188.8.131.52 Damage due to HCF
Damages in the HCF range (crack initiation at load cycle numbers above 105) are a major problem in engine construction. This is related to the high utilization of strength reserves required by the demands for minimum weight. Primarily affected parts are compressor and turbine blades. This is due to their characteristic geometry and the wide variety of powerful potential incitements (e.g. flow, rotor, rubbing). Unlike LCF damage (Chapter 12.6.2), HCF damage can only be caught in very rare cases, through periodic inspections and/or life span limiting. This is due to the fact that it is usually not known when, and for how long, critical dynamic loads (usually resonance) act on the part. Therefore, the damaging load cycle number is also unknown. If crack initiation has occurred, then the typical (not necessarily) high vibration frequencies lead to uncontrollably rapid cyclical crack growth. An additional problem is the appearance of crack-initiating weak points during operation. These include erosion notches, corrosion marks, fretting damage, and foreign object damage. HCF damage has already been covered in the chapters dealing with different components. For this reason, the following is a selection of examples of typical damages intended to demonstrate the multiplicity of damage symptoms.
Serious HCF damages began occurring in the very first days of engine development ( to 184.108.40.206-2). These were usually vibration excitements in the resonance of the parts, caused by flow disturbances (Ill. 220.127.116.11-2). These disturbances can also be caused by external influences, such as at the compressor inlet. These include icing and ingested foreign objects (Fig. "Vibrations by flow disturbance at compressor inlet").
There is a large number of other excitement mechanisms (Chapter 18.104.22.168). Fig. "Imbalance by fan blade dynamic fracture" shows a mechanism in which a coupled disk-blade vibration acted in combination with a shaft deflection due to centrifugal loads to cause HCF damage to fan blades.
Shaft vibrations can be controlled in the design phase by avoiding resonances. However, if unforeseen changes in stiffness occur during operation of the rotor system, such as a thermal strain-induced loosening of the rotor assembly (Ill. 22.214.171.124.6), it can cause serious damage.
A typical cause of vibration in thin-walled sheet metal constructions (Fig. "Identifyig dynamic fatigue in metal sheet parts") is a sheet thickness that is less than the specified thickness (even if this is only local). This has a pronounced effect on the eigenfrequencies through the section modulus, and can result in resonance.
Even apparently massive parts such as cast turbine disks in small gas turbines are not immune to dangerous high-frequency vibrations ( to -16).
Damage-relevant air vibrations in ring-shaped spaces ( ) occur when the system creates a resonator or if a clearance gap flow acts as an excitement. This is the situation found especially in labyrinths (Volume 2, Chapter 7.2).
The damage process in dynamic fatigue fractures begins with crack initiation (Fig. "Material changes during crack development") due to dynamic fatigue. This is followed by crack growth and, after the critical crack length is reached (Fig. "Phases of fatigue fracture"), a forced residual fracture. The cause should always be sought in unexpectedly high dynamic loads or damage. The damage acts as a notch and/or weak point, around which the fatigue strength is below the dynamic operating loads.
Parts have specific stressed zones in which fatigue cracks can be expected (Fig. "Lifespan limiting by fatigue cracking").
Figure "Historic case of disk fracture by vibration" (Ref. 126.96.36.199-4): The early Whittle engine had a double-flow radial compressor (middle diagram). At 14,000 RPM it began to make a screaming sound. If the engine was operated for even a short time while making this sound, radial cracks would occur in the transition from the blades to the disk (bottom right diagram). Emitted sounds can still be used for verification of dangerous vibrations today, albeit in an altered form. During dangerously intense vibrations, the parts emit characteristic sonic frequencies. Frequency analysis is capable of using these sounds to detect fatigued parts.
The bottom right diagram shows the damage symptoms after the blade failure. The blades of the axial turbine also fractured due to high-frequency vibrations (also see Fig. "Vibration dampers at roots of turbine rotor blades"). In order to avoid this damage, important methods for analyzing excitement possibilities were developed at the time. These include the Campbell diagram (Fig. "Development preventing vibrations"). In addition to the typical expected fractures at the blade transition to the root platform (fundamental flexural modes), dynamic fatigue fractures also occurred in unusual/unexpected areas of the turbine rotor blades, such as the center or tip of the blade. Whittle identified the cause as a thermal element that was located in the exhaust gas flow roughly one meter behind the turbine. The realization that flow disturbances can be located far behind the excited part is always startling. After the disturbance was removed, the dynamic fatigue fractures no longer occurred.
