aerodynamic gas bearing design and performance gas bearing design and performance foil bearing has...

Aerodynamic gas bearing design and performance


Dynamic viscosity of common gases is 2 or 3 orders lower than viscosity of mineral oils at operating temperature. It follows, that also load carrying capacity and friction loss of gas bearing of the same dimensions as hydrodynamic bearing is 2 or 3 orders lower. Aerodynamic bearings should be therefore used for light rotors operating with very high speeds, while aerostatic ones could be used for somewhat higher loads and no limit exist as concerns minimum speed (it can be zero). Low viscosity of gases results also in relatively low damping of gas bearings, and that is why it is practically impossible to run through rotor bending critical speed. The rotors for gas bearing support should be therefore designed as rigid. Moreover, based on experience, the 1st bending critical speed of the rotor should be at least 60% above maximum operating speed. If this condition is not fulfilled and if the residual unbalance is distributed evenly along the rotor, incipient rotor bending will increase vibration amplitudes in bearings to unacceptable level. In some cases even satisfying the above condition is not sufficient and it is necessary to carry out correction of the 1st bending mode of the rotor [1] to achieve acceptable relative vibration amplitudes. As field of aerostatic bearings application is relatively limited and the subject is covered by other publication, we will hereafter deal with design and performance of aerodynamic journal and thrust bearings. 1.0 Aerodynamic journal bearings

Principle of aerodynamic bearing function is the same as that of hydrodynamic bearing, the only difference being much lower dynamic viscosity of gas as compared to liquid. Circular or lemon profile is practically impossible to use for aerodynamic bearings due to very bad stability properties of such bearings. To ensure rotor stability more complicated bearing geometry should be used. In less demanding applications it is possible to use pocketed or spiral groove bearings. Bearing with 2 or 3 pockets around sliding surface circumference (Fig. 1) is a modification of lemon or three-lobbed bearing; instead of sliding surfaces with preload, i.e. with radius of curvature greater than bearing radius, the pockets have in fact negative preload (radius of curvature lower than bearing radius). Two or three pressure peaks generated around the bearing circumference push the journal to the bearing centre thus increasing stability.

Fig. 1 Two/three pocket bearing Fig. 2 Spiral groove bearing

Spiral groove bearing (Fig. 2) is equipped with two rows of grooves, pumping gas into the bearing, thus increasing load carrying capacity and stability. Relatively good dynamic properties of the above bearing types are obtained at the expense of high demands on precision of manufacture, because bearing clearance had to be very small (usually only several micrometers). Operation with such a small gap can cause problems due to temperature


dilations; in extreme cases bearing clearance can be reduced to extent causing seizure of sliding surfaces.

For applications characterized by very high speed or demanding operating conditions it is necessary to use bearings not only with excellent dynamic properties, but also with some possibilities of adaptation to changed conditions. Two types of bearings fulfil these requirements, namely foil bearings and tilting pad bearings.

1.1 Foil bearings

Sliding surface of foil bearing is composed of one or more flexible elements, the geometry of which can be changed by aerodynamic pressure generated in gas film. The original types of foil journal bearings are shown in Figs. 3 and 4; in both types with pulled or bent foils the shape of working gap is far from optimum and start-up of rotor in such a bearing is difficult due to friction of foils pressed to the journal surface.

Fig. 3 Bearing with pulled foil Fig. 4 Bearing with bent foils

Foil bearings underwent relatively long development and their operation properties were significantly improved. The best solution seems to be so called Hydresil foil bearing (Figs. 5 and 6), which is composed of two foils; one foil (bump foil) is made of thicker material and supports thinner bearing foil, constituting sliding surface. Both foils deform in operation due to pressure generated in the gas film and create favourable geometry of bearing gap. In case of bearing clearance reduction elasticity of supporting foil enables restoration of original clearance.

