Christopher DellaCorte

Glenn Research Center, Cleveland, Ohio

Oil-Free Shaft Support System Rotordynamics: Past, 

Present, and Future Challenges and Opportunities

NASA/TM—2011-217003

April 2011
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Christopher DellaCorte

Glenn Research Center, Cleveland, Ohio

Oil-Free Shaft Support System Rotordynamics: Past, 

Present, and Future Challenges and Opportunities

NASA/TM—2011-217003

April 2011

National Aeronautics and

Space Administration

Glenn Research Center

Cleveland, Ohio 44135

Prepared for the

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Seoul, Korea, September 12–15, 2010
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NASA/TM—2011-217003 1 

Oil-Free Shaft Support System Rotordynamics: Past, Present, and 

Future Challenges and Opportunities 

Christopher DellaCorte 

National Aeronautics and Space Administration 

Glenn Research Center, 

Cleveland, Ohio 44135 

Abstract 

Recent breakthroughs in Oil-Free technologies have enabled new high-speed rotor systems and 

turbomachinery. Such technologies can include compliant-surface gas bearings, magnetic bearings, and 

advanced solid lubricants and tribo-materials. This presentation briefly reviews critical technology 

developments and the current state-of-the-art, emerging Oil-Free rotor systems and discusses obstacles 

preventing more widespread use. Key examples of “best practices” for deploying Oil-Free technologies 

will be presented and remaining major technical questions surrounding Oil-Free technologies will be 

brought forward. 

Nomenclature 

DN bearing surface velocity parameter 

W shaft load 

 fluid viscosity 

WLC load capacity at speed 

D journal bearing performance coefficient (lbs/in.3

/krpm) 

Dt thrust Bearing performance coefficient (lbs/in.3

/krpm)

L bearing axial length (inches) 

D shaft diameter (inches) 

Dm thrust runner median diameter (inches) 

w thrust foil radial extent, width (inches) 

Krpm shaft speed in thousands of rpm 

Introduction 

Oil-lubricated rotor support technologies, based upon fluid film and rolling element type bearings, 

have been an industrial mainstay for centuries and have served society well. Their reliability and high 

performance with respect to extreme load capacity and their demonstrated long life have enabled them to 

appropriately dominate most mechanical systems. Ubiquitous examples include: internal combustion 

engines, power plant turbines, fluid compressors, and electric motors. Decades of productive research, 

development and experience have resulted in well-understood and accepted rules for successful 

application and a myriad of specific designs of these machine elements for particularly challenging 

applications (Ref. 1). This positive experience with conventional rotor support technologies has also had a 

secondary effect; the diminished awareness of the vital role rotor support technologies play in the overall 

success and performance of rotating systems. Such a lack of understanding for and appreciation of 

conventional rotor support technologies is to be expected and to a limited extent acceptable provided that 

new machines do not push bearing technologies beyond the well established and proven design space. 

When new machines, however, go beyond the norm with respect to rotor speed, temperature or other 

factors, closer attention must be paid to the rotor support system including the consideration of alternate 

bearing technologies. In this paper, Oil-Free rotor support technologies are examined within a framework 
NASA/TM—2011-217003 2 

of their ability to enable the design and deployment of new machines with expanded capabilities. Recent 

advances, particularly for compliant-surface foil-gas bearings, high temperature solid lubricants and 

hybrid electromagnetic-gas bearings, are reviewed and a recommended technical path is given for risk 

mitigation in the development of new machines. Lastly, the remaining technical challenges facing the 

deployment of Oil-Free technologies are examined. 

Foil Gas Bearings-Operating Principles 

Compliant-surface foil-gas bearings are a class of hydrodynamic bearings that use the ambient gas as 

their working fluid and thus require no dedicated lubrication system. Their hydrodynamic fluid film is 

formed between the moving shaft surface and a flexible bearing surface typically made from several 

layers of sheet metal foils. Figure 1 depicts the common bump type of foil bearing. 