Figure "Vibrations by remote engine copmponents" (Ref. 188.8.131.52-5): In January 1941, the engine Jumo 004A was successfully accelerated to full thrust during development. At this time, dynamic fatigue fractures of the rotor blades were a problem (also see Fig. "Vibrations by shaft caused flow disturbance"). The affected blades were compressor stator vanes made of sheet metal and without shrouds.
In the summer of 1943, several dynamic fatigue fractures occurred in turbine rotor blades (also see Fig. "Vibration dampers at roots of turbine rotor blades"). They were incited through the six can-type combustion chambers and the second harmonic of the three housing braces behind the rotor. The 36 stator vanes between the combustion chamber and turbine exacerbated the problem. The eigenfrequency of the blades was estimated by a musician by the sound they made when rubbed with a violin bow. This was an early form of modal analysis. The problem was solved by tapering and shortening (1 mm) the rotor blades and reducing the maximum RPM.
Figure "Vibrations by flow disturbance at compressor inlet": Vibration-inciting flow disturbances in the compressor inlet can be caused by external influences in many different ways. Typical causes are:
- FOD (e.g. bird strikes)
- Uneven distribution of temperature, pressure, and velocity in the inlet flow (Chapter 184.108.40.206).
- Shifting of the inlet cross-section through foreign objects (Example "Flow disturbance"),
- Icing (Volume 1, Ill. 5.1.4-2).The poor visibility into inlet ducts favors the accidental disposition of foreign objects (e.g. tools).
Excerpt 1 (Ref. 220.127.116.11-6): “…The …incident involved ingestion of a cardboard food container. the material blocked a compressor inlet, and in the course of 10 hours operation, the blockage “excited blade failures”, apparently due to excessive heat.”
Excerpt 2 (Ref. 18.104.22.168-14): “…sustained compressor damage from ingestion of a cardboard food container at an earlier time…“
Comments: Although both excerpts evidently concern the same incident, their statements are remarkably different. Overheating of the compressor blades due to flow disturbance would only be thinkable following a surge and expansion of the combustion chamber into the compressor. The destruction of the compressor after about 10 operating hours does not indicate the sudden entry of the foreign object into the compressor after this time. Rather, it can be assumed that the flow disturbance ahead of the compressor led to HCF blade damage.
Excerpt: ”…(The OEM) is making additional modifications to the second stage of its…turboprop engines following a compressor blade failure…Aircraft operated by (the same Operator) were involved in three incidents in a four-day period…The …blade failure was the fifth incident involving …(this type of) engines….(One) incident…occurred during a landing approach on a training flight… and apparently stemmed from a blade failure in the second stage…Following the first three blade failures,…(the OEM) instituted a fix that increased the compressor blade clearance by machining material from the engine casing. The objective was to eliminate blade scrubbing as a source of blade stress…
Since the engines that sustained blade failures (on two different aircraft types) had been fitted with machined cases for the increased blade clearance,…(the OEM) has to implement other fixes aimed at improving the vibratory stress margins in the second stage blades.
The failures have occurred at various power settings, but the maximum aerodynamic excitation effect on the blades seems to be occurring at off-design intermediate settings in the 80-90% range. The fixes, which have been designed to have the most effect at that operating point, include:
- Reducing the blade angle of incidence by about 2 deg. at the tip to ease aerodynamically induced vibration on them.
- Shot-peening all second-stage blades to reduce residual material stresses in the blades and tailor the metal to withstand higher levels of vibration.
- Changing the blade anti-corrosion coating from one that is put on by the deposition method to a painted coating.
The shift in the method of coating the blades was necessary because it was found that the deposition process of applying the anti-corrosion material negated the conditioning effects of the shot-peening treatment…Testing has verified, that none of the blade fixes, including the increase in blade clearance, has had an adverse effect on the engine efficiency..(the OEM said).”