Fig. 5 Hydresil foil bearing Fig. 6 Variant of Hydresil foil bearing

Rotor vibrations due to residual unbalance cause foil movement accompanied by friction of supporting foil on bearing casing and friction between supporting and bearing foil. The friction constitutes additional damping, which adds up to the damping of gas film. Hydresil


foil bearing has therefore excellent dynamic properties. A variant of Hydresil bearing (Fig. 6) is provided with bearing foil wound several times around bearing periphery, thus bringing more damping. It is just this thin bearing foil, which constitutes “weak point” of Hydresil bearing, because it is very vulnerable element, both for bearing assembly and operation.

1.2 Tilting pad bearings (TPJB)

Tilting pad journal bearings have excellent dynamic properties stemming from very small cross-coupling terms of stiffness matrix. Cross-coupling stiffness terms generate tangential forces “driwing” the journal around bearing centre, thus promoting rotor instability. In contrast to bearings with fixed sliding surfaces, which have cross-coupling terms of the same order as the main one, in TPJB the cross-coupling terms are two or three orders lower than main terms. It means, that TPJB is inherently stable and possible rotor instability can be caused only by external destabilizing forces, generated e.g. in labyrinth seals.

The above facts are valid for hydrodynamic, as well as for aerodynamic bearings. The only difference between hydrodynamic and aerodynamic bearings is concerned with number of pads; while hydrodynamic bearings have usually 4 or 5 pads (in hydro-generator guide bearings even 12 pads), aerodynamic bearings have generally only 3 pads. Low viscosity of gasses is the reason, why longer sliding surface is needed to generate sufficient pressure in gas film to carry some load. Another reason is magnitude of bearing clearance, which puts heavy demands on precision of manufacture. Aerodynamic TPJB are therefore in most cases designed in such a way, that bearing clearance can be adjusted during assembly and higher number of pads would make the process of adjustment very difficult.

1.2.1 Typical examples of TPJB design

Typical examples of aerodynamic TPJB design are presented in Figs. 7 and 8.

Fig. 7 Simple aerodynamic TBJB Fig. 8 TPJB with the system of self-compensation

In the simplest type of TPJB (Fig. 7) the pads 2 are supported on spherical ends of pins 3, the position of which could be adjusted by means of stoppers 4. Apart from bearing clearance, also position of pads within the bearing (bearing centre) could be set up. Wear of contact surfaces occurs due motion of pads caused by residual unbalance of the rotor, which limits bearing life. There is no possibility of the bearing to adapt itself to change of operating conditions and the original clearance may significantly change due to different working temperatures of the rotor and bearing. This disadvantage can be corrected by system of bearing clearance self-compensation, used in more sophisticated types of TPJB.


TPJB designed according to Czech patent specification (Fig. 8) utilizes another variant of bearing support enabling pad tilt both in peripheral and transverse directions without excessive wear. The pad 2 can tilt in peripheral direction on a pin resting on a supporting member 3, while the pin can tilt in transverse direction. As most of the pad motion takes place in peripheral direction when the pad rolls on the pin, there is minimum of friction and wear. Two pads have fixed position adjusted by stoppers 5. The third pad, with horizontal rotation axis usually the upper one, is equipped with the system of self-compensation, enabling the pad to shift in radial direction in case, that the clearance was reduced to dangerously small value. The condition, when the pad will shift, is adjusted by preload of the spring 7, which should be lower than pad load capacity. Basic bearing clearance and spring preload are adjusted by mutual position of screw 9 and stop 6. By these elements it is also possible to adjust maximum possible shift of the pad, so that there is no possibility to damage machine parts with small gaps (e.g. labyrinth seals or impeller blades).

Above described TPJB proved, apart from excellent operating properties, very high reliability and long life. Expansion turbines for liquefaction of helium were operated with speeds up to 350.000 rpm for periods exceeding 15.000 hours without any failure. This type of bearing was successfully used also on bigger machines, e.g. circulators with the output up to 9 kW and operating speed up to 85.000 rpm, or even on 100kW turbo-blower with operating speed of 18.000 rpm. Nevertheless, the bearing design is rather complex and its manufacture was relatively expensive; that is why simpler solution with similar properties was sought for.

