23:231:231

23.1 Anti friction bearings/roller bearings

23.1.1 Basics and failure mechanisms

There is a multitude of specialist literature and articles (Lit. 23.1.1-22), which deal with the identifikation of failures on anti friction bearings (here named bearings) and its causes. Thereby the macroscopic and microscopic failure mode/appearance is correlated in a systematic with the most probable cause (Ill. 23.1.3-6). These publications are often in the practical work only limited useful, because the failures propagated already so widly, that satisfactory conclusions are no more possible. In extensive publications, especially emphasized, are failure modes/appearances. Expert contributions rather deal with failure mechanisms and its influence, Both do not satisfy the practician of the aero engine technology.
In this chapter those informations are combined. The typical failure mechanisms of bearings are described and explained, as well as correlated with failure modes/appearances. This shall enable to identify potential dangers in time and be aware to avoid those. Are changes observed at the bearings, the evaluation and assessment correspondent with specifications/instructions and manuals are suppiorted with a deeper understanding.
In aeroengines, two types of bearing are distinguished depending from the sort of application. Main bearings (Lit. 23.1.1-1), which carry the radial and axial (thrust bearing) loads of the rotors and bearings at the connection to accessory devices and inside these (e.g., radialdrive/tower shaft to the accessory gear).
Aeroengine bearings differ basically from those of the common mechanical engineering. What distinguishs them are the demanding operation conditions like high service temperatures and high `n x D' values (high circumferential speeds). To damp vibrationes from unbalances and prevent dynamic overloads on the supporting structures, damped or undamped radial movable (elastic or sliding guided), suspended bearings are used (Ill. 23.1.1-2).
Besides this chapter, bearing problems and failures are also discussed under separate aspects in the further volumes of these publication:

  • With the reconstruction of the temporal failue sequence deals volume 1, Ill. 4.5-13.
  • The monitoring and identification in time of developing bearing failures is in detail discussedin chapter 22.3.3.1 (oil investigations and evaluation of particles) and chapter 22.3.4 (oilmonitoring).
  • The fracture of shafts as result of bearing failures, is discussed in volume 1, chapter 4.5.
  • Bearing failures by lightnings can be found in volume 1, chapter 5.1.3.
  • Deterioration by „brinelling“ during transport, see volume 4, Ill. 18-8 and Ill. 8-10.
  • Bearing problems by fretting (vibration wear) at beating seats volume 2, Ill. 6.2-16 and Ill. 6.2-17.
  • Bearing contaminatuion by particles from bearing chamber seals (labyrinths) can be found in volume 2, Ill. 7.3.2-12.

Ill. 23.1.1-1 (Lit. 23.1.1-38 bis Lit. 23.1.1-40): 'Aviation bearings' are subjected the loading types of the common machine engineering. However the especially demanding service conditions with highest safety requirements are a big challenge. So weight and assembly space are of unusual importance. This show specific designs. So the outer rings also form an application specific flange (sketch). Gear shafts are provided with integral bearing inner rings (Ill. 23.2.1-3). Damped and elastic designs (Ill. 23.1.1-2) guarantee an optimal dynamic behaviour of the rotors. Typical for aeroengines are:
High outer loads: To these belong operation loads and stresses, according to special service conditions. Often shock forces are concerned. In radial direction act e.g., landing bumps and unbalances. Typical are failures on rotor blades or because of rotor bow (volume 2, Ill. 7.1.2-9). Thrust bearings must take unusual, high axial forces from compressor surges (volume 3, Ill. 11.2.1.2-1).
Lubrication must be guaranteed also under high g-loads, also during maneuvers, especially in military missions (volume 2, Ill. 7.1.2-12). Even the short time loss of the oilsupply must be tolerated. Aggravating acts during this, because of the decreased cooling, the oil properties in the limit range, because of the high temperatures. A special problem is the lubrication and cooling of the sliding guided, high speed rotating cages (Ill. 23.1.1-7) and rolling elements in the cage pockets.
Extreme rotation speeds have, because of the large main bearing diameter (D x n values) of big aeroengines, a special significance (Ill. 23.1.1-10.1 and Ill. 23.1.1 -10.3). For aeroengines with little power, the rotation speeds lay above 50 000 U/min, in ranges where small damages act catastrophic within seconds. The centrifugal forces of the rolling elements demand a high part of the bearing loading capacity. Especially high demands are claimed a the dimensional accuracy, respectively the balance condition of the rotating elements (Ill. 23.1.1-19). Effects like cage slippage (skidding 23.1.1-14.1) or roller flutter (roller weaving, Ill. 23.1.1-19) urge into the foreground. This is also caused by the high resistance of the oil film. Three shaft aeroengines own so called intershaft bearings. The inner ring and the outer ring of these are operating with different rotation speeds and or counter-rotating. A special problem is the supply and scavenge of the lubrication oil.

High service temperatures demand as well from the bearing materials (Ill. 23.1.1-11.1), as also from the oil a load bearing capacity in the limit range. With this, the ignition of an oil fire gets more and more probable (volume 2, Ill. 9.2-2). especially demanding is the temperature load on the hot part (e.g., turbine). The problem are high temperatures and coking of the hot part bearings after shut off of the engine (heat soaking, Ill. 22.3.2-6.1 and Ill. 23.3.2-6.2).
Environmental influences: Developing of condensation during the shut off periods, as well as sea salt act corrosive. This can demand the application of special materials (Ill. 23.1.1-11.1). Here, a corrosion cell formation with the cage material (e.g., bronze, silver plating) acts corrosion promoting .
Assembly conditions like module design can demand different combinations of the bearing inner ring with the outer ring and the cage. To this belongs also the damaging danger of the roller bearing during the pushing together during the assembly of the modules (Ill. 20.1-5 and volume 1, Ill. 4.3-6).
Wear/fretting must be especially considered at oil damped bearings (Ill. 23.1.1-2), caused by the principal determined oscillating movements. At bearing seats, relative movements caused by the light and elastic design of shafts and casings, can promote fretting. This must be considered with a suitable material selection of the bearing seat. For example the bearing innr ring must not have a direct contact with a titanium alloy. In this case a TC (tungsten carbide) coating at the shaft has proven (volume 2, Ill. 6.2-16 and Ill. 6.2-17).

Ill. 23.1.1-2 (Lit. 23.1.1-2 up to Lit. 23.1.1-4): Anti friction bearings can be damped as well mechanically (friction) as with an oil cushion Today generally oil damped systems prevailed (frame). The oil cushion is located in a circumferential gap at the outer ring or in a supporting casing ring. The gap is rather sealed at military applications with sidewise, axially pressed on O rings. If the oil cushion is connected with a feeding and a drainage (sketch above right), the oilstream can dissipate heat. Is this not the case, it must taken care that the O rings get not overheated (chapter 23.4).

At one design the bearing is centered by an elastic spring element („squirrel cage”). This also serves as locking against rotation. Such a spring usually has a shape similar a drum, with axial bars. At smaller bearings we find one piece designs (detail above left), larger bearings are supported by assembled „cages“ (bars as rods with threaded heads).
You also find only oil dampened designs. Here, without an elastic centering, the outer ring is secured against rotation with fixing lugs in radial slots.
Dampened bearings are preferably used near turbine disks. Compared with the stiff rotordrums of a compressor, this turbine shafts are elastic. Operation caused unbalances (thermal distortion, oxidation of the blading) often can only be controlled with damped bearings. With a centering elastic suspension, natural frequencies can be lowered from the opreation range to non-dangerous rotation speeds.

