Samuel A. Howard, Christopher DellaCorte, and Robert J. Bruckner

Glenn Research Center, Cleveland, Ohio

Integration Methodology For Oil-Free Shaft Support 

Systems: Four Steps to Success

NASA/TM—2010-216827

November 2010
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Samuel A. Howard, Christopher DellaCorte, and Robert J. Bruckner

Glenn Research Center, Cleveland, Ohio

Integration Methodology For Oil-Free Shaft Support 

Systems: Four Steps to Success

NASA/TM—2010-216827

November 2010

National Aeronautics and

Space Administration

Glenn Research Center

Cleveland, Ohio 44135

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NASA/TM—2010-216827 1

Integration Methodology For Oil-Free Shaft Support Systems: 

Four Steps to Success

Samuel A. Howard, Christopher DellaCorte, and Robert J. Bruckner

National Aeronautics and Space Administration

Glenn Research Center

Cleveland, Ohio 44135

Abstract

Commercial applications for Oil-Free turbomachinery are 

slowly becoming a reality. Micro-turbine generators, highspeed electric motors, and electrically driven centrifugal 

blowers are a few examples of products available in today’s 

commercial marketplace. Gas foil bearing technology makes 

most of these applications possible.

A significant volume of component level research has led to 

recent acceptance of gas foil bearings in several specialized 

applications, including those mentioned above. Component 

tests identifying such characteristics as load carrying capacity, 

power loss, thermal behavior, rotordynamic coefficients, etc. 

all help the engineer design foil bearing machines, but the 

development process can be just as important. As the 

technology gains momentum and acceptance in a wider array 

of machinery, the complexity and variety of applications will 

grow beyond the current class of machines. Following a robust 

integration methodology will help improve the probability of 

successful development of future Oil-Free turbomachinery.

This paper describes a previously successful four-step 

integration methodology used in the development of several 

Oil-Free turbomachines. Proper application of the methods put 

forward here enable successful design of Oil-Free 

turbomachinery. In addition when significant design changes 

or unique machinery are developed, this four-step process 

must be considered.

Nomenclature

Cαα damping (kNs/m, α = matrix index)

D shaft diameter (m)

Kαα stiffness (MN/m, α = matrix index)

L bearing axial length (m)

S′ modified Sommerfeld number in air

( )

DL

W

D D

S

ω

′ =

Wapplied applied bearing load (N)

WLC load capacity at speed (N)

D journal bearing performance coefficient (N/m3

/krpm)

ω shaft speed (krpm)

ωlo loft-off speed (krpm)

ωmin minimum speed (krpm)

Introduction

In the realm of traditional turbomachinery, there is a wealth 

of historical, analytical, and experimental experience to bring

to bear when designing a new machine. Many “new”

turbomachines with conventional bearing technology are 

really improvements of existing designs involving small, 

incremental, and evolutionary changes. As a result, often the 

approach is to use a design that has worked well in the past, or 

to slightly modify an existing design to enhance its 

capabilities. As long as the new application does not extend 

beyond the bounds of prior art, design guidelines and tools can 

be used with success. With Oil-Free turbomachinery, this is 

typically not the case. Usually, these machines are unique with 

little historical knowledge from which to draw. In addition, 

there is less analytical and experimental data in the literature 

than with conventional bearings. Thus, when developing new 

foil bearing turbomachines, it stands to reason that risk 

minimization would necessitate a well-defined method of 

analysis combined with experimental validation.

That necessity is at the heart of the integration methodology 

outlined in this paper. Experience has taught that all

development programs negotiate through a similar process, but 

the most successful plan for it from the outset. The basic 

methodology hinges on using analysis to conduct preliminary 

design and layout trade studies followed by experimental 

verification of the resulting design with iteration possible 

anywhere along the way. Formally, the process breaks down 

into four distinct steps: 1) Preliminary feasibility and design 

layout, 2) Component level bearing testing, 3) System level 

rotor simulator testing, 4) Final design full-scale demonstration.