1.2.2 TPJB with pads supported on elastic elements

To achieve free pad tilt and at the same time simple bearing design is not an easy task. Inspiration was found in foil bearings and the final solution joins together advantages both of tilting pad and foil bearings. The principle of the bearing is presented in Fig. 9.

Fig. 9 TPJB with pads supported on elastic elements

Pads 2 are supported on elastic elements 4, which are deformed to required shape by means of bolts 3 with nuts 6. The difference between inner radius of bearing casing 1 and outer radius of the pad should ensure, that the pad can freely roll on elastic element inner surface. The elastic element had to be preloaded to such extent, that the pad load capacity


exceeds the force necessary for elastic element deformation. Basic bearing clearance is adjusted by nuts 6, so that the journal can move in one direction in the range given by manufacturing clearance and bearing preload. Unlike in foil bearings, the shape of bearing gap is given by the difference between pad and journal radii and can be therefore optimised. Friction between elastic elements and bearing body contributes to damping of gas film, similarly as in foil bearings. Moreover, overall damping is further increased by squeeze effect of gas pushed out from the gap between elastic elements and bearing body.

Bearing of above described design was successfully tested with rotor of 0,3 kg mass to speeds exceeding 180.000 rpm.

1.2.3 Calculation of TPJB static and dynamic characteristics

No simple relations can be found for calculation of TPJB operating characteristics. Solution of Reynolds equation, governing flow in bearing gap, had to be carried out numerically. Load carrying capacity, friction loss and minimum film thickness are the basic bearing static characteristics necessary for machine design. However, the same or even greater importance should be given to stiffness and damping coefficients, needed for rotor dynamic calculation. For hydrodynamic TPJB it is sufficient to compute 4 stiffness and 4 damping coefficients, while for aerodynamic bearings 8 stiffness and 8 damping coefficients may be needed, comprising transversal movement as well as tilting of journal in the bearing.

Reynolds equation for journal bearing with compressible medium

��

��

�� 2. Λ ��.�� .�� , (1)

where �... dimensionless coordinate in direction of bearing periphery. � � � �⁄ … dimensionless coordinate in direction of bearing width, L … bearing width, � � � �⁄ … dimensionless film thickness, c … radial clearance,

Λ � .!.�"# �$%

�, &� � ' '(⁄

pa … ambient pressure,

Equation (1) is evidently non-linear in P and some form of linearization is needed for its solution. Suitable method is to introduce substitution ) � �&�. ��, through which we get

��*�� *�� +

,� �

��

-*

�*�� +

�� ) � .-

�.�./*�*�� (2)

Although equation (2) is still non-linear, it can be easily solved numerically by iteration methods. By discretization of above equation on rectangular mesh N x M points by replacing the derivatives by finite differences

�*��

*0,2345*0,264�∆� ,

��*��

*0,2345�*0,28*0,264�∆�� , (3)

��*��

*034,25�*0,28*064,2�∆�� ,

we get

*0,2345�*0,28*0,264

�� *034,25�*0,28*064,2�� + ,

� 9�� +

-/*0,2:

*0,2345*0,264�;� + �

�� < (4)

For steady state solution of equation (4) RHS = 0. Boundary conditions are as follows


)��, = >⁄ � � �� ... pressure at bearing periphery is equal to ambient pressure, i.e. &� � 1,

�*�� , 0� � 0 distribution of pressure is symmetrical relative to central plane of the bearing,

(therefore usually only one halve of bearing length is calculated)

)��,, A� � )��, A� � ��… pressure at inlet and outlet edge of sliding surface is equal to ambient pressure.

Using boundary conditions for equation (4) we arrive at matrix equation

BCDE. F)DG BHDE. F)D5,G BIDE. F)D8,G � F�DG , (5)

where [A], [B], [C] square matrix of order M (M … number of mesh points in ζ, direction)

{Q} … column vectors of dependent variable of order M,

{R} … column vectors of right hand side of order M. Boundary conditions are

Q1,j=H j2,

QM+1,j=QM-1,j, (6)

Qi,1=H12 and Qi,N=HN

2.