Comments: This example conveys the impression that only the HCF failure of the compressor rotor blade in stage two is not disputed. If one goes by the implemented remedies, the rubbing of the blade tips on the housing does not seem to have been the only cause of the dangerous blade vibrations. Evidently there is a resonance when the engine is started up. Because this is given quite specifically at 80-90 %, the flow disturbance caused by the shaft that sticks out the front of this turboprop engine. It is also plausible that a rotating stall occurs at the blade tips in this RPM range and excites vibrations. This would explain the correction in the angle of incidence of the blade tips. The other measures are also generally an indication of what the OEM “really thinks”. These are measures to increase dynamic strength, even when no rubbing occurs. The change in the corrosion protection shows that the blade material is probably a type of high-alloy heat-treated steel.
Figure "Vibrations by shaft caused flow disturbance" (Example "Dynamic fatigue fracture in a fan blade"): Evidently vibration problems in blades are not always eliminated in the development phase. This is especially likely if the application of the engine changes. For example, if a helicopter engine is used as a turboprop, then forces from the propellor can act on the filigreed engine. Unexpected problems with clearances and rubbing occur due to housing deformations that are dependent on the peculiarities of operation and the aircraft type. If the clearances are too tight, rubbing can excite unallowably strong vibrations in the blades. If the blade tip clearances are too large and/or uneven, they can excite blade and shaft vibrations (Fig. "Self exciting rotor vibration"). Another excitement mechanism that is directly related to the tip clearance is the rotating stall (Fig. "Damages by rotating stall"). If there are flow disturbances in the compressor inlet, such as a power takeoff shaft (top diagram), they can have a damage-promoting effect in combination with the described influences.
Figure "Imbalance by fan blade dynamic fracture" (Example "Increased dynamic loads under climbing flight conditions"): In this case, a commercial aircraft crashed due to a dynamic fatigue fracture in a fan blade. The blade vibrations were explained by coupled, self-increasing vibrations of blades and disk in connection with shaft deflection (whirl, also see Ills. 22.214.171.124-9 and 126.96.36.199-15). In this case, it was evidently not possible to recognize this damage mechanism in testing rig tests on the ground during engine certification before serial use.
Excerpt 1 (Ref. 188.8.131.52-2): “…as the aircraft was climbing…the crew experienced moderate to severe vibration and smell of fire. The area microphone for the cockpit voice recorder (CVR) picked up a sound of vibration or `rattling' at this time and the flight data recorder (FDR) showed significant fluctuations in lateral and longitudinal accelerations. ..Replay of the FDR showed that severe vibration had occurred in the No 1 (left) engine at this time, accompanied by marked fluctuations in fan speed (N1), a rise in exhaust gas temperature (EGT) and low, fluctuating, fuel flow….
Ground witnesses who saw the final approach saw clear evidence of fire associated with the left engine. The intake area of the engine was filled with yellow/orange fire, and flames were observed streaming aft from the nacelle, pulsating in unison with `thumping noises'. Metallic `rattling' was also heard, and flaming debris was seen falling from the aircraft….
The fracture surfaces of all fan blade failures were examined metallurgically and,with the exception of the failure of blade No 17 above the mid-span shroud, were all found to be overload failures, consistent with either bending or a combination of tension and bending. The failure of blade No. 17 was observed to be very flat over the forward half chord, and exhibited several features consistent with fatigue, but was covered by a carbonaceous deposit which was very resistent to chemical cleaning. This deposit was analysed in the scanning electron microscope (SEM) and samples taken for chemical analysis. These analyses showed that the deposit was consistent with an amalgam of the materials present in the fan case, intake, acoustic liners and the fan abradable liner….
Metallurgical examination of blade No. 17: The characteristics of the fracture were those of high cycle fatigue which was deduced to have originated from a point on the concave (pressure) face of the blade, about 1 to 1.5 mm aft of the original leading edge (LE). The fatigue extended aft to about mid chord (ie above the mid-span shroud). were the fracture mechanism had changed to overload tension…
If this fracture had occurred first, the initiation of Event 1 would have been the sudden separation of a single fan blade outer panel. This would have induced a localised disturbance in the air flow, coupled with severe imbalance, both of which can cause the onset of a fan stall, and subsequently a booster and core stall. The severe imbalance would also have caused blade tip and air-seal rubbing throughout the engine, with a consequent degradation of engine stall margin…
Cause of the fatigue initiation in fan blade No. 17:..the fracture had propagated initially by fatigue...as a result of inter-blade clashing, made it impossible to identify any surface feature which may have led to the fatigue initiation. However, the depth of the material removed from the pressure face indicated, that if there had been an anomaly at the origin of the fracture it must have been very small...