Foil bearings can be configured as journal or thrust bearings supporting radial or axial shaft loads, 

respectively as shown in Figure 2. 

Figure1.—Cross section view of a simple bump type radial foil bearing. 

Figure 2.—Journal and thrust foil bearings used to 

control radial and axial shaft motion, respectively. 
NASA/TM—2011-217003 3 

The dominant characteristic of foil-gas bearings is that they float on a self-generated fluid film during 

normal operation but experience a brief period of dry sliding contact during low speed operation (i.e., 

start-up, shut-down). The bearing geometry and fluid film thickness depend upon the establishment of 

equilibrium between the gas film pressure and the elastic deflection of the foil surface and its underlying 

spring support structure (Ref. 2). During operation, dynamic rotor motions cause fluctuations in the film 

pressure and commensurate, albeit minute, motions in the foils. This causes internal sliding contact within 

the bearing structure and contributes a coulomb damping effect to the system (Ref. 3). Thus, in a sense, a 

foil bearing is both a bearing and a damper. The compliant nature of the bearing also allows it to respond 

to and accommodate thermal and mechanical distortions much more so than traditional rigid gas bearings. 

These salient features of foil-gas bearings have long been known. 

In a 1969 review of gas bearings, W.R. Gross wrote “Foil bearings were evolved to minimize 

instability problems, reduce manufacturing tolerances, and permit adaptation of the bearing to changes in 

shaft diameter caused by centrifugal force or temperature gradients. The fluid film between the shaft and 

foil is likely to have a greater stiffness than the foil itself” (Ref. 4). The subsequent four decades of 

research and development in foil bearings has largely reinforced Gross’ summary and must be kept first 

and foremost in mind when integrating foil bearings into new rotor systems. 

With this as a backdrop, it is apparent that foil bearing load capacity and dynamic properties are 

governed largely by the low viscosity nature of the lubricant. This brings both advantages and challenges 

to systems built using such rotor supports. The primary advantage of gas bearings is the simplification 

possible through the elimination of the oil system combined with the ability to operate bearings at higher 

speeds and temperatures. The primary challenge is to develop innovative machine designs that can take 

advantage of the performance characteristics of foil bearings while accommodating their performance 

shortcomings (Ref. 5). 

Foil bearings provide very modest load capacity, stiffness and damping compared to conventionally 

supported systems. One can expect stiffness and damping to be an order-of-magnitude lower than 

comparably sized oil-lubricated bearings. On the other hand, friction can be lower especially at high 

rotational speeds and foil bearings have no intrinsic DN speed limitations, as do rolling element type 

bearings. Foil bearing load capacity is heavily influenced by speed as well. At high speeds, foil-gas 

bearings exhibit comparable or even higher load capacity than rolling-element bearings but have limited 

capability at low speeds (Ref. 6). 

These characteristics of foil-gas bearings dictate that their successful application occurs when a rotor 

system is designed around the bearing capabilities in contrast to the now common practice of first 

designing the aero-components and determining speeds and loads from which the rotor design and 

bearing requirements follow. Later in this paper, a recommended four-step rotor system development path 

will be described which has been effective in adopting Oil-Free technologies for new high speed 

machines (Ref. 7). 

Foil-Gas Bearings-Solid Lubricants 

Foil bearings experience brief sliding contact during start-up and shutdown when surface velocity and 

hydrodynamic gas film pressure is insufficient to promote a fully developed lubricant film. Solid 

lubricants are needed to mitigate wear of the foils, reduce friction and maintain smooth surfaces required 

by gas bearings. Solid lubricants also provide a secondary function as a sacrificial, non-structural material 

that enables the development of conforming surfaces through the wear-in process (Ref. 8). This 

characteristic is vitally important for bearings with non-optimal geometries or undesirable manufacturing 

variances. Towards this end, soft solid lubricants are typically applied to the foil surface and hard wearresistant coatings are used on shaft surfaces. An alternative approach employs composite shaft coatings 

that are simultaneously reservoirs for solid lubricants and offer wear protection by also containing harder 

phases. Figure 3 depicts the system nature of foil and shaft coatings used for the solid lubrication of foilgas bearings. 
NASA/TM—2011-217003 4 

Figure 3.—System of solid lubricants is used for foil bearings analogous to the system used for conventional 

oil-lubricated bearings (Ref. 8). 

