bump foils) to remove heat. This approach was also effective but results suggested that air flow 

rate and direction were important and, if not done with care, could exacerbate the development 

of an axial gradient to the point of bearing failure. This technique also runs the risk of increasing 

the bearing’s preload and thermal runaway as direct cooling shrinks the bearing while the journal 

is expanding from the viscous heat, albeit at a slower rate. To minimize the risks, it is 

recommended that air be introduced into the hotter end of the bearing and at a flow rate that 

produces a more even axial temperature distribution. Applying any of these methods requires a 

thorough thermal analysis on the system to understand the temperature distribution in the bearing 

during operation, especially if it is located near a heat source. Given the choice of the three 

methods, direct journal cooling provides the most inherent benefits, but unfortunately, may be 

difficult or impossible to implement in most systems. 

The operating principle behind hydrostatic gas bearings may provide an alternate, more useful 

approach to thermal management in foil air bearings. Hydrostatic bearings, whether radial or 

axial, depend upon a pressurized fluid injected into the bearing cavity to separate surfaces and 

provide load support. The pressure driven flow is also responsible for maintaining moderate 

bearing temperatures by removing the viscous heat produced in the fluid-film. Applying this 

concept to foil bearings operating in a gas turbine engine environment would require injecting air 

from one of the compressor stages directly into the middle of the circulating thin air film. 

Expected to be set-up as an open-loop arrangement, the heated air expelled from the bearing 

would then be piped through the engine to exit in the rear. This concept is similar to the 

approach reported by Kim and Park who integrated a hydrostatic capability into a radial foil 

bearing to alleviate the high start-up torque associated with large systems and to increase bearing 

load capacity (18). The hydrostatic function also demonstrated the ability to reduce bearing 

temperatures. However, there is no intent with this concept to reproduce a hydrostatic load 

carrying effect. Its only purpose is to provide thermal management control and is expected to be 

applicable to foil bearing designs using single or multiple top foils where space between foil(s) 

exist. 
9 

This report details a novel thermal management technique capable of reducing axial thermal 

gradients and bearing bulk temperatures by injecting air directly into the middle of the bearing to 

promote mixing with the hot circulating air film. The study was conducted on a Generation III 

bearing instrumented with three, type K thermocouples and an access hole for the injection air. 

The tests were conducted at room temperature and at various speeds while the bearing was 

supporting 222N. The technique was evaluated at two different injection mass flow rates, 0.017 

kg/min and 0.051 kg/min.

2. Test Apparatus 

Shown in figure 1 is the Generation III bearing used for the tests (19). The bearing has a 

nominal 50 mm inner diameter and a length of 41 mm. Temperatures were measured with three, 

type K thermocouples that were placed in the same axial plane, one in the middle (#2) and one 

on either side (#1 & #3) with a distance of 19 mm between #1 and #2 and 14 mm between #2 

and #3. This distribution allowed monitoring of the axial temperature gradient that develops 

across the bearing’s half-width. For a close approximation of the top foil temperature, the 

thermocouple was glued to the backside of a bump that was in contact with the top foil. The hole 

for air injection was located in the space between the free and fixed ends of the top foil and 

coincident with the middle of the bearing, placing it directly opposite from the load support 

region of higher pressure. Its orientation also directed the air to enter the bearing perpendicular 

to the circulating air film. No attempt was made, either analytically or experimentally, to 

ascertain the proper size, shape, or orientation of the hole to optimize results. In addition, a 

small amount of the top foil at the free end was removed with a high-speed grinding tool to 

provide an unobstructed path for the injection air to enter the bearing. 

Figure 1. Schematic of the test bearing showing locations 

of thermocouples and inlet hole for air injection.
10 

The set-up for the injection air delivery system is shown in figure 2 and consisted of steel and 

flexible plastic tubing and a 0.228 m3/min full scale flow meter supplied by 861.8 kPa shop air. 

The steel tube had an inner diameter of 2.9 mm and, with the smallest cross-sectional area, set 

the maximum amount of air deliverable to the bearing. A threaded fastener at the end of the steel 

tube screwed into the bearing holder and its tip extended into the injection hole in the bearing 

shell. Plastic tubing with an inner diameter of 4.3 mm was used to connect the flow meter to the 

air source and steel tube. A thermocouple inserted at the plastic tube/steel tube junction 

measured the air temperature before entering the bearing. With the flow meter control valve in 

its fully open position, the maximum flow rate was 0.068 kg/min 

Figure 2. Hardware layout for injecting air into the bearing.

The normal, in-house procedure for preparing the test journal is to first plasma spray a high 

temperature, solid lubricant coating onto the journal surface and then heat treat it at 520 °C for 

100 hr. Once cool, the journal is installed on the test rig and sized, using an in-place grinding 

system, to a final nominal diameter of 50.8 mm. After performing in-place, two-plane balancing 

to reach 50 krpm, the bearing and journal are together subjected to approximately 700 start-stop, 

break-in cycles under a 6.89 kPa load at 520 °C to produce a lubricious oxide surface layer and 

conforming surfaces (20). 