However, after metallurgical examination of the two fan blade fractures suffered by G-BNNL and G-OBMG (two similar incidents on other aircraft) had revealed a number of features common to each other and to that from ME, it had become clear there was a generic problem affecting the (…variant of the engine type). Both of the later fractures had originated at the base of the shot-peened layer on the blade pressure surface, close to the mid-chord position…
Source of vibratory stresses: After the two later fan blade fatigue fractures, it became suspected that the fan was being subjected to higher vibratory stresses than were thought to exist….(the similarities) indicated, that there was a unique vibratory mode involved which was excited regularly. Following the test bed research, the strain-gauged blade tests conducted on a flight test engine confirmed the presence and mode of the vibratory response and showed that it could produce stress levels approaching the endurance limit.
Excerpt 2 (Ref. 184.108.40.206-3): ”..Initial reports indicated a single engine failure, but subsequent information now shows that one engine did indeed fail and the good engine was also shut down. Investigators believe that the pilot shut down the right engine when a fire indicator in the aircraft controls flashed an alarm. However, it was the left engine that actually failed. The right…(engine) may in fact have had problems as well but the probability of two engines with less than 500 flight hours each failing at the same time is astronomical: more than 1 million to one.
Comments: Several comparable fractures of the top section of a fan blade occurred. In all cases, the fractures were HCF fractures that were evidently not traced back to flaws in the crack incitement zone. The damage cause was increased dynamic loads under certain climbing flight conditions. These vibrations (Fig. "Friction damping influencing vibrations") evidently did not occur during the certification tests of the affected engine, which were conducted on the ground. They were only verified under the specific flight conditions in engines with fan blades that were outfitted with DMS. The catastrophic crash occurred because the remaining undamaged engine was accidently also shut down.
Figure "Overheating and fusing at shafts by vibrations": The components (compressor and turbine disks) of this rotor from a small shaft-power engine are axially fixed with two centrical clamping bolts. The turbine-side clamping bolt had a stiff spring at the free end (top diagram). This was intended to ensure a sufficient axial pre-load. In addition, thermal strain differences between the turbine disks and the centrical clamping bolt had to be controlled, especially during fast instationary processes. During testing runs in the development phase, strong vibrations occurred during instationary operating states such as startup and shutdown. At almost constant rotor RPM, the frequency of the externally measurable vibrations changed. This appearance was explained by unmatched stiffnesses and thermal strain in the clamping bolt and compressor disks. Evidently, the assembly periodically loosened and changed the rotor eigenfrequency. This concerned very intense rotor vibrations, during which the contact surfaces between the compressor rotors experienced large relative movements. Overheating in the form of local welding beads occurred. Evidently large, high-frequency relative movements of the contact surfaces had caused this overheating (bottom diagrams). The overheated zones were evenly distributed around the circumference. Therefore, this at least temporarily involved a wave with several nodal diameters that travelled around with the rotor (Fig. "Vibration models of disks").
The solution was a spring on the free end of the clamping bolt on the compressor side (middle diagram).
Figure "Vibration dampers at roots of turbine rotor blades" (Example "Dampening vibrations to tackle fatigue cracks"): Dynamic fatigue fractures in turbine rotor blades have been a problem from the very beginnings ( ) of engine development. Unfortunately, in this case of a helicopter engine (middle diagram), the available literature gives no information regarding the source of the excitement, nor does it mention the turbine stage in which the fractures occurred. The excitement evidently occured via a flow disturbance and its scope was not recognized in the development phase. It is possible that the rotor blades of the first HPT stage (bottom right diagram) were excited by the turbine stator vanes in a certain RPM range. The detailed diagram at bottom left is a schematic depiction of the dampers made of metal sheeting that were introduced as a remedy for the problem (Fig. "Influences on damping of parts").
