EP3155225A1 - Turbine and method for expanding an operating fluid - Google Patents

Turbine and method for expanding an operating fluid

Info

Publication number
EP3155225A1
EP3155225A1 EP15726328.6A EP15726328A EP3155225A1 EP 3155225 A1 EP3155225 A1 EP 3155225A1 EP 15726328 A EP15726328 A EP 15726328A EP 3155225 A1 EP3155225 A1 EP 3155225A1
Authority
EP
European Patent Office
Prior art keywords
stages
turbine
operating fluid
expansion
centrifugal
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Granted
Application number
EP15726328.6A
Other languages
German (de)
French (fr)
Other versions
EP3155225B1 (en
Inventor
Roberto Bini
Mario Gaia
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Turboden SpA
Original Assignee
Turboden SpA
Turbodem SRL
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Turboden SpA, Turbodem SRL filed Critical Turboden SpA
Publication of EP3155225A1 publication Critical patent/EP3155225A1/en
Application granted granted Critical
Publication of EP3155225B1 publication Critical patent/EP3155225B1/en
Active legal-status Critical Current
Anticipated expiration legal-status Critical

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01D—NON-POSITIVE DISPLACEMENT MACHINES OR ENGINES, e.g. STEAM TURBINES
    • F01D1/00—Non-positive-displacement machines or engines, e.g. steam turbines
    • F01D1/02—Non-positive-displacement machines or engines, e.g. steam turbines with stationary working-fluid guiding means and bladed or like rotor, e.g. multi-bladed impulse steam turbines
    • F01D1/06—Non-positive-displacement machines or engines, e.g. steam turbines with stationary working-fluid guiding means and bladed or like rotor, e.g. multi-bladed impulse steam turbines traversed by the working-fluid substantially radially
    • F01D1/08—Non-positive-displacement machines or engines, e.g. steam turbines with stationary working-fluid guiding means and bladed or like rotor, e.g. multi-bladed impulse steam turbines traversed by the working-fluid substantially radially having inward flow
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01D—NON-POSITIVE DISPLACEMENT MACHINES OR ENGINES, e.g. STEAM TURBINES
    • F01D1/00—Non-positive-displacement machines or engines, e.g. steam turbines
    • F01D1/02—Non-positive-displacement machines or engines, e.g. steam turbines with stationary working-fluid guiding means and bladed or like rotor, e.g. multi-bladed impulse steam turbines
    • F01D1/12—Non-positive-displacement machines or engines, e.g. steam turbines with stationary working-fluid guiding means and bladed or like rotor, e.g. multi-bladed impulse steam turbines with repeated action on same blade ring
    • F01D1/14—Non-positive-displacement machines or engines, e.g. steam turbines with stationary working-fluid guiding means and bladed or like rotor, e.g. multi-bladed impulse steam turbines with repeated action on same blade ring traversed by the working-fluid substantially radially
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01D—NON-POSITIVE DISPLACEMENT MACHINES OR ENGINES, e.g. STEAM TURBINES
    • F01D5/00—Blades; Blade-carrying members; Heating, heat-insulating, cooling or antivibration means on the blades or the members
    • F01D5/02—Blade-carrying members, e.g. rotors
    • F01D5/04—Blade-carrying members, e.g. rotors for radial-flow machines or engines
    • F01D5/043—Blade-carrying members, e.g. rotors for radial-flow machines or engines of the axial inlet- radial outlet, or vice versa, type
    • F01D5/048—Form or construction
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01D—NON-POSITIVE DISPLACEMENT MACHINES OR ENGINES, e.g. STEAM TURBINES
    • F01D9/00—Stators
    • F01D9/02—Nozzles; Nozzle boxes; Stator blades; Guide conduits, e.g. individual nozzles
    • F01D9/04—Nozzles; Nozzle boxes; Stator blades; Guide conduits, e.g. individual nozzles forming ring or sector
    • F01D9/045—Nozzles; Nozzle boxes; Stator blades; Guide conduits, e.g. individual nozzles forming ring or sector for radial flow machines or engines
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F05—INDEXING SCHEMES RELATING TO ENGINES OR PUMPS IN VARIOUS SUBCLASSES OF CLASSES F01-F04
    • F05D—INDEXING SCHEME FOR ASPECTS RELATING TO NON-POSITIVE-DISPLACEMENT MACHINES OR ENGINES, GAS-TURBINES OR JET-PROPULSION PLANTS
    • F05D2210/00—Working fluids
    • F05D2210/40—Flow geometry or direction
    • F05D2210/44—Flow geometry or direction bidirectional, i.e. in opposite, alternating directions

Definitions

  • the present invention refers to a turbine in which the fluid expands in centripetal and centrifugal directions, and in case axially, and to a method for expanding an operating fluid in such a turbine, in particular an organic fluid in a Rankine cycle.
  • the isentropic enthalpy drop provided by the expansion in the turbine i.e. the maximum work per mass unit the expanding working fluid can produce, expressed for example in kJoule/kg, most of all depends from fluid characteristics and, generally, is a function of the difference between evaporation and condensation temperatures of the fluid.
  • characteristics of the working fluid itself have a great influence on the enthalpy drop that is higher in fluids with simple molecule and low value of molecular mass.
  • expansion ratio both defined as the ratio between the inlet pressure and the exhaust pressure, and as the volume expansion ratio, i.e. the ratio between the volumetric flow rate at the exhaust zone and that one at the turbine inlet zone.
  • ORC Organic Rankine cycle
  • ORC plants are used for the combined production of electrical and thermal power starting from solid biomass; alternatively waste heats of industrial processes, heat recovery from prime movers or geothermal heat sources are used.
  • an ORC plant fed with biomass usually comprises:
  • a heat exchanger provided to give part of the heat of combustion fumes/ gases to a heat-transfer fluid, such as a diathermic oil, delivered by an intermediate circuit;
  • the heat-transfer fluid for example diathermic oil
  • the heat-transfer fluid circulates in a closed loop, passing through the afore mentioned heat exchanger in which the organic operating fluid evaporates.
  • the steam of the operating fluid expands in the turbine, producing mechanic power, that is then transformed into electrical power by the generator connected to the shaft of the turbine itself.
  • the steam of the operating fluid condenses in an appropriate condenser, giving heat to a cooling fluid, usually water, used downstream of the plant as thermal carrier at about 80°C - 90°C, for example for the district heating.
