Note: Descriptions are shown in the official language in which they were submitted.
TURBINE WITH CENTRIPETAL AND CENTRIFUGAL EXPANSION
STAGES AND RELATED METHOD
***
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,
to 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
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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
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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
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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
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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
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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
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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
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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:
(1) k(10 = Akis) / (u2/2),
where Ak,$) 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'(6) = Ah(s, rid) / (u12/2),
where Ah(i9. 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 ul 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
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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,
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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
100/0 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.
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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
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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 lithe 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 'Obis 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
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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.
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