The high-speed rig used to perform the tests is shown in figure 3 and is described in greater 

detail in reference 10. The rig consists of a main shaft supported by two ceramic, oil lubricated 

ball bearings in an overhung arrangement. An air impulse turbine attached to one end of the 

shaft is capable of driving the journal to a maximum rotational speed of 60 krpm. Radial loading 

of the foil bearing is accomplished via a vertical cable system with one end attached to the 

bearing housing in a stirrup configuration and the other to a pneumatic load cylinder mounted 

below the test rig. The rod extending from the bearing housing acts as a moment arm to relay the 

bearing torque to a load cell. The torque reading, however, was only used for health monitoring 
11 

as the air delivery system had an influence on the output. The furnace consists of two halves that 

surround the bearing for heating and also as a scatter shield for operational safety. 

Figure 3. High-speed, high-temperature, journal foil air bearing test rig.

3. Procedure 

Test preparation began by placing the bearing on the journal, attaching the air delivery system 

steel tube and radial loading cable to the bearing and then connecting the thermocouple leads. 

Once the scatter shield was in position and the data acquisition system started, the rig was 

accelerated to the test speed and the load applied. This condition was maintained for about 10 

min to allow bearing temperatures to reach steady-state before air introduction. The first flow 

rate of 0.017 kg/min was injected into the bearing and held until the temperatures reached 

steady-state. The injection air flow rate was then tripled to 0.051 kg/min and the operating 

conditions maintained until temperatures again reached steady-state, at which time the data 

acquisition was stopped and the test concluded. This procedure was followed under room 

temperature conditions (23 °C) for speeds from 20 to 50 kprm while supporting 222N. 

4. Test Results and Discussion 

Overall test results are shown in table 1. For each test, the bearing temperatures and thermal 

gradients for injection flows of 0 (baseline), 0.017 and 0.051 kg/min are given. As demonstrated 

in the table the highest bearing temperatures were in the middle with the lower temperatures 
12 

occurring towards the edges. This temperature distribution is very common in this and other foil 

bearing designs along with the thermal gradient acting across the bearing’s half width. As 

expected, bearing temperatures rose with increasing speed due to higher viscous shear. In fact, 

every temperature reading almost doubled when the speed went from 20 to 40 krpm. For each 

test condition, the response to air injection at 0.017 kg/min was almost immediate with a 

decrease in both bearing temperatures and thermal gradient but the change was small. The 

reaction was much more pronounced when the flow was tripled to 0.051 kg/min. An example of 

this behavior can be seen in figure 4 for the bearing operating at 30 krpm and supporting a 222N 

load. In this case, there was only a 3% and 2% decrease in the middle and edge temperatures, 

respectively, with an injection flow of 0.017 kg/min. However, at 0.051 kg/min, the decrease 

was 20% and 17%, respectively, demonstrating that bearing temperatures and thermal gradients 

can be managed through proper control of the injected air flow rate. The results also 

demonstrate the method’s effectiveness is not localized to the middle of the bearing (coincident 

to the injection hole’s location) but influences the entire bearing. It is not clear if the physical 

mechanism responsible for these results is due to displacing the retained slug of air in the middle 

of the bearing with a fresh, cooler supply or from fluid mixing. Additionally, it is unknown if 

the injected air is penetrating the air film sufficiently to directly cool the journal. Only through 

further in-depth experimental research or computer-based analytical modeling can these 

questions be answered. 

Table 1. Bearing temperature results for the two injection air flows at three different operating speeds while 

supporting 222N. Injection air temperature = 23 °C. 

Injection Flow, 

kg/min T1, °C T2, °C T3, °C Max Gradient, 

°C/cm 

20 krpm 

0 77.2 85.6 77.2 6.0 

0.017 77.4 83.9 77.4 4.6 

0.051 67.4 70.0 65.7 3.1 

30 krpm 

0 104.0 116.6 102.4 10.2 

0.017 102.2 112.7 100.6 8.7 

0.051 87.3 93.5 85.1 6.0 

40 krpm 

0 148.4 169.7 148.4 15.2 

0.017 146.4 164.5 146.4 12.9 

0.051 122.9 136.5 123.0 9.6 
13 

Figure 4. Time history plot of bearing temperatures and maximum axial thermal gradient showing the effects of 

applying air injection flows of 0.017 and 0.051 kg/min. Speed = 30 krpm, load = 222N. 