Foil coating development and performance is well documented in the literature and can be broadly 

classified depending upon temperature requirements of an application. For low temperatures up to 300 C 

thin polymer films made from PTFE (polytetrafluoroethylene), polyimide and polyamides are traditionally 

employed. These films are often doped with solid lubricants such as graphite, molybdenum disulfide and 

boron nitride to further reduce friction and enhance life. For temperatures above 300 C, foil coatings can 

either be strictly hard wear-resistant ceramics such as aluminium oxide and titanium carbide or metalceramic-solid lubricant composites deposited by spraying, dipping or through electrochemical techniques 

followed by high temperature curing (Ref. 9). Such films are often mechanically burnished (polished) to 

achieve more uniform thickness and smooth surface finishes. Whatever techniques and materials used, the 

polymer coatings yield low friction and long life at a low cost and are thus preferred whenever temperatures 

are not extreme. 

For good solid lubrication to occur, the foil’s mating surface must be engineered appropriately. When 

operating against polymer foil coatings, the mating surface is typically coated with chrome, either hard 

chrome or thin dense chrome. Chrome plating results in a smooth and durable finish that resists galling 

with the foil in the event of a foil coating wear through. When high temperatures are encountered, either 

bare or hard coated foils are operated against ceramic or composite shaft coatings like the NASA 

developed PS300 and PS400 plasma sprayed coatings. Such approaches offer moderate friction levels and 

the ability to operate foil bearings to 650 C or higher (Refs. 10 and 11). 

Foil-Gas Bearings-Performance Characteristics 

Experimental and analytical data combined with experience with operating Oil-Free turbomachinery 

has given rise to estimation rules for foil bearing performance. These “Rules-of-Thumb” (ROT) can be 

used as a starting point from which to develop a new machine design. The ROT for foil bearing load 

capacity was clearly established by comparing bearing structural design to measured performance and the 

following linear relationship between bearing size, speed, design and load capacity follows (Ref. 6): 
NASA/TM—2011-217003 5 

 WLC = D(LD)DKrpm (1) 

The symbols in Equation (1) are defined in the nomenclature section. A few specific examples follow to 

illustrate the current understanding of journal bearing scalability effects (Ref. 12). 

A 100 mm diameter (D = 1, L/D = 1) bearing operating at 20,000 rpm (2MDN) would be expected to 

carry about 1,280 lbs (~6 kN) maximum load. Bearings of this size have been produced and tested. 

Experimental data from the literature (Ref. 21) corroborates the load capacity ROT estimates for such 

bearings and is considered within current state-of-the-art. 

Extending the ROT approach, a bearing twice the diameter and twice the length (200 mm) would 

support eight times the load (10000 lbs, 4800 kN) and be operating at approximately the previously 

established DN speed limit (4 million). Based upon these static load (14 to 55 kPa) considerations alone, 

the 100 mm diameter journal bearings used in the above example (two per rotor) can support rotors from 

300 to 1200 N (64 to 256 lbs). Rotors of this size are typical for air compressors in the 75 to 400 kW 

(100 to 500 hp) range and small business class turbofan engines (1,000 lbs, 5000 N thrust levels). 