Individual elements of matrices [A], [B], [C] and those of vector {R} are given by relations

C�J,D�,D � +2 � ,�;��

,�;K��

,�� , for L M N2,OP,

C�,,,�,D � 1, C�J,D8,�,D � ,�∆K�� for L M N2,OP, C�,,��,D � 0 ,

C�J,D5,�,D � ,�∆K�� for L M N2,O + 1P, C�Q,Q5,�,D � �

�∆K�� ,

C�J,D�,D � 0 for 1 ≤ l ≤ i-2, i+2 ≤ l ≤ M,

H�J,D�,D � ,�;K��

,� R

��

-/*S

,�;� , for L M N2,OP,

H�,,,�,D � 0 , H�J,T�,D � 0 for i ≠ l,

I�J,D�,D � ,�;K�� +

,� R

��

-/*S

,�;� , for L M N2,OP,

I�,,,�,D � 0 , I�J,T�,D � 0 for i ≠ l, �,,D � �D� , �J,D � 0 . Equation (5) can be solved by so called column method, which is very effective and rapid procedure for small values of M. Solution of equation (5) is assumed in the form

F)D5,G � BUD5,EF)DG FVD5,G (6)

where [Ej] … generally complex square matrix of order M,

{Fj} … generally complex vector of order M.

{Qj} … generally complex vector of dependent variable of order M. In case of steady state solution matrices [Ej] and vectors {Fj},{Qj} are real. By substitution into (5) we get recurrent relations

BUDE � WBCDE BHDEXUT5,YZ5,BIDE , (7)

FVDG � WBCDE BHDEBUD5,EZ5,WF�DG + BHDEFVDGZ (8) Boundary condition, ambient pressure at the sliding surface inlet edge, is fulfilled by boundary conditions

[E1]=0, {Fj}=Q1.


As a first step matrices [E1] and vectors {Fj} for [ M N2, \ + 1P are calculated, followed by calculation of vectors{Qj} for [ M N\ + 1,2P. Above described procedure usually converges very rapidly. Sufficient density of the mesh for tilting pad bearing is 7x15 points. Present day PC enable to select much higher number of mesh points, but the difference in calculated results for mesh with double number of points is less than 5%. Calculation of static and dynamic characteristics for one value of load and several speeds takes several second on.

For TPJB with pads supported on elastic elements it is necessary to calculate also stiffness and damping of the pad support. However, the 1st step is to determine optimum thickness of supporting element.

1.2.4 Some results of TPJB calculation

Results of numeric calculation are shown on the example of aerodynamic tilting pad bearing 30 mm in diameter, which was used for gas turbo-blower with the output of 6,3 kW for research centre CERN [2] (Fig. 10).

Fig. 10 Turbo-blower with aerodynamic bearing

The rotor with maximum operating speed of 77.000 rpm is supported in two tilting pad journal of the design described in chapter 1.2.1. Bearings have 3 pads with angular span of 110°, l/D ratio of 0,93, manufacturing radial clearance 0,045 mm and preload 0,5. Fig. 11 presents variation of bearing relative eccentricity and power loss with speed.

Fig. 11 Relative eccentricity and friction loss vers. speed

It can be seen, that with increasing speed the relative eccetricity decreases, because pad load carrying capacity grows, while friction losses increase progressively with speed. Main stiffness and damping element of TPJB as a function of speed are shown in diagram in Fig. 12.