The true effectiveness of injecting air as a thermal management technique was demonstrated 

when attempting to conduct tests at the highest speed. When testing the bearing operating at 

50 krpm and supporting 89N (excessive test rig vibrations prohibited higher radial loads at this 

speed), the bearing temperatures did not reach steady-state but continued to rise along with the 

thermal gradient that neared 30 °C/cm, which is a dangerous level. A plot of bearing 

temperatures is shown in figure 5. Concurrently, there was also a slow increase in bearing torque 

and a corresponding drop in operating speed, symptoms of impending failure. An attempt was 

made to prevent the failure by first injecting 0.051 kg/min into the bearing at the 3.5 min mark, 

then 0.068 kg/min of air a minute later to reduce the gradient. Unfortunately, this was 

unsuccessful and the test was discontinued at the 5 min mark to avert failure and possible bearing 

and journal damage. Once the bearing had cooled, a second attempt at completing the test was 

made by injecting the maximum flow rate of 0.068 kg/min at the start to control the gradient. 

This approach was successful as it prevented the gradient from growing to excessive levels. As 

shown in figure 6 bearing temperatures reached steady-state with the middle temperature at 

192 °C and the other two were approximately 174 °C. The thermal gradient remained low and 

fluctuated around 12 °C/cm. This test demonstrated that injecting air into the circumferential 

flow stream is an effective thermal management technique that allows foil bearings to operate at 

conditions that would otherwise result in failure. 

0

3

6

9

12

15

18

0

20

40

60

80

100

120

0 5 10 15 20 25 30 35 40 45

Bearing Temperature, °C

Thermal Gradient, °C/cm

Time, min

TC #1 TC #2 TC #3 Max Gradient

0.017kg/min 0.051kg/min 
14 

Figure 5. Bearing temperatures demonstrating an excessive thermal gradient situation. Injection air at 

0.051 kg/min was delivered at the 3.5 min mark followed by 0.068 kg/min at the 4.5 min mark. 

Test was discontinued once the thermal gradient surpassed 25 °C/cm. Speed = 50 krpm. 

Load = 89N.

Figure 6. Repeat of operating conditions that led to failure (figure 5) except a constant air injection of 

0.068 kg/min was applied at the beginning of the test. Speed = 50 krpm. Load = 89N.

0

10

20

30

40

50

0

50

100

150

200

250

012345678

Bearing Temperature, °C

Thermal Gradient, °C/cm

Time, min

TC #1 TC #2 TC #3 Max Gradient

0.051Kg/min 0.068Kg/min

Termination of test

0

10

20

30

40

50

0

50

100

150

200

250

0 5 10 15 20

Bearing Temperature, °C

Thermal Gradient, °C/cm

Time, min

TC #1 TC #2 TC #3 Max Gradient
15 

Unlike testing in the lab where the radial load remains fixed in the same circumferential location, 

a foil bearing operating in a real application, such as in a gas turbine engine on a helicopter, is 

required to support dynamic loads in any direction based on the resultant vector summation of 

rotor weight and gyroscopic forces produced during any flight maneuver. To address this issue, 

tests were devised to evaluate if the angular distance between the load support region and the 

point of injection either contributes to or deters from the method’s effectiveness. Changes to this 

distance affect the time available for the incoming air to either mix or displace the existing air 

film. Two tests were performed at room temperature with the bearing operating at 20 krpm and 

supporting 89N. Adjusting the hole location was accomplished by unscrewing the threaded 

fastener for the steel tube, rotating the bearing in its deadweight housing to line up the injection 

hole with pre-drilled, threaded holes, and reinstalling the fastener (see figure 2). The first test 

was performed with the injection hole 90° before the high pressure zone (upstream) and the 

second test with the hole 90° after the high pressure zone (downstream). Results are shown in 

table 2. For the upstream case the bearing pre-test temperature was a uniform 23.4 °C but, after 

operating for 10 min, bearing temperatures rose with the middle reaching approximately 76.1 °C 

and 69.2 °C at the edges (nos. 1 & 3), resulting in a 5.0 °C/cm thermal gradient. Injecting air at 

0.051 kg/min dropped the middle temperature to 59.8 °C and the edges to approximately 56.0 °C 

with the gradient leveling off at 3.1 °C/cm. At the beginning of the downstream test, the bearing 

temperature was a uniform 20.8 °C but rose to 71.9 °C in the middle, 68.0 °C at the edges with a 

2.8 °C/cm gradient, close to the preceding test. After 10 min of injecting 0.051 kg/min of air the 

bearing temperatures reached steady-state with the middle at 57.3 °C and the edges 55.6 °C. The 

gradient was 1.2 °C/cm. A comparison of the upstream and downstream test results does not 

show any significant differences in bearing temperatures, therefore suggesting that the method’s 

effectiveness is not a function of injection location. 

Table 2. Effect of air injection location with respect to the load support region on bearing temperatures and thermal 

gradient. Speed = 20 krpm. Load = 89N. Injection air temperature = 22 °C. 