Advantages against an undampened mount:

  • Vibrations of the rotor and the aeroengine are lowered during unbalances. * Resonances will be suppressed, natural frequencies can be shifted to more suitable ranges.
  • Lower vibration fatigue stresses of supporting structures like bearing chambers respectively casings.
  • Shock loads from the outside do less act at the rotor.
  • Enhanced dynamic behaviour of the rotor, especially during changes of the rotation speed.
  • Suppression of the tendency for skidding (Ill. 23.1.1-14.1 and Ill. 23.1.1-15).
  • Bearing rings are cooled from a, design depending, sufficient oil flow.


There are also some disadvantages and potential problems:
Fretting wear: The function caused relative movement of the outer bearing ring against the supporting structure can develop fretting at sealing and contact surfaces. This is especially serious for a direct contact between titanium alloys and the bearing ring from steel (volume 2, Ill. 6.2-16 and Ill. 6.2-17). Is the oil film „penetrated” under high radial forces, fretting wear also forms in the circumferential region.

Cavitation is a sort of material fatigue in connection with the formation of vapour bubbles in the oil (volume 1, Ill. 5.3.1-11.2). The radial amplitudes of the bearing outer ring lead inside the oilgap to heating and locally high pressure fluctuations. In the phase of the pressure drop the formation of vapour bubbles occurs. Those implode in the pressure phase and deteriorate the material surface. Cavitation is promoted by oil contaminations as water or fuel (Ill. 22.3.3.2.1-1). Expecially dampers without elastic centering tend to cavitation.

Changes of the friction conditions: Is the oil film sealed from the side surfaces, arises during the relative movement an additional friction damping. In such cases it is understandable, that the function of this system also is governed from the acting tribological system. So can changes like a wear coating as repair have astonishing negative consequences. Therefore such measures/remedies outside the instructions in the manual require a proving with the OEM.
For civil applications wth its typical long operation periods, radial tight piston rings are used (sketch below right). Its leakage guarantees a cooling oil flow. An additional damping effect has the friction of the piston rings.

Heating of the oil by absorbing damping energy leads to the warming of the oil film. However, this film thould also act cooling at the bearing. With the oil temperature, the risk of cavitation increases (passage before). If the oil temperatures rise very high, an accelerated aging of the oil occurs (chapter 22.3.1 and chapter 22.3.2).

Fractures of the struts from the elastic suspension (squirrel cage) can occur during unnormal high shock loads (forced fractures) or dynamic circumferential forces (fatigue). Skidding (Ill. 23.1.1-14.1) or unbalances trigger high dynamic friction forces in the bearing, when it comes to the break through of the lubrication film, Extreme unbalances can overload bearings with a shock like circumferential force. Than the struts break in a forced fracture mode.
Fracture of the fixing lugs: Dampings without elastic reset take the circumferential forces of the bearing outer ring with lugs, which slide in flutes. Extreme unbalances can penetrate the oil film on the races. Thereby high shock like circumferential loads occur. With this, the lugs can crack or break out (Ill. 23.1.1-16), up to the fracture of the outer bearing ring.

Worse detectability of failures caused from unsuitable at the casing positioned acceleration pick ups: Vibrations are monitored at modern aeroengines with ascceleration sensors, which are fixed at the casings (Ill. 25.2.1-5). A damped bearing, from experience can make it difficult (Ill. 25.2.1-7) to register „lesser“, however for the rotor components (e.g., fatigue of the bearings) dangerous unbalances. This concerns also the partial fracture of a turbine rotor blade.

Abrasion: Damping systems without elastic resetting let the rotor sink about the thickness of the oil film (0,1-0,5 mm), as a result of the declining oil pressure during shut down of the engine. This seems indeed at the first sight little, but can markedly effect the tip gap of the blades (rubbing) and at labyrinth leaks (leakage of the oil during stand still) during start.

Ill. 23.1.1-3 (Lit. 23.1.1-5 up to Lit. 23.1.1-9): The friction in a non friction bearing, influences the stress distribution inside the races. Increases the friction coefficient between rolling element and race/ring caused by mixed friction (e.g., during cage slipping = skidding, Ill. 23.1.1-14.1 and Ill. 23.1.1-15), the subsurface failure effective stress maximum from tension and shear dislocates to the surface (sketch above).
The lower diagram shows the stress distribution perpendicular to the surface. It applies to hydrodynamic lubrication (Ill. 23.1.1-6) and mixed friction. During the mixed friction the high stress maximum can be identified at a locally, very narrow contact point. With this, also the deteriorating effect of foreigen objects/particles or damages of the race (e.g., indentations with `bell mouth', fusion craters) gets understandable. In its area fatigue pittings form and with this the bearing failure.

Ill. 23.1.1-4 (Lit. 23.1.1-9, Lit. 23.1.1-10 and Lit 23.1.1-15): The „normal” lifetime limiting failure at anti friction bearings are break outs ( fatigue pits/pittings, detail left) on the races (sketch above). They are caused of vibration fatigue due to rolling contact. Its expansion takes place in rolling direction of the rolling elements. The origin of the cracking is usually located in the highest stressed zone below the pitch surface/race (Ill. 23.1.1-3). Under dynamic overload (e.g., unbalances), serve the material specific weak points (volume 1, Ill. 3-1)as crack starter. Under normal/design according service loads the early origin of fatigue cracks (Ill. 23.1.1-9) is located at damages of the races/pitch surfaces.
Concerned are corrosion pits, foreigen objects, respectively its indentations or other deteriorations like fusion craters (electric arc) and deterioration by brinelling (Ill. 23.1.1-12).
So the fatigue of the race is rather a consequence but not a cause of a bearing failure. Because of this, bearings with corrosion pits must be scrapped during overhaul. Therefore corrosion is the most frequent cause for the scrapping of bearings (Lit. 23.1.1-21). If a metallographic cross section is taken in circumferential direction perpendicular to the race (sketch below right), difficult etching zones (white etching areas = WEAs) at the crack origin can be found. Because of its geometry, they are also called butterflies, though this is not the only mode (Lit. 23.1.1-15). Obviously they are in connection with a change in the material structure, caused by extreme heating (up to softening) in the micro region under the dynamic load. WEAs may apply as signs of dangerous high dynamic loads. Its hardness is above this of the base material. This can be explained with a fast cooling from high temperatures (new hardening). The frequency of the WEA depends from load and time. It is about an irreversible material change/deterioration by fatigue.
Are there already break outs/pittings, the failure will accelerate. Thereby the notches and spalled race particles act as foreign objects (Ill. 23.1.1-9). In spite of this, the failure propagation is in the most cases relatively slow. So it is possible, to identify with sensors (magnet plugs, magnet sondes, chapter 22.3.4) in the oil circuit sufficient early a catastrophic failure of the component. The experience shows. that merely bearings of small, very high speed aeroengines (helicopter engines, APUs), develop failures too fast, so that they can not be intercepted.

Ill. 23.1.1-5.1, -5.2 (Lit. 23.1.1-10 ): The fatigue of races from anti friction bearings depends from very many influences, which can also combine. In spite of the offer of systematic specialist literature it is not easy to suggest, sufficient certainfrom the failure mode/appearance of a bearing at the causative influences of a fatigue. This is further aggravated from secondary failures (e.g., roll overs), which have changed the informative beginning/primary failure.
The summary shall give an impression of the variety possible failure relevant influences without claiming completeness.

Material (Lit. 23.1.1-21): In the bearings of aeroengines the fatigue strength/ rolling fatigue strength and/or wear (during mixed friction like in the case of skidding), is the life governing role. As main cause for the exchange of bearings during overhaul, the expert literature however mentions corrosion damages of the races. But there is no single optimal material property for all service influences.
The experience shows, that fatigue failures of aeroengine bearings rather are in connection with race deteriorations. Less frequent is fatigue at coarse impurities in the material structure. Therefore high rolling fatigue strength demands especially for bearings of aeroengines, which usually run at EHD conditions, an insensitivity as high as possible against particle indentations (Lit. 23.1.1-23). For this is, contrary to bearings with high hertzian compression of the usual machine engineering (Lit. 23.1.1-6), a little less hardness in favour of a better ductility is more beneficial. So the notch effect can be lowered by plastic deformation. The slightly lower hardness also leads to a slower growth of the fatigue cracks. This increases the chance, to prevent a catastrophic bearing failure (Ill. 23.1-5.2). A disadvantage of lower hardness is a higher sensitivity for `brinelling' (Ill. 23.1.1-12) and wear/abrasion at mixed friction.