  • the operating fluid is fed to the heat exchanger crossed by the heat-transfer fluid, completing the cycle in closed loop.
  • the present invention is applied in Rankine cycles indifferently of ORC or steam types, in Kalina cycles and, in general, in industrial processes where the expansion of an operating fluid is provided, in cases in which the isentropic enthalpy drop of the turbine is high in relation to the square of the turbine rotation speed - i.e. in the present Application in cases with a drop higher than 40 kJ/kg -, for a turbine rotating at 1500 revolutions per minute (therefore adapted to be directly coupled with a four- pole electric generator at 50 Hz) or 160 kJ/kg for a turbine rotating at 3000 revolutions per minute and so on, and in particular can be used in cycles characterized by high volume expansion ratio of the operating fluid, i.e. higher than 50 in the cases of the present application.
  • a so-called “cantilevered” solution is preferably adopted, meaning that the bearings supporting the shaft are at the same part with respect to the rotor in which the produced power is taken out.
  • Patent applications WO 2010/106569 and WO 2010/106570 in the name of the Applicant, describe cantilevered solutions.
  • Patent application EP 2699767 in the name of Exergy S.p.A., describes a radial centrifugal turbine for applications in ORC Rankine cycles.
  • WO 2013/108099 describes a third solution than can be considered the closest prior art for the present invention.
  • WO 2013/108099 describes a turbine with only one shaft in which the fluid expands in radial centrifugal stages and in axial stages, in succession.
  • At least one array of stator or rotor blades, name angular blades is arranged between the radial stages and the axial stages in order to divert the operating fluid.
  • the enthalpy drop of the operating fluid expanded through the angular blades is equal to at least 50% of the average enthalpy drop provided for completing the fluid expansion in the whole turbine.
  • radial stages can be assembled on the shaft at one end, and axial stages can extend substantially in a cantilever way, so that the turbine has extremely compact size compared with other known solutions, and the bearings, the reduction gear and the electric generator are on the same side, being easily accessed for maintenance operations. From the thermodynamic point of view this solution provides a greater enthalpy drop occurring at the assembly of angular blades and subsequent axial blades.
  • the present invention in a first aspect thereof, relates to a turbine according to claim 1.
  • the present invention refers to a turbine for expanding a compressible operating fluid, for example gas or steam, comprising a plurality of stages defined by arrays of stator blades and arrays of rotor blades.
  • the turbine has only one shaft supporting the arrays of rotor blades.
  • the stator blades are supported by a stationary portion of the turbine, for example a casing thereof.
  • the shaft has a longitudinal axis X-X being the rotation axis, and is radially supported by at least two bearings. If necessary, there can be also one or more axial thrust bearings.
  • At least one group of stages extends in a direction substantially radial with respect to the axis X-X to carry out the centrifugal expansion of the operating fluid.
  • the turbine comprises a group of stages, named centripetal stages, extending in a radial direction with respect to the axis X-X.
  • centripetal stages the operating fluid undergoes a first expansion in the centripetal direction.
  • all the arrays of rotor blades are constrained to the shaft at an end thereof, anyway not in the area between the bearings, i.e. according to a so-called “cantilevered" configuration and particularly advantageous to carry out maintenance operations.
  • the proposed solution allows obtaining high efficiencies without complicating the turbine design, which remains simply to be maintained and can be manufactured with held down costs.
  • the addition of centripetal stages, upstream of the centrifugal stages, where the volumetric flow rate of the operating fluid is typically moderate allows carrying out a first expansion from an outer radial position but anyway not much exceeding that one of the first centrifugal stages, maintaining a high efficiency as the moderate value of the insertion diameter of blades causes a relatively great blade height.
  • centripetal stages are placed on the outside, i.e. on a diameter larger than the assembly constituted by shaft and bearings; this allows the turbine to be partially disassembled, for example by partially taking out the shaft and/or the bearings in order to access to other rotor disks, to carry out inspections or maintenance operations without the need of disassembling the turbine itself completely.
  • centripetal stages have little effect on the axial thrust applied to the shaft, i.e. they do not increase it appreciably compared to what other turbine stages do. This allows the structure of thrust bearings to remain simple.
  • centripetal stages are in the number of 1 to 10, according to the turbine size.
  • rotor arrays of centripetal stages are assembled on a first supporting disk in its turn constrained to the shaft, and rotor arrays of the centrifugal stages are assembled on a second supporting disk.
  • the second supporting disk is constrained to an end of the shaft and the first supporting disk is constrained to the second supporting disk and borne by it.
  • This arrangement is not only particularly compact: it allows the afore described cantilevered configuration to be made. Practically, the first supporting disk rests on the second supporting disk by cantileverly projecting on the shaft portion where the bearings are.
  • the second supporting disk and the shaft, and the second supporting disk and the first supporting disk are coupled by a self-centering toothing of Hirth type, obtained on these components.
  • the second supporting disk is coupled at one end of the shaft having an increased section, in an intermediate position between the same end and the bearings.
  • a channel substantially being U-shaped in the meridian section, is provided between the centripetal stages and the centrifugal stages.
  • the U- shaped channel is partially defined by the first supporting disk and partially by a turbine casing or another stationary component.
  • the operating fluid reverses its own expansion direction.
  • one or more stages are provided and named axial stages, which extend in the axial direction with respect to the axis X-X to carry out an axial expansion of the operating fluid.
  • centrifugal stages upstream of the centripetal stages with respect to the expansion direction, additional centrifugal stages are provided, for example one or more stages.
  • the rotor arrays of the centrifugal stages downstream of the centripetal stages can be assembled, for example, to the second supporting disk.
  • rotor arrays of such stages are supported by the first supporting disk, i.e. the same disk on which rotor blading of the centrifugal stages is assembled.
  • One or more rotor arrays of the axial stages might be supported by a third supporting disk constrained to one end of the shaft having an increased section, at the side opposite to the first supporting disk.
  • the adduction or the extraction of a flow rate of operating fluid can be provided.
  • adductions or extractions of operating fluid can be provided.
  • Another object of the present invention is to provide a method for expanding an operating fluid in a turbine, which allows optimizing the distribution of enthalpy drops of the fluid among different turbine stages, keeping the structure of the turbine compact and easy for maintenance access.
