Gas turbine system and method of manufacturing

Abstract: Power generation turbines having different power outputs for different power grid frequency applications have modular second and third stages, rotors, bucket wheels for all stages and other ancillary parts. The first and second turbines have sizes not geometrically scaled according to speed. Four-stage turbines having different outputs for different power grid frequencies not geometrically scaled have an identical annulus through the first, second and third stages, different geometry in the first and fourth stages, and identical geometry in the second and third stages.

Description: The present invention relates to gas turbines for operation at different frequency applications more especially having a high degree of hardware commonality and particularly relates to gas turbines for land use operation at 50Hz and 60Hz power grid frequencies using common modular components. Gas turbines, when used for land use electrical power generation, are typically required for both 50Hz and 60Hz applications, depending upon power grid frequency. The costs involved in developing and producing machines for both frequencies are quite significant. For example, components for a turbine designed for each different frequency application are typically unique to that turbine. This results in higher investment costs for tooling and virtually no commonality of hardware as between the two turbines, which would beneficially impact turbine costs. One approach commonly used for developing gas turbines for 50Hz and 60Hz frequency applications is simple geometric scaling of one design to a second frequency. Scaling is based on the principal that one can reduce or increase the physical size of a machine while simultaneously increasing or decreasing rotational speed to produce aerodynamically and mechanically similar compressors and turbines for the different frequency applications. Application of scaling techniques has enabled the development of turbines for both frequency applications which, while reducing development costs, still results in turbine components unique to the turbine for a particular frequency application. For example, components for a turbine designed for a 50Hz application are scaled geometrically by the frequency ratio 50/60 = 0.833 to yield similar turbine performance at 60Hz frequency. With this fixed geometric scaling, power output scales by the inverse square of the frequency, i.e., (50/60) 2 = 0.694. Thus, a turbine sized at 50Hz to provide a power output of 100 megawatts would, when geometrically scaled by a factor of 0.833, provide a power output of 69.4 megawatts at 60Hz. More generally, there is a fixed relationship or ratio between output power at one speed and power output at another speed when turbine designs are geometrically scaled. The advantage of this scaling approach is that components sized at one frequency can be readily redesigned at the scaled frequency. However, the output of the turbine is fixed by the scale factor and thus one or the other of the turbines may not be optimum for a particular application. That is, market demands may require turbines for operation at different frequencies and the power output of one turbine at one frequency may not result in the desired output of the other turbine at the other frequency when the first turbine is geometrically scaled to afford the second turbine. Equally important, the components (hardware) for a base turbine for one frequency application have virtually no commonality with the components (hardware) of the scaled turbine for the different frequency application, resulting in increased tooling and component parts costs as well as other disadvantages. With the use of the present invention, there may be provided gas turbines which can be used for 50Hz and 60Hz applications, respectively, with substantial and significant commonality of hardware with minimum or negligible loss in turbine performance for each application whereby substantial reductions in costs are realized by a commonality of design, hardware and tooling. Additional economic benefits may be realized in terms of reduced design cycle times and resources necessary to design and manufacture the turbines for use at different power outputs and frequencies. Moreover, the present invention breaks the relationship between geometric scaling and power output whereby the output of the 50Hz and 60Hz machines can be set independently of the turbine by setting compressor mass flow and adjusting the turbine accordingly. In short, the design of turbines with different power outputs at different frequencies, according to this invention, is no longer constrained by the geometric scaling factor. For the design of one or the other of the two turbines for different frequency applications, e.g., 50Hz or 60Hz, and considering the desirability of having an identical hot gas flowpath to the extent possible for the two turbines, a turbine exit mach number is initially set such that the pressure loss in the diffuser downstream of the turbine and its mechanical performance are