Fatigue processes at surfaces/races of of rolling faces (Ill. 23.1.1-7) may be the reason for the recommended hardness in the range of 62-64 HRc (Lit. 23.1.1-28).
Also abrasive wear can be minimized with a hardness of the race as high as possible.
In contrast adhesive wear (galling/seizing) is less influenced from the hardness. This, for example, occurs during overload of the bearing with a break through of the oil film or during lack of lubricant.

For the rollig fatigue strength, according to the design, the cleanliness of the bearing material is of highest importance (Ill. 23.1.1-11.1). Good fatigue properties demand a fine and uniform material structure. Unfortunately this can be only difficult realized with corrosion insensitive bearing materials. Here larger Cr carbides disturb, which are traced back to the addition of corrosion protecting chromium. This dilemma seems to be solvable with highly nitridated bearing steels.
Gear shafts in auxilary gears exist, at which integral bearing race rings come to application (Ill. 23.2-3). For this the inner ring is joined with the shaft by electron beam welding or friction welding. A sufficient hardness of the race is realized with case hardening (Ill. 23.1.1-11.1). Influence of dimension and shape: Good measurable, even if requirements for very tight tolerance and low roughness guarantee service properties, according to the design. Perhaps therefore, there are no reports about failures caused by dimensional production deviances. However, it is absolutely possible, that the combination of shape and measure variations inside the prescribed tolerances trigger failures (Ill. 23.1.1-10.1 and Ill. 23.1.1-10.2). Thereby dynamic instabilities of the roller movement are of importance (flutter/ weaving, Ill. 23.1.1-19).
An elastic deformation of the bearing rings is possible by the seat, as a result of stiffness differences at the circumference or distortion during service. But those obviously are controlled by dimensioning and design. To prevent skidding (Ill. 23.1.1-14.1), the outer ring is elastic deformed (oval, trisectioned, Ill. 23.1.3-12). With testing it must be proved, that locally mixed friction can not occur and damage the race.
Important is the control of unbalances of the rolling elements (Ill. 23.1.1-10.1, Ill. 23.1.1-10.2 and Ill. 23.1.1-18) and of the cage (Ill. 23.1.1-17).

Deteriorations during operation are frequently traced back at aeroengines to foreign objects /particles in the oil (chapter 22.3.3.1). To this belong auxilary materials like blasting grit/shot (Ill. 22.3.3.1-2).
Further potential causes for the deterioration of races:

  • Abrasive and adhesive wear because of skidding (Ill. 23.1.1-14.1).
  • Brinelling as result of vibrations during stand still (Ill. 23.1.1-12).
  • Deterioration of the cage and the races by Rollerweaving (flutter, Ill. 23.1.1-19).
  • Lightning strike (Ill. 23.1.1-21 and volume 1, Ill. 5.1.3-4.1).
  • Electrical circuit from the generator (micro arcs, Ill. 23.1.2-2).

Assembly, repair and transport can deteriorate anti friction bearings in different ways:
Plastic deformation of the bearing during the assembly process (Ill. 20.1-25).
Brinelling during transport (Ill. 23.1.1-13 and volume 4, Ill. 18-10).
Design/dimensioning: Here are unexpected axial loads at thrust bearings to mention Concerned can be too high or too los loads. They are in connection with unusual gaps, respectively lekages of the air seals. These disturb the balance of the design conform axial forces (piston forces, volume 2, Ill. 7.2.1-2).
Was a unsuitable cage guidance/lubrication choosen (Ill. 23.1.1-17 ), this can lead under certain circumstances to an almost spontaneous destruction of the bearing.
Too weak cage material or riveted cages (Ill. 23.1.1-18 and Ill. 23.1.1-19) are especially sensitive for unexpected load changes (e.g., unbalances, increased forces of the rolling elements).
Oil problems are deeper discussed in chapter 22.3. These are great challenge, especially for bearings of aeroengines with ecxtreme circumferential velocities and temperatures.

Ill. 23.1.1-6 (Lit. 23.1.1-10 and Lit. 23.1.1-18): For roll off surfaces in an oil three different lubrication conditions are known.:
At hydrodynamic lubriaction the lubricating film about 2µ thick is sufficient, to fully separate the roll off surfaces. During this pressures of several thousand bar develop in the lubrication film. Thereby no noteworthy elastic deformation of the surfaces will occur. However the high shear forces in the lubrication film produce much heat. It heats the lubrication film and so promotes the aging (chapter 22.3).
Boundary lubrication/mixed friction has a 0,001-0,05µm thickness of the lubrication film. A touch of the roughness tips from the pitch surfaces is not avoided. This condition can be expected at low rotation speeds during start and shut down. In this case the lubrication is determined from chemical and physical properties of the lubricant, as well as from the wetted surfaces. Here the viscosity is not of significance.
Predominantly at bearings of aeroengines is the so called elastohydrodynamic lubrication (=EHL). The thickness of the lubricant film lays here in range of 0,05-2,0µm. The pressures in the lubrication gap are with about 20 000 bar extremely high. This is sufficient for markedly elastic deformations of the pitch surfaces (see marked region in the upper sketch). With this a lifetime relevant fatigue load of the races develops. The viscosity of the oil (no Newtonian fluid) markedly increases by the pressure. This improves the bearing strength of the lubrication film.

In the range of the roughness tips, where material typical hard structure components act (sketch above), it comes to peak stress under EHL-conditions (extreme pressurese, high oil viscosity). Thereby so called running-in pittings (Ill. 23.1.1-7) in the size of some 'µm' can form (detail below left).

For the lifetime of the bearing, the thickness of the oil film h0 and the accumulated roughness of both opposing surfaces respectively its ratio S (detail below right) are of importance (Ill. 23.1.1-8).
Does it come in case of boundary lubrication (low rotation speeds during start and shut off, overload of the bearing) to the contact of the roughness tips of the pitch surfaces, the deteriorating effect can be compared with the bridging by contamination particles (Ill. 23.1.1-8).

Ill. 23.1.1-7 (Lit. 23.1.1-10 , 23.1.1-16 and Lit. 23.1.1-17): The high magnifications of today usual SEM investigations (= Scanning Electron Microscopy, volume 4, Ill. 17.3.2-7) make the phenomenom assessable. In contrast, photo-optical misinterpretations and a false evaluation of the failure relevance are rather probable.
Minimum lubrication thickness leads at tiny roughness tips („A“) after many roll-overs, to normal microscopic wear of the race. Such a roughness is typical for grinded and polished races of anti friction bearings. This wear does not influence the design life time. It shows as tiny fatigue break outs in the surface region. If not deeper cracks occur we speak about run in pittings.
During this mechanism, first the roughness tips are plastically deformed to a layer in the µm-range („B”). This work-hardened zone covers the structure of the base material with its typical fine dispersed carbides, necessary for the function.
Rises the load and/or decreases the thickness of the lubricant film (Ill. 23.1.1-6), the strain hardened layer separates („C“). Micro-fatigue cracks form, during the direct contact with the carbides, which cause the development of further break-outs in the size of few µm („D”). However these microspalls (micropittings) don't cause in this run-in period larger, dangerous fatigue break-outs (fatigue pittings, Ill. 23.1.1-4).
At high speed roller bearings and very low unbalances, the danger exists that the rolling motion passes into a sliding motion (slip, Ill. 23.1.1-8). Thereby the cage is decelerated (cage slip, skidding , Ill. 23.1-14.1). This can lead up to the short time standstill of the cage. Thereby, at the beginning the mentioned surface symtoms arise. Increasing this effect, a markedly lifetime reducing deterioration µm thin plane particles of several tenth millimeter size can develop.