  • the present invention concerns a method, according to claim 14, for expanding a compressible operating fluid, for example gas or steam, in a turbine.
  • the method comprises the steps of:
  • the operating fluid is organic and its expansion happens in a Rankine cycle, or else in a Kalina cycle or, in general, in a thermodynamic cycle providing the expansion of the operating fluid.
  • the method can concern the expansion of every fluid in a process, for example within a process of liquefaction and/or regasification of natural gas.
  • the solution hitherto described in its various embodiments has to be intended as characterized, in addition to the exceeding of the overall enthalpy drop available, also to the exceeding of a threshold characterizing the not-axial part of the turbine, i.e. the threshold of "isentropic k" as described in the following.
  • Ah ( i S ) is the isentropic enthalpy drop available for the stage and (u) is the peripheral speed of the rotor array of the same stage, considered at the average diameter of the same array.
  • Ah ( i S; rad ) is the overall enthalpy drop performed in the radial stages of the turbine, calculated as the difference between the overall enthalpy drop of the turbine and the enthalpy drop performed in the axial portion downstream of the radial portion, and where ui is the peripheral speed at the average diameter of the first axial stage.
  • the calculation of the threshold value can be made with known techniques by taking account of the specific operating fluid of the turbine, the respective operative parameters and the inlet and exhaust conditions of the turbine (which are measurable), whereas the overall enthalpy drop of the axial part can be calculated from an accurate survey of the geometry of the axial arrays themselves, or from the respective CAD files (the array of angular blades assigned to the flow rotation from radial to axial is included, if present).
  • FIG. 1 is a partial section view of a first embodiment of the turbine according to the present invention
  • FIG. 2 is a partial section view of a second embodiment of the turbine according to the present invention.
  • FIG. 3 is a partial section view of a third embodiment of the turbine according to the present invention.
  • FIG. 4 is a partial section view of a fourth embodiment of the turbine according to the present invention.
  • figure 4A is a perspective view of a detail shown in figure 4.
  • FIG. 5 is a partial section view of a fifth embodiment of the turbine according to the present invention.
  • FIG. 6 is a partial section view of a sixth embodiment of the turbine according to the present invention.
  • FIG. 7 is a partial section view of a seventh embodiment of the turbine according to the present invention.
  • FIG. 8 is a partial section view of a eighth embodiment of the turbine according to the present invention.
  • FIG. 9 is a partial section view of a ninth embodiment of the turbine according to the present invention.
  • FIG. 10 is a partial section view of a tenth embodiment of the turbine according to the present invention.
  • Figure 1 is a partial view, in an axially symmetrical section, of a turbine 1 according to the present invention for the expansion of a compressible operating fluid, for example an organic fluid in a Rankine cycle.
  • a compressible operating fluid for example an organic fluid in a Rankine cycle.
  • the turbine comprises a shaft 2, whose longitudinal axis of rotation is shown as X-X, an outer case 3, or volute, and a plurality of expansion stages.
  • the turbine 1 comprises a group of centripetal stages 4 designed to carry out a first expansion of the operating fluid in the radial direction towards the axis X-X, and a group of centrifugal stages 5 designed to carry out a second expansion of the operating fluid in the radial direction, this time moving away from the axis X-X.
  • centripetal stages 4 and the centrifugal stages 5 are defined by arrays of stator blades and arrays of rotor blades.
  • stator and rotor blades of the two stage groups 4 and 5 are shown, respectively.
  • the centripetal stages 4 are characterized by an increasing trend, on the average, of the blade height of the respective arrays, as the axis X-X becomes nearer. In this way, the speed of the transport component speed of the mass is limited at the reversal of the expansion direction of the flow before it comes into the centrifugal stages 5. In particular, at first the transport component is radial centripetal, then axial at the middle of the reversal, at last centrifugal.
  • stator blading can be provided and has the function of straightening the flow before the reversal.
  • a rotor or stator reversing blading can be provided too, in which the flow rotation, from centripetal to centrifugal, happens in an appropriate array, which is rotating or fixed.
  • the arrays being both of rotor or stator type, is characterized by such an expansion ratio, in terms of pressure, to exceed the 10% of the average expansion ratio of all the radial stages, including the reversing arrays herein considered, so that the rotation of the flow is aided by the expansion, in order to reduce the losses.
  • the rotor arrays of the centripetal stages 4 and centrifugal stages 5 are assembled on respective supporting disks 8 and 7 according to the cantilevered configuration now described.
  • the shaft 2 is supported by at least two bearings 9, so that an end 21 of the shaft, which has a section thickened with respect to the central portion of the shaft, extends cantileverly with respect to the bearings 9.
  • all the rotor arrays 42, 52, etc., of the different stages 4, 5 are supported by the end 21 of the shaft, by interposition of the supporting disks 7 and 8.
  • the supporting disk 7 is coupled to the end 21 of the shaft 2 by a self-centering toothing of Hirth type, and the disk 8 is coupled to the disk 7 still by a self-centering toothing of Hirth type.
  • This configuration allows the partial disassembly of the turbine 1 in a practical way, by taking out the shaft 2 from the bearings 9 and "opening" the stages 4 and 5.
  • the disk 8 is provided with a labyrinth seal 10 in the direction of the volute 3, so that to confine the fluid having a higher pressure and make a chamber A, the latter being able to be connected with the other parts of the turbine 1 or plant in which the turbine 1 operates (e.g., the exhaust duct of the turbine or else the condenser in a Rankine cycle), with a convenient lower pressure in order to achieve a compensation of the axial thrust on the disks 7 and 8 and, therefore, on the respective rotor arrays.
  • a labyrinth seal 10 in the direction of the volute 3, so that to confine the fluid having a higher pressure and make a chamber A, the latter being able to be connected with the other parts of the turbine 1 or plant in which the turbine 1 operates (e.g., the exhaust duct of the turbine or else the condenser in a Rankine cycle), with a convenient lower pressure in order to achieve a compensation of the axial thrust on the disks 7 and 8 and, therefore, on the respective rot
  • connection of the chamber A can be of direct type, through convenient ports such as the ports H or K indicating different possible solutions, or else it can happen through one or more ducts that can be also valve-controlled to modulate the compensation effect (controlling input valves can preferably be the inlet and exhaust pressures of the turbine, a measure of the thrust on the shaft, a measure of the axial load on the bearings, the present value of the produced power).