acceptable. For a given firing temperature, the turbine pressure ratio and quantity of cooling air introduced into the turbine airfoils and ancillary parts such as shrouds and into the gas path determine the metal temperature of the last-stage bucket. With the selection of an appropriate alloy for the last-stage bucket, the maximum allowable centrifugal stress can be determined, for example, for the 60Hz machine. This centrifugal stress is directly proportional to AN 2 where A is the annulus area formed by the last-stage buckets and N is the speed of rotation. By limiting the exit mach number, the maximum allowable flow through the turbine can be determined and hence its power output. For a given initial design, e.g., either 50 or 60Hz, and using as an example, 60Hz, the hub (inner) radii of the flowpath can be set considering turbine performance, rotor length and weight, leakages and the like. With the hub radii and last-stage annulus area set, bucket tip radii can be set. Because of the N 2 term in the centrifugal stress calculation, the bucket lengths are limited by the higher speed 60Hz turbine. To provide the additional turbine power output necessary for a 50Hz machine, given the constraints for the design of the 60Hz machine, and assuming identical firing temperature and the same gas flow properties, as well as substantially similar pressure ratios, the mass flow through the constant area flowpath of the turbine must be increased. To provide for this increased mass flow while maintaining an acceptable exit mach number, the height of the exit annulus is increased to afford increased exit area. This increase in height of the last-stage nozzles and buckets for a 50Hz turbine is accommodated in the tip area, while maintaining a common hub radius with the 60Hz turbine. Consequently, the last-stage, e.g., the fourth stage in a four-stage turbine, has increased nozzle and bucket tip radii. To maintain turbine pressure ratio while accommodating increased mass flow, the first-stage nozzles and buckets are changed to increase their throat areas, i.e., the area available for passage of flow. The cross-sectional area of the annulus forming the first-stage flowpath remains the same, although its flow area increases due, e.g., to the change in the orientation of its buckets and partitions. Importantly, the intermediate stages, e.g., the geometry of tne second and third stages in a four-stage turbine, according to the present invention, remain unchanged as between the turbines of different power outputs at different frequencies. While the speed and mass flow change between the 50Hz and 60Hz turbines causes the incidence angle of gas flowing onto the airfoils of the second and third stages to change slightly, those changes in incidence angle can be accepted by the airfoil design for those stages. Further, while the gas pressure within the flowpath changes with turbines of different outputs, cooling flow and purge flow source pressures can be selected to ensure adequate backflow margin is maintained in both machines to preclude hot gas in the flowpath from entering the rotor cavities and damaging the rotor structure, or entering coolant passages within the gas path components. It will be appreciated that bucket airfoils are normally oriented so that centrifugally generated bending loads counteract those generated by gas pressure. In the intermediate stages, e.g., the second and third stage of a four-stage turbine hereof, the airfoils are required to operate at both power outputs and frequencies, resulting in centrifugal bending loads which differ at the two speeds. It has been found, however, that the airfoils can be leaned circumferentially and axially at an intermediate position to reduce net resultant bending stress to acceptable levels at both speeds. It will be further appreciated from the foregoing that the resulting turbines of different power outputs at different frequencies, for example, 50 and 60Hz turbines, share a high degree of hardware commonality. Specifically, the rotor, rotor wheels for the buckets for all four stages, the spacers between the stages, the impeller plate, the aft shaft, the forward shaft, seal plates, the buckets for the second and third stages, the second and third-stage nozzles, the diaphragms, the shrouds for the second and third-stage buckets, as well as for the first-stage buckets, the inner shell and the outer shell are common hardware components for both the 50 and 60Hz turbines. Stated somewhat differently, the items unique to the individual 50 and 60Hz machines are principally the nozzles and buckets of the first and last stages, the shrouds for the last stage and the diffuser fairing at the exit annulus. The invention can therefore be characterized as having a high degree of modularity among the component turbine parts for use with turbines at different frequency applications, e.g., 50Hz and 60Hz applications. Various design considerations for turbines of the kind employed in the generation of electricity are discussed in the documents D. Kalderon: "Design of large steam turbines.", GEC