Note: Descriptions are shown in the official language in which they were submitted.
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HYBRID COMPRESSOR
TECHNICAL FIELD
[0001] The technical field relates generally to air compressors and/or
turbines, for
example of the type employed in aero gas turbine engines and the like.
BACKGROUND
[0002] Centrifugal compressors achieve compression primarily through the
increasing
radius of their impellers. High compression can be achieved in a single
compact stage,
but since the velocities increase continuously through the impeller,
supersonic velocities
can be experienced at the exit of the impeller, which can increase the losses
and decrease
the efficiency of the compressor. Axial compressors, on the other hand,
typically employ
multiple stages of relatively lower compression to cumulatively achieve high
compression. A stationary blade row, often referred to as stator vanes, after
each rotary
blade row reduces the velocity before entering the next rotor. Thus, although
total
compression increases with each stage, velocity does not increase at the same
rate. The
result is that multi-stage axial compressors are generally more efficient, for
a given
pressure ratio, than centrifugal compressors. The trade-off, however is that
axial
compressors tend to be longer and, in many cases, the compact size and low
cost of a
centrifugal compressor can outweigh the efficiency penalty, where weight/size
are
critical, such as in small gas turbine engines in prime mover applications.
Nonetheless
the trade-off is not always satisfactory, and the desire for better solutions
remains.
[0003] Although, it has been generally recognized that high efficiency may be
realized
by providing a series of compression-diffusion stages along the flow path of
constant or
near constant diameter (i.e. in conventional axial compressors), prior art
like US Patent
Nos. 2,350,839 and 4,428,715, and French patent publication 972751A
respectively
teach employing multiple compression-diffusion stages along the flow path of
increasing
diameter. Both teach limiting the pressure ratio, radius change and or air
speeds in each
stage, such that a series of axial stages are provided in what otherwise looks
like the
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envelope of a centrifugal compressor. However, these compressor are relatively
complicated and heavy relative to the single stage centrifugal compressor.
Such
compressors would also tend to exhibit poor surge/stall characteristics if
they
incorporated into an aero gas turbine engine. Thus, there remains a need for
improved
efficiency, packaging, weight, cost, durability and/or operability, to name
but a few, in
compressor/turbine rotor design.
SUMMARY
[00041 In one aspect, there is provided a gas compressor comprising a rotor
rotatable
about a compressor axis and cooperating with an annular compressor shroud to
define a
gas path extending from an upstream compressor inlet to a downstream
compressor
outlet, the inlet substantially parallel with the compressor axis and the
outlet being non-
parallel with the compressor axis, the rotor having first and second
compression stages,
the first compression stage comprising a circumferential array of blades
extending from a
rotor hub towards the shroud, the second compression stage downstream of the
first and
comprising a circumferential array of centrifugal-like compressor vanes
extending from
the hub towards the shroud, each compression stage having a mean outlet radius
greater
than its mean inlet radius, the second compression stage having a pressure
ratio
exceeding 2:1, the compressor further comprising a diffusion stage interposed
between
said compressor stages, the diffusion stage comprising a circumferential array
of vanes
extending from the shroud toward the rotor hub, the diffusion stage vanes
disposed
between trailing edges of the first stage blades and leading edges of the
second stage
vanes, diffusion stage vanes inclined oppositely to the compression stages to
decrease
tangential velocity of gas exiting the first compression stage.
[0005] In a second aspect, there is provided a gas compressor comprising a
single axial-
like compression-diffusion stage followed by a single centrifugal-like
compression stage,
the compressor having an axial inlet and a substantially radial outlet, the
axial-like
compression-diffusion stage including an a circumferential array of axial-like
blades
extending from a rotatable rotor of the compressor immediately downstream of
the inlet
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and extending towards a shroud, the array of axial-like blades having a
pressure ratio not
greater than 2:1, the axial-like compression-diffusion stage further including
a
circumferential array of diffusion vanes located downstream of trailing edges
of the
axial-like blades and extending from the shroud toward the rotor, the
centrifugal-like
compression stage including a circumferential array of centrifugal-like
compressor vanes
extending from the rotor towards the shroud, the array of centrifugal-like
compressor
vanes having a pressure ratio exceeding 2:1, the array of centrifugal-like
compressor
vanes having trailing edges immediately upstream of the outlet and leading
edges
downstream of trailing edges of the diffusion vane, the diffusion vanes
inclined
oppositely to the axial-like blades to decrease tangential velocity of gas
exiting the axial-
like blade array.
