"Illustration 2.5-3": There are many recommendations, standards and specifications for the use of vibration measurements, monitoring and the analysis in industrial applications, especially for gas turbines. They are published by national and private organisations as well as manufacturers (OEM). In the USA there are:

• ISO (International Organization for Standardization). In this context belongs the TC 108 for ‘mechanical vibration and shock’. The ISO/7919 and ISO/10816 refer amongst others to turbo machines as well rotating parts (rotors) as also static components (casings).
• NEMA (National Electrical Manufacturers Association).
• API (American Petroleum Institute), especially for test runs and acceptance tests of new and overhauled engines under the aspect of the oil industry.
• NBC (National Building Code) and Steel Design Code deal with the evaluation of resonance problems of components and basements.

A german spezifikation is the DIN 1016-4. It concentrates on the measurement at non rotating parts of gas turbines and the evaluation of the results. To selects the suitable standard the scheme is recommended. It considers two effects:

• The displacement/deflection of the shaft relative to the fit of a bearing outer race respectively the sleeve of a journal bearing. Anti- friction bearings without a damping oil (damping is expected in modern derivates) have a high rigidity. This benefits vibration measurements on struts of main bearing cases or outer casings. In contrast engines with main journal bearings or with a damping oil film require a mesurement direct at on the shaft.
• The ratio of stiffness of the bearing support to the bearing itself. It determines the vibrations of the shaft, related to the casing.

As parameter (vibration magnitude, rms) for the intensity of a vibration, the maximum vibration velocity is applied (rms=0,707 x maximum deflection) in the design relevant frequency range. Also the deflection, acceleration and the maximum values can be used. The estimation of the vibration severity considers maximum values and the time rate of change.

Introductions for vibration measurements as condition monitoring and diagnosis are included in the ISO 13373. There are included:

• Measuring methods and Parameters,
• Selection of the sensors, positioning and fastening.
• Dealing with the measurement data, collection and documentation.
• Operating conditions of the engine.
• Monitoring procedure (time intervals, continuously) and the monitoring system.
• Connection to the analysis system. The ISO 18436-2 addresses to the education/ training of the personnel.

"Illustration 2.5-4": (Lit. 2-16): Depending from the cause, the vibration monitoring of stationary gas turbines can have different goals ( "Ill. 2.5-5").

• Identification of unusual operation conditions like unstable combustion ( "Ill. 3.2.2-5") or a compressor surge ( "Ill. 3.1.1-1").
• Occurrence, development and cause of problems and damages like main bearing failures or crack formation at supporting casing sections ( "Ill. 2.5-6").
• Guard of the engine against over speed.
• Avoidance of dangerous intensive vibrations with secondary damages as result of problems like rotor bow ( "Ill. 2.2-2"), rubbing or blade damages

The vibration behaviour of a gas turbine depends strongly on the bearing stiffness. The oscillations of the rotor are damped by an oil film between the outer race of the bearing and the seat in the supporting casing structure. This was primarily applied to heavy frame engines with sleeve bearing mounted rotors. Even when derivates used as gas generator have direct in the casing supported antifriction bearings, it is possible that only the hot gas coupled power turbine has sleeve bearings. Derivates of modern aero engines are equipped with antifriction bearings, however often damped with an oil film against the supporting casing.

In both cases the oil film can prevent the early registration of dangerous rotor vibrations. So, when, e.g., a smaller fragment of a blade arises, there is the danger that the damage is not recognized because the trigger was not overstepped. Anyway high dynamic bearing loads can cause extensive secondary damages.

The sketch shows schematic typical sensors for the vibration measurement and their positions in and on the engine. The chart at the right contains a survey for heavy frame engines with sleeve brearings to the rotor.

A particular problem is the possibility of a failure (short term total breakdown) of the measuring system ( "Ill. 3.6.2-1"). That complicates the identification of a true failure. Including damages on the axial thrust bearings. To avoid spurious shutdowns sufficient certain redundant systems with two sensors are used. Their indications will be compared.