  • the labyrinth Z can be absent, and in this case the chamber A will be connected directly to the exhaust through the port S.
  • FIG 1 the presence of a chamber B is reported as comprised between the labyrinths Q and R and fed by the port Y in connection with an appropriate position of the expansion path.
  • the purpose of the chamber B is to generate an effective thrust compensation on the machine shaft and, hence, the bearings.
  • Figure 2 shows a second embodiment having in addition, compared with the turbine 1 of figure 1, the axial stages 11 placed downstream of the centrifugal stages 5. Between the centrifugal stages 5 and the axial stages 11 the blades 12 are provided, named angular, stator or rotor blades, preferably equivalent to those described in the patent application WO 2013/108099.
  • Figure 3 shows a variation having one more axial stage 13 compared to the turbine 1 of figure 2.
  • the axial stage 13 is supported by another disk 14 constrained directly to the end 21 of the shaft 2 by means of a Hirth toothing, at the opposite side with respect to the disk 7.
  • Figure 4 shows still another embodiment having in addition, compared to that one visible in figure 3, a stator array 15 upstream of the centripetal rotor stages 4.
  • the stator array 15 is provided with nozzles with variable pitch angle, according to known techniques, with the purpose of changing the areas of the channels among the blades, in order to affect the fluid flow rate through the turbine.
  • stator array with variable pitch angle can be the quick stop of the flow rate of operating fluid in case of sudden load interruption, for example on the alternator connected to the turbine.
  • a blade array with variable pitch angle can be added upstream of a centripetal stator blading, instead of a rotor blading.
  • Figure 5 shows an embodiment having in addition, compared to that one visible in figure 3, a chamber P of adduction or extraction of the operating fluid added or taken away, through the adduction or extraction duct 16, downstream of the centrifugal stages 5 and upstream of the angular blades 12.
  • Figure 6 shows a sixth embodiment of the turbine 1, comprising (in the example of Figure), five centrifugal stages 5, angular blades 12, axial stages 13 and radial exhaust of the operating fluid.
  • the shaft 2 extends on the opposite side with respect to the adduction of the operating fluid, which is frontal i.e. axial, in the turbine.
  • a partition F isolates a chamber L (no indication in Figure) placed in communication with a low-pressure point, so that to compensate the axial thrust, similarly to what referred for the previous versions.
  • Figure 7 shows a seventh embodiment in which the fluid enters the turbine frontally, in the axial direction, and additional centrifugal stages 18 are provided upstream of the centripetal stages 4.
  • Figure 8 shows an eighth embodiment provided with centripetal stages 4, centrifugal stages 18 and 5, and axial stages.
  • additional connections extracting or adducting fluid with intermediate pressures are reported as shown with the letters M and N, in addition to the already considered connection P.
  • Figure 9 shows a ninth embodiment differing from the first one in that a stator blading Si is provided in the channel 6 as constrained to the volute 3 and has the function of reversing the expansion way of the operating fluid from radial centripetal one to radial centrifugal one.
  • Figure 10 shows a tenth embodiment differing from the first one in that a rotor blading Ri is provided in the channel 6 as constrained to the disk 8 and has the function of reversing the expansion way of the operating fluid from radial centripetal one to radial centrifugal one.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Engine Equipment That Uses Special Cycles (AREA)
  • Turbine Rotor Nozzle Sealing (AREA)

Abstract

A turbine (1), at least partially centrifugal, for the expansion of a compressible operating fluid, for example gas or steam. At least one group of stages (5), named centrifugal stages, extends in a radial direction with respect to the axis X-X to carry out the centrifugal expansion of the operating fluid. Advantageously, the turbine comprises a group of stages, named centripetal stages (4), extending in a radial direction to carry out a first expansion of the operating fluid centripetally in the radial direction. Moreover, all the arrays of rotor blades are constrained to the shaft (2) at an end thereof, anyway not in the area between the bearings (9), i.e. according to a so-called "cantilevered" configuration and particularly advantageous to carry out maintenance operations. The proposed solution allows high efficiencies to be achieved by a compact turbine. A corresponding method for expanding the operating fluid is further described.

Description

TURBINE AND METHOD FOR EXPANDING AN OPERATING FLUID
***
Field Of The invention
The present invention refers to a turbine in which the fluid expands in centripetal and centrifugal directions, and in case axially, and to a method for expanding an operating fluid in such a turbine, in particular an organic fluid in a Rankine cycle. State of the Art
In every Rankine cycle and for a given operating fluid in the subcritical zone, the isentropic enthalpy drop provided by the expansion in the turbine, i.e. the maximum work per mass unit the expanding working fluid can produce, expressed for example in kJoule/kg, most of all depends from fluid characteristics and, generally, is a function of the difference between evaporation and condensation temperatures of the fluid. On the other hand, characteristics of the working fluid itself have a great influence on the enthalpy drop that is higher in fluids with simple molecule and low value of molecular mass.
Computing the isentropic enthalpy drop is a well known point when studying Rankine cycles; the turbine designer carries out the turbine design by using, as starting values, the composition of the operating fluid, the inlet values of temperature, flow rate, pressure and titer of the fluid, as well as the value of exhaust pressure. From these data the value of the isentropic enthalpy drop can be easily calculated with known methods, therefore such a value has to be interpreted as a characterizing parameter in turbine design.
The same data take part of the turbine design also in case in which the turbine itself is included in a power cycle different from the Rankine cycle (for example a Kalina cycle or a Brayton cycle), or even if it is not part of a cycle but belongs to a thermodynamic process of different nature (the example of an expander of natural gas placed at the end of a distribution duct of the gas itself counts).
Another aspect calculable from afore mentioned data is the expansion ratio, both defined as the ratio between the inlet pressure and the exhaust pressure, and as the volume expansion ratio, i.e. the ratio between the volumetric flow rate at the exhaust zone and that one at the turbine inlet zone.
The acronym ORC "Organic Rankine cycle" , as everyone knows, identifies the thermodynamic cycles of Rankine type that use an organic operating fluid preferably provided with high molecular mass, much higher than that of the water vapor used in most of the Rankine power cycles.
For example, ORC plants are used for the combined production of electrical and thermal power starting from solid biomass; alternatively waste heats of industrial processes, heat recovery from prime movers or geothermal heat sources are used.