Injection Flow, 

kg/min T1, °C T2, °C T3, °C Max. Gradient, 

°C/cm 

90° Upstream 

0 23.4 23.4 23.4 - 

0 69.2 76.1 69.2 5.0 

0.051 56.6 59.8 55.4 3.1 

90° Downstream 

0 20.8 20.8 20.8 - 

0 68.0 71.9 68.0 2.8 

0.051 55.6 57.3 55.6 1.2 
16 

5. Conclusion 

Test results suggest that injecting air into the circulating thin film of a foil bearing is a viable 

thermal management technique at controlling bearing bulk temperatures and thermal gradients. 

It is not, at the injection flow rates used here, as effective as other methods demonstrated by the 

author in the past, but room for improvement exists and may be more easily implemented. Test 

results also suggest that injection location, with respect to the load support region, does not 

impact the effectiveness of the technique. 

Because there was no attempt to optimize the technique’s effectiveness, additional research on 

method of air delivery should be performed before implementing this approach. Specifically, the 

hole quantity, size, geometry (oval, round), and orientation (angled to deliver the air with or 

against the circulating air film) should be investigated. Lastly, foil thrust bearings use a number 

of individual, sector-shaped pads spaced evenly around a circular plate. With ample space 

between the pads foil thrust bearings would likely benefit from this type of thermal management 

approach. 
17 

6. References 

1. Blok, H.; van Rossum, J. J. The Foil Bearing-A New Departure in Hydrodynamic 

Lubrication. Lubrication Engineering 1953, 9 (6), 316–320. 

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

Environmental Control Systems. in Proc. of the ASME Intersociety Conference on Environmental Systems, San Diego, CA, Paper No. 780-ENAS-18, 1978. 

3. Advanced Gas Turbine (AGT) Technology Development, NASA CR-165175, 1979. 

4. Suriano, F. J.; Dayton, R. D.; Woessner, F. G. Test Experience with Turbine-End Foil 

Bearing Equipped Gas Turbine Engines. ASME 28th International Gas Turbine Engine 

Conference, Phoenix AZ, March 27–31, 1983. 

5. Lubell, D.; DellaCorte, C.; Stanford, M. Test Evolution and Oil-Free Engine Experience of a 

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

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

6. Larue, G.; Kang, S.; Wick, W. Turbocharger with Hydrodynamic Foil Bearings. Patent 

#7,108,488, 2006. 

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

Using an Air Foil Bearing; ASME Paper GT-2005-68404; presented at the ASME Turbo 

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

8. Salehi, M.; Swanson, E.; Heshmat, H. Thermal Features of Compliant Foil Bearings-Theory 

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

9. Heshmat, H.; Walowit, J.; Pinkus, O. Analysis of Gas-Lubricated Foil Journal Bearings. 

ASME Journal Lubrication Tech Trans. 1983,105 (4), 638–646. 

10. DellaCorte, C. A New Foil Air Bearing Test Rig for Use to 700 °C and 70,000 rpm; NASA 

TM 107405; 1997. 

11. Howard, S.; DellaCorte, C.; Valco, M.; Prahl, J. Steady-State Stiffness of Foil Air Journal 

Bearings at Elevated Temperatures. STLE Tribology Transactions 2001, 44 (3), 489–493. 

12. Radil, K.; Howard, S.; Dykas, B. The Role of Radial Clearance on the Performance of Foil 

Air Bearings. STLE Tribology Transactions 2002, 45 (4), 485–490. 

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

Compliant Foil Air Bearing. STLE Tribology Transactions 2004, 47 (4), 470–479. 
18 

14. Carpino, M.; Talmage, G. A Fully Coupled Finite Element Formulation for Elastically 

Supported Foil Journal Bearings. STLE Tribology Transactions 2003, 46 (4), 560–565. 

15. Dykas, B.; Howard, S. Journal Design Considerations for Turbomachine Shafts Supported 

on Foil air Bearings. Trib. Trans. 2004, 47, 508–516. 

16. DellaCorte, C.; Valco, M. Load Capacity Estimation of Foil Air Journal Bearings for OilFree Turbomachinery Applications. STLE Tribology Transactions 2000, 43 (4), 795–801. 

17. Radil, K.; Zeszotek, M. Thermal Management Techniques for Oil-Free Turbomachinery 

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

18. Kim, D.; Park, S. Hydrostatic Air Foil Bearings: Analytical and Experimental Investigation. 

Tribology International 2009, 42, 413–425. 

19. Heshmat, H. High Load Capacity Compliant Foil Hydrodynamic Journal Bearing. U.S. Patent 

No. 5,988,885, 1999. 

20. Radil, K. C.; DellaCorte, C. The Effect of Journal Roughness and Foil Coatings on the 

Performance of Heavily Loaded Foil Air Bearings; NASA TM-2001-210941; 2001. 
19 

No. of 

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