Thrust foil bearings ascribe to the same underlying physics as journal bearings and can be modeled in a 

similar fashion (Ref. 13). Using an analogous approach, the thrust bearing load capacity ROT is as follows: 

 WLC = Dt (wDm)DmKrpm (2) 

where the value w represents the radial pad extent (outer radius-inner radius) and the Dm represents the 

median bearing diameter. Though load capacity performance data is more limited than journal bearing data, 

when using Equation (2), one finds performance coefficients ranging from 0.02 to 0.13 or about 5 to 10 

times smaller than those for journal bearings. Such modest thrust bearing load capacity levels may be due to 

developmental immaturity or possibly due to thermal management challenges inherent in thrust bearings. 

One such challenge is that lubricating fluid ejected from a pad feeds directly into the following pad making 

bearing cooling difficult. It may also be that thermal and centrifugal distortion is more pronounced in thrust 

runners because their size is often larger than journal bearings. Whatever the reasons, thrust-bearing 

performance remains a characteristic that must be accommodated in machine designs. 

Though not as well established, similar rules (ROT’s) are being developed for bearing stiffness and 

damping. Foil-gas bearing stiffness and damping are characterized by non-linearity, frequency and 

temperature dependence and no well-accepted methods for measurement. Predictive tools are becoming 

available that enable dynamic property estimation based upon bearing structural design and operating 

parameters but these are in their early development. For preliminary machine design an experimentally 

based “first-order” approximation for stiffness and damping can be made by assuming that broad ranges 

for these parameters exist depending upon bearing loading and operating condition. 

Stiffness for foil bearings can be expected to be similar for all bearings when normalized for bearing 

size (projected area) because the gas film properties (hydrodynamic pressures) are similar. Hydrodynamic 

gas pressures are typically one to five atmospheres. To function properly, a foil bearing structure must be 

sufficiently deflected by such low pressures and this necessitates rather modest structural stiffness. In 

practice, this unit stiffness is around 6 MN/m for each square cm of projected bearing area (D*L). Lightly 

loaded bearings exhibit a reduction in this value of about 50 percent and heavily loaded bearings see this 

value increased by about 50 percent. Thus the range one can expect for stiffness is about 3 to 9 MN/m/cm2

(2500 to 7500 lbs/in./in.2

) (Ref. 14). Bearings operating near their load capacity limit may provide static and 

dynamic stiffness values one to two times higher than this range but such operation is not typical and can 

suffer from thermal instability and rapid failure (Ref. 15). 

Damping in foil bearings is due mostly to a sliding friction dissipation mechanism between foil layers 

and between the sleeve and foils where contact occurs. Since such coulomb damping is predicated on the 

low amplitude rubbing motion between compliant, lightly loaded components (foils) the dampening effect 

is modest compared to fluid film (oil) dampers. Further, the energy dissipation that results must be 

managed via conduction or convective cooling to maintain thermal equilibrium. For these reasons, the 

observed damping action of a foil gas bearing is typically one to two orders-of-magnitude lower than oillubricated sleeve bearings and squeeze film dampers but higher than a hard mounted ball bearing. 
NASA/TM—2011-217003 6 

Experimental values range from 100 to 1800 N-s/m/cm2 (0.5 to 10 lbs/in./in.2) with the most typical 

values being around 200 N-s/m/cm2

 (Ref. 16). Again, these values have been normalized by the bearing 

size (D*L) in cm2

. The large scatter of damping values is partly due to the lack of uniformity in the 

measurement technique and the wide range of support structures utilized by foil bearings (i.e., bump foils, 

lever springs, wire mesh, visco-elastic layers, etc.). 

Utilizing these rules, one can layout (design) a rotor system effectively sizing the bearings and 

determining general performance characteristics. Bearing load capacity margin, lift-off speed, shaft orbits 

and critical speed margins can then be roughly assessed. This type of first order analysis is considered to 

be the first step in a four-step process used for machine development as outlined below. 