Excerpt (Ref. 220.127.116.11-8): “The two year, $56 million retrofit of more than 1,000 engines is to correct a chronic malady…(of the engine type that powers a military helicopter) and a related model…The engine failure that preceded the fatal crash of a …(helicopter) last year had occurred about 150 times previously in other army helicopters, army officials said. No other deaths have been attributed to the problem, and the twin-engine helicopters are designed to land safely with one engine operating. Service officials have known for at least six years that the engines that power early …(helicopter) versions…were prone to fatigue cracks of the gas generator blades…The gas generator is at the front of the turbine. Thus some cracked and weakened blades would break entirely and be sucked through the engine, damaging other parts.
The problem occurs in engines built before 1989. Prior to the…crash, the malady was being corrected during scheduled depot overhauls by attaching 2-centimeter-long metal springs to the rotor blades to control vibration that lead to cracking, service officials said. Engines made after 1989 were fitted with the dampers during production…“
Comments: HCF fractures (dampers as a remedy) evidently occurred following resonant vibrations of the turbine blades. The effectiveness of the dampers indicates that the vibrations only occurred briefly, most likely when the turbine was starting up. The dampers shown in Fig. "Vibration dampers at roots of turbine rotor blades" are only intended to illustrate the principle, and may be different from the ones actually used.
According to the literature, a fatal helicopter crash occurred after about 150 parallel damage cases. In this case, operation continued as allowed after one engine failed. However, the “improbable” situation of the second engine failing occurred. This can be seen as an impressive example of the problems with arguments that hold that a frequent damage type without previous catastrophic consequences proves that it is innocuous.
Figure "Identifyig dynamic fatigue in metal sheet parts": Thin, lamellar sheet metal constructions under high-frequency vibrations (sheet vibrations) typically exhibit branching cracks and material break-outs (top diagram, Volume 2, Ill. 7.2.2-24). In a turboprop engine type, these dynamic fatigue cracks occurred in the exhaust pipe flange after years of damage-free operation. This pipe was made from sheet metal that was about 1 mm thick.
The damage investigation revealed that the damage only occurred in newly installed parts. Their sheet metal thickness was within the specified tolerances, but was at the lower limit, unlike the older parts. Dynamic inspections revealed that the lower stiffness of the somewhat thinner metal sheets promoted dangerous vibrations.
Similar damage was observed in thin-walled labyrinth cones of another engine type (Volume 2, Ill. 7.2.2-24)
These examples demonstrate the influence of the sheet thickness on the stiffness and therefore also on the vibration-sensitivity of a part. This is influenced in two ways:
Changes in the eigenfrequencies: While the cube of the thickness goes into the stiffness, the frequency only increases slightly overproportionally when the vibrating mass does not become considerably larger, as well. This is the case if the change in wall thickness is limited to the fastening point (bottom right diagram). Contrary to expectations, this situation occurs relatively often in practice. It occurs in flange areas, for example, if the welded-on sheet metal wall has been deep-drawn for the necessary geometry.
Elastic deflection and stresses under excitement forces: This changes considerably with changes in stiffness. Excitement forces in thin-walled sheet metal structures are generally pressure vibrations in the gas flow and mechanical vibrations.
Therefore, it must be ensured that sheet metal thicknesses are correct during new part manufacture, repair, or when considering replacement parts. If in doubt, the guiding policy should be to use configurations that have been proven in operation.
Figure "Vibrations excited in ring shaped spaces 1" (Ref. 18.104.22.168-10): The fighter aircraft engine depicted in the middle diagram showed a rotor-dynamic instability when it went into serial operation in the mid-1980s. Up to 50% of the engines were rejected during the acceptance phase due to very strong non-integral vibration (NIV). Only the introduction of an oil-damped high-pressure turbine (HPT) bearing (bottom diagram) solved the problem.
The predecessor of the affected engine had already been in operation for years without similar problems. The rotor instability coincided with the introduction of the new HPT, even though the rotor dynamic was not noticeably affected by this.
Comprehensive investigations of over 30 engine configurations were conducted in order to understand the damage mechanism and the remedy and its verification. Differences between the cooling systems of the two engine variants were especially interesting. In the affected variant (top diagram), the cooling air (“D”) for the HPT rotor blades jumps over a tangential onboard injector (TOBI; “E”) on the front side of the first rotor disk (“F”). The labyrinth seals “A” and “B” form a ring-shaped space “C” ahead of the turbine blade, which distributes the cooling air onto the blades. The ring chamber “C” directs the air to the blades through holes in the plate that carries the rotor of the labyrinth “A”. Measurements in the ring chamber “C” showed a pressure vibration with the NIV frequency. This revealed that changes in the fan inlet pressure (engine inlet) influenced the amplitude of the vibrations. This led to the damage hypothesis of a cross-coupled self-increasing excitement/instability in the HPT system.