[00061 In a third aspect, there is provided a gas turbine engine comprising a
compressor
having a rotor and shroud defining a compressor gas path extending from a
substantially
axial inlet to a substantially radial outlet, the compressor including at
least two
compression stages and one diffusion stage between the inlet and outlet, the
compression
stages comprising respective circumferential arrays of blades extending from
the rotor,
the compression stages arranged serially relative to one another between the
inlet and
outlet, the diffusion stage comprising a circumferential array of vanes
located serially
between said compression stages, the diffusion stage vanes extending from the
shroud
toward the rotor, the compressor having at least one bleed outlet located
between the first
and second compressor stages configured for bleeding from the gas path.
[00071 Further details of these and other aspects will be apparent from the
detailed
description and figures included below.
DESCRIPTION OF THE DRAWINGS
[00081 Reference is now made to the accompanying figures, in which:
[00091 Fig. 1 schematically shows a gas turbine engine incorporating a
compressor
according to the present description;
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[0010] Fig. 2 is a schematic cross-sectional view of the upper half of the
compressor of
Figure 1;
[0011] Fig. 3 is a schematic cross-sectional view of the upper half of another
example
of the compressor of Fig.
[0012] Fig. 4 is a representative view taken along line IV-IV of Figs. 2
and/or 3,
illustrating the configurations and relative inclinations of successive
stages; and
[0013] Fig. 5 is a schematic cross-sectional view of the upper half of another
example
of the compressor of Fig. 1;
[0014] Figs. 6-8 schematically show gas turbine engines incorporating a
compressor
according to other aspects of the present description.
DETAILED DESCRIPTION
[0015] Fig. 1 schematically illustrates a turboshaft gas turbine engine 10,
generally
comprising in serial flow communication an inlet 12 through which ambient air
is drawn,
a compressor 20 for pressurizing the air, a combustor 16 in which the
compressed air is
mixed with fuel and ignited for generating hot combustion gases, a turbine
section 18 for
extracting mechanical energy from the combustion gases, and a mechanical
output, such
as a reduction gearbox 19 in this example. The compressor 20 has an
centrifugal
"hybrid" design, as will now be further described.
[0016] Referring now to Fig. 2, the "hybrid" compressor 20 has a generally
axially-
oriented inlet 22 (i.e. airflow generally parallel (0 ) to axis M) and in this
example a
generally radial outlet 24 (i.e. airflow generally perpendicular ( 90 ) to
axis M),
communicating via a gas path 26. A first stage of axial-like compressor blades
30 is
provided extending from an inducer rotor 70, each blade having a leading edge
32 and a
trailing edge 34. The leading edges 32 of the blades 30 are generally adjacent
to the inlet
22. A second stage of centrifugal-like compressor vanes 40 (also known as
impeller
blades, or the like) is provided extending from an exducer rotor 72, each vane
40 having
a leading edge 42 and a trailing edge 44. The trailing edges 44 of the
impeller-like vanes
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40 are generally adjacent to the outlet 24. In this example, there is no
diffusion stage
downstream of the centrifugal-like stage 40 within the hybrid compressor
itself but
rather, as shown in Fig. 1, a conventional diffuser 90, such as a pipe or vane
diffuser 90
of the type typically provided in gas turbine engines to diffuse centrifugal
compressors,
is provided downstream of the outlet 24. The diffuser 90 may be any suitable
diffusion
stage, however, and may be located in any suitable location, including within
the hybrid
compressor if desired.
[0017] The compressor rotor 20 in this example is, as mentioned, in the form
of an
inducer 70 and exducer 72, with bolts 74 (or other means of joining together
the inducer
and exducer parts) joining the inducer and exducer together. A shroud 50 is
disposed
around the rotor 20 and cooperates with the inducer/exducer hubs to define the
gas path
26. Extending inwardly from the shroud 50 in this example is a row of
stationary stator
vanes 60 between the first and second compression stages. A small space 62 is
present
between the compression stages 30/40 and the stationary vanes 60, as is
generally typical
between compression and diffusion stages.