While success can never be guaranteed, the four-step 

process maximizes its probability and helps identify 

applications that have high risk of failure at the full-scale 

level. Many proprietary examples exist of attempts to develop 

new turbomachines without following all four steps, only to 

result in failure that could have been avoided. This paper 

outlines the four steps of the development process with the 

goal of showing how they can be utilized when developing

hardware.

Foil Gas Bearings

The four-step integration methodology can be readily 

applied to any development program, but it is especially 
NASA/TM—2010-216827 2

important when a new application is outside the previous 

design space. This is especially true in the development of OilFree turbomachinery. Since many Oil-Free applications utilize 

gas foil bearings, an introductory discussion on their operation 

and construction follows.

Gas foil bearings (Fig. 1) are gas lubricated, hydrodynamic 

bearings with a compliant inner sleeve surface. The compliant 

surface is usually comprised of several layers of sheet metal, 

called foils. One or more layers of the sheet metal foil are

formed into bumps (bump-type foil bearing) or other structural 

element (overlapping leaf, etc) that behaves as a spring. The 

compliant layer is covered by one or more sheets of flat metal 

foil, called the top foil. The top foil acts as the bearing inner 

surface, against which the hydrodynamic pressure acts to 

support the shaft weight.

When the rotor spins, pressure is developed in the gas film

between the shaft and top foil. As the speed increases, the 

pressure increases, and eventually equals that of the weight of 

the rotor, suspending the rotor on a film of air. Once the rotor 

is suspended, there is nominally no contact between the rotor 

and the top foil surface.

The foil bearing operating principle just described makes

clear some of the advantages of foil bearings over other types 

of bearings. The advantages of gas foil bearings make them 

attractive for application to gas turbine engines and other 

turbomachinery (Ref. 1). Owing to the low viscosity of air 

relative to other lubricants, the torque (or power loss) in gas 

foil bearings is low compared to most other bearing types. The 

gas film does not degrade at temperatures up to the 

temperature limits of the bearing materials, as liquid lubricants 

do, resulting in high temperature capability. Once the bearing 

is lifted off, there is no wear giving foil bearings long life and 

minimal maintenance. The compliance enables the foil bearing 

to expand and conform to the pressure generated in the 

lubricating film. While rigid gas bearings can be vulnerable to 

seizure due to thermal and centrifugal distortion and growth, 

the compliance of gas foil bearings allows them to tolerate 

shock loads, misalignment and thermal and centrifugal growth

(Ref. 2). In addition, the performance characteristics of gas 

foil bearings can be tuned by varying the different geometry 

parameters associated with the foil structure, such as bump 

pitch, foil thickness, etc (Refs. 3 and 4).

As rotor support elements in turbomachinery, foil bearings 

also have a few limitations in comparison to other bearing 

types. The low viscosity and compressibility of the gas 

lubricant results in low load capacity compared to liquidlubricant hydrodynamic bearings and rolling element bearings 

of similar size. However, the load capacity does increase with 

a linear relationship to speed (to a point), resulting in useful 

load capacities at high speed. Therefore, foil bearings are 

generally well-suited to high-speed, lightly-loaded 

turbomachines (Ref. 5). In addition, the damping due to 

viscous energy dissipation is low compared to oil-lubricated 

bearings or squeeze film dampers. However, there is evidence 

that frictional damping in the foils adds stability to foil bearing 

systems, and typically low cross-coupled stiffness results in 

less need for high damping (Ref. 6). These unique 

characteristics of gas foil bearings increase the importance of 

exercising a well-defined development methodology when 

designing new machinery.

Four-Step Integration Methodology:

Introduction and Definition

The four-step method for integrated design of a new 

turbomachinery system is outlined in flowchart form in 

Figure 2. The flowchart is intended to help visualize all the 

process steps, and the potential for multiple iterations within 

and across all four steps.

The first step, preliminary feasibility and design layout, is 

to determine if a proposed application is plausible from the 

standpoint of bearing performance and system rotordynamics.

During this step, trade studies of various rotor configurations

are performed to determine bearing loads and sizing, bearing 

positions (straddle mount, overhung, etc.), calculation of 

critical speeds, and any other basic information needed to 

down-select a rotor configuration for further consideration.