Turbine Generators Ltd., Rugby, England XP002007571; DE-A-2 408 641; and CH-A-85 282. According to the present invention, there is provided a turbine system for providing gas turbines for operation at two or more different power outputs and rotational speeds for use in electrical power systems having different power grid frequencies, the gas turbines including first, intermediate and final stages, each stage comprising a fixed diaphragm having stationary partitions and a rotatable turbine wheel having buckets, characterised by: the system comprising sets of turbine stages in which respective turbines for use at different power grid frequencies comprise first stages having different geometries from one another; corresponding final stages having different geometries from one another; and intermediate stages having identical geometries. The turbine having said first power output may be rotatable at a first speed of 3600 RPM for a 60Hz power grid and the second turbine having said second power output may be rotatable at a speed of 3000 RPM for a 50Hz power grid. The final stage for said first turbine may have an exit annulus of a cross-sectional area less than the cross sectional area of the exit annulus of the final stage for said second turbine. The first stages of each of said first and second turbines may have different geometries from one another. Accordingly, it is a primary object of the present invention to provide turbines and methods of constructing turbines wherein non-geometrically scaled turbines have different power outputs at different frequencies with substantial significant commonality of hardware as between the turbines and negligible impact on turbine performance. BRIEF DESCRIPTION OF THE DRAWINGS FIGURE 1 is a schematic illustration of a gas turbine according to the present invention; FIGURE 2 is a schematic diagram of a combined cycle system employing the gas turbine and heat recovery steam generator for greater efficiency; FIGURE 3a is a schematic cross-sectional view of a four-stage turbine having a predetermined power output and frequency constructed in accordance with the present invention; FIGURE 3b is a view similar to Figure 3a illustrating a second turbine having a different power output and frequency than the turbine illustrated in Figure 3a; and FIGURE 4 is a schematic representation of the flowpath of the two turbines illustrated in Figures 3a and 3b. Figure 1 is a schematic diagram for a simple cycle, single-shaft heavy-duty gas turbine 10 incorporating the present invention. The gas turbine may be considered as comprising a multi-stage axial flow compressor 12 having a rotor shaft 14. Air enters the inlet of the compressor at 16, is compressed by the axial flow compressor 12 and then is discharged to a combustor 18 where fuel such as natural gas is burned to provide high-energy combustion gases which drive the turbine 20. In the turbine 20, the energy of the hot gases is converted into work, some of which is used to drive the compressor 12 through shaft 14, with the remainder being available for useful work to drive a load such as a generator 22 by means of rotor shaft 24 for producing electricity. A typical simple cycle gas turbine will convert 30 to 35% of the fuel input into shaft output. All but 1 to 2% of the remainder is in the form of exhaust heat which exits turbine 20 at 26. Higher efficiencies can be obtained by utilizing the gas turbine 10 in a combined cycle configuration in which the energy in the turbine exhaust stream is converted into additional useful work. Figure 2 represents a combined cycle in its simplest form, in which the exhaust gases exiting turbine 20 at 26 enter a heat recovery steam generator 28 in which water is converted to steam in the manner of a boiler. Steam thus produced drives a steam turbine 30 in which additional work is extracted to drive through shaft 32 an additional load such as a second generator 34 which, in turn, produces additional electric power. In some configurations, turbines 20 and 30 drive a common generator. Combined cycles producing only electrical power are in the 50 to 60% thermal efficiency range using the more advanced gas turbines. In both the applications illustrated in Figures 1 and 2, the generator is typically supplying power to an electrical power grid. The power grid is conventionally either 50Hz or 60Hz, although the scope of the present invention may include turbine power applications at frequencies other than 50Hz and 60Hz. As alluded to earlier, conventional practice in supplying turbines for land-based power generation require a unique turbine for each frequency application and rated power output resulting in a lack of commonality of hardware as between the various turbines. While geometric scaling has been applied to design various turbines for use in applications at different frequencies, thus reducing costs, still each turbine is unique. The present invention affords turbines which break the relationship between power output at the different frequencies and the scaling factor, thereby enabling maximization of common