[0018] The gas path 26 has a generally increasing inner hub radius, and a
diminishing
cross-sectional area (i.e. between hub and shroud), from the inlet 22 to the
outlet 24. In
Fig. 2, the rotor assembly has a somewhat regularly increasing inner. hub
radius. It
should be noted that there is no requirement to have a "circular arc" shaped
shroud (i.e.
when viewed in 2D cross-section, as in Fig. 2), but rather the profile can be
arbitrarily
shaped to optimise the aerodynamic loading in the stator vane and/or reduce
the Mach
number (Mn) at the impeller tip, etc., in light of the teachings herein. The
skilled reader
will appreciate that the same may apply to the hub profiles.
[0019] The hybrid compressor may also include an apparatus for bleeding
secondary air
from an intermediate position within the compressor - it will be understood
that
intermediate bleed is not generally practical with conventional centrifugal
impellers. In a
first example, in the multiple rotor arrangement of Fig. 2, a bleed outlet 80
may be
provided through the rotor, in this case through the exducer. In a second
example shown
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in Fig. 3 and discussed further below, a bleed outlet 82 may instead/also be
provided on
the shroud housing 50. Bleed can be controlled using a controlling orifice
(not shown)
or any other suitable arrangement.
[0020] Referring now to Fig. 4, as will be understood by the skilled reader,
the
compressor blades 30 and impeller vanes 40 increase the air flow velocity as
the blades
turn, and the stator vanes 60 are be provided with an opposite orientation, or
inclination,
relative to the compressor axis (i.e. when viewed in a 2-D projection, as in
Fig. 4), as
compared to the blades 30 or vanes 40, in order to convert tangential velocity
energy in
the air into static pressure head energy prior to entry into the stage 40.
[0021] Thus, a hybrid compressor 20 is provided having a first, axial-like
rotary
compression stage, followed by a stationary stator vane diffusion stage,
followed by a
second, aerodynamically highly loaded centrifugal-like rotary compression
stage, which
thus yields the present example of an axial-centrifugal hybrid compressor. In
use, air
entering the inlet is accelerated and compressed by the axial-like stage 30,
diffused and
slowed by the stator vane row 60, and then further compressed and accelerated
by the
centrifugal-like stage 40, before exiting the compressor to be diffused
downstream of the
compressor by the diffuser 90, for delivery to the combustor 16 and,
ultimately, to the
turbine 18.
[0022] As mentioned above, the centrifugal-like stage may have a substantially
radial
exit, i.e. the exit angle of substantially 90 to the compressor axis M, while
the axial-like
stage will have a substantially axial inlet, i.e. the inlet angle is
substantially parallel to
the compressor axis M. The relative inlet and outlet angles of the axial-like
stage and the
centrifugal-like stage will now be discussed in more detail.
[0023] The skilled reader will appreciate that axial (or axial-like)
compressor stages are
considered to have low aerodynamic loadings, while mixed-flow and centrifugal
stages
are considered to have aerodynamic loadings, with centrifugal stages typically
having
higher blade loadings than mixed-flow stages. While some grey areas may exists
when
defining boundaries between them, a number of characteristics may be defined
which
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distinguish a stage as either axial or centrifugal, or whether it has a low or
high
aerodynamic loading. For example, axial compression stages typically have a
pressure
ratio (PR) of less than about 2:1, whereas mixed-flow and centrifugal stages
typically
have pressure ratios greater than 2:1, and often well in excess of 2:1 (e.g.
typically 4:1 or
5:1 or more typically even higher, such as 7:1 or higher, for a typical aero
gas turbine
centrifugal compressor). Therefore, it is to be understood that the axial-like
stage 30 of
the hybrid compressor of Fig. 2 may have a pressure ratio of less than 2:1 per
stage,
while the centrifugal-like stage 40 will have a pressure ratio in excess of
2:1, and perhaps
well in excess of 2:1, such as 4:1 or 5:1, or higher, such as 7:1 or higher.