For smaller engines some manufacturers (OEMs’) offer as standard equipment only seismic probes on the casing. In those cases also proximity probes as an option are available. The simulteneously use of both sensor types generally is recommended for an engine monitoring.

The API-standard 616 (see also "Ill. 2.5-3") demands as components af a monitoring/guard system proximity probes. This is applied as well for an acceptance run as for the operation. The probes and the monitoring system often conform to the API 670 standard.

Generally measurements of the elastic strain with strain gauges are possible. But they are not mentioned. Apparently the failure probability (mechanical damage, thermal damage, debonding) is to high for a gas turbine specific long time use.

"Illustration 2.5-5": In this illustration, typical causes for unusual heavy vibrations of a gas turbine are collected.

Oscillations in the combustion chambers (combustion instabilities, "Ill. 3.2.2-5"): They normally vibrate with the relatively low frequency (‘buzz’) of the gas stream during the combustion. Those vibrations should be considered and mastered by the designer. However it’s possible that due to water or steam injection ( "Ill. 3.2.2-4") or problems with the fuel injection, abnormal high vibrations occur. Especially frequent oscillations are observed at Dry-Low-Nox combustion chambers which operate with a very lean "Ill. 2.5-5" fuel/air mixture. Even gas from an other source can stimulate instabilities.

The pre mixing with a large quantity of air favors such a self amplifying instability of the combustion ( "Ill. 3.2.2-5"). Those violent vibrations, e.g. can lead to unexpected heavy abrasive wear (fretting) of the assembly connections and the break away of chamber shingles ( "Ill. 3.2.1-4").

Surge in the compressor can in the formation phase (eg. rotating stall, "Ill. 3.1.1-5") trigger vibrations and fatigue failures of blades and vanes. A complete stall of all blades (surge) with the disruption of the air stream leads to heavy abrupt impulses with the threat of a short-term dynamic overload of

the blading and/or intense rub ( "Ill. 3.1.1-6"). Not dissipated friction heat with the air (‘blunger loss’) in the momentary lack of an air stream can short-term overheat and damage the rotor blades. The lack of cooling air in the hot parts, together with a shortage of combustion air, can lead in very short time to grave overheating damages.

Rotor bow and rub events: The differentthermal inertia of the rotor and the casings, even hours in the cooling phase can lead to gap bridging and jam of the rotor ( "Ill. 2.2-2" and "Ill. 3.1.2.4-2"). The warm air in the shut down, standing still engine rises up and heats the upper part of the rotor more intensive than the zones beneath. The consequences are self reinforcing vibrations in the frequency of the rotor speed with dangerous damages. This includes the weakening of the rotor by abrasive wear and the overheating as well as overloading the blading.

Air seals: Vibrations of labyrinths can be excited in manifold ways. Examples are excitations by the leakage stream or gas oscillations in the ring chambers around the rotor shaft, formed by the seals. The results can be fatigue cracks and break-outs (plate vibrations) at thin walled components. Endangered are e.g., baffles or support cones of the labyrinths.

Casings: The stiffness of supporting casing structures can change as result of a crack formation. Normally this leads to the weakening of the cross section with a drop in the natural frequency. That permits increased vibrations ( "Example 2.5-2"), especially in case of resonance. Corresponding with the crack growth, normally the vibration level rises with the operation time.

Turbine vanes, turbine nozzles: Damages of the turbine nozzles (high pressure turbine entrance vanes) can excite the rotor via the blades to vibrations that also show outside the engine.

Turbine rotor blades: In the turbine the breaking of a blade can quite happen, especially if only a part of the blade which is not at once noticed, is concerned. On the other hand in a compressor substantial secondary failures can be expected when a blade breaks. This means the immediately breakdown of the engine ( "Example 2.5-2").

Smaller turbine defects can be recognized by the unbalances. This is not easy when the engine has elasic suspended, damped main bearings. We find such a design in aero engine derivates. In this case the unbalance forces are only little noticeable at the acceleration probes on the outside of the engine. Heavy damages can be expected from an operation over longer time with internal unbalances. Therefore it is important that the trigger intensity settings of the probes are satisfactory sensitive without activation of false alarm.