For example an ORC plant fed with biomass usually comprises:
- a combustion chamber fed with fuel biomass;
- a heat exchanger provided to give part of the heat of combustion fumes/ gases to a heat-transfer fluid, such as a diathermic oil, delivered by an intermediate circuit;
- a heat exchanger provided to give part of the heat of the intermediate heat- transfer fluid to an operating fluid to be evaporated;
- a turbine fed with an operating fluid in the vapor state; and
- an electric generator activated by the turbine for the production of electrical power.
In the combustion chamber the heat-transfer fluid, for example diathermic oil, is heated up to a temperature usually equal to about 300°C. The heat-transfer fluid circulates in a closed loop, passing through the afore mentioned heat exchanger in which the organic operating fluid evaporates. The steam of the operating fluid expands in the turbine, producing mechanic power, that is then transformed into electrical power by the generator connected to the shaft of the turbine itself. As the respective expansion in the turbine will end, the steam of the operating fluid condenses in an appropriate condenser, giving heat to a cooling fluid, usually water, used downstream of the plant as thermal carrier at about 80°C - 90°C, for example for the district heating. The operating fluid is fed to the heat exchanger crossed by the heat-transfer fluid, completing the cycle in closed loop.
Generally, the present invention is applied in Rankine cycles indifferently of ORC or steam types, in Kalina cycles and, in general, in industrial processes where the expansion of an operating fluid is provided, in cases in which the isentropic enthalpy drop of the turbine is high in relation to the square of the turbine rotation speed - i.e. in the present Application in cases with a drop higher than 40 kJ/kg -, for a turbine rotating at 1500 revolutions per minute (therefore adapted to be directly coupled with a four- pole electric generator at 50 Hz) or 160 kJ/kg for a turbine rotating at 3000 revolutions per minute and so on, and in particular can be used in cycles characterized by high volume expansion ratio of the operating fluid, i.e. higher than 50 in the cases of the present application.
In case of turbines with shaft power of about 20 MW, a so-called "cantilevered" solution is preferably adopted, meaning that the bearings supporting the shaft are at the same part with respect to the rotor in which the produced power is taken out. In fact it is an easier solution from the implementation point of view, needing only one rotating seal, it is cost-effective and can be maintained more easily than a solution with a rotor comprised between the bearings.
Patent applications WO 2010/106569 and WO 2010/106570, in the name of the Applicant, describe cantilevered solutions.
Patent application EP 2699767, in the name of Exergy S.p.A., describes a radial centrifugal turbine for applications in ORC Rankine cycles.
The International patent application WO 2013/108099, in the name of the Applicant, describes a third solution than can be considered the closest prior art for the present invention. In particular WO 2013/108099 describes a turbine with only one shaft in which the fluid expands in radial centrifugal stages and in axial stages, in succession. At least one array of stator or rotor blades, name angular blades, is arranged between the radial stages and the axial stages in order to divert the operating fluid. The enthalpy drop of the operating fluid expanded through the angular blades is equal to at least 50% of the average enthalpy drop provided for completing the fluid expansion in the whole turbine. From the structural point of view radial stages can be assembled on the shaft at one end, and axial stages can extend substantially in a cantilever way, so that the turbine has extremely compact size compared with other known solutions, and the bearings, the reduction gear and the electric generator are on the same side, being easily accessed for maintenance operations. From the thermodynamic point of view this solution provides a greater enthalpy drop occurring at the assembly of angular blades and subsequent axial blades.
Object and Summary of the Invention
It is an object of the present invention to provide a turbine having little size and cantilevered configuration, simple from the structural point of view and characterized by the splitting of the enthalpy drop among an optimal number of stages and by high efficiency also in first expansion stages of the operating fluid, where the volumetric flow rate of the operating fluid is typically minimal and a good efficiency can be achieved with more difficulty.
It is another object of the present invention to provide a low cost turbine, which is mechanically robust and allows the effective containment of axial thrust the rotor apply on the shaft.
Therefore the present invention, in a first aspect thereof, relates to a turbine according to claim 1.
In particular, the present invention refers to a turbine for expanding a compressible operating fluid, for example gas or steam, comprising a plurality of stages defined by arrays of stator blades and arrays of rotor blades. Preferably the turbine has only one shaft supporting the arrays of rotor blades. Obviously, the stator blades are supported by a stationary portion of the turbine, for example a casing thereof.
The shaft has a longitudinal axis X-X being the rotation axis, and is radially supported by at least two bearings. If necessary, there can be also one or more axial thrust bearings.
At least one group of stages, named centrifugal stages, extends in a direction substantially radial with respect to the axis X-X to carry out the centrifugal expansion of the operating fluid.
Advantageously, the turbine comprises a group of stages, named centripetal stages, extending in a radial direction with respect to the axis X-X. In centripetal stages, the operating fluid undergoes a first expansion in the centripetal direction. Moreover, all the arrays of rotor blades are constrained to the shaft at an end thereof, anyway not in the area between the bearings, i.e. according to a so-called "cantilevered" configuration and particularly advantageous to carry out maintenance operations.
The proposed solution allows obtaining high efficiencies without complicating the turbine design, which remains simply to be maintained and can be manufactured with held down costs. As a matter of fact the addition of centripetal stages, upstream of the centrifugal stages, where the volumetric flow rate of the operating fluid is typically moderate, allows carrying out a first expansion from an outer radial position but anyway not much exceeding that one of the first centrifugal stages, maintaining a high efficiency as the moderate value of the insertion diameter of blades causes a relatively great blade height. Hence, the so-called secondary losses and the losses due to leakage at the blade end, or in the corresponding labyrinth, are acceptable; the turbine is still anyway compact and robust, since the centripetal stages extend substantially in a radial way and have a minimal bulk in the axial direction. Exploiting the enthalpy drop available at the turbine is more efficient than what can be ascertained in known solutions, where the expansion ratio per stage is excessive and/or the aerodynamic load on blades is excessive; the proposed solution allows distributing the enthalpy drop on a greater number of stages, almost with the same bulks with respect to known solutions, to efficiency advantage.
These advantages are particularly evident in case in which the turbine is included in a thermodynamic cycle characterized by high enthalpy drops, with relation to the rotation speed, greater than 40 kJ per kg of operating fluid for a machine rotating at 1500 revolutions per minute.