The final stage
of the hybrid compressor is thus highly loaded, as will be discussed further
below. As is
known in the art, the loading of a stage may be selected, within limits,
through selection
of blade shape, angle, number, radius change, gas path shape, etc. for a given
rotor speed.
[0024] Some prior art has expressed aerodynamic blade loadings in terms of
diffusion
factor or ratio, and used this characteristic to distinguish between axial and
centrifugal
compressors. Typically, = the concept of diffusion ratio applies only to axial
flow
compressors and is not easily applied to a centrifugal compressor. However, US
Patent
No. 4,428,715, the full contents of which are incorporated herein by
reference, teaches an
equation which it applies to centrifugal compressors to determine the
diffusion loading
on the centrifugal-like stage. US Patent No. 4,428,715 also teaches that, as
the diffusion
ratio increases, power losses rise at a relatively low rate until a diffusion
ratio of about
0.55 is reached, after which losses increase much more sharply. US Patent No.
4,428,715 proposes to minimize losses by keeping each stage below a 0.55
diffusion
factor limit. Similar teachings are found in other prior art, because the
conventional
wisdom is that in order to have low overall losses, losses within the
compressor must be
small everywhere. Hence, the diffusion factor, and thus pressure ratio, is
taught to be
limited in each stage. However, as discussed further below, the present
inventors have
found that although losses may increase sharply above some critical point
(such as 0.55
in the example taught in US Patent No. 4,428,715), the effect of such losses
on
compressor performance can be tolerated as long as the losses within the
compressor are
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commensurate with the pressure ratio. That is, a compressor having a stage
with a high
aerodynamic blade loadings may "tolerate" higher losses and yet still achieve
a good
efficiency. In general, the higher the aerodynamic loadings, the higher the
losses
tolerated..
[00251 Thus, in contrast to US Patent No. 4,428,715, in the present design the
centrifugal-like stage preferably has a pressure ratio greater than 2:1 and
thus would
necessarily have a diffusion ratio above 0.55.
[00261 Radius change is another characteristic which may be used to
distinguish
between axial and centrifugal compressors. US Patent No. 4,428,715 sets a
radius
change (as defined in that reference) of 15% as a maximum for axial stages,
and the
skilled reader will agree that this limit is reasonably accepted in the art.
Again, as with
diffusion factor, US Patent No. 4,428,715 and similar prior art teach
minimizing losses
in each stage, by minimizing the pressure ratio across the stage below 2:1,
and hence
minimizing radius change to less than 15%. Therefore, a 15% radius increase,
or rn+i/rn
ratio of 1.15, may provide a practical upper limit to an axial-like stage, and
thus also
served as a practical lower limit to radius change for a centrifugal-like
stage.
[00271 In the present hybrid compressor (e.g. as shown in Fig. 2), the
centrifugal-like
stage 40, by definition, has a radius change in excess of 15%. The radius
change of the
centrifugal-like stage 40 may typically be greater than the axial length of
the stage, or in
other words, have an angle of more than 45 degrees (i.e. the slope of the gas
path, or the
right triangle with the gas path axial length and radius change as its
perpendicular sides).
[00281 The hybrid compressor of Fig. 2 has a single axial-like stage 30,
followed by a
single centrifugal-like stage 40, with a single vane row 60 interposed in
between.
However, as shown in Figs. 3 and 5, in which like elements are indicated with
the same
reference numerals described above, in other embodiments one or more stages of
various
designs may be provided upstream of the centrifugal-like stage 40 .
Specifically, Fig. 3
depicts a hybrid compressor in which the first stage 30 is a "mixed flow"
stage, rather
than an axial-like stage as in Fig. 2. The skilled reader will appreciate that
a mixed flow
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stage may be defined as a stage having a pressure ratio of 2:1 and greater
(and thus may
have a diffusion ratio of greater than 0.55 and/or a radius increase in excess
of 15%, i.e.
in a similar fashion to a centrifugal stage). It will also be understood that,
as between a
mixed-flow stage and a centrifugal stage, the mixed-flow stage will tend to
have lower
pressure ratios and/or radius increases (e.g. a mixed stage may be in the
range of, say 2:1
to 4:1 pressure ratio, whereas a pressure ratio of 5:1 or greater (e.g. 7:1 or
higher, for
example) may generally be considered a centrifugal stage, and so on). The
skilled reader
will also appreciate that the line between what is an axial, mixed-flow and
centrifugal
stage is not always clear and may be open to interpretation in various
circumstances.