In step two, component level testing, the bearings chosen 

in step one are fabricated and subjected to a battery of tests to

verify their capability to perform as needed in the system.

Examples of tests include load capacity, power loss, start-stop 

durability, stiffness and damping, thermal management, etc. 

Step three is a rotordynamic simulation intended to ensure 

that the system designed in steps one and two performs as 

expected. This is the step that usually gets skipped to the 

detriment of a development program. Rotordynamic 

simulation is an important step because it helps to identify 

system level integration issues that cannot be vetted in steps 

one and two. It has been the authors’ experience that 

bypassing this step was the cause of many failures experienced 

in development programs that went from design straight to 
NASA/TM—2010-216827 3

full-scale demonstration. The authors have witnessed program 

cancellations due to failures of demo hardware that could have 

been avoided by operating a simulated rotor system. It seems 

as if there is generally less tolerance of failures when real 

hardware is involved compare to simulated test hardware. Step 

three is intended to identify and correct any of these potential 

unknown issues in a low-cost, simplified test rig before 

developing in a full-scale demonstration unit.

Step three is implemented by building a rotor system that 

matches the layout of the actual machine, but with simulated 

aerodynamic components. Building a rotor with nonfunctional turbine and compressor components affords an 

opportunity to test the rotor system without the complications 

of the thermodynamic process. The benefits of simulated rotor 

components include easy and low-cost production and testing,

if failures occur they are usually non-critical and less costly to 
NASA/TM—2010-216827 4

repair, changes can be made faster and at less cost, and more.

Additionally, the rotordynamic behavior of the simulated 

system can be checked against the predicted behavior and 

changes can be made to optimize the system before any final 

designs are made. If warranted, the simulated rotor system can 

be intentionally pushed to failure to identify failure modes and 

true limits of operation. 

The last step is full-scale demonstration. This step involves 

building a highly instrumented, production-like machine and 

running it through a series of qualification tests. By the time a 

development program gets to the fourth step most, if not all, of 

the design changes should be complete, leaving only to verify 

that the machine performance meets the design criteria. Small 

changes may still be needed to realize the desired operation of 

the new machine in the final step, but most of the integration 

issues should have been identified and corrected. In the

following sections, a simple example of each of the first three 

steps is given. 

Four-Step Integration Method: Example 

Application—Simulated Turbocompressor

An existing test rig that simulates a small turbocompressor 

is used as the surrogate turbomachine in this example in order 

to model realistic hardware. The first three steps are illustrated 

with sample calculations to demonstrate how one might utilize 

the four-step method to develop a piece of hardware. The 

fourth step, demonstration of the final design, is self-evident 

and only applies to a production machine, and therefore is not 

included here. It is hoped by navigating through the first three 

steps in this example readers will understand how this method 

can be used to methodically progress through a design process 

to minimize the risk involved in developing new hardware.

Also, the example provides an opportunity to demonstrate an 

array of design tools for gas foil bearings that have emerged 

from research in the past decade and how one can utilize these 

tools in practice.

The first objective of step one is to size the bearings. Many 

factors play a role in dictating the size of the shafting, and 

therefore the bearing size in turbomachines. Often times, the 

desired function of the machine dictates much of the design. For 

example, the process fluid, flow rate, pressures, temperatures, 

power, etc. are all considered in designing the aerodynamic 

components. Once the aero components are roughly sized for 

the desired performance, a rough estimate of the mass of the 

machine is known. Additionally, the aero components generally 

dictate the rotational speed and the power transmitted through 

the shaft. With that information, one can begin to characterize 

the design space. For this discussion, it is assumed the aero 

components are designed a priori to do a specific function, and 

are not part of the four-step integration effort. If, however, that 

is not the case for a particular application, the aero component 

design can be incorporated into the four-step process and can 

evolve iteratively along with the rotor and bearing designs. For 

instance, if the speed required by the compressor is too low, one 

could use a smaller wheel at higher speed. Once the weight and 

torque constraints are known, the bearing size can be selected.