The Four-Step Process 

The integration of Oil-Free foil air bearing technology into turbomachinery systems is best done via a 

four-step process consisting of: (1) Rotor layout (arrangement) and concept design feasibility assessment; 

(2) bearing design and performance verification testing; (3) simulated rotor system testing; and (4) 

turbomachinery system demonstration. This multi-step approach enables one to tailor the rotor system 

design to meet the needs of the machine (power levels, mass flow, pressure ratio, etc.) while staying 

within the limitations of the rotor support technologies (Refs. 7, 17, and 18). Since Oil-Free bearing 

technologies differ significantly in their characteristics compared to conventional bearings, they cannot be 

directly retrofitted into existing machines without an unacceptable risk of failure. 

The first step (Step 1) involves conceptual design of the turbomachinery system and development of 

layouts of the rotating group including identification of major components, bearing locations, and 

approximate bearing specifications. The feasibility of the concept is analytically assessed through 

comparison of the required bearing performance with the available foil bearing technology level. Also, 

there are opportunities for conceptual design iteration to adjust the design layout for system level 

improvements (Ref. 19). If the concept cannot be shown to be feasible with foil bearings then this step is 

the end of the process. Figure 4 shows a typical rotordynamics model prepared in Step 1 to assess system 

stability and feasibility during a design layout trade study. 

Once the feasibility of the system is established a foil bearing development and test verification 

(Step 2) is started. Foil bearings are sized and designed to meet the performance requirements established 

by the conceptual design (Ref. 14). The bearings are tested and evaluated as individual bearing 

components in laboratory test rigs to verify adequate load support at the required speed and temperature 

conditions. In addition to experimental verification of bearing performance, a number of computer-based 

tools are emerging for foil bearing performance prediction (Ref. 20). Though in their infancy, these tools 

are useful in exploring the bearing design space prior to bearing manufacture and testing. If the bearings 

cannot meet the performance requirements then either the turbomachinery system must be reconfigured or 

more development is needed. 

Figure 4.—Rotordynamics layout model used to assess 

system feasibility based upon critical speeds, required 

damping, bearing loads and orbits. 
NASA/TM—2011-217003 7 

Figure 5.—Foil bearing shaft sustained no damage 

even during extended operation below the lift-off 

speed under heavy load and good solid lubrication 

(Ref. 21). 

When extending Oil-Free technologies to larger rotors other challenges surface like starting heavilyloaded, large diameter foil bearings that may undergo extended low speed operation during start-up. Such 

issues are best addressed in (Step 2 of the four-step process) bearing characterization research. A good 

example of this Step 2 bearing characterization is captured in a recent paper on the initial starting 

considerations for foil bearings intended for rotorcraft engines (Ref. 21). In that work, a large (76 mm 

diameter, 89 mm long) journal bearing was subjected to a 270 N static load and slowly accelerated to 

15,000 rpm over a 30 sec period to simulate the starting of a rotorcraft engine. Despite concerns that such 

low speed operation, below and just above the lift-off speed, would physically damage the bearing 

surfaces, no problems were encountered. Extending the operation (10 min) at a constant speed (3000 rpm) 

below the lift-off speed to simulate engine windmilling did not result in damage as long as good solid 

lubrication was provided. Figure 5 shows the condition of the shaft following this testing. These results 

add valuable and relevant information to the system design process in that start torque and shaft 

acceleration requirements can be assessed prior to system demonstration. 

Following verification of individual bearing performance, the next step (Step 3) is a simulated rotor 

system test (Refs. 22 and 23). A simulated rotor is fabricated to incorporate representative masses and 

inertias of the actual rotating group aerodynamic components (compressors and turbines). The simulated 

rotor system is tested using the prototype foil air bearings to verify rotordynamic stability and 

performance. This step is also an opportunity to explore operational limits such as bearing misalignment 

and sensitivity to rotor imbalances in a controlled test environment. Figure 6 depicts a test rig used to 

experimentally examine shaft rotordynamics to support Step 3 activities. 

Finally, after completing concept feasibility, bearing testing, and rotor simulator testing, the Oil-Free 

turbomachinery system is ready for a system level technology demonstration (Step 4) of the actual 

hardware assembly. 