Furthermore the centripetal stages are placed on the outside, i.e. on a diameter larger than the assembly constituted by shaft and bearings; this allows the turbine to be partially disassembled, for example by partially taking out the shaft and/or the bearings in order to access to other rotor disks, to carry out inspections or maintenance operations without the need of disassembling the turbine itself completely.
Another advantage of the proposed solution is that the centripetal stages have little effect on the axial thrust applied to the shaft, i.e. they do not increase it appreciably compared to what other turbine stages do. This allows the structure of thrust bearings to remain simple.
For example, the centripetal stages are in the number of 1 to 10, according to the turbine size.
Preferably, rotor arrays of centripetal stages are assembled on a first supporting disk in its turn constrained to the shaft, and rotor arrays of the centrifugal stages are assembled on a second supporting disk. The second supporting disk is constrained to an end of the shaft and the first supporting disk is constrained to the second supporting disk and borne by it. This arrangement is not only particularly compact: it allows the afore described cantilevered configuration to be made. Practically, the first supporting disk rests on the second supporting disk by cantileverly projecting on the shaft portion where the bearings are.
In the preferred embodiment, the second supporting disk and the shaft, and the second supporting disk and the first supporting disk, are coupled by a self-centering toothing of Hirth type, obtained on these components.
Preferably, the second supporting disk is coupled at one end of the shaft having an increased section, in an intermediate position between the same end and the bearings.
In an embodiment a channel, substantially being U-shaped in the meridian section, is provided between the centripetal stages and the centrifugal stages. The U- shaped channel is partially defined by the first supporting disk and partially by a turbine casing or another stationary component. In the U-shaped channel the operating fluid reverses its own expansion direction.
Different embodiments of the turbine can be envisaged. For example, in an embodiment downstream of the centrifugal stages with respect to the expansion direction, one or more stages are provided and named axial stages, which extend in the axial direction with respect to the axis X-X to carry out an axial expansion of the operating fluid.
As an alternative or in addition, upstream of the centripetal stages with respect to the expansion direction, additional centrifugal stages are provided, for example one or more stages. In this case, the rotor arrays of the centrifugal stages downstream of the centripetal stages can be assembled, for example, to the second supporting disk.
In an embodiment providing axial stages, rotor arrays of such stages are supported by the first supporting disk, i.e. the same disk on which rotor blading of the centrifugal stages is assembled. One or more rotor arrays of the axial stages might be supported by a third supporting disk constrained to one end of the shaft having an increased section, at the side opposite to the first supporting disk.
For example, among the centripetal stages and the centrifugal stages, the adduction or the extraction of a flow rate of operating fluid can be provided. Similarly, among the centrifugal stages and the axial stages, adductions or extractions of operating fluid can be provided.
Another object of the present invention is to provide a method for expanding an operating fluid in a turbine, which allows optimizing the distribution of enthalpy drops of the fluid among different turbine stages, keeping the structure of the turbine compact and easy for maintenance access.
Therefore, in its second aspect the present invention concerns a method, according to claim 14, for expanding a compressible operating fluid, for example gas or steam, in a turbine.
Particularly the method comprises the steps of:
- prearranging a turbine according to the present invention, i.e. having the afore described characteristics;
- feeding the operating fluid to the turbine and carrying out at least one first expansion in the centripetal direction, reversing the direction of the operating fluid and carrying out a second expansion in the centrifugal direction.
Advantages offered by the method are the same described with relation to the turbine.
Preferably, the operating fluid is organic and its expansion happens in a Rankine cycle, or else in a Kalina cycle or, in general, in a thermodynamic cycle providing the expansion of the operating fluid. Alternatively, the method can concern the expansion of every fluid in a process, for example within a process of liquefaction and/or regasification of natural gas.
The solution hitherto described in its various embodiments has to be intended as characterized, in addition to the exceeding of the overall enthalpy drop available, also to the exceeding of a threshold characterizing the not-axial part of the turbine, i.e. the threshold of "isentropic k" as described in the following.
In general, with "isentropic k" of a stage is meant the ratio:
(l) k(is) = Ah(is) / (u2/2),
where Ah(iS) is the isentropic enthalpy drop available for the stage and (u) is the peripheral speed of the rotor array of the same stage, considered at the average diameter of the same array.
For what concerns the above characterizing threshold, the referred ratio is, on the contrary:
(2) k'(is) = Ah(is, rad) / (ui2/2),
where Ah(iS; rad) is the overall enthalpy drop performed in the radial stages of the turbine, calculated as the difference between the overall enthalpy drop of the turbine and the enthalpy drop performed in the axial portion downstream of the radial portion, and where ui is the peripheral speed at the average diameter of the first axial stage.
Well, by using the formula (2), the threshold value rendering the proposed solution as considerably advantageous is 7 (seven).
The described criterion does not take account, to all intents and purposes, of the single radial stage but of the overall behavior of the radial stages, the physical limits of the number of radial stages having been taken account, which cannot be arranged in succession. Reference enthalpies have to be intended as overall enthalpies and not as static enthalpy.
If the proposed solution is introduced as retrofit in an existing radial-axial turbine, the calculation of the threshold value can be made with known techniques by taking account of the specific operating fluid of the turbine, the respective operative parameters and the inlet and exhaust conditions of the turbine (which are measurable), whereas the overall enthalpy drop of the axial part can be calculated from an accurate survey of the geometry of the axial arrays themselves, or from the respective CAD files (the array of angular blades assigned to the flow rotation from radial to axial is included, if present).
Brief Description of the Drawings
Further details of the invention will be evident anyway from the following description course made with reference to the attached drawings, in which:
- figure 1 is a partial section view of a first embodiment of the turbine according to the present invention;
- figure 2 is a partial section view of a second embodiment of the turbine according to the present invention;
- figure 3 is a partial section view of a third embodiment of the turbine according to the present invention;
- figure 4 is a partial section view of a fourth embodiment of the turbine according to the present invention;
- figure 4A is a perspective view of a detail shown in figure 4;
- figure 5 is a partial section view of a fifth embodiment of the turbine according to the present invention;
- figure 6 is a partial section view of a sixth embodiment of the turbine according to the present invention;
- figure 7 is a partial section view of a seventh embodiment of the turbine according to the present invention;
- figure 8 is a partial section view of a eighth embodiment of the turbine according to the present invention;
- figure 9 is a partial section view of a ninth embodiment of the turbine according to the present invention;
- figure 10 is a partial section view of a tenth embodiment of the turbine according to the present invention.