Ill. 23.1.1-8 (Lit. 23.1.1-25 ): With EHL-lubrication (Ill. 23.1.1-6) the lifetime of an anti friction bearng is in relation with the roughness of the race (diagram above left).
We see, that the lifetime of roller bearings reacts especially susceptible on the roughness. Roller bearings with the typicaql larger lubricat on gap surfaces, compared with the ball bearings, reach the maximum lifetime already above2 µm lubrication film thickness. Obviously ball bearings reach this condition only at much thicker lubrication films.
The diagram above right shows the influence of the lubrication gap on the lifetime of a roller bearing. A wideness of the lubrication gap, which correlates the double roughness „h0 (S =2,0), enables maximum lifetime. This optimal thickness of the lubrication gap is determined by the bearing clearance. So it must be considered during design and production.
Occurs in the lubrication gap of a roller bearing skidding (e.g. during cage slip) between the rolling surfaces, withn a breach of the oil film, „S” approximates 0. With this in correlation to the example in the diagram below left, the lifetime drops by the factor 100.
To avoid skidding of the cage, minimum friction between the rolling surfaces is necessary. It must guarantee the impellent of the cage, to overcome the resistance of the oil film against the friction of the rolling elements and the cage (Ill. 23.1.1-14.2). So, to avoid slip, in correlation to the diagram below right a minimum friction coefficient respectively traction coefficient is necessary. Understandably this is for a narrow lubrication gap (S =0,5) much higher and is reached markedly sooner than for a wide gap (S = 2,0). A minimum unbalance can be enough to prevent slip/skidding.

Ill. 23.1.1-9 (Lit. 23.1.1-7): In aeroengines, particles in the oil must be expected. So the typical high temperatures promote coke formation (chapter 22.3.2). Additionally contaminations of different origins can appear in the oil (chapter 22.3.3.1). Particles become failure effective, if they are thicker than the bearing oil film (Ill. 23.1.1-6). The thickness of the lubrication film under EHL conditions is in the range of 0,05 -2,0 µm (Ill. 23.1.1-8). In contrast a human hair is about 100 times thicker (sketch above left).
Does a dangerous particle (size, hardness) get between the races and rolling elements, in every raceway an indentation forms (sketch above right). Typical is a ring shaped bulging around the indentatation. This can break through the lubrication film. Mostly particles or its fragments are centrifuged or washed away. However, in some cases at least particle fragments got stuck in the raceway. At beginning failures they allow an identification. This is the precondition for targeted, prevention of such contaminations. Later the bulge will be plastically flattenend by rolling.
The bulge can be considered as the main cause for failures of antifriction bearings. The metallic contact disturbs the rolling cinematic and triggers sliding effects as well as stress peaks (Ill. 23.1.1-8). After some time from them a fatigue failure will develop (sketch below right). A shortening of the lifetime, up to a factor 1000 must be expected.
This failure effectivity of the bulge serves as justification for a polishing rework of the races. Tests showed: Are the bulges polished away, in spite of this the indentations will remain, the lifetime can be increased again about the factor 10-100. Naturally such a regeneration effect can be only used, if not already a markedly fatigue deterioration has occurred. To evaluate this is a problem which can not be underestimated. This may be the cause, why repolishing, if at all, can only be recommended for, from experience known sufficient low loaded, used bearings (e.g., in gears). Naturally such a rework seems especially for the expensive main bearings lukcrative. Basically rework must be carried out only according to the manual respectively instructions.

Note: The repolishing of the races, if necessary from used anti friction bearings, must be close to the specifications/instructions, respectively the manual. This appies as well for the process as also for the bearings (location of the application, type of the bearing).

Ill. 23.1.1-10.1, Ill. 23.1.1-10.2 (Lit. 23.1.1-8, Lit. 23.1.1-27 and Lit. 23.1.1-31): Demands for higher performance concentration and higher pressure ratios with as few compressor stages as possible, cause always higher rotorspeeds (Ill. 23.1.1-10.3). This trend affects the aeroengine bearings. This high-speed (diagramm right) with according high dm x n values (dm = diameter of the pitch circle from the center of the rollers) leads to special effects. Also little, in the usual machine engineering, unnoticed unbalances of the rollers (100 mg.mm are considered as an acceptable limit), now get failure effective. The rolling elements itself reach with this rotation speeds around their own axis in the range of 100 000 rpm. Because of this, for aeroengines the unbalance of the rolling elements is limited/specified and checked for every single one. Unbalances can be caused by little summarizing inhomogenties of the material structure (Ill. 23.1.1-10.2, sketch above) and geometrical unbalances respectively dimensional deviations within the specified tolerances (so no failure!). To this belong especially the edges of the rollers, which are also changed by wear during service (Ill. 23.1.1-19).

Unbalances of cylinder roller bearings: Under a static unbalance (frame above right), a symmetric dual mass unbalance at the roller axis in the resting condition is understood. The masses can be positoned at the same side or 180° offset. The static unbalance does not directly affect the operation behaviour of the bearing. However it can influence corrosion during stand still, at the contact surfaces of the rolling elements.
Of high importance is a dynamic unbalance during operation. It can trigger high loads at the edges of the rollers and flutter movements (skewing, weaving) around a radial axis (Ill. 23.1.1-19). Because of nonuniform load distribution, rollerweaving can cause high stresses at the contact surfaces to the races. Thereby the roller edges rub increasingly at the skirt of the bearing ring and the edges of the cage pockets. This leads to relatively high friction forces with a brake effect at the cage. The increased driving power at the cage, needed for slip free run, must be delivered from the rollers in the circumferential region of the load transfer. This causes high shear stresses at the load transferring rolling surfaces. Is such a penetration of the lubrucation film triggered, it comes to extremely local bearing temperatures and adhesive wear. The consequence is a high material stress with markedly shortened life time. Are radial loads too small for the drive forces of slipping rollers, skidding will occur (Ill. 23.1.1-14.1 and Ill. 23.1.1-15). In an extreme case the cage will stop for a short time. In this time period the anti friction bearing functions as a friction bearing.

Dynamic unbalances of the balls (sketch below in Ill. 23.1.1-10.2) are also dengerous. A ball should find for every revolurion a new race track. The corresponding pattern of the race track is called „ball of wool effekt“. However an unbalance can cause a stabilisation of the axis of rotation. Then the ball rotates aroud its axis of inertia. The most frequent overruns in the new fixed track shorten the fatigue lifetime during a adequate high load. This can be an explanation for the fatigue failure of a single ball in a bearing.
So the chance of a failure clarification with a microscopic follow-up investigatiion of a not too badly damaged ball surface exists. This chance should be given after test runs or certification runs of bearings and can point at unbalance problems.
The unbalances of a ball are caused by density deviations from enrichments of alloy elements/segregations, a special orientation of the material structure (sketch above in Ill. 23.1.1-10.2) as result of the forging process for the ball shape, or other inhomogenities of the material structure.