turbine hardware for different power and speed or frequency combinations than presently allowed by pure geometric scaling. Referring now to Figures 3a and 3b, there is illustrated a pair of turbines T a and T b for use in the above-identified systems. Turbine T a , for example, illustrated in Figure 3a, may be for use with 60Hz applications, whereas turbine T b , illustrated in Figure 3b, may be for use with 50Hz applications. Suffice it to say that the two turbines T a and T b are designed for different power outputs for the 60Hz and 50Hz applications. Referring to Figure 3a, turbine T, includes an outer shell 40a forming the structural outer shell or housing of the turbine, an inner shell 42a and a rotor Ra. Rotor Ra mounts a plurality of bucket wheels 44a, as well as spacer wheels 46a between adjoining bucket wheels 44a, all bolted together between forward and aft shafts 48a and 50a, respectively, by a plurality of bolts 52a arranged about the longitudinal axis of the rotor Ra. The turbine T a includes a first stage, at least one intermediate stage (preferably two) and a last stage, each stage comprising a diaphragm mounting a plurality of circumferentially spaced partitions or nozzle vanes between inner and outer rings and a plurality of buckets mounted on the turbine wheels. In the illustrated form, a four-stage turbine is provided, with first-stage nozzles 54a, buckets 56a; second-stage nozzles 58a and buckets 60a; third-stage nozzles 62a and buckets 64a; and fourth-stage nozzles 66a and buckets 68a. The nozzles 54a, 58a, 62a and 66a form part of diaphragms mounting the partitions extending between the inner and outer diaphragm rings in the usual manner. Additionally, the inner shell 42a carries shrouds 70a and 72a about the outer tips of buckets 56a and 60a of the first and second stages, respectively. Shrouds 74a and 76a are carried directly by the outer shell 40a about the tips of the third and fourth-stage buckets 64a and 68a. Thus, the nozzles, the shrouds and the outer surfaces of the bucket wheels define an annular flowpath through the turbine which receives the hot gases of combustion for expansion through the various stages, thereby imparting work to the buckets and rotor. The turbine T b illustrated in Figure 3b has like parts similarly arranged and designated by like reference numerals, followed by the letter "b." As discussed, the turbine T a illustrated in Figure 3a is designed for a specified power output at a certain rotational speed and power grid frequency, e.g., 3600 rpm for 60Hz applications, while the turbine of Figure 3b is designed for a specified power output at a different rotational speed and power grid frequency, e.g., 3000 rpm for 50Hz applications. In accordance with the present invention, the turbines have a high degree of hardware commonality whereby the common hardware parts can be interchangeably used in either of the two turbines having the different power outputs at the different frequencies. As indicated previously, the cross-sectional area of the annulus defining the flowpath through the first, second and third stages is identical through the two turbines. However, to obtain different power outputs for a common flowpath, it is necessary to adjust the mass flow through the turbine at the different speeds of the two turbines. The flowpath inner radius is set to be common in the two turbines. The last-stage annulus can likewise be set for a given firing temperature, turbine pressure ratio and quantity of cooling air introduced, thus determining the bucket tip radius of the last stage. However, because of the high centrifugal stresses on the last stage, and the need to select an appropriate alloy for the last-stage bucket, the bucket length is limited by the higher frequency machine, e.g., a 60Hz turbine. Consequently, to provide the increased mass flow necessary for a 50Hz turbine, while maintaining an acceptable exit mach number and with a constant flow cross-section at least through the first, second and third stages, the height of the exit annulus of the final stage is increased to afford an increased exit area. The inner radius of the last-stage diaphragm and buckets, however, remains the same and consequently, the radius of the last-stage partitions and buckets are enlarged at the outer radius of the flowpath to meet the increased mass flow and slower speed requirements of the 50Hz turbine as compared with the 60Hz turbine. Further, to maintain turbine pressure ratio while accommodating increased mass flow, the first-stage nozzles and buckets are restaggered to increase their throat areas while maintaining the annulus area constant as between the two turbines. Thus, the orientation of the buckets and partitions in the first stage of the 60Hz turbine is changed when a 50Hz turbine is undergoing fabrication. The profiles of the airfoils of the first stage are also changed to accommodate this increase in mass flow. It has been found, however, that the speed and mass flow changes as between the 60 and 50Hz turbines can be accommodated by a particular (and common) airfoil design in the