Detailed Description Of The Invention
Figure 1 is a partial view, in an axially symmetrical section, of a turbine 1 according to the present invention for the expansion of a compressible operating fluid, for example an organic fluid in a Rankine cycle.
The turbine comprises a shaft 2, whose longitudinal axis of rotation is shown as X-X, an outer case 3, or volute, and a plurality of expansion stages.
In particular, the turbine 1 comprises a group of centripetal stages 4 designed to carry out a first expansion of the operating fluid in the radial direction towards the axis X-X, and a group of centrifugal stages 5 designed to carry out a second expansion of the operating fluid in the radial direction, this time moving away from the axis X-X.
The centripetal stages 4 and the centrifugal stages 5 are defined by arrays of stator blades and arrays of rotor blades. For example, with numeral references 41, 42 and 51, 52, stator and rotor blades of the two stage groups 4 and 5 are shown, respectively.
The centripetal stages 4 are characterized by an increasing trend, on the average, of the blade height of the respective arrays, as the axis X-X becomes nearer. In this way, the speed of the transport component speed of the mass is limited at the reversal of the expansion direction of the flow before it comes into the centrifugal stages 5. In particular, at first the transport component is radial centripetal, then axial at the middle of the reversal, at last centrifugal.
The reversal of the expansion direction happens at the channel 6. Immediately upstream of the channel 6 a stator blading can be provided and has the function of straightening the flow before the reversal.
In the channel 6 a rotor or stator reversing blading can be provided too, in which the flow rotation, from centripetal to centrifugal, happens in an appropriate array, which is rotating or fixed. In this case the arrays, being both of rotor or stator type, is characterized by such an expansion ratio, in terms of pressure, to exceed the 10% of the average expansion ratio of all the radial stages, including the reversing arrays herein considered, so that the rotation of the flow is aided by the expansion, in order to reduce the losses.
The rotor arrays of the centripetal stages 4 and centrifugal stages 5 are assembled on respective supporting disks 8 and 7 according to the cantilevered configuration now described. The shaft 2 is supported by at least two bearings 9, so that an end 21 of the shaft, which has a section thickened with respect to the central portion of the shaft, extends cantileverly with respect to the bearings 9. Well, all the rotor arrays 42, 52, etc., of the different stages 4, 5, are supported by the end 21 of the shaft, by interposition of the supporting disks 7 and 8.
In particular, the supporting disk 7 is coupled to the end 21 of the shaft 2 by a self-centering toothing of Hirth type, and the disk 8 is coupled to the disk 7 still by a self-centering toothing of Hirth type.
This configuration allows the partial disassembly of the turbine 1 in a practical way, by taking out the shaft 2 from the bearings 9 and "opening" the stages 4 and 5.
The disk 8 is provided with a labyrinth seal 10 in the direction of the volute 3, so that to confine the fluid having a higher pressure and make a chamber A, the latter being able to be connected with the other parts of the turbine 1 or plant in which the turbine 1 operates (e.g., the exhaust duct of the turbine or else the condenser in a Rankine cycle), with a convenient lower pressure in order to achieve a compensation of the axial thrust on the disks 7 and 8 and, therefore, on the respective rotor arrays.
The connection of the chamber A can be of direct type, through convenient ports such as the ports H or K indicating different possible solutions, or else it can happen through one or more ducts that can be also valve-controlled to modulate the compensation effect (controlling input valves can preferably be the inlet and exhaust pressures of the turbine, a measure of the thrust on the shaft, a measure of the axial load on the bearings, the present value of the produced power).
As an additional alternative the labyrinth Z can be absent, and in this case the chamber A will be connected directly to the exhaust through the port S.
In figure 1 the presence of a chamber B is reported as comprised between the labyrinths Q and R and fed by the port Y in connection with an appropriate position of the expansion path. The purpose of the chamber B is to generate an effective thrust compensation on the machine shaft and, hence, the bearings.
Figure 2 shows a second embodiment having in addition, compared with the turbine 1 of figure 1, the axial stages 11 placed downstream of the centrifugal stages 5. Between the centrifugal stages 5 and the axial stages 11 the blades 12 are provided, named angular, stator or rotor blades, preferably equivalent to those described in the patent application WO 2013/108099.
Also the rotor arrays of the axial stages 11 are assembled on the supporting disk 7.
There is also a labyrinth lObis downstream of the axial stage, and the course of the channel downstream of the axial stage is provided with an inner ring W to allow an effective diffusion recovering part of kinetic energy present at the outlet of the axial array. The pressure in the chamber C is maintained about the exhaust one through the ports J.
Figure 3 shows a variation having one more axial stage 13 compared to the turbine 1 of figure 2. The axial stage 13 is supported by another disk 14 constrained directly to the end 21 of the shaft 2 by means of a Hirth toothing, at the opposite side with respect to the disk 7.
Figure 4 shows still another embodiment having in addition, compared to that one visible in figure 3, a stator array 15 upstream of the centripetal rotor stages 4. The stator array 15 is provided with nozzles with variable pitch angle, according to known techniques, with the purpose of changing the areas of the channels among the blades, in order to affect the fluid flow rate through the turbine.
An additional purpose of the stator array with variable pitch angle can be the quick stop of the flow rate of operating fluid in case of sudden load interruption, for example on the alternator connected to the turbine. For the same purpose a blade array with variable pitch angle can be added upstream of a centripetal stator blading, instead of a rotor blading.
This solution allows avoiding the increase of the rotation speed of the turbine shaft due to the load disjunction, before the conventional valves upstream of the turbine can cut-off the flow and before the flow rate already canalized in the turbine completes the expansion.
Figure 5 shows an embodiment having in addition, compared to that one visible in figure 3, a chamber P of adduction or extraction of the operating fluid added or taken away, through the adduction or extraction duct 16, downstream of the centrifugal stages 5 and upstream of the angular blades 12.
Figure 6 shows a sixth embodiment of the turbine 1, comprising (in the example of Figure), five centrifugal stages 5, angular blades 12, axial stages 13 and radial exhaust of the operating fluid. In this variation, the shaft 2 extends on the opposite side with respect to the adduction of the operating fluid, which is frontal i.e. axial, in the turbine.