Ill. 23.1.1-11.1 (Lit. 23.1.1-28): There are two main groups of bearing materials in aeroengines. Through hardened, which are especially used in main bearings. Case haredned materials are used in smaller bearings of gears. These are even in some cases an integral element (Ill. 23.2.1-3) of gear shafts. To reach the material pureness necessary for a sufficient fatigue strength, the melting must take place in the vacuum (`V…' processes in the diagram). Indeed, such materials show no contaminations. But under very high dynamic loads (specified, i.e. weak points), inhomogenities like carbides determine the designed lifetime. Then they are the origin of fatigue failures.
Generally a historical trend to higher fatigue lifetimes can be noticed.
In both cases, the minimum hardness of the rolling surfaces from a bearing is abbout 58 HRc (Ill. 23.1.1-5.2).
Through hardened materials have been developed for aeroengine main bearings in the direction to high service temperatures. Concerned are alloys with molybdenum, tungsten, chromium, vanadium, cobalt, aluminium and silicium, which enable sufficient hardness. A representative of this type is `M-50' with 0,8% C, 4,0% Cr, 4,25% Mo for survice temperatures up to 315°C. Indeed the lifetime of the bearing rises at elevated operating temperatures, compared with lower alloyed materials. However, at lower service temperatures the bearing life time drops with the amount of alloy components (Ill. 23.1.1-11.2, Lit. 23.1.1-32).
In the last decades, primarily improvements have been achieved with material structure optimizations like grain orientation and alloying. To this belongs a residual austenite under 3% . For the use in corrosive atmosphere, high chromium contenting materials like`440C' with 1%C and 17%Cr are accessed. But this `corrosion -resistant' material (diagram) can only be used up to about 170°C. It also shows a weakness in the fatigue strength.
Even more corrosion-resistant bearing materials can be expected from highly nitridated Cr alloys (e.g., 15%Cr, 0,35%N, 1%Mo; Lit. 23.1.1-21). These show no weakness of the fatigue lifetime (Lit. 23.1.1-21).
Case hardened materials usually have a carburised hardening zone of > 0,4 mm thickness. Its hardness is in the range between 58 and 63 HRc. The core hardness lies under 48 HRc. They have a service temperature limit of about 170°C. From the high impact strength and fatigue strength is expected. This is especially interresting for parts with loads from outside (shafts). So these materials are predestined for the use in auxilary gears.

Ill. 23.1.1-12 (Lit. 23.1.1-28, Lit. 23.1.1-29 and Lit. 23.1.1-30): During stand still anti friction bearings can be deteriorated by vibrations. This is called „brinelling”. Such vibrations primarily occur during the transport of the whole aeroengine or its modules (Ill. 23.1.1-13). Because the, during service existing separating, damping oil film is displaced, during stand still the rolling elements under pressure of the rotor weight or a mass force get in direct contact with the race rings. So the not damped impact forces and micro relative movements act wearing (fretting).

We distinguish two variants:
'True brinelling' (upper frame) is a race deterioration at which the force transferring rolling elements produce plastic deformations in the race of the bearing ring. They require a lifting and pitching of the rolling elements. Such overloads during stand still, can be expected from shocks like shunting of wagons, vibration caused by rail joints or road holes. Are assembly devices with modules or aeroengines moved over an uneven hall floor (e.g., contact faces of concrete slabs), the danger of brinelling exists. Obviously also during ultrasonic cleaning, true brinelling was observed (Lit. 23.1.1-16).
`False Brinelling' (frame below) is an unevenness of the race at the contact zone of the rolling elements, caused by wear (fretting). They are the result of micro movements, caused from vibrations. During this process the rolling elements don't lift off. This deterioration does not require vibration forces, corresponding the true brinelling. A `normal' transport or vibrations during the cleaning process (acting of ultrasonic sound, volume 4, Ill. 16.2.2.5-7.1) are sufficient. Here the extreme case of standing contact fatigue (= SCF) should be mentioned. In this case a dynamic load produces through a contacting ball ring shaped fatigue cracks. This load is used in test devices for materials, especially of case hardened surfaces.

Ill. 23.1.1-13: Vibrations can cause true brinelling (Ill. 23.1.1-12) during the transport of not suitable supported aeroengines or unsufficient tensioned rotors. From experience, such deteriorations trigger, especially on small aeroengines with typical high speed rotors (e.g., APU, helicopter engine) already during start and run up catastrophic failures.

Ill. 23.1.1-14.1 (Lit. 23.1.1-10, Lit. 23.1.1-20, Lit. 23.1.1-25, Lit. 23.1.1-26 and Lit. 23.1.1-32): Cage slippage (skidding) primarily occurs at roller bearings. But there are also single cases, when this phenomenon concerns ball bearings. During skidding, the rotation speed of the cage is slower than the roll off of the rolling elements would require. So there are no clean kinematic conditions. Therefore sliding (slippage) of the roller elements on the race occurs. In an extreme case, the cage stands still and the bearing acts as a friction bearing.
The conditions for skidding are given, when the driving forces are not sufficient to equalise fhe braking forces which act on the cage (friction on the guiding faces, and the cage pockets). This situation can have several causes:

  • High hydrodynamic friction between guiding surface of the cage and bearing ring (outboard guidance, Ill. 23.1.1-17).
  • High mechanical friction in the guidance of the cage due to unsufficient lubrication or too high cage temperature during outboard guidance (clamping).
  • At too low radial load (Ill. 23.1.1-14.2) the driving forces on the rollers are not sufficient. This is the case without the required minimum unbalance.
  • High friction in the cage pockets. This can also arise at ball bearings, e.g., with thick silver sliding cooatings (shaping).
  • Weaving respectively flutter of rollers (skewing, Ill. 23.1.1-19). Thereby a contact between the rollers, the skirts of the bearing ring and the cage pocketds act with high brake forces at the cage.Because the centrifugal force of the rolling elements acts additionally against the contact pressing force at the inner ring, this is especially endangered from skidding failures.

The slipping rollers are extremely accelerated in the load transferring zone, during thre break through of the lubrication film (details below). The metallic contact at the face to the race ring takes place under high relative velocity. Thereby weldings (galling, seizing) occur, plastic deformations (detail above) and overheatings. At race rings and rollers forms a typical „dart pattern“ which grows over the circumference into extensive rough deteriorated areas (Ill. 23.1.1-15).
The driving forces pulsating with the circulation of the rollers, cause high stresses in the cage pockets (Ill. 23.1.1-18). The result is heavy wear/abrasion and fractures of the bars/lands.
As remedy against skidding serve measures, which guarantee every time sufficient radial forces. Usually this is a sufficient minimum unbalance. In especially critical cases, slight elastic deformations (e.g., ovalisation) of the outer bearing ring (Ill. 23.1.3-12) can be used.

Ill. 23.1.1-14.2 (Lit. 23.1.1-32): This schematic diagram shows the dependence of the cage slip (skidding) from the radial load of the bearing. These are the driving forces of the cage. The necessity of a minimum load can be seen . The slip rises with the rotation speed. This is explainable on the one hand with increased hydrodynamic friction forces, which act on the cage, on the other hand also the centrifugal forces may become noticeable. These add at high circumferential speeds remarkedly to the decrease of the contact pressure at the inner ring.

Ill. 23.1.1-15 (Lit. 23.1.1-10, Lit. 23.1.1-16, Lit. 23.1.1-20, Lit. 23.1.1-32 and Lit. 23.1.1-33): Failures caused by cage slip (skidding) have in the development phase a typical appearance. Dartlike dull zones form (sketch left). These are more rough than the unhurt race surface. In the micro range, with the help of the SEM evidence for a deterioration from the contact of the metallic surfaces under high relative velocity can be identified. Concerned are flat micro beak-outs (microspalls, Ill. 23.1.1-7), microcracking and plastic deformed regions (smeared grinding structure).
In the advanced stadium, the dull structures at the circumference grow to circumferential ring areas (sketch right), till they cover large areas of the bearing ring race (mostly the innwer ring, Ill. 23.1.1-14.1).
At the extension of the skidding deterioration small tongues („streamers”) can be identified. They follow the usual fine machining marks (Ill. 23.1.1-7). This appearance can be seen as an affirmation of the described failure mechanism.