An analytical investigation dealt with the possible causes of instability (middle left diagram) in the high pressure rotor.
One characteristic was fretting wear in the flange threads (“G”) between the turbine disk and the shaft. This wear indicates that internal friction in this flange connection promoted rotor vibrations (Ills. 22.214.171.124-10 and 126.96.36.199-16). Internal friction of the flange can promote instability if the shaft that rotates in itself begins deflected rotation (“whirl”, Ills. 188.8.131.52-13 and 184.108.40.206-15).
It must be noted that the cross-coupled stiffness incites a force in the direction of rotation of the deflected shaft (orbiting movement of the whirl), creating an instability. In order to quantify the forces responsible for the instability, it takes a good “engineer`s sense” (!) in addition to the mathematical analysis. It was discovered that the frequencies of the system react sensitively to the stiffness of the suspension of the first turbine stator assembly.
A second possible instability-promoting influence was the different labyrinths around the HPT. Labyrinth seals affect the stiffness and damping of the system both directly on integral rotor components and via gas coupling. In contrast, the Alford force (Fig. "Tip clearance influencing 'orbiting'") only acts via gas coupling. The stiffness and damping of a labyrinth depend on clearances, pressure drops at the labyrinth, geometry, and rotating speed (see Volume 2, Chapter 7.2.1). They are not easy to accurately adjust due to the many influences.
A third suspected influence was a reciprocal aerodynamic influence between the instability and TOBI (top diagram). The pressure measurements in the ring duct “C” revealed a considerable influence of the turbine on the temporal pressure progression in the gas flow. This observation led to the assumption of a sinusoidal pressure vibration in the ring duct (Fig. "Vibrations excited in ring shaped spaces 2"). In this way, a moment was able to act on the turbine.
The individual instability-promoting influences were analyzed in tests in which specific design characteristics were changed for each test.
The final analysis of the analytical observations and the tests resulted in the following conclusion:
- It is extremely difficult to determine the cause of a non-integral vibration. Estimates demand iterative analytical and test-technical steps.
- The primary cause for the vibrations was an aeromechanical coupling of the cooperation of three instability-promoting propulsive forces in roughly equal amounts: Alford forces (Fig. "Tip clearance influencing 'orbiting'"), influence of TOBI, forces related to the labyrinth seals of the ring duct “C” and pressure vibrations in the cooling air inflow. Other realizations were:
- Variation in the influences can lead to variously strong vibrations (Fig. "Cyclic spin test disk fatigue crack ").
- Instability-promoting influences can be surpressed with sufficient safety through the use of a specifically designed damped bearing (middle right diagram, Fig. "Estimating allowable fatigue loads").
Figure "Vibrations excited in ring shaped spaces 2" (Refs. 220.127.116.11-10 and 18.104.22.168-12): This diagram shows the results of tests to determine the damage mechanism of the case described in Ill. 22.214.171.124-9.
In the ring duct between the TOBI and the turbine disk (top left diagram), a gas vibration forms that has six nodal diameters and rotates non-integrally relative to the shaft (top right diagram, Refs. 126.96.36.199-12 and 188.8.131.52-13). This is excited by a labyrinth vibration at low fan inlet pressure, and therefore low pressure levels around the TOBI, via an increased clearance gap flow (also see Volume 2, Chapter 7.2).
The instability-promoting influences in the various test variations of the engine, ring duct, and the surrounding labyrinths are assigned to the involved components in the bottom diagram:
“I” highest fan inlet pressure
“II” reduced fan inlet pressure
“III” inner static labyrinth carrier removed
“IV” opposite facing labyrinth tips (detail)
“V” as “III”
“VI” complete inner labyrinth seal removed
Figure "Rotor deflection during flexural mode" (Ref. 184.108.40.206-20): This diagram shows an example of the movement of a deflected rotor (orbiting). The available literature does not state which forces caused orbiting in this case.