(52)

(38)

0,00

0,10

0,20

0,30

0,40

0,50

0,60

0,70

0 20000 40000 60000 80000 100000

rela

tive

ecce

ntric

ity(1

)

speed (rpm)

Relative eccentricity vers. speed

0

5

10

15

20

25

30

35

40

0 20000 40000 60000 80000 100000

fric

tion

loss

(W)

speed (rpm)

Bearing friction loss vers. speed


Fig. 12 Stiffness and damping of TPJB vers. speed

Unlike hydrodynamic TPJB, aerodynamic TPJB has Kxx stiffness element lower than Kyy element (x coordinate is in direction of static load, y coordinate is perpendicular to x in direction of rotation). This is consequence of relatively big angular span of the pads, so that lower pads extend up to upper part of the bearing. The same holds also for the main damping elements Bxx, Byy. For rotor with vertical rotation axis, both main stiffness and damping elements are the same, i.e. Kxx = Kyy, Bxx = Byy. While stiffness grows almost continually with speed, the damping progressively decreases with growing speed. For the case without consideration of pad mass cross-coupling stiffness (Kxy, Kyx) and damping elements (Bxy, Byx) are zero.

Fig. 13 shows distribution of aerodynamic pressure on pad sliding surface for two values of speed; both upper and lower pad are included.

Fig. 13 Distribution of aerodynamic pressure on pad surface; pad load capacity vers. speed

As can be seen from Fig. 13, aerodynamic pressure generated in gas film is dependent on rotational speed. However, due to compressibility of gas the dependence is not so steep, as in hydrodynamic bearings. For 4times higher speed the pressure increased only about 50%. The pad load capacity grows almost linearly with increasing speed, both for lower and upper pad,

0,0E+00

5,0E+05

1,0E+06

1,5E+06

2,0E+06

2,5E+06

3,0E+06

3,5E+06

4,0E+06

0 20000 40000 60000 80000 100000

stiff

ness

(N

/m)

speed (rpm)

Bearing stiffness vers.speed

Kxx Kyy

0,0E+00

2,0E+02

4,0E+02

6,0E+02

8,0E+02

1,0E+03

1,2E+03

1,4E+03

1,6E+03

0 20000 40000 60000 80000 100000

dam

ping

(N

.s/m

)

speed (rpm)

Bearing damping vers. speed

Bxx Byy

1,0

1,1

1,2

1,3

1,4

1,5

1,6

1,7

1,8

1,9

0 5 10 15

pres

sure

rat

io (

1)

number of mesh point

Aerodynamic pressure along pad length

upper pad - 20.000 rpm lower pad - 20.000 rpmupper pad - 80.000 rpm lower pad - 80.000 rpm

0

5

10

15

20

25

0 20000 40000 60000 80000 100000

load

cap

acity

(N)

speed (rpm)

Pad load capacity vers. speed

lowerpad upper pad


which is demonstrated in right diagram of fig. 13. Fig. 14 presents diagrams of the pad stiffness and damping coefficients vers. speed.

Fig. 14 Pad stiffness and damping elements (for translation journal movement)

By comparing of Figs. 12 and 14 one can see, that dependence of the bearing stiffness and damping on speed is similar to that of the pad main stiffness and damping elements Kxx, Bxx. Main elements Kyy, Byy, as well as cross-coupling terms Kxy, Kyx, Bxy, Byx are relatively small. Stiffness and damping elements Kij, Bij apply for translational journal movement. For analysis of pad movement are important stiffness and damping elements Gij, Dij, which apply for the case of journal tilting. Fig. 15 shows stiffness and damping coefficients Gij, Dij for the same bearing as in Fig. 14.

Fig. 15 Pad stiffness and damping elements (for journal tilting)

It can be seen, that the dependence on speed is quite similar, but stiffness and damping elements for journal tilting are several orders lower than those for journal transverse motion.

0,0E+00

5,0E+05

1,0E+06

1,5E+06

2,0E+06

2,5E+06

3,0E+06

3,5E+06

0 20000 40000 60000 80000 100000

stiff

ness

t (N

/m)

speed (rpm)

Pad stiffness matrix elementsfor journal translation movement

Kxx Kyx Kxy Kyy

-2,0E+02

0,0E+00

2,0E+02

4,0E+02

6,0E+02

8,0E+02

1,0E+03

1,2E+03

1,4E+03

0 20000 40000 60000 80000 100000da

mpi

ng(N

.s/m

)speed (rpm)