[0029] Fig. 5 depicts a hybrid compressor in which a plurality of axial-like
stages 30
are provided upstream of the centrifugal-like stage 40, each followed by a
respective
stator vane row 60.
[0030] Since the addition of stages within the hybrid compressor increases the
weight
and complexity of the compressor, and may introduce operability issues such as
surge
margin problems and so on, it may be desirable to limit the number of stages
within the
hybrid compressor to 3 or less (i.e. two upstream stages and a final
centrifugal-like stage
mounted to the same rotor, e.g. as depicted in Fig. 5).
[0031] Therefore, the hybrid compressor of Figs. 2,5 may have a low number,
e.g. one
or two, stages at the front of the compressor, followed by a single
centrifugal-like stage
40 as the final stage. The upstream stage(s) may have small pressure ration
(<2:1) (and
may also have a radius change (<+15%) and/or low diffusion loading (<0.55)).
The
centrifugal-like stage, on the other hand, is a highly loaded stage which has
a a pressure
ratio greater than 2:1 and may also have a radius increase of more than 15%
and/or a
diffusion ratio of more than 0.55. In practice, the centrifugal stage 40 will
have a PR
much greater than 2:1, such as a typical PR of 4:1, and more typically 7:1 or
higher, as
demonstrated in the Example below. The centrifugal-like stage may have
supersonic
inlet flow and/or a supersonic exit flow (Mn ~4.0), without detrimental
effects on
efficiency, especially if the final stage is relatively highly loaded, as
mentioned. The
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skilled reader will appreciate that, in contrast, prior art such as US Patent
Nos. 2,350,839
or 4,428,715 or French publication FR972751A which teach limiting internal
airflow to
subsonic speeds in and between compression stages by defining a larger number
of
compressions stages with lower diffusion loadings and/or smaller radius
changes per
stage, all with a goal of improving efficiency by keeping losses small
everywhere. The
inventors have found, however, that efficiency does not require losses to be
small
everywhere, and that the compressor stage(s) may have a supersonic flow,
locally in the
region of the centrifugal blades, without detrimental effects on efficiency,
particularly if
the stage has a relatively high pressure ratio. It has also been found that
that providing
keeping the number of stages low (i.e. three or less, and most preferably only
two) and
by providing diffusion between the stages, within a gas path of what would
otherwise be
an envelope similar to a purely centrifugal stage, may reduce the exit Mach
number and
increase the efficiency for a given total pressure ratio across the
compressor, relative to a
comparably sized purely centrifugal compressor. This may also provide an
opportunity
to achieve a given pressure ratio compressor at a reduced exit diameter, and
therefore tip
speed, as compared to a conventional centrifugal compressor.
100321 Example: The hybrid compressor may be further understood with respect
to the
following example. A meanline analysis may be done for the compressors of Fig.
2 and
Fig. 3, and compared with a reference impeller for a given mass flow, speed,
etc. A
leading edge position for the axial-like stages of Figs.2 and 3 is assumed
identical to the
reference compressor. If desired, for example to limit the Mach number within
the
compressor, some swirl may be retained at the exit of stator vanes 60. For the
conditions, speed (Nc) = 55,000 rpm, mass flow (We) = 5.15 lb/sec, and the
term
(N*rt(W)) = 124,800, the results shown in Table 1 may be obtained. Thus, it
may be seen
that a hybrid compressor of the type depicted in Figs. 2 or 3 may offer
improvement
relative to the reference impeller (i.e. a conventional centrifugal
compressor).