Orbiting is a process in which a rotor is deflected by flexure, similar to situations involving a bent flexible shaft (Fig. "Orbiting mechanism and cause"). In contrast to orbiting, the deflection of an unbalanced shaft occurs at RPM rate of the shaft. The frequency of the flexure (orbiting frequency) is considerably below the rotor RPM, so that the rotor rotates around its bent axis during orbiting (Ill. 220.127.116.11-15).
In an 18-stage axial compressor (top diagram is only suggestive), rotor deflections of up to half a millimeter occurred in the eighth stage (bottom diagram). The rotor was running at 7700 RPM (128 s-1), which is 100% RPM. The vibration frequency was attributable to the fundamental flexural mode of the system: rotor/bearing support. The tip clearance was measured at the 12 - and 3-o`clock positions (top right diagram). The orbiting frequency was in the range between 50 and 65 Hz in several tests, which is equal to 30% to 50% of the rotor RPM. The greatest deflections were recorded at the higher frequencies.
The primary cause was determined to be an overly soft rotor. Stiffening the rotor and the housing solved the problem.
Figure "Cyclic spin test disk fatigue crack ": The vibration excitement of a rotor disk can also occur with flexural modes of the rotor. In the depicted example, a specimen disk was cyclically stressed in a vacuum vertical centrifugal testing rig by an annulus that had been “drained” by keyhole bores (to simulate blade loads). After a short operating time, penetrative crack detection discovered several cracks (right diagram) in the annulus area that were primarily radially oriented and periodically distributed (bottom left diagram). The findings from the crack surface inspection indicated dynamic loads in the HCF range. This was plausibly explained by a disk vibration with three nodal diameters. The vacuum in the testing rig made vibration excitement due to aerodynamic forces impossible. However, there was also no effective air damping. Therefore, the only possible excitement was from the testing rig spike that powered the disk. The spike is elastically suspended in the testing rig on the free end, together with the disk. In the case of imbalance, it runs into a safety bearing. Evidently the rotor vibrated within itself in a way that no externally noticeable imbalances occurred.
Illustrations 18.104.22.168-13 and 22.214.171.124-14: This case occurred during the development phase of a small, low-performance turbine for a helicopter (top diagram). Without any indication of imbalances or other anomalies in operation, the integral cast turbine disk of the last stage burst at full RPM in a steady operating state. Fragments of the annulus had broken off the disk. The entire fracture surface could be seen as a combination of two fractures that ran together at an angle to one
another that was tangential to the balancing ring (bottom diagram). The fracture surface analysis revealed a bow-shaped limited fracture zone in the middle of each of the annulus fragments near the balancing ring. It also showed stage 1 characteristics (cleavage cracks) that are typical for this Ni-based cast material (Fig. "Stage 1 at dynamic fatigue cracks"). A testing-technical vibrational analysis (Fig. "Experimental determinating disk vibration modes" ) with heaps of powder resulted in a distribution pattern that corresponds with the bottom left diagram. This Chladni figure showed that a three nodal diameter vibration (Fig. "Vibration models of disks") is excited exactly at full RPM by the three bearing braces of the exit housing located behind the disk.
A successful provisional solution was the installation of a disturbing brace with no load-bearing function. It was only intended to dampen the vibration in the right rhythm with a disturbing pulse. The final solution was four equally-spaced supporting braces, from which no dangerous excitement was expected (middle diagram). The effectiveness of this solution was confirmed by later experience.
Figure "Fixing of centered rotor damping bolts": “Mysterious” fractures of the first stage turbine disk occurred several times during the development phase of a small shaft-power engine (middle diagram). Every failure occurred after very few operating hours. The fracture split the integral turbine disk, which was made of a Ni-based cast alloy with no central bore (!), completely in half (bottom diagram). At the base of the thread for the compressor-side clamping bolt, a zone with facet-like fracture structures was located almost in the center. These fracture structures are a material-specific sign of a dynamic fatigue crack. In the short operating times before damage occurred, only a few startup/shutdown cycles had taken place. Therefore, a high-frequency dynamic load in the shape of a one-nodal-diameter vibration of the turbine disk was suspected to be the cause of damage (bottom left diagram).
The causal vibration excitement of the very solid disk most likely occurred through flexural modes of the relatively thick and, therefore, heavy compressor-side clamping bolt. These vibrations can induce very strong axial forces on the ends where the bolt is fixed (top diagram and bottom example). This effect is even used in dynamic testing machines. The flexural mode of the bolt probably resonated with the vibration of the turbine disk. However, with such an intense excitement mechanism, it is also thinkable that there were other dangerously large loads apart from the resonance.