Pad damping matrix elementsfor journal translation movement

Bxx Bxy Byx Byy

0,0E+00

2,0E+01

4,0E+01

6,0E+01

8,0E+01

1,0E+02

1,2E+02

1,4E+02

0 20000 40000 60000 80000 100000

stiff

ness

(N.m

/rad

)

speed (rpm)

Pad stiffness matrix elementsfor journal tilting

Kxx Kyx Kxy Kyy

-1,0E-02

0,0E+00

1,0E-02

2,0E-02

3,0E-02

4,0E-02

5,0E-02

0 20000 40000 60000 80000 100000

útlu

m(N

.m.s

/rad

)

speed (rpm)

Pad damping matrix eleme nts

for journal tilting

Bxx Bxy Byx Byy


2. Aerodynamic thrust bearings

There is a great number of aerodynamic thrust bearing types. Some examples are presented in Figs. 16 to 19.

Fig. 16 Stepped thrust bearing Fig. 17 Tilting-pad thrust bearing

Fig. 18 Tapered land bearing Fig. 19 Pocket thrust bearing

All above shown bearing types can be used for not too difficult operating conditions, because with the exception of tilting-pad bearing they have no possibility to accommodate to changed operating conditions. For all these bearing types it holds, that their load carrying capacity is moderate and therefore not sufficient for some applications.

That is why, as well as in case of journal bearings, we will further concentrate on two bearing types – one with the best possibilities of adaptation to changes of operating conditions, and the other with maximum load capacity.

2.2 Foil thrust bearings

The original aerodynamic foil thrust bearing was composed of thin foils, which were fastened to flexible membrane (spring) with ribs supporting the foils (Fig. 20). Although this bearing had some possibility of adaptation to operating conditions, its function properties could be further improved.

Hydresil foil thrust bearing (Fig. 21) developed by MTI is equipped, as well as journal bearing, with two kinds of foils – the thicker supporting bump foil and the thinner foil, composing pad sliding surface. This design enables the geometry of bearing gap to adapt itself to generated aerodynamic pressure and operating conditions (speed, load). Elasticity of supporting foil enables also to damp vibrations, to which the machine can be exposed. Hydresil foil thrust bearing is therefore suitable for difficult operating conditions, but its load


carrying capacity is relatively low due to sliding surface divided into several pads. At the periphery of sliding surface the pressure in bearing gap falls to ambient pressure; maximum load carrying capacity can be therefore reached with annular bearing surface.

Fig. 20 Original foil thrust bearing Fig. 21 Hydresil foil thrust bearing

. Fig. 22 Photograph of supporting foil and partly disassembled bearing [6]

Fig. 22 shows the shape of supporting foil and the whole bearing with one bearing foil removed. Supporting foil can have different thickness and various shape and span of the bumps, by which bearing properties can be tuned according to operating conditions.

2.2 Spiral groove thrust bearings

The highest load carrying capacity can be achieved with spiral groove thrust bearings – Fig. 23.

Fig. 23 Spiral groove bearing with through hole for the shaft and with blocked centre


Especially so called “blocked centre” bearing (Fig. 23 right) provides extremely high load capacity, because gas pumped into the bearing by spiral grooves cannot escape from central part of the bearing. Fig. 24 illustrates load capacity of individual bearing types, characterized by value of load coefficient θ. Bearing load capacity then can be calculated according to formula

]̂ (_ � !�$��`a0b� c (N),

where µ ... dynamic viscosity of gas (Pa.s),

ω ... angular velocity of rotation (s-1), R2 ... bearing outer radius (m),

hmin ... minimum film thickness (m).

1 2 3 4 5 6 7 8

Fig. 24 Load capacity of aerodynamic thrust bearings [3]

Fig. 24 includes following bearing types and their load coefficients θ:

1 – tapered land (Michell) bearing, θ = 0,042,

2 – stepped bearing, θ = 0,047,

3 – spiral groove bearing (blocked centre), θ = 0,366,

4 - spiral groove bearing (with through hole, inward pumping) , θ = 0,125,

5 - spiral groove bearing (with through hole, outward pumping) , θ = 0,085,

6 – herringbone grooved bearing, θ = 0,106,

7 - spiral groove bearing (inward pumping, without sealing ring) , θ = 0,045,

8 – spiral groove hemispherical, θ = 1,106.

