Table of Contents
3.7.2 Stationary gears
Without a gear we can not reasonable use the motor performance. It distinguished itself normally by inconspicuousness. It serves over a long time without problems. Because the lubrication does not happen with motor oil because of the necessary higher ‘load capacity’, gear oil must be refilled in time intervals.
Does however the gear fail this will be usually expensive and the repair time consuming.
Similar is the behaviour stationary gears of a gas turbines, even though it has no feasibility to shift. The output speed and the torque of the gas turbine must be optimal adjusted to the powered machine.
If the output speed of the power delivering rotor does not fit to the required driving speed of the machine that must be driven a stationary gear must be interconnected.( "Ill. 3.7.2-1"). This needs itself special components like an own oil supply ( "Ill. 3.7.2-2"). Additional it can couple the gas turbine with the starter and drive further aggregates like fuel pumps, hydraulics, and control devices . Contrary to accessory gears of derivates ( "Ill. 3.6.1-1") in stationary gears mainly sliding bearings for gear pinions and shafts are used. They need obout 80 % from the supply of the pressure oil.
Usually so called ‘parallel schaft gears’with two coging gear wheels. Today they are single or double angular teethed ( "Ill. 3.7.2-3"). The single agular gearing produces contray to the double angular teethed an axial thrust. This can be carried by a thrust bearing and/or used for the balance of the axial thrust of the turbine rotor. For this purpose serve axial sliding bearings ( "Ill. 3.5-14").
To avoid catastrophic secondary damages due to oil starvation or foreign objects in the oil in causative connection with the gas turbine, a stationary gear usually has its own oil system ( "Ill. 3.7.2-2"). This supplies individual every bearing and gear wheel. The requirements for the oil (e.g., viscosity, high pressure additives) differ between the gear and the gas turbine. Here primarily anti friction bearings are lubricated. Therefore when indicated, the OEM will prescribe a suitable oil. The gear for a 80 MW engine requires about 1 m3 lubrication oil per minute to dissipate the heat of about 1 MW produced by the power loss. At an aspired oil temperature under ca. 60°C at the entrance, this gear experiences a temperature increase of about 25 °C. For this an sufficient dimensioned oil cooler is needed, which transfers the heat to the air or water. Naturally the power dissipation is also a cost factor that can not be ignored. Under these aspects a gear efficiency as high as possible is aspired. At a conventional construction up to 99% can be realized.
At the gear tooths develop the main part of the losses. Thereby it is astonishing that not the friction between the tooth flanks accounts, but the generation of foamed squeezing oil. Also the swirl losses (‘blunger losses’) of the air in the gearbox may be not to ignore. All these losses can be decreased by about 80% in a sealed and evacuated gear (High Efficiency Turbogear = HET). The flooding with a light inert gas (helium) offers itself also. So the oxidation as cause for an oil aging (deterioration) and sealing problems should be minimized.
There exists a certain danger of fatigue cracks and fractures with catastrophic results by stresses from torques and centrifugal forces also in the hub of the high loaded high volume gear wheels. To avoid propagation capable from production (material inhomogenities) and dangerous tensile stresses (hardening, forging) spin tests (constant and cyclic) are carried out. During operation with acceleration pick ups vibrations that point to a beginning failure, can be detected and identified. For this monitoring respectively a vibration analysis and conclusions at the failure cause ( "Ill. 3.7.2-5.1" , -5.2, -5.3) there are specifications/norms like DIN 4979 oder ISO 8579-2. They enable the operator among other things to check if the vibration monitoring of his engine is complying the state of the art and give him a better understanding of the vibration causes.
"Illustration 3.7.2-1": (Lit. 3.7-9): There are plenty concepts of gas turbine plants. They optimize customer specific components and its arrangement for the power delivery, respectively receiving the power. As basic principle a parallel shaft gear with two coging gear wheels can be seen.
For a transfer performance up to ca. 10 MW planetary gears are used. They serve especially with small, fast running gas turbines. Those gears require little space. Also favorable is that drive shaft and driven shaft are aligned to the same axis.
"Illustration 3.7.2-2": (Lit. 3.7-12): In stationary gears in gas turbine plants an own oil system is used. The oil flow is primarily needed to dissipate the heat losses. The typical oil entrance temperature into the oil cooler(„C“) lies at ca. 55 °C. There a cooling to ca. 40°C takes place. For this purpose air/oil or water/oil heat exchangers are used.
Typical oil filters („B“) of stationary gears have a mesh size of 0,04 mm. Not to cross the oil film between the tooth flanks and so to trigger wear and fatigue ( "Ill. 3.7.2-5.1"), the particles must be smaller than 0,01 mm (see also "Ill. 3.5-2").
To avoid an oil shortage/starvation by a blocked filter a bypass is destined („H“). The gear oil prescribed by the OEM must be necessarily used.
"Illustration 3.7.2-3": (Lit. 3.7-11): Contact patterns/tooth bearers show geometric dependent damaging flank loads. There is the possibility to suggest from the contact pattern created with a ‘master wheel’ at the type of failure respectively the failing component. Near a coupling due to twisting and bending of very slender gear wheels other contact patterns than those shown can be formed.