For gas foil bearings, the correct size is based on design 

criteria related to load capacity and power loss. First, it is 

necessary to ensure the load capacity of the bearings is greater 

than the respective load on each of the bearings. This can be 

accomplished utilizing the foil bearing load capacity rule-ofthumb (ROT) developed by DellaCorte and Valco (Ref. 7).

The ROT states that a foil bearing can support a load 

proportional to its projected area multiplied by its surface 

velocity. Specifically, Equation (1) gives the empirically 

derived relationship.

WLC = D × (L × D)× (D ×ω) = D × LD2ω (1)

where D, the performance coefficient, depends upon the 

bearing design and other parameters such as process fluid, 

temperature, ambient pressure, etc.

For our simple example, assume a system study predicted a 

rotor weight of about 67.2 N, with a desired speed range of 

10,000 to 30,000 rpm. The load capacity ROT is used to pick a 

bearing size as follows:

For simplicity, the rotor is assumed to be symmetric with half 

the weight supported by each of two journal foil air bearings.

The best available foil bearing technology will be used,

meaning the performance coefficient is around 2.71×105

N/m3

/krpm (1.0 lb/in.

3

/krpm) (Ref. 7). At this stage of the 

design, it is reasonable to choose an L/D ratio for the bearings of 

1. The load capacity ROT predicts that the minimum bearing 

size needed to support the weight of the rotor is about 23-mm 

diameter by 23-mm long (0.90- by 0.90-in.) using 10,000 rpm 

(ω = 10 krpm) as the minimum speed.

However, in addition to providing enough load capacity, the 

bearings should operate in a low power loss state as defined by 

the performance map first explained in DellaCorte, et.al. 

(Ref. 8), and later expanded in Radil and DellaCorte (Ref. 9). 

A sample performance map is shown in Figure 3. 

The performance map plots power loss versus a modified 

Sommerfeld number (speed/load parameter), or versus load 

and speed in three dimensions, and makes clear two distinct 

operating regimes. As shown in Figure 3, there is a point in 

the operating space of a foil bearing where the power loss is at 

a minimum. To the right of the minimum power loss (higher

Sommerfeld number = higher speed and/or lower load), the 

bearing operates in a lightly loaded, thick film regime. To the 

left of the minimum (lower speed and/or higher load), the 

bearing is operating in a highly loaded, thin film regime. It is 

desired to be to the right of the minimum power loss, in the 

thick film regime because this region of the operating 

envelope is stable, meaning that small changes in running 

conditions correct themselves in time. Operating near the 

minimum power loss, or to the left of it is thermally unstable 

because small perturbations tend to grow until failure is 

experienced. A more detailed discussion is given in (Ref. 9). 
NASA/TM—2010-216827 5

Without having a bearing to test, which is likely to be the 

case in step one, it is not possible to generate a performance 

map, and therefore, it is not possible to know where the 

minimum power loss occurs. In the absence of the power loss 

map, the load capacity ROT can be used to estimate the 

minimum power loss speed for a given load. A loose 

correlation has been experimentally observed that the power 

loss minimum occurs at roughly 2 to 6 times the lift-off speed 

for a given load. The load capacity ROT can be used in 

reverse to estimate a lift-off speed for a known load because 

the lift-off speed is the speed at which the applied load is the 

load capacity. Rearranging Equation (1) gives:

2

applied

DLD

W

ωLO = . (2)

Rearranging to solve for the bearing size, and making the 

substitutions: D = L (from L/D = 1), and D =2.71×105

, gives:

LO

W

D × ω = 5

applied 3

2.71 10 . (3)

In order to satisfy the power loss requirement of 2 to 6 times 

the lift-off speed as a minimum operating speed, use:

( ) min 6 to 1

2 ωLO = 1 ω . (4)

Then, for the above example with Wapplied= 33.6 N per bearing 

(7.55 lb), ωLO = 1/2 to 1/6 (10 krpm) = 5.0 to 1.67 krpm, the 

approximate desired minimum bearing size ranges from 

29 mm (~1.2 in.) to 42 mm (~1.7 in.) diameter. A designer 

might use his experience to determine a bearing size based on 

this range. Many times, a trade-off between competing design 

criteria will feed into this decision. For example, one may 

choose to pick a bearing at the large end of the range to ensure 

operation to the right of the power loss minimum. Alternately, 

there might be a maximum size limitation for the given 

application, or a minimum rotor diameter to transmit the 

necessary torque. The usefulness of the ROT is to provide a 

reasonable range for a starting point. The example test rig 

utilizes 50.8-mm diameter by 40.6-mm long (2.0- by 1.6-in.) 