Apart from extreme load capacity of hemispherical bearing, which is very difficult to manufacture, the maximum load capacity can be achieved with spiral groove bearing without trough hole, i.e. blocked centre type. Unfortunately both these bearing types without through hole can be used only in very limited number of applications, because axial force is generally directed to compressor or turbine impeller and that is why the thrust bearing had to be designed with through hole.

As can be seen from Fig. 24, the highest load capacity from bearings with through hole has inward pumping spiral groove bearing, which is therefore the most widely used type of


aerodynamic bearing. Bearing of this type, with inner/outer diameter of 13,6/25 mm, achieved at 290.000 rpm load capacity of 55 N, which corresponds to specific load 0,16 MPa. To ensure even higher load capacity it was necessary to run in the bearing surface by thrust runner. However, after finishing of this procedure the machine must not be dismantled in order not to dislocate perfect alignment of sliding surfaces. Aerodynamic thrust bearings operate with extremely thin gas film, in some cases with film thickness lower than 1 µm, and the highest precision of manufacture is therefore essential.

For bearings of bigger diameter it is difficult to achieve necessary alignment between the bearing and thrust runner sliding surfaces. In such cases it is necessary to use some kind of flexible support, which is in most cases constituted by gimbal suspension. An example of gimbal suspension of thrust bearing 250 mm in diameter for turboblower with operating speed of 18.000 rpm is shown in Fig. 25.

Fig. 25 Turboblower with aerodynamic bearing support with detail of thrust bearing gimbal mounting

Conclusions Aerodynamic bearings are destined for specific operating conditions, namely for high speeds and low loads. The best way is to design the machine for aerodynamic support from the beginning. Reconstruction from other bearing type is sometimes possible (e.g. turboblower in Fig. 24 originally with ceramic ball bearing support), but it can never be optimal. Aerodynamic tilting pad journal bearings and spiral groove thrust bearings proved high reliability and long term durability in operation of helium expansion turbines [e.g. 4] or gas circulators [2, 5]. New generation of foil bearings (Hydresil type) looks also very promising even for very demanding applications, such as automobile turbochargers, power generators or air cycle machines, providing cooling, heating and pressurization in airplanes [e.g. 7]. As aerodynamic bearings operate with contact of sliding surfaces during run-up and run-down, very important role play also material properties, which can be important even in a short time contact of sliding surfaces at operating speed. Research of suitable sliding materials for the pads, foils and linings of spiral groove bearings is not less important part of their development than calculation and design, but it is outside the scope of this paper.


References

[1] Šimek, J.: Balancing of the 1st bending mode of high-speed rotors www.techlab.cz, section „Sliding bearings and rotor support“ - „Useful

information“

[2] Šimek, J.: Design of aerodynamic bearing support of Ar circulator with the output of 6,3 kW. Technical report TECHLAB No. 05-407.

[3] Wiemer, A.: Luftlagerung. VEB Verlag Technik, Berlin, 1969

[4] Šimek, J. - Kozánek, J. - Šafr, M.: Some interesting features of gas bearings. 7th IFToMM-Conference on Rotor Dynamics, Vienna, Austria, 2006.

[5] Šimek, J. - Svoboda, R.: Design of aerodynamic bearing support of argon circulator rotor for research centre CERN. Technical report TECHLAB No. 00-402.

[6] Dykas, B. – Bruckner, R. – Prahl, J: Design fabrication and performance of foil gas thrust bearings for microturbomachinery applications. NASA/TM-2008-215062.

[7] Agrawal, L.: Foil air/gas bearing technology – an overview. ASME publication 97-GT-347.

aerodynamic gas bearing design and performance gas bearing design and performance foil bearing has...

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