Ill. 23.1.1-16 : During suddenly occurring unusual intense unbalances (Ill. 23.1.1-1) like:

  • the centrifuging of a rotorblade (sketch below) or of a rotor fragment (e.g., containment case, Ill. 2, chapter 8.2).
  • Extreme rub processes or compressor surge (volume 3, Ill. 11.2.1.2-1), it can come to shock loads on the bearing.

For a forced overload of the bearing two effects may get effective:

  • The penetration of the damping film, if a damped bearing is concerned, can elastically deform the outer ring oval. Also the lubrication film can be penetrated. Then a local metallic contact between the rolling elements and the bearing rings occurs. So high circumferential forces act at the roller elements and with this, at the cage and the bearing outer ring.
  • The elastic bending of the shaft and with this, the deflection of the axis of the inner, ring can cause the tilting of the rolling elements contact with the race. This triggers a sudden skewing of the rolling elements with a jamming effect (Ill. 23.1.1-19). Such an abrupt breaking/deceleration of the cage and the outer ring is similar to an overrunning clutch.


From experience, damped supported anti friction bearings (Ill. 23.1.1-2) without elastic anti-rotation device, show after such an incident forced failures. To these belongs (sketch above)

  • crack formation in notch endangered regions of the outer ring, like the transition of the fixing lugs to the outer ring.
  • Fracture of the lugs.
  • Fracture of the cage, respectively the lands/bars of the cage.


In such a case, the lands of the elastic fixing of a dampened bearing are highly endangered by a forced fracture.

Ill. 23.1.1-17 (Lit. 23.1.1-34 up to -36): For the operation behavior of a high speed anti friction bearing, the type of the radial cage guidence is of special interrest.
At the circumferential surface (bord), which serves the guidence of the cage, friction heat is procduced in the oil film and/or during mixed friction in direct metallic contact. This leads to heating and thermal expansion of the cage. It heats the cage, especially at a local sliding contact. The expansion, limited at a fraction of the circumference, triggers unbalances.
An inbord guidance hinders by centrifuging the supply of lubrication oil to the sliding surfaces of the cage. The heating expands the cage and so it takes off. This widening of the gap improves with an increasing oil entrance the cooling. At the same time the sliding surface, which must divert the friction heat, gets smaller. This acts stabilising at the cage temperature. Adversely is a worsened guidance, which causes an increasing unbalance. This can locally overload the cage. Outbord guided cages don't get unbalanced, even during heating.

The lubrication is promoted by the centrifugal forces and presses oil from the race into the lubrication gap. However, the centrifugal forces incease the friction forces and with this the heating. This leads to thermal expansion with the increase of the diameter and so the slow down of the cage, up to jamming during stand still. With this it can be expected, that bearings with outbord guided cage are also more susceptible for cage slip (skidding, Ill. 23.1.1-14.1). The increase of the friction forces is interactive connected with equivalent heat procuction. It comes to the deformation of the cage and high local loads. With the friction force on the cage, the necessary driving forces, which act from the rolling elements increase. This causes in the cage pockets high forces (Ill. 23.1.2-5) and friction heat.

Ill. 23.1.1-18 (Lit. 23.1.1-35 and Lit. 23.1.1-37): The sketches above show a summary of the operation forces on the cage and the rollers. This gives an impression of their complex interaction. In detail the following forces/loads and its causes are concerned:

  • At the inner and outer race rings, elastohydrodynamic forces (EHL, Ill. 23.1.1-6) act at the roller.
  • Centrifugal forces from the circulation in the center of gravity of the roller.
  • Unbalances of the rollers (Ill. 23.1.1-10.1).
  • Air drag of the rolling elements and the cage.
  • Loads which are transterred from outside through the bearing rollers.
  • Gravity of the rollers. This may rather be a stand still problem (Ill. 23.1.1-10.1). Corrosion/condensate aroud the gap at the contact poit of the ball/roller.
  • Circumferential orintated acceleration forces on rollers and balls (cage slip/skidding, Ill. 23.1.1-14.1).
  • Friction forces at the cage, which develop in the cage pockets, centering surfaces/guiding faces (Ill. 23.1.1-17 and Ill. 23.1.1-18) and by shear of the lubrication oil.

The forces act at the cage, rolling elements and race rings, whereat they influence each other. The oscillation of the balls, during passing the loaded circumference area (frame below), pumps dirt and aging products of the oil (chapter 22.3) between the two parts of a riveted cage. This can cause the expansion and loosening of the rivets up to their fracture.

Ill. 23.1.1-19 (Lit. 23.1.1-26, Lit. 23.1.1-27 and Lit. 23.1.1-37): Under roller skewing (gyration, weaving, flutter) a random circular movement is understood around an axis, which is not parallel to the bearing axis. This leads to a load at the roller ends/crowns with icreased wear at the roller faces, the cage pockets and the bords of the race rings. In an extreme case, the cage pocket gets so heavily worn, that the roller can rotate inside, fully around the radial axis (sketch in the middle).
We distinguish, accordant to the cause, the following types of roller skewing:
Misaligned skewing can already occur at a misalignment of 0,0002 mm/mm. Such misalignments appear at not sufficient careful assembled shafts or during operation durig the flexing of the shaft (unbalances).
Roller gyration skewing is caused by unbalances of the rollers (Ill. 23.1.1-10.1; frame below, left sketch). Typical shape determined causes are:

  • Wear of the edge radii/roller crowns.
  • Deviation of the axis of the roller crown from the axis of the roller (frame below, sketch in the middle).
  • The roller face is not sufficient rectangular to the bords of the race rings.

These influences must act diagonal at the opposite roller ends (frame below, left sketch).

Ill. 23.1.1-20 (Lit. 23.1.1-41): It is known, that electrostatic and electromagnetic caused currents, especially in industrial used turbo-machines can trigger failures (Ill. 23.1.1-24). Also aeroengines may have a potential for such problems. These effects can develop in two manners:

  • Elektrostatic charge by friction of non conducting rotor surfaces like ceramic coatings (interstage rings/spacers) and/or plastic components (e.g., fan blading) with the gasstream (principle of the Van-de-Graaf-Generator).
  • Elektromagnetic produced currents by a rotating magnetic field inside a metallic casing in which the rotor runs. Such conditions can exist at magnetised bearings, gears and shafts. Possible cause is a magnetic crack inspection without sufficient demagnitisation. Also electrical continuity `current flow' (pulsating) during lightning strike or unsuitable welding processes can magetise components. Electromagnetic produced currents can also originate from startergenerators of the aeroengines. During malfunction dangerous currents can emanate from them (Ill. 23.1.2-2).


This occurs without , eye-catching syptoms at the outside. The electric voltages between rotor and casing, respectively the suspension, get only alarming above a certain limit. It is very machine individual and depends on many influences. In such a case, only the OEM if at all, can give an advice.
Experience (without engagement):

  • Voltages below 1 Volt are uncritical.
  • At voltages above 5 Volt bearing failures must be expected.
  • Voltages above 20 Volt, trigger also gear failures and clutch/coupling failures.


With targeted voltage measurements (oscilloscope) at rotors, conclusions can be drawn if there is suspicion. Measurements have to be carried out at every single system which has a coupling (e.g., spline coupling). Thereby it is assumpted, that a current circuit is not always guaranteed. If there is accessibility, a suitable metal brush, like it is used in in mechanical rotating transducers (collector ring), can serve as probe/sensor.
The waveform of the voltage can enable conclusions at the cause. Voltage peaks let evaluate the danger.
Electrostatic shows sharp rectified voltage peaks in constant time periods. The signal depends from service conditons like pressure and temperature.
Electromagnetic produced voltages show a much more regular wave form. The frequencies can be assigned the harmonics/natural frequencies of the shaft rotation speeds. From experience here exist no comparable dependence of the electrostatic effects from the operation conditions. Potential results of circuit continuity are overheating/weld puddles from electric sparks at bearing races, gear teeth flanks and coupling faces. There was for the author no literature available about the endangering of the functions from always more frequent used electric components at aeroengines. However, it can be supposed, that unsufficient protected control units and sensors can react sensitive, respectively are endangered.