bearings. This bearing size is well above the minimum for 

load capacity, and also above the minimum acceptable range 

of diameter based on power loss. The L/D ratio < 1 was a 

matter of availability of bearings. A 50.8- by 40.6-mm bearing 

was available as a standard size from the bearing supplier. 

Using the ROT, the 50.8- by 40.6-mm bearing yields a lift-off 

speed of about 1200 rpm. Thus, the bearings may be oversized 

for the application. Regardless, the bearing design chosen in 

this manner then becomes the preliminary input to the 

feasibility study. If the choice does not satisfy the design 

criteria as the process progresses, it can be modified in a 

future iteration.

Once a tentative bearing size is chosen, dynamic 

coefficients can be predicted using one of the techniques 

outlined in various papers (Refs. 10 and 11), or one’s own inhouse analysis techniques. Then, a rotordynamic analysis is

done to assess the feasibility of the proposed design and 

generate a rotor layout. The rotordynamic layout trade study 

with the bearing characteristics for the tentative bearing size, 

and an idea of what the rotor will look like, is used to predict 

the system critical speeds, stability threshold, unbalance 

response, etc. to ensure the design criteria are met. If not, the 

layout or the bearing size might be modified and another 

iteration performed until the rotor behaves as desired.

Using the test rig as an example, the 50.8- by 40.6-mm 

bearings are analyzed using the technique developed by 

San Andres and Kim (Ref. 10) to determine dynamic 

coefficients. The resulting bearing coefficients are given in 

Table 1.

TABLE 1.—STIFFNESS AND DAMPING COEFFICIENTS 

FOR 50.8- BY 40.6-mm FOIL BEARING. 

[Evaluated at frequency = synchronous 

speed,static load = 33.6 N (in Y direction)]

Speed

(rpm)

Kxx

(MN/m)

Kxy Kyx Kyy Cxx

(kNs/m)

Cxy Cyx Cyy

5000 2.5 –1.6 –2.5 5.9 2.6 –2.5 0.63 5.9

10000 2.8 –1.1 –2.0 4.8 1.8 –1.2 0.62 2.6

15000 3.0 –0.78 –1.6 4.4 1.4 –0.75 0.56 1.6

20000 3.3 –0.54 –1.3 4.2 1.1 –0.55 0.51 1.2

25000 3.6 –0.38 –1.0 4.2 1.0 –0.45 0.45 0.91

30000 3.8 –0.26 –0.80 4.2 0.86 –0.38 0.39 0.76

Using the bearing data from an analysis code, one can 

conduct a rotordynamic trade study and system layout. The 

process can involve iteration back and forth between bearing 

design and rotor layout in the search for a feasible design.

Since the test rig already exists, this step is used to illustrate 

what one iteration might look like in a design program. The 

rotor geometry is modeled in a rotordynamic analysis package 
NASA/TM—2010-216827 6

using the bearing data from Table 1. A sketch of the test rig is 

shown in Figure 4. The primary result of step one of the 

integration method is a bearing design and rotor layout that 

can be carried forward into the later steps for testing. As such, 

the rotordynamic analysis of the system is compared to the 

desired operating envelope to determine if the system will 

behave acceptably.

As an example, the test rig model in Figure 5 was analyzed, 

and its unbalance response is shown in Figure 6. As seen in 

Figure 6, there are predicted resonances at around 5,000 rpm 

and 12,000 rpm. Since the desired speed range is 10,000 rpm to 

30,000 rpm, another iteration of bearing design and/or rotor 

layout could be performed to try to move the 12,000 rpm 

resonance out of the operating range. In this case, the test rig is 

already built, but this illustrates how the iterative nature of the 

four-step process can be beneficial. In this case, it could have 

been used to improve the frequency response of the test rig. 