Figure "Fraction of turbine disk by nodal cycle vibration": This case concerns the fracture of an integral cast turbine disk from the second stage of a low-performance helicopter engine (top diagram). The fracture occurred following the initiation of several cracks at the circumference in the transition radius of a balancing ring (bottom diagram). The explanation for the damage was a nodal circle-vibration with a nodal diameter (Fig. "Vibration models of disks"). This vibration could have been excited by a neighboring disk that is connected to the damaged part by the centrical tie rod. The turbine disks of the first and third stages of the same engine type had fractured several times due to disk vibrations ( ).
Figure "Vibrations at seal membrane": In the area of the labyrinth bracket below the stator assembly of the first stage of the low-pressure turbine (top right diagram) of a large shaft-power engine, several damaging dynamic fatigue fractures occurred in the HCF range following a constructive change. These were dynamic cracks along the circumference (bottom right diagram) on the front wall of the ring duct that is formed by the seal bracket (top right detail).
The constructive change connected with the damage was a wire seal ring between the stator vanes and the labyrinth bracket. The seal effect against leakage air flow from the space with higher
pressure levels ahead of the wall to the ring duct merely relied on the resilient force of the disk-shaped front wall of the labyrinth bracket. After several hundred operating hours, the seal ring and the contact surfaces of the seal groove in the labyrinth bracket showed heavy fretting wear (Volume 2, Chapter 6.2).
The damage cause was thought to be vibration excitement of the resilient seal wall following sufficiently advanced fretting wear. This excitement principle (hydrodynamic paradox, Ref. 126.96.36.199-11 and Fig. "Vibrations of flat seals") is based on the resilient action of the cover and the pressure drop in the leakage air flow (Bernoulli) and is viewed as an exemplary physical laboratory experiment (bottom left diagram).
The damage hypothesis was confirmed with the aid of a simple demonstratory experiment. A stator assembly with a seal ring was set up and realistic pressure ratios similar to those in operation were simulated. As predicted, with sufficient wear and reduced resiliency, a strong vibration excitement of the covering plate with a corresponding pulsating leakage air flow could be observed.
After the damage mechanism was understood and reproduced using original parts, the requirements for a guaranteed solution were met.
Figure "Fatigue crack at compressor stator": In a small shaft-power engine (right diagram), compressor damage occurred during a trial run in the development phase. The inlet edge of the radial compressor disk of the second stage had broken out following an HCF dynamic crack (left detail). The dynamic crack did not have any pronounced lines of rest, which indicated constant crack growth. In addition, the stator (sheet metal) located behind the damaged disk had several dynamic fatigue fractures in its vanes and fastening bolts. This damage had already occurred frequently without cracks in the compressor disk. This indicated the following damage sequence:
First, the vanes of the compressor stator broke due to dynamic overstress (bottom diagram). Break-outs evidently created strong flow disturbances that acted against the airflow into the compressor disk. Here, it resulted in vibration excitement in the blade in a manner typical of the fundamental flexural modes of radial compressor blades, with crack initiation in the inlet edge area.
Figure "Gear excited fatigue fractures at cooler disk": After blades of an oil cooler disk made from a high-strength forged Al alloy (about 12 cm diameter, 20,000 RPM) were frequently affected by dynamic fatigue fractures, disks made of fiber-reinforced synthetic material were introduced. These disks, which are made from short-fiber molding material (fiberglass/epoxide), ran for a long period without problems. In one case, however, a disk suffered a dynamic fatigue fracture (top left diagram) despite the high inner damping of the synthetic material. The disk was replaced by an identical new part. This disk also fractured in a similar way within about 100 hours. This made a damage-causing flaw in the disk unlikely.
The damage anlysis revealed that a toothed gear in the drive chain (bottom right diagram) was probably damaged during installation. The damage was a hardly noticeable deformation of the edge of the roll-off surface. This “minor” tooth damage evidently led to the extreme dynamic loads. After the gear was replaced, no further dynamic fatigue fractures occurred in the oil cooler disk even after long run times. This example demonstrates the danger of vibration excitement due to tooth frequencies.
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