- „A“ Corner bearer are traced to faulty flank orientation, beveling or offset of the axis.
- „B“ Wobbling failure show after a rotation of 180° mostly at the tooth width a change of the bearing area (changing corner bearer).
- „C“ Head bearer („C1“) and root bearer („C2“) are results of a failure in the engagement (base circle failure).
- „D“ Waviness in the toot height („D1“) direction or over the tooth width („D2“) are a reason of the tooth manufacturing
"Illustration 3.7.2-4": (Lit. 3.7-11): The contact pattern of gear teeth is of particular information value. It should be controlled during approval in the factory and putting into service. So assembly caused elastic deformations of the gear and its attachments/connections are also covered. Such an influence is unevenness of the base plate. At first a determination of the contact pattern during engine idle is carried out. For this purpose touch up color is used. Because of the not complete elastic shaping we get contact patterns equal to the sketch at the right. At nominal load the whole tooth flank should bear.
The shown contact patterns comply as example the specification British Standard 1807. It applies for unloaded tooth flanks. Prescribed is, independent from the contact pattern, only the bearing area. The classes A1 and A2 apply for a pitch circle velocity of more than 50 ms.
"Illustration 3.7.2-5.1", "Illustration 3.7.2-5.2" and "Illustration 3.7.2-5.3": (Lit. 3.7-11 and Lit. 3.7- 12): In those pictures will be tried to give the practitioner the chance of a first evaluation of tooth failures in gears. This facilitates the understanding of descriptions in overhaul manuals and specifications. The shown failure types can also be found in accessory gears ( "Ill. 3.6.1-1"). In cases of doubt, basically a specialist e.g., the OEM should be consulted to follow the specifications.
Literature from chapter 3.7
3.7.1 T.Zaba,P.Lombardi,BBC&Co Ltd.,.“Experience in the Operation of Air Filters in Gas Turbine Installations“, ASME Paper 84-GT-39 (1984).
3.7-2 A.W.Anderson,R.G.Neaman,“Field Experience with Pulse-Jet Self-Cleaning Air Filtration on Gas Turbines in Desert Environment“,ASME Paper 82-GT-283 (1982).
3.7-3 H.J.Willcocks,P&W Aircraft,“Icing Conditions on Sea Level Gas Turbine Engine Test Stands“, AIAA-82-1237 (1982).
3.7-4 J.Dickson,Trans Canada Pipelines,Toronto,“Extreme-cold-weather operation gas-turbines show keyproblems“, The Oil And Gas Journal - April 26,1976, Page 104.
3.7-5 J.Dickson,Trans Canada Pipelines,Toronto,“Problems Associated with Cold Weather Operation of Gas Turbines“, ASME Paper 76-GT-129.
3.7-6 T.L.Bowen,D.P.Guimond,R.K. Muench,“Experimental Investigation of Gas Turbine Recuperator Fouling“, ASME Paper 87-GT97 (1987).
3.7-7 M. Sauer-Kunze, „Auswahl geeigneter Luftfiltersysteme zur Optimierung des Wirkungsgra- des von Gasturbinen bei gleichzeitiger Verminderung der Lebens-Zyklus-Kosten“, aus „Gastur- binen in Praxis und Entwicklung“, VDI-Gesellschaft Energietechnik, VDI-Berichte 1721, ISBN 3-18-091721-0, Page 115 up to 127.
3.7-8 A.Rossmann, „Die Sicherheit von Turbo-Flugtriebwerken“, Band 3, ISBN 3-00-017733-7, 2003, Axel Rossmann Turboconsult, Bachweg 4, 85757 Karlsfeld.
3.7-9 T. Deeg,“Getriebe“, in „Stationäre Gasturbinen“, Herausgeber C.Lechner, J.Seume, ISBN 3-549-42831-3, Springer Verlag Berlin Heidelberg, 2003, Page 113-134.
3.7-10 H.-G. Brummel ,“Abgasstrecke und Abhitzedampferzeuger in GUD-Anlagen“, in „Stationäre Gasturbinen“, Herausgeber C.Lechner, J.Seume, ISBN 3-549-42831-3, Springer Verlag Berlin Heidelberg, 2003, Page 135-206.
3.7-11 „Handbuch der Schadensverhütung“, Allianz Versicherungs-AG München und Berlin 1972, Kapitel „Stationäre Getriebe“, Page 388-426.
3.7-12 P.Lynwander ,“Gear Drives for Turbomachinery“, in „Sawyer’s Turbomachinery Maintenance Handbook, Volume III“, Turbomachinery International Publications ISBN 0-937506-02-8, 1980, Page 11-1 up to 11-22.
3.7-13 G.G.Ostrand, “Gas Turbine Inlet Air Filtration“, in „Sawyer’s Turbomachinery Maintenance Handbook, Volume III“, Turbomachinery International Publications ISBN 0-937506-02-8, 1980, Page10-1 up to 10-20.