In a similar manner, the rotordynamic model can be used to 

predict stability, and the design can evolve on paper until both 

the unbalance response and the stability are acceptable. 
NASA/TM—2010-216827 7

In step two of the integration methodology, the bearing 

design chosen in step one is tested to verify the desired 

performance is achieved. As seen in the flowchart, the first 

action of step two is to fabricate or procure the bearings 

designed in step one. Next, the bearings are tested for 

acceptable performance. The testing performed in step two can 

depend upon the application, but may include measurements 

of load capacity, power loss, dynamic coefficients, thermal 

behavior, wear, etc. As an example, a power loss plot for the 

bearings used in the test rig is shown in Figure 7. The test was 

done for a fixed load, which means the only variable in the 

modified Sommerfeld number is speed. Thus, plotting with 

speed on the abscissa is qualitatively the same, and it makes it 

easier to see that for speeds between 10,000 and 30,000 rpm, 

the test rig will indeed be operating to the right of the power 

loss minimum, i.e., the lightly loaded regime. 

Any number of other component level tests can be done to 

determine that the bearings will behave as designed. If they do 

not, sometimes modifications are identified during testing, and 

bearings can be altered and retested. Otherwise, the designer 

can back up and re-iterate through step one to change the 

bearing or rotor design to correct for shortcomings uncovered 

in step two.

In step three, rotordynamic simulation testing serves to 

verify that a rotor/bearing system with the same bearings and 

mass properties as the desired turbomachine will behave as 

designed. A rotor with matching mass, geometry, and 

gyroscopic properties is built and tested using the bearing 

design chosen in step one and verified in step two. Here, the 

predicted behavior from the feasibility study can be verified, 

and the rotor system can be tested for overall desirable 

dynamic response. Other integration issues such as build 

techniques and procedures can be tested out as well as the 

fitment of various components. Lingering questions about any 

design or build-up issues can be answered in step three, such 

as shaft interference fits, bearing mounting (alignment), 

material pairs, etc. Once the design successfully completes 

step three there should be little doubt that the full-scale demo 

will be successful. Again, like the other steps, step three 

affords an opportunity to identify and make changes to the 

bearings or rotor and re-test, or major design flaws can be 

worked out by stepping back into the previous steps and reiterating.

Again referring back to the test rig, an example of the types 

of tests one can perform in step three is a comparison of the 

unbalance response to the predicted response in step one.

Figure 8 shows the measured unbalance response curves from 

the same bearing location on the test rig. In this example case, 

one can see that a reasonable prediction was made during step 

one. This also illustrates that foil bearing design tools are 

currently becoming quite capable for design of 

turbomachinery. 

Step three is followed by fabrication of a full-scale 

demonstration unit. As with the previous steps, this can be an 

iterative process to fine tune the design, or revert back to a 

previous step if major issues surface. Once successful, the 

demonstration unit can lead into a production turbomachine.

Summary Remarks 

As Oil-Free turbomachinery applications gain momentum 

in the marketplace, more new machines will be developed to 

take advantage of the unique capabilities of gas foil bearings.

A four-step integration methodology is outlined in this paper 

that, if followed during development, will increase the 

probability of successful design of new turbomachines. A 

structured approach to development such as this is especially 

important for applications that are a significant departure from 

machines that currently exist. Many times, development 

efforts will skip one or more of the steps outlined here and 
NASA/TM—2010-216827 8

rush to prototype demonstration only to fail prematurely.

Following all the steps provides opportunities to gauge the 

performance of the design and make changes along the way to 

maximize the chance of success without expensive hardware 

failures.

Several examples are given to illustrate the types of testing 

and analysis that can be done to feed into the design and to 

measure the chances of successful development. While the 

examples and descriptions given are based on gas foil bearing 

development, the process can be applied to any development 

program to enhance the potential for success.

References

1. Klaass, R.M., and DellaCorte, C., “The Quest for Oil-Free 

Turbine Engines,” Report number 2006–01–3055, (2006), SAE 

Technical Papers, Warrandale, PA.

2. Howard, S.A., “Misalignment in Gas Foil Journal Bearings: An 

Experimental Study,” Journal of Engineering for Gas Turbines 

and Power, Vol. 131, March 2009.