A partition F isolates a chamber L (no indication in Figure) placed in communication with a low-pressure point, so that to compensate the axial thrust, similarly to what referred for the previous versions.
Figure 7 shows a seventh embodiment in which the fluid enters the turbine frontally, in the axial direction, and additional centrifugal stages 18 are provided upstream of the centripetal stages 4.
Figure 8 shows an eighth embodiment provided with centripetal stages 4, centrifugal stages 18 and 5, and axial stages. In the drawing relating to this variation, also additional connections extracting or adducting fluid with intermediate pressures are reported as shown with the letters M and N, in addition to the already considered connection P.
Figure 9 shows a ninth embodiment differing from the first one in that a stator blading Si is provided in the channel 6 as constrained to the volute 3 and has the function of reversing the expansion way of the operating fluid from radial centripetal one to radial centrifugal one.
Figure 10 shows a tenth embodiment differing from the first one in that a rotor blading Ri is provided in the channel 6 as constrained to the disk 8 and has the function of reversing the expansion way of the operating fluid from radial centripetal one to radial centrifugal one.

Claims

1. A turbine (1) for the expansion of a compressible operating fluid, for example gas or steam, comprising a plurality of stages (4, 5) defined by arrays of stator blades (41, 51) and arrays of rotor blades (42, 52), and a supporting shaft (2) to support the arrays of rotor blades (42, 52), wherein the shaft (2) has a longitudinal rotation axis X-X and is, in its turn, supported by at least two bearings, and wherein at least one group of stages (5), named centrifugal stages, extends in a radial direction with respect to the axis X-X to carry out a centrifugal expansion of the operating fluid,
characterized in that it comprises a group of stages (4), named centripetal stages, extending in a radial direction with respect to the axis X-X to carry out the initial expansion of the operating fluid in the centripetal direction and
in that all the arrays of rotor blades (42, 52) are constrained to the shaft at an end thereof, anyway not in the area between the bearings.
2. Turbine (1) according to claim 1, wherein the shaft (2) is a single shaft and the number of centripetal stages (4) is comprised between 1 and 10.
3. Turbine (1) according to any one of claims 1-2, wherein the rotor arrays (42) of the centripetal stages (4) are assembled on a first supporting disk (8) and the rotor arrays (52) of the centrifugal stages (5) are assembled on a second supporting disk (7), and wherein the second supporting disk (7) is constrained to an end of the shaft (2) and the first supporting disk (8) is constrained to the second supporting disk (7) and borne by it.
4. Turbine (1) according to claim 3, wherein the second supporting disk and the shaft, and the second supporting disk and the first supporting disk, are coupled by a self-centering toothing of Hirth type.
5. Turbine (1) according to claim 3 or claim 4, wherein the second supporting disk (7) is coupled at one end of the shaft (2) having an increased section, in an intermediate position between the same end and the bearings.
6. Turbine (1) according to any one of claims 3-5, wherein a channel (6), substantially being U-shaped in the meridian section and being defined by the first supporting disk (8) and partially by a turbine casing (3), is provided between the centripetal stages (4) and the centrifugal stages (5), the operating fluid reversing its own expansion direction along such a channel.
7. Turbine (1) according to claim 6, wherein a stator blade assembly (Si) or a rotor blade assembly (Ri) is provided in the channel (6).
8. Turbine (1) according to any one of claims 1-7 wherein, downstream of the centrifugal stages (5) with respect to the expansion direction, one or more stages are provided and named axial stages, which extend in the axial direction with respect to the axis X-X to carry out an axial expansion of the operating fluid, and/or
wherein, upstream of the centripetal stages with respect to the expansion direction, one or more additional centrifugal stages are provided.
9. Turbine (1) according to claim 8, wherein the rotor arrays of the axial stages are supported by at least one second supporting disk (7) and, in case, by a third supporting disk constrained to one end of the shaft having an increased section, at the side opposite to the second supporting disk (7).
10. Turbine (1) according to claim 8 or claim 9, wherein also the rotor arrays of the centrifugal stages downstream of the centripetal stages are assembled to the second supporting disk.
11. Turbine (1) according to any one of preceding claims 1-10, wherein among the centripetal stages (4) and the centrifugal stages (5) the adduction or the extraction of a flow rate of an operating fluid is provided.
12. Turbine (1) according to any one of preceding claims 8-11, wherein k'(iS) is greater or equal to 7 (seven), where
k'(is) = Ah(iS; rad) / (ui2/2),
and where Ah(iS; rad) is the overall enthalpy drop occurring in the radial stages (4, 5), calculated as the difference between the overall enthalpy drop of the turbine (1) and the enthalpy drop occurring in the axial stages (11) downstream of the radial stages (5), and where ui is the peripheral speed at the average diameter of the first axial stage.
13. Turbine (1) according to any one of preceding claims 1-12 further comprising, upstream of the first stage, an adjusting stator array to adjust the flow rate of the operating fluid, wherein the adjusting stator array is provided with blades (15) having a variable pitch angle which have the function of varying the passage section available for the operating fluid through the turbine and thereby adjusting the flow rate between a minimum value, near to zero, and a maximum value.
14. A method for expanding a compressible operating fluid, for example gas or steam, in a turbine (1), comprising the steps of:
- prearranging a turbine according to any one of claims 1-13;
- feeding the operating fluid to the turbine and carrying out at least one first expansion in the centripetal direction, reversing the direction of the operating fluid and carrying out a second expansion in the centrifugal direction.
15. Method according to claim 14, further comprising one or more of the following steps:
- injecting or extracting a flow rate of operating fluid at the area (6) of direction reversal;
- upstream of the centripetal expansion, carrying out a centrifugal expansion of the operating fluid in corresponding centrifugal initial stages;
- downstream of the last centrifugal expansion, diverting the operating fluid and carrying out an expansion in an axial direction;
- downstream of the last centrifugal expansion, diverting the operating fluid and carrying out an expansion in an axial direction so that the following conditions are fulfilled:
k'(is) = Ah(iS; rad) / (ui2/2),
where Ah(iS; rad) is the overall enthalpy drop performed in the radial stages (4, 5), calculated as the difference between the overall enthalpy drop of the turbine (1) and the enthalpy drop performed in the axial stages (11) downstream of the radial stages (5), and where ui is the average peripheral speed of the first axial stage.
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