Ill. 23.1.1-21 (Lit. 23.1.1-11 and Lit. 23.1.1-12): Lightning strike represents a danger, especially for turboprop engines (volume 1, Ill. 5.1.3-4). Concerned are especially antifriction bearings and friction bearings as well as gear tooth flanks in the propeller gear (sketch above). Typical feature is a series of small weld puddles respectively in the advanced condition, fatigue out breaks (pittings). The failure mechanism is shown in the frame below. For the existence of deteriorating electric sparks, respectively arcs, a separating oil film is necessary, which is penetrated by them in short time periods.

Ill. 23.1.1-22 (Lit. 23.1.1-36 and Lit. 23.1.1-37): A good sliding behaviour of the numerous sliding surfaces (Ill. 23.1.1-18) from the cage of an antifriction bearing a silver plating provides. This is especially important for military aeroengines. Here, few minutes oil shut-offs are required. These consider hindered oilintake and oil hiding (Ill. 22.3-9) during especial manoeuvres in flight (e.g., upside-down flight). For such extreme´flight conditions, former cage materials are no more sufficient (bronze with platings of lead and silver). Therefore today cages from steel with a 0,025-0,05 mm thick silver plating are used. Beyond expectations, tests have shown, that a thicker silver plating (0,1-0,2 mm, diagram) acts adverse. Reason is the total moulding of the ball at the cage land. It comes to smearing of the silver and high friction forces. Obviously no sufficient oil film can develop under such conditions. The decelerating of the balls prohibits a kinematic rolling and so triggers `skidding' (Ill. 23.1.1-14.1). Further consequences are an extreme cage temperature and the fracture of the cage.

Ill. 23.1.1-23: Magnetised antifriction bearings can get in different ways failure effective.

  • Through attracted steel chips, which deteriorate the races as foreign objects.
  • Acting as generator and procucing currents which also deteriorate other components (Ill. 23.1.1-20).
  • Disturbance of electronics.To avoid such problems, it is necessary to identify the causes of the magnetisation.


Magnetic crack detection: If the necessary demagnetisation was imperfect, critical magnetic fields will remain.
Lightning strike: Pulsating cocurrents magnetise the bearing (Ill. 23.1.1-21).
Cocurrent flows caused from external currents like from welding. A further possibility is the magnetising by magnetic fields of electric driven transport vehicles. Therefore may an OEM explicit forbid such transports.

Ill. 23.1.1-24 Lit. 23.1.1-42 op to Lit. 23.1.1-45): Electrical continuity is a big problem of bearings. The currents can be associated with external/parasitic currents from electric machines (motors, generator, starters, Ill. 23.1.2-2). But they can be also produced by rotating magetised components (gears, shafts, bearings, Ill. 23.1.1-20). The danger may rise with the concept of the so called „more electric engine“. Here the starter-generator sits on a main shaft (at two shaft engines the high pressure shaft) and is driven by it (sketch above). Thereby the generator power of bigger fan aeroengines can be up to some hundred kW. The accessory equipment is no more driven by a radial shaft (tower shaft), but electric. A power line seves as connection. So a conventional accessory gear lacks. This configuration has some advantages:

  • Smaller face of the eroengine nacelle.
  • More favourable pisitioning of the accesory equipment.
  • Less maintenance effort.

Is alternating current with variable frequencies concerned (variable-frequency drive = VFD), what can be expected for the spectrum of the rotation speeds, an epecially high danger of electric bearing deterioration exists. This is connected with the voltage peaks of the electronic frequency control (thyristors).
The increased introduction of plastic components like rotorblades (fan) out of fibre reinforced plastic, can promote the electric charging in the airstream. Does the discharge take place through the bearings these are endangered. Therefore is the knowledge of the typical failure modes for current continuity important (Ill. 23.1.1-21).

'Frosting': This term comes from an macroscopic appearance of the failure zone, similar to frost (sketch below left). It is used for different failure causes and failure mechanisms like rolling fatigue (Ill. 23.1.1-7), circuit continuity or cavitation. To the naked eye the dull surface seems like dry blasted. Concerned is the beginning phase of the deterioration. Under the elektron scanning microscope (SEM) a multitude of tiny weld puddels can be seen. It is important not to confuse during macroscopic evaluation this failure with rather harmless running-in pittings (Ill. 23.1.1-7).

'Fluting': Applied is the so called „ripple formation”, an advanced failure mode. It indicates from the outside with intense vibrations. During the elektroerosion (electric discharge), longitudinal ribs have formed (flutes, sketch below right), which also are covered with tiny weld puddles.

References

23.1.1-1 I.E.Traeger, „Aircraft Gas Turbine Engine Technology, Second Edition“, Verlag : Glencoe/McGraw-Hill 1994, ISBN 0-07-065158-2, page 559-562

23.1.1-2 „The Jet Engine”, Rolls-Royce.plc., ISBN 0-902121-2-35, Ausgabe 1996, page 81.

23.1.1-3 A.F.Storace, S.J.Cline, „NASA-General Electric Energy Efficient Engine, High Load Squeeze Film Damper-System Analysis and Test Results“, Paper AIAA-84-1217 der AIAA/ASME 20th Joint Propulsion Conference, June 11-13, 1984/Cinncinnati, Ohio, page 1-9.

23.1.1-4 R.W.Shende, S.K.Sane, „Squeeze Film Damping for Aircraft Gas Turbines”, Zeitschrift Def.Sci.J. (Defense Science Journal), Vol.38, No.4, October 1988, page 439-456.

23.1.1-5 H.K.Lorösch, „Die Lebensdauer des Wälzlagers bei unterschiedlichen Lasten und Umweltbedingungen“. Zeitschrift „Wälzlagertechnik”, Heft 1981-1, page 17-21.

23.1.1-6 H.Schlicht, „Werkstoffeigenschaften, abgestimmt auf die tatsächlichen Beanspruchungen im Wälzlager“, Zeitschrift „Wälzlagertechnik”, 1981-1, page 24-29.

23.1.1-7 H.J.Böhmer, „Wälzverschleiß und -ermüdung von Bauteilen und Maßnahmen zu ihrer Einschränkung“, Zeitschrift „Materialwissenschaft und Werkstofftechnik”, 29 (1998), page 697-713.

23.1.1-8 H.Schlicht, E.Schreiber, O.Zwirlein, „Ermüdung bei Wälzlagern und deren Beeinflussung durch Werkstoffeigenschaften“, Zeitschrift „Wälzlagertechnik”, 1987-1, page 14-22.

23.1.1-9 O.Zwirlein, H.Schlicht, „Werkstoffanstrengung bei Wälzlagerbeanspruchung-Einfluss von Reibung und Eigenspannungen“, Zeitschrift „Werkstofftechnik”, 11, 1-4 (1980), page 1-14.

23.1.1-10 L.Engel, H.Klingele, „Rasterelektronenmikroskopische Untersuchungen von Metallschäden“, 2. Auflage, Carl Hanser Verlag München Wien 1982, ISBN 3-446-13416-6, page 160-168.

23.1.1-11 NTSB Identification: NYC961A036, „Incident Dec-07-95, Boeing 747-240”, page 1. (351-114)

23.1.1-12 NTSB Identification: DEN911A028 microfiche number 43589A, „Incident Dec-19-90, Boeing 727-22“, page 1.