3. Ku, C.P.R, Heshmat, H., “Compliant Foil Bearing Structural 

Stiffness Analysis: Part I-Theoretical Model Including Strip and 

Variable Bump Foil Geometry,” Journal of Tribology, Vol. 114,

April 1992, pp. 394–400.

4. Ku, C.P.R, Heshmat, H., “Compliant Foil Bearing Structural 

Stiffness Analysis: Part II-Experimental Investigation,” Journal 

of Tribology, Vol. 115, July 1993, pp. 364–370.

5. Heshmat, H., “Advancements in the Performance of 

Aerodynamic Foil Journal Bearings: High Speed and Load 

Capability,” Journal of Tribology, Vol. 116, April 2004, 

pp. 287–295.

6. Rubio, D., and San Andres, L., “Bump-Type Foil Bearing 

Structural Stiffness: Experiments and Predictions,” Journal of 

Engineering for Gas Turbines and Power, Vol. 128, 2006, pp. 

653–660.

7. DellaCorte, Christopher and Valco, Mark J., “Load Capacity 

Estimation of Foil Air Journal Bearings for Oil-Free 

Turbomachinery Applications,” Tribology Transactions, 

Vol. 43, No. 4, 2000, pp. 795–801.

8. DellaCorte, C., Radil, K.C., Bruckner, and Howard, S.A., 

“A Preliminary Foil Gas Bearing Performance Map,”

NASA/TM—2006-214343, 2006, NASA Glenn Research 

Center, Cleveland, OH. 

9. Radil, K.C., and DellaCorte, C., “A Three-Dimensional Foil 

Bearing Performance Map Applied to Oil-Free 

Turbomachinery,” ARL–TR–4473, 2009, Army Research 

Laboratory, Cleveland, Ohio.

10. San Andrés, L., and Kim, T.H., “Thermohydrodynamic Analysis 

of Bump Type Gas Foil Bearings: A Model Anchored to Test 

Data,” Journal of Engineering for Gas Turbines and Power, InPress (ASME Paper No. GT2009–59919).

11. Carpino, M, and Talmadge, G., “Prediction of Rotor Dynamic 

Coefficients in Gas Lubricated Foil Journal Bearings with 

Corrugated Sub-Foils,” Tribology Transactions, Vol. 49, 2006, 

pp. 400–409.
REPORT DOCUMENTATION PAGE Form Approved 

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01-11-2010 

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Technical Memorandum 

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4. TITLE AND SUBTITLE

Integration Methodology For Oil-Free Shaft Support Systems: Four Steps to Success 

5a. CONTRACT NUMBER

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6. AUTHOR(S)

Howard, Samuel, A.; DellaCorte, Christopher; Bruckner, Robert, J. 

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WBS 877868.02.07.03.01.01.04 

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13. SUPPLEMENTARY NOTES

14. ABSTRACT

Commercial applications for Oil-Free turbomachinery are slowly becoming a reality. Micro-turbine generators, high-speed electric motors, 

and electrically driven centrifugal blowers are a few examples of products available in today’s commercial marketplace. Gas foil bearing 

technology makes most of these applications possible. A significant volume of component level research has led to recent acceptance of gas 

foil bearings in several specialized applications, including those mentioned above. Component tests identifying such characteristics as load 

carrying capacity, power loss, thermal behavior, rotordynamic coefficients, etc. all help the engineer design foil bearing machines, but the 

development process can be just as important. As the technology gains momentum and acceptance in a wider array of machinery, the 

complexity and variety of applications will grow beyond the current class of machines. Following a robust integration methodology will 

help improve the probability of successful development of future Oil-Free turbomachinery. This paper describes a previously successful 

four-step integration methodology used in the development of several Oil-Free turbomachines. Proper application of the methods put 

forward here enable successful design of Oil-Free turbomachinery. In addition when significant design changes or unique machinery are 

developed, this four-step process must be considered. 

15. SUBJECT TERMS

Bearings; Foil bearings; Gas bearings; Turbomachinery; Rotor dynamics 

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