Using this process Oil-Free technology was successfully demonstrated in an Oil-Free turbocharger for 

a heavy-duty truck diesel engine (Refs. 24 and 25). The Oil-Free turbocharger demonstration project was 

triggered by the confluence of four factors: (1) the doubling of foil bearing load capacity via better 

structural design; (2) the emergence of high temperature solid lubricant coatings tailored for foil bearings; 

(3) the availability of foil bearing test rigs to verify bearing performance; and (4) having had established a 

four step hardware development process for risk mitigation. 
NASA/TM—2011-217003 8 

Figure 6.—Oil-Free rotordynamics shaft test rig used to verify rotor support technology prior to system level 

demonstration (Step 3) (Ref. 22). 

 
NASA/TM—2011-217003 9 

Figure 7.—Cutaway photograph of the NASA OilFree turbocharger (Ref. 25). 

NASA’s Oil-Free turbomachinery team lead the industry effort that combined the foil bearings of 

Mechanical Technology Incorporated (later the effort was transferred to Mohawk Innovative 

Technologies), the turbocharger technology of Schwitzer Corporation (now Borg-Warner Automotive), 

and the diesel technology of Caterpillar Corporation. Under this project, long-life foil bearings were 

demonstrated at temperatures over 650 C, and to speeds of 120,000 rpm. The Oil-Free turbocharger, 

shown in Figure 7, was demonstrated in early 1999 and well publicized through magazine articles, peerreviewed journal papers and conference proceedings (Ref. 25). Further, the successful outcome validated 

the use of the Four-Step Process for managing risk in deploying Oil-Free technologies to new machines. 

The pathfinder turbocharger spurred industrial development programs in the US and abroad. To help meet 

these program needs, reports have been written that document foil bearing design and manufacturing 

techniques (Refs. 13 and 14). Figure 8 captures the current state-of-the-art very well by detailing the first 

known production patent for an Oil-Free turbocharger (Ref. 26). Further turbocharger development in 

Korea has quantified that Oil-Free turbochargers exhibit a significant reduction in friction losses versus 

conventional machines, with up to 10 percent overall efficiency improvements (Ref. 27). The 

development of turbochargers signifies that many important fundamental milestones for Oil-Free 

turbomachinery have been realized. Foil bearing design tools, bearing manufacturing methods, and solid 

lubricant coatings with long-life at low and high temperatures have been proven. Practical commercial 

rotordynamic modeling tools and software packages exist that predict, reasonably well, the critical speeds, 

bearing properties and shaft dynamic behavior for foil bearing supported systems (Ref. 20). In addition, 

the successful commercialization or demonstration of large compressors (300 kW class), microturbines 

(30 to 200 kWe), cryogenic turbines and air cycle machines signals that the technology base for Oil-Free 

turbomachinery is strong. More development is needed to apply the technologies to new applications. 
NASA/TM—2011-217003 10 

Figure 8.—Production patent for an Oil-Free turbocharger (Ref. 26). 
NASA/TM—2011-217003 11 

Future Challenges and Directions 

As rotors become heavier and bearings become larger, several characteristics of foil bearings present 

challenges. Among these are high-start torque requirements, limited damping and limited low speed load 

capacity (Ref. 21). Recent work with large (75 mm diameter) journal bearings operating under high static 

loads (75 kPa) has shown that durable low friction coatings and careful minimization of spring preload 

forces can allow bearings to survive even low shaft acceleration during prolonged start-up periods. 

Notwithstanding, long-term applications of large foil bearings for heavy rotors could benefit from 

hydrostatic or electromagnetic load sharing, especially during low speed operation. 

Perhaps more important than load capacity and durability, gas bearings lack the rigidity of conventional 

bearings and provide orders-of-magnitude lower damping than oil-lubricated technologies. Emerging design 

rules-of- thumb for foil bearing stiffness and damping will certainly aid in designing systems that can 

accommodate technological shortcomings. Nonetheless, some high performance machine designs may not 

be possible without the additive benefits of hybridization of rotor support technologies. 