23.1.1-13 NTSB Identification: FTW951A064, „Incident Dec-12-94, Aerospitale ATR 72-212”, page 1.

23.1.1-14 „Investigation Team Identifies Causes of CF6-80 Problem“, Zeitschrift „Aviation Week & Space Technology”, February 7, 1983, page 32.

23.1.1-15 H.Schlicht, „Über die Entstehung von White Etching Areas (WEA) in Wälzelementen“, Zeitschrift „Härtereitechnische Mitteilungen = HTM” 28 (1973) Heft 2, page 113-123.

23.1.1-16 „Metals Handbook, Ninth Edition, Volume 11, Failure Analysis and Prevention“, American Society for Metals (ASM), ISBN 0-87170-007-7, 1986, pages 1, 4, 490-513.

23.1.1-17 HK.Lorösch „Die Gebrauchsdauer von Wälzlagern hängt nicht nur von der Tragzahl ab”, FAG-Berichte aus der Firmengruppe „Wälzlagertechnik . Industrietechnik“, Heft 503DA, 1992, page 15-21.

23.1.1-18 L.Chang, C.Cusano, T.F.Conry, „Analysis of High-Speed Cylindrical Roller Bearings Using a Full Elastohydrodynamic Lubrication Model, Part 1: Formulation and Part 2: Results”, Zeitschrift „Tribology Transactions“ Volume 33, Number 2, April 1990, page 274-291.

23.1.1-19 V.P.Povinelli Jr., „Current Seal Design and Future Requirements for Turbine Engine Seals and Bearings”, Zeitschrift „Journal of Aircraft“ Vol. 12, No.4, April 1975, page 266-273.

23.1.1-20 B.A.Tassone., „Roller Bearing Slip and Skidding Damage”, Zeitschrift „Journal of Aircraft“ Vol. 12, No.4, April 1975, page 281-287.

23.1.1-21 F.J.Ebert, W.Trojan, H.W.Zoch, „Hochaufgestickter, martensitischer Wälzlagerstahl ist ermüdungsfest und korrosionsbeständig”, Zeitschrift „Wälzlagertechnik-Industrietechnik“ Fa. FAG Kugelfischer, Ausgabe 1992-503DA, page 9-21

23.1.1-22 L.Engel, H.Winter, „Wälzlagerschäden”, Zeitschrift „ANT-Antriebstechnik“ 18 (1979 Nr. 3, page 71-74.

23.1.1-23 H.Schlicht, O.Zwirlein, „Werkstoffeigenschaften und Überrollungslebensdauer”, Zeitschrift „ZWF“, 76 (1981) 6, page 298-303.

23.1.1-24 P.Dreschmann, K.Lorösch, R.Weigand, „Das Verhalten von Wälzlagern aus unterschiedlichen Werkstoffen bei ungünstiger Schmierung-ein Forschungsbericht”, Zeitschrift „Wälzlagertechnik“, 1983-1, page 34-39.

23.1.1-25 J.H.Brahney, „Film Thickness: the key to bearing performance”, Zeitschrift „Aerospace Engineering“, June 1987, page 19-22.

23.1.1-26 F.Greby, „What turbine technology is teaching us about High-Speed Roller Bearings”, 1972, 6 Seiten.

23.1.1-27 P.Nicolich, „Geräte zur Bestimmung der Unwucht von Rollkörpern der Fluglager“, Zeitschrift „Wälzlagertechnik-Industrietechnik” Fa. FAG Kugelfischer, Ausgabe 1990-501, page 39-43.

23.1.1-28 H.M.Flower, „High Performance Materials in Aerospace“, Verlag „Chapman & Hall”, page 163-173.

23.1.1-29 B.Alfredsson, M.Olsson, „Standing contact fatigue“, Verlag „Blackwell Science Ltd.”, Zeitschrift „Fatigue Fract Engng Mater Struct“ 22, 2. November 1998, page 225-237.

23.1.1-30 R.Errichello, „Another Perspective: False Brinelling and Fretting Corrosion”, Zeitschrift „Tribology & Lubriction Technology “, April 2004, page 34-36.

23.1.1-31 B.Frisch, V.Wigotsky, „Computer Programs, Guidelines für High Speed Bearing Design”, Zeitschrift „ Astronautics & Aeronautics“ April 1983, page 6-9.

23.1.1-32 K.T. O'Brien, C.M.Taylor, „Cage Slip in Roller Bearings” , Zeitschrift „ J.mech.Engng. Science“,15 (1973) , Nr. 5, page 370-378.

23.1.1-33 P.F.Brown, „Bearing Retainer Material for Modern Jet Engines” , Paper des „25th ASLE Annual Meeting“ in Chicago, May 4-8, 1970, Zeitschrift „ ASLE Transactions”,13 (1973) , page 225-239.

23.1.1-34 R.P.Shevchenko, „Lubricant Requirements For High Temperature Bearings“ , Paper 660072 des SAE „Automotive Engineering Congress” Detroit, Michigan, January 10-14, 1966, page 1-12.

23.1.1-35 R.G.Edge, A.T.B.P.Squires, „Lubricant Evaluation and Systems Design for Aircraft Gas Turbine Engines“ , Paper 690424 des SAE „National Air Transportation Meeting” New York, April 21-24, 1969, page 1-21.

23.1.1-36 T.Tauqir, I.Salam, A.ul Haq, A.Q.Khan, „Causes of fatigue failure in the main bearing of an aero engine“, „Engineering Failure Analysis 7”, (2000) Pergamon Verlag, page 127-144.

23.1.1-37 P.F.Brown, L.J.Dobek, M.J.Carrano, R.A.Valori, R.D.Daytonn, „Increase the Wear Life of Gas Turbine Engine Roller Bearings“, Paper der Proceedings AGARD-CP-323, „Problems in Bearings and Lubrication”, chapter 3-1.

23.1.1-38 F.J.Ebert, W.Trojan, H.W.Zoch, „Hochaufgestickter, martensitischer Wälzlagerstahl ist ermüdungsfest und korrosionsbeständig“, Zeitschrift „Wälzlagertechnik-Industrietechnik” Fa. FAG Kugelfischer, Ausgabe 1992-503DA, page 9-21.

23.1.1-39 E.Broszeit, F.Schmidt, H.J.Schröder, „Werkstoffanstrengung im Hertz'schen Kontakt infolge Last -und Eigenspannungen“, Zeitschrift „Werkstofftechnik”, Nr. 9, page 210-214.

23.1.1-40 J.Schmidt, W.K.Hank, A.Klein, K.Maier, „The Oil/Air System of a Modern Fighter Aircraft Engine“, MTU-München GmbH, page 1-20.

23.1.1-41 J.L.Caggianor, „Shaft Voltage Measurement-A Technique for Detecting `Electrical' Problems in Turbomachinery”, Zeitschrift „Turbomachinery International“,July-August 1983, page 26-31.

23.1.1-42 H.Boyanton, „Bearing Damage Due to Electric Discharge”, „Shaft Grounding Systems, Inc.“,1995, page 1-25.

23.1.1-43 „Stray currents can damage turbine bearings”, Zeitschrift „Power Engineering“,August 1997, page 1 and 2.

23.1.1-44 D.J.Powell, G.W.Jewell, J.D.Ede, D.Howe, „An integrated starter-generator for a large civil aero-engine”, Paper AIAA 2005-5550 der „3rd International Energy Conversion Engineering Conference“ 15.-18 August 2005, San Francisco, California, page 1 f.f.

23.1.1-45 R.R.Secunde, „Integrated Engine-Generator for Aircraft Secondary Power”, 1972, page 1 f.f.

© 2021 ITTM & Axel Rossmann
23/231/231.txt · Last modified: 2021/03/16 22:06 (external edit)

Page Tools