Main bearings:
Anti friction bearings (chapter 3.5.2.1): Failures cause remarkable, mostly high frequency vibrations. There are failures which show early enough before a catastrophic vibration. To this belong fatigue outbreaks (fatigue pittings, pittings) in the bearing races. Is there the suspicion of a bearing failure, magnet plugs and oil filters must be controlled. On the other hand, in case of indications at a magnet plug, a very close look for vibrations is recommended ( "Ill. 5.1-1").

There is quite a realistic chance to recognize at big engines a bearing failure in time with a relatively low number of rotation. In contrast smaller engines with higher speed sustain catastrophic bearing failures in short time. Small gas turbines may suffer this in seconds during the observable end phase. Comprehensible, an engagement to capture such a failure in time is hardly possible.

An intermittent oil supply can also provoke vibrations in the bearing area. Thereby also exists the danger of the breaking of the oil jet with extensive second failures.

Friction bearings/journal bearings (Chapter 3.5.2.2): In gas turbines of the heavy type, normally journal bearings are used. They have specific failure modes (Lit. 2-10 und Lit 2-11). Typical causes of damage are:

• Production failures of the babbitt (sliding layer/ sliding material).
• Contaminations from the production e.g. (machining chips), overhaul (eg., blasting abrasives), unsatisfactory maintenance (problems with the cleanness) and damages in the oil system (e.g., bearing or gear failures). They are washed in by the oil.
• Abrasion by mixed friction.This happens from lubrication problems (e.g,. the oil supply, over load or tilting). - Over load of the babbitt leads to local plastic deformation (creep). Causes are too high contact pressure in connection with the operating temperature.
• Dynamic fatigue (fatigue cracks in the babbitt, ‘paving stone’ break-outs) by pressure pulsation in the lubrication gap.
• Cavitation leads to small break-outs by vibration fatigue of the babbitt. This is favored by water in the oil (vapor bubble formation) and entrained air.
• Chemical attack develops by corrosion and/or reactions of the babbitt with the oil (e.g., with aggressively deterioration products).
• Electro erosion/electric discharge erosion from an electric current.

Accessory equipment (Chapter 3.6.1) can be excited to vibrations by the failure of components like gears (Chapter 3.7-2), couplings/insert shafts ( "Ill. 3.6.1-7") or bearings (Chapter 3.5-2). In an extreme case this leads to the fracture of the mounts and/or destruction of the device. It is also possible that vibrations are routed into the engine and produce damages. So, e.g., a case emerged, at which a small inaccuracy of a gear wheel lead to the fatigue fracture of the central interlocking bolt of the turbine rotor. Emanating from the teeth of gear wheels are extremely high frequent vibrations (up to the ultrasonic region) because the number of teeth and the high rotation speed. At the arising vibration modes even smallest amplitudes in the region of 0.1 mm are enough for a damage by vibration fatigue.

Subsequent installed accessory drives with alternating torsion moment like cardan shafts or universal couplings with elastomers can excite vibrations through a gear in the engine and produce failures. In such a case it came several times to the fracture of a drive shaft in a control unit which was driven by a gear. This potential transmission of vibrations could easily be overlooked. It was therefore discovered not until several failures.

Vibrations of the drive train: Insufficient aligned shafts and couplings can trigger vibrations. The neccessity to drive through resonances has the risk of high amplitudes with an over load of the schaft up to a fracture. Also vibrations of the powered aggregates (e.g., pump or generator) can be fed into the engine by couplings. Are vibrations of the rotor excited, an increased damage hazard exists. Therefore special regard to the correct funktion of the drive side is needed.

"Illustration 2.5-6": Through crack formation or too high operation temperatures, a casing can become more elastic and can trigger the drop of the natural frequency and get in resonance with the rotor speed. This will be shown by the vibration warning system. Such effects (increase of the amplitude with the operation time) are to be followed exactly, even if the amplitudes are in a acceptable range. So it should be possible to take remedy measures if necessary (see "Example 2.5-2").