An early demonstration of such a hybrid approach included the side-by-side demonstration of a large 

(100 mm diameter) foil-magnetic bearing (Refs. 28 and 29). Such concepts have been patented (Refs. 30 

and 31) and recent demonstrations indicate that this technology marriage is more capable than either 

bearing technology alone. Hybridization leads to a “smart” bearing in which the rotor static weight loads 

can be relieved and active damping can be added via electromagnetics (Ref. 28). Figure 9 depicts the 

hybrid “smart” bearing approach. 

Thrust bearing development will also be needed to handle more challenging Oil-Free machine 

applications (Ref. 32). Thrust foil bearings operate under the same basic principles as journal bearings but 

have received generally less investigative and development effort. This is partly because many machines 

can be designed with minimal thrust load requirements via aerodynamic pressure balancing. Thus, foil 

thrust bearing performance has not greatly limited machine design as has journal bearing performance. 

Also, more emphasis has been placed on enhanced rotordynamics that are largely controlled by journal 

bearings not thrust bearings. It is perhaps for these reasons that thrust bearing load capacity lags that of 

journal bearings by approximately a factor of 5 to 10. Since designing machines in which axial shaft 

(thrust) loads are minimized can be an obstacle to efficiency and performance, improved thrust bearing 

technology remains a formidable challenge to the extension of Oil-Free Turbomachinery. 

Figure 9.—Early hybrid foil-magnetic bearing concept 

that nests the foil bearing inside electromagnetic coils 

to create a “smart” bearing (Ref. 30). 
NASA/TM—2011-217003 12 

Summary Remarks 

Oil-Free rotor support technology has progressed greatly in the five decades since the first compliant 

surface gas foil bearings were developed. While the reliance on thin gas films as their working fluid leads 

to modest performance characteristics, doing so also enables design of machines with higher speed and 

temperature capabilities and reduced or even eliminated rotor system maintenance requirements. 

To date, Oil-Free rotor support system technology has enabled a broad range of high-speed machinery 

from air cycle machines, compressors and blowers to turbochargers, microtubines and small gas turbine 

engines. A four-step process for rotor system design has been articulated and proven to mitigate 

development risk. Such a formalized development process is vital during the current stage of Oil-Free 

technology deployment. As the technology becomes better understood, it is expected that more conventional 

evolutionary extensions of existing rotor system designs will again become the normal practice. 

Lastly, it is expected that future systems will involve more complex technology approaches. Hybrid 

bearing approaches utilizing external pressurization, active thrust load management and electromagnetic 

devices will evolve to meet the needs for higher performance and larger machinery. Until that time, the 

current technology will be implemented in ever increasing corners of the high-speed machinery market. 

References 

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Performance Map,” Presented at the STLE Annual Meeting, Calgary, Canada, May 2006, 

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Coating for High Temperature Wear Applications,” NASA/TM—2009-215678, August 2009. 

12. DellaCorte, C. and Bruckner, R.J.: “Remaining Technical Challenges and Future Plans for Oil-Free 

Turbomachinery,” IGTI paper GT-2010-22086, June 2010. 

13. Dykas, B., Bruckner, R.J., DellaCorte, C., Edmonds, B.J., and Prahl, J.M.: “Design, Fabrication and 

Performance of Foil Gas Thrust Bearings for Microturbomachinery Applications,” NASA/TM—
NASA/TM—2011-217003 13 

2008-215-062, ASME Paper GT2008-50377, and ASME Journal of Engineering for Gas Turbines 

and Power, vol. 131, 012301-1 to 012301-8, January 2009. 

14. DellaCorte, C., Radil, K.C., Bruckner, and Howard, S.A.: “Design, Fabrication, and Performance of 

Open Source Generation I and II Compliant Hydrodynamic Gas Foil Bearings,” NASA/TM—2007-

214691, ARL-TR-4102, and STLE Tribology Transactions, vol. 51, pp. 254–264, 2008. 

