Turbogroup of a power generating plant

Details Turbogroup of a power generating plant Turbogroup of a power generating plant

2002-08-08

2004-03-23

Kim, Ted (Department: 3746)

Power plants

Combustion products used as motive fluid

Having mounting or supporting structure

C060S039183

Reexamination Certificate

active

06708500

ABSTRACT:

FIELD OF THE INVENTION
The invention relates to a turbogroup of a power generating plant, in particular a gas-storage power plant, comprising a turbine unit and a generator unit.
BACKGROUND OF THE INVENTION
A turbine unit normally has a turbine and a further fluid-flow machine on a common turbine shaft. In a conventional power generating plant, this further fluid-flow machine may be formed by a compressor which is driven by the turbine via the turbine shaft. In a gas-storage power plant, in particular an air-storage power plant, this further fluid-flow machine is formed by an additional turbine, to which the gas of a gas reservoir of the gas-storage power plant is admitted, so that the additional turbine likewise transmits drive output to the turbine shaft. As a rule, a generator unit has a rotor of a generator on a generator shaft and serves to generate electricity. The turbine unit serves to drive the generator unit, so that accordingly the turbine shaft is in drive connection with the generator shaft.
During operation of the turbogroup, relatively large masses rotate at relatively high speeds. In order to be able to control the dynamic vibration behavior of the turbogroup, in particular of the turbine unit, a high-capacity bearing system is necessary. Such a bearing system normally comprises at least four radial bearing units, with which the shafts are radially mounted and at least supported at the bottom, and at least one thrust bearing unit, which normally absorbs the thrust of the turbine, or possibly of the turbines, in the axial direction at the turbine shaft. For this purpose, a first radial bearing unit is arranged on a side of the turbine which faces away from the generator unit, whereas a second radial bearing unit is arranged on a side of the further fluid-flow machine which faces the generator unit. A third radial bearing unit is arranged on a side of the generator which faces the turbine unit, and a fourth radial bearing unit is arranged on a side of the generator which faces away from the turbine unit. In this case, the thrust bearing is expediently arranged axially between the generator and the further fluid-flow machine of the turbine unit. It is possible here in principle to arrange the thrust bearing unit next to the second radial bearing unit. If the further fluid-flow machine is a compressor, the thrust bearing unit can be integrated in an air-feed casing which serves to feed air to the compressor.
Thrust bearings work optimally when the bearing axis runs coaxially to the rotation axis of the shaft to be supported. Thrust bearings react in a sensitive manner to changes in inclination and misalignments; in particular, friction, the generation of heat, and wear increase. If the turbine unit has an annular combustion chamber for firing the turbine and if the further fluid-flow machine of the turbine unit is formed by a compressor, the changes occurring during operation in the relative position between the bearing axis of the thrust bearing unit and the rotation axis of the turbine are relatively small. However, if a combustion chamber lying at the top, a “silo combustion chamber”, is used instead of an annular combustion chamber, temperature differences in the outer casing of the turbine unit from top to bottom cannot be ruled out. This different temperature distribution in the outer casing may lead to the outer casing arching convexly upward—“banana formation”. While the casing bends, the rotation axis of the turbine shaft remains invariable. Since the thrust bearing unit is normally integrated in the casing of the turbine unit next to the second radial bearing unit, the relative position between the bearing axis of the bearing unit fixed to the casing and the rotation axis of the turbine shaft may change to a relatively pronounced degree due to the asymmetrical thermal expansion of the casing, as a result of which a proper thrust bearing arrangement is put at risk.
If the turbogroup is now to be used in a gas-storage power plant, the further fluid-flow machine used is an additional turbine instead of the compressor. Such an additional turbine has a radial gas feed with optional additional gas inlets or gas discharges compared with the conventional compressors. Accordingly, the thermal expansion effects referred to appear to a greater extent, as a result of which the loading of the thrust bearing unit in particular additionally increases. Furthermore, such an additional turbine inside a gas-storage power plant works on the inlet side with considerably higher pressures and temperatures in the fed gas flow than a conventional compressor. This may also intensify the thermal expansion effects. At the same time, the outlay for the oil supply to the thrust bearing unit increases considerably on account of a large axial thrust.
During operation of the turbogroup, the radial bearing units and the thrust bearing unit absorb not only inertia forces or thrust forces but also vibrations which are caused, for example, by out-of-balance of the rotating masses. In this case, both the turbine unit and its bearing system in each case form vibratory systems which are coupled to one another and have natural frequencies or resonant frequencies. For reliable operation of the turbogroup, it is necessary that natural vibrations in the turbine unit and in the bearing system do not occur within an attenuation range of the turbine-shaft operating speeds which extends, for example, from −10% to +15% of the rated operating speed of the turbine shaft. On account of the highly complex coupling of the vibration systems and on account of a multiplicity of boundary conditions which cannot be determined exactly, it is presently not possible to be able to predict the vibration behavior of the turbine unit and of the associated bearing system in a sufficiently reliable manner at a justifiable cost. Measures are therefore sought which make it simpler or make it possible to subsequently influence the vibration system. Of particular interest in this case are measures which involve minimum interference with the design and the construction of the turbine unit.
SUMMARY OF THE INVENTION
The invention is intended to provide a remedy here. The invention, as characterized in the claims, deals with the problem of showing how, for a turbogroup of the type mentioned at the beginning, to make it possible or easier to influence the vibration behavior of the turbine unit and/or of the bearing system.
This problem is achieved according to the invention by the subject matter of the independent claim. Advantageous embodiments are the subject matter of the dependent claims.
In the inventive embodiment of the turbogroup, the first radial bearing unit and/or the second radial bearing unit have pendulum supports which are in each case supported on a bearing pedestal. The present invention is now based on the general idea of supporting the pendulum supports, at least at one radial bearing unit of the turbine unit, on the associated bearing pedestal in each case via a spring element. Such a spring element changes the vibration properties of the respective radial bearing unit and thus of the entire vibration system coupled thereto. By suitable selection of this spring element, the desired tuning of the entire vibratory system can be carried out to the effect that the critical natural frequencies are clearly outside the attenuation range for the operating speeds of the turbine shaft. In this case, it is perfectly possible to adapt the spring element by the “trial-and-error principle”, since this selection of the suitable spring elements for the respective turbogroup type need only be made once before the initial commissioning of the first turbogroup of a new series. The spring element configuration found once may then be adopted for all subsequent models of this type.
According to an especially advantageous development, the bearing pedestal may have a top side extending essentially in a planar manner, the spring element then being formed by a metal plate which extends essentially parallel to the bearing ped

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