second and third stages without substantial performance loss. Consequently, the second and third stages, including the partitions, buckets, wheels and shrouds, are sized and dimensioned identically to permit interchangeability of the second and third stages in either one of the two turbines of different power outputs and frequency applications. That is, the intermediate stages of the turbine design can be modularized for installation in either one of the two machines of different power outputs at the different frequencies. Thus, as illustrated by the common stippling in Figures 3a and 3b, the partitions and buckets of the second and third stages of the two machines are identical. Further, the rotor wheels for all buckets, e.g., the first, second, third and fourth-stage buckets, the spacers between the stages, the impeller plate, the aft and forward shafts, and seal plates constitute common hardware as between the 60Hz and 50Hz machines. Note also that the shrouds for the first, second and third-stage buckets, as well as the inner and outer shells are common between the 60 and 50Hz turbines. Importantly, the rotors Ra and Rb are also common. As illustrated by the different shading of the first and last stages upon comparing Figures 3a and 3b, the uniqueness of the 50 and 60Hz turbines is manifested primarily in the first and last stages. Particularly in the first stage, the throat area between the partitions for the 50Hz turbine is opened to accommodate the greater mass flow as compared with the 60Hz turbine. With respect to the last or fourth stage, the buckets and partitions are increased in radius at their tip ends to accommodate the increased mass flow for the 50Hz machine. Referring to Figure 4, the difference in the flowpath through the two turbines of different outputs for 60Hz and 50Hz applications is illustrated. The first, second, third and fourth stages ST1, ST2, ST3 and ST4 are illustrated with each having nozzles and buckets designated by the letter N and B, respectively, followed by a number indicating the turbine stage. It will be appreciated that the cross-sectional area of the annulus for both the 50Hz and 60Hz turbines is identical for the first, second and third stages and that the flowpath through the second and third stages is identical. With respect to the fourth stage, the lower mass flow, higher speed 60Hz machine, has an outer annulus wall 80, illustrated by the dashed line, while the larger mass flow, lower speed 50Hz machine has an outer wall 82. The increase in the radius of the nozzles N4 and buckets 84 of the fourth stage at their tips is thus indicated by the solid line 82 for the larger mass flow lower speed 50Hz machine.

Claims: A turbine system for providing gas turbines for operation at two or more different power outputs and rotational speeds for use in electrical power systems having different power grid frequencies, the gas turbines including first (54a, 56a), intermediate (58a, 60a, 62a, 64a) and final stages (64a, 66a), each stage comprising a fixed diaphragm having stationary partitions and a rotatable turbine wheel (44a etc.) having buckets, characterised by : the system comprising sets of turbine stages in which respective turbines for use at different power grid frequencies comprise first stages having different geometries from one another; corresponding final stages having different geometries from one another; and intermediate stages having identical geometries. A turbine according to claim 2 wherein said turbine having said first power output is rotatable at a first speed of 3600 RPM for a 60Hz power grid and the second turbine having said second power output is rotatable at a speed of 3000 RPM for a 50Hz power grid. A turbine according to claim 1 wherein said final stage for a first turbine has an exit annulus (80) of a cross-sectional area less than the cross sectional area of the exit annulus of the final stage for a second turbine (82). A method of manufacturing first and second turbines having substantially identical firing temperatures and pressure ratios for use with gas flows having substantial identical properties wherein each turbine has first (54,56), intermediate (58, 60, 62, 64) and final (66, 68) stages with each stage including partitions and buckets, characterised by the steps of: forming a pair of first stages for installation in said first and second turbines, respectively, wherein said first stages have geometries different from one another; forming a pair of final stages for installation in said first and second turbines, respectively, wherein said last stages have geometries different from one another; forming a pair of intermediate stages having geometric characteristics identical to one another for installation in said first and second turbines, respectively; and, installing the stages in said first and second turbines, respectively.

Inventors: STALEY TERRIL K, WALKER ALAN, STALEY, TERRIL K., WALKER, ALAN
Applicants: GEN ELECTRIC, GENERAL ELECTRIC COMPANY
Application year: 1996
Main classification: F01D5
IPC codes: F01D15, F01D1, F02C6, F01D5, F02C7