15. Radil, K.C., and Zeszotek, M.: “An Experimental Investigation Into the Temperature Profile of a 

Compliant Foil Air bearing,” NASA/TM—2004-213100 and ARL-TR-3200, May 2004. 

16. DellaCorte, C.: “Stiffness and Damping Coefficient Estimation of Compliant Surface Gas Bearings 

for Oil-Free Turbomachinery,” NASA/TM—2010-216924, ASME/STLE IJTC 2010-41232. 

17. Valco, M.J., and DellaCorte, C.: “Emerging Oil-Free Turbomachinery Technology for Military 

Propulsion and Power Applications,” Proceedings of the ARMY Sciences Conference, Ft. 

Lauderdale, FL. February 2003. 

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November 1975. 

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Optimized Oil-Free Rotorcraft Engine Concept,” NASA/TM—2008-215064, ARL-TR-4398, 

Proceedings of the American Helicopter Society, AHS Forum 63, Virginia Beach, Virginia, 

May 2007. 

20. Howard, S.A., and San Andres, L.: “A New Analysis Tool Assessment for Rotordynamic Modeling 

of Gas Foil Bearings,” ASME GT-2010-22508, June 2010. 

21. Radil, K.C., and DellaCorte, C.: “Foil Bearing Starting Considerations and Requirements for 

Rotorcraft Engine Applications,” Army Research Report ARL-TR-4873, August 2009. 

22. Howard, S.A.: “A New High-Speed Oil-Free Turbine Engine Rotordynamics Simulator Test Rig,” 

NASA/TM—2007-214489, April 2009. 

23. Howard, S.A.: “Misalignment in Gas Foil Journal Bearings: An Experimental Study,” Journal of 

Engineering for Gas Turbines and Power, volume 131, March 2009. 

24. Howard, S.A.: “Rotordynamics and Design Methods of an Oil-Free Turbocharger,” NASA CR-1999-

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Turbocharger Demonstration Paves Way to Gas Turbine Engine Applications,” presented at the 

International Gas Turbine & Aeroengine Congress & Exhibition sponsored by the American Society 

of Mechanical Engineers, Munich, Germany, May 8–11, 2000, 2000-GT-620. 

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Bearings for SUV, SAE paper 08SFI-0083, 2008. 

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Journal of Engineering Gas Turbines Power, vol. 124, issue 2, pp. 375–383, April 2002. 

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Combined Smart Bearings,” Proceedings of the 4th World Tribology Congress 2009, Kyoto Japan, 

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30. Nadjafi, R.H.: “Hybrid Foil/Magnetic Bearing,” US Patent #6,135,640, October 2000. 

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#6,353,273, March 2002. 

32. Kim, K.S.; “An Experimental Study on the Performance of Air Foil Thrust Bearing for Application to 

Turbomachinery,” Ph.D. Dissertation, Korean Advanced Institute of Science and Technology, 

Daejon, Korea, December 2006. 
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Oil-Free Shaft Support System Rotordynamics: Past, Present, and Future Challenges and 

Opportunities 

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6. AUTHOR(S)

DellaCorte, Christopher 

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14. ABSTRACT

Recent breakthroughs in Oil-Free technologies have enabled new high-speed rotor systems and turbomachinery. Such technologies can 

include compliant-surface gas bearings, magnetic bearings, and advanced solid lubricants and tribo-materials. This presentation briefly 

reviews critical technology developments and the current state-of-the-art, emerging Oil-Free rotor systems and discusses obstacles 

preventing more widespread use. Key examples of “best practices” for deploying Oil-Free technologies will be presented and remaining 

major technical questions surrounding Oil-Free technologies will be brought forward.

15. SUBJECT TERMS

Oil-free turbomachinery; Foil bearings; Gas turbines; Propulsion; Hybrid foil-magnetic bearings; Rotordynamics 

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