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Table 1
Fig. 2 Configuration Fig. 3 Configuration Reference Impeller
PR Efficiency Mn PR Efficiency Mn PR Efficiency Mn
Axial 1.71 0.862 -- 1.66 0.870 -- -- -- --
Stage
Cent. 5.27 0.819 -- 6.04 0.829 -- 8.00 0.796 --
Stage
Total 8.99 0.812 1.04 9.99 0.820 -- 8.00 0.796 1.14
[00331 As has been discussed previously, US Patent No. 4,428,715 teaches
providing a
plurality of internal compression-diffusion stages, each with a low
aerodynamic blade
loading (i.e. diffusion ratio <0.55, and a radius change <15%). That is, each
stage is
sought to be kept as axial-like as possible, in order to maintain the
beneficial efficiency
aspects of the axial compressor. However, such an approach presents a number
of
practical problems, such as the fact that the relatively large number of
stages necessarily
leads to stages of small dimension, which may increase manufacturing problems,
for
example, by making it difficult to provide the required blade edge radii and
tip
clearances with sufficient precision to achieve the desired efficiency gains.
As well, the
small blade rows have low Reynolds numbers, which may presents problems
relating to
controlling surge margin and the operability of the compressor, particularly
in an aircraft
prime mover engine application, and even more particularly at altitude. The
inventors
have found that is generally more desirable to much fewer stages and higher
pressure
ratio per stage, as will be discussed further below.
[00341 As mentioned, prior art such as US Patent Nos. 2,350,839 or 4,428,715
attempts
to provide efficiency through minimizing losses within each stage of the
compressor,
however as also mentioned the present inventors have found that a high
efficiency
compressor may have high internal losses (e.g. may tolerate supersonic
internal air
speeds), as long as the losses are commensurate with the pressure ratio of the
stage under
consideration. That is, a compressor with a stage having a high aerodynamic
blade
loading(i.e. a high pressure ratio) can tolerate higher losses and yet still
achieve a good
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efficiency. Hence, presented herein is a compressor with significantly fewer
stages than
generally contemplated in prior art such as US Patent No. 4,428,715, and
having a higher
pressure ratio final stage than contemplated in other prior art, through
application of the
design criteria described herein.
[0035] In another embodiment, not depicted, the final stage 40 of the hybrid
compressor may be a mixed-flow-like compressor stage, rather than a
centrifugal-like
stage. The final mixed-flow-like stage may have an exit angle of 30 , or
greater, to the
compressor axis M. Like the centrifugal-like final stage 40, the mixed-flow-
like stage
will have pressure ratio above 2:1, and perhaps 4:1 or greater. The mixed-flow-
like
stage radius change may be greater than 15%, and likewise the diffusion ratio
for the
mixed-flow-like stage may be greater than 0.55, as with a centrifugal stage.
As
mentioned above, the skilled reader will appreciate that the line delimiting a
mixed-flow
stage from centrifugal stage may sometimes be blurry - it will generally be
understood
that while both types have similar lower limits for pressure ratio, diffusion
ratio and
radius change, the lower ranges for these parameters will tend to be defined
as a mixed-
flow stage by those skilled in the art, while the higher ranges will generally
be defined as
centrifugal stages.
[0036] Referring to Figs. 6-8, the engine 10 may employ a hybrid compressor
stage 20
in addition to upstream (or downstream) compressor stages, comprising
compressor
rotors 14 and vane stages 15. For example, Fig. 6 depicts an engine 10 with a
"l A+l H"
compressor arrangement, having a conventional axial stage upstream of the
hybrid
compressor. The hybrid compressor 20, however, is provided on a single-piece
rotor 71,
rather than the inducer-exducer design described above. Similar to Fig. 6,
Fig. 7 depicts
an engine 10 with an "nA + I H" compressor arrangement. Fig. 8 depicts an
engine 10
with a "1M + 1H" arrangement, wherein a mixed-flow stage is provided upstream
of a
hybrid compressor.
[0037] In summary, the hybrid compressor combines the benefits of interrupted
compression (i.e. intermediate diffusion) with the increasing radius of a
conventional
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centrifugal stage to provide a lower exit velocity than is otherwise
obtainable with a
centrifugal stage alone. This, in turn, may result in greater compression than
would be
obtained from a centrifugal stage alone, or greater efficiency, or both.
[00381 This compressor design may yield a compact enough design that it is
possible to
replace an existing centrifugal stage with a present design occupying the same
physical
space or envelope, but which provides greater compression and efficiency. When
comparing the present compressor with a conventional centrifugal impeller, if
the present
compressor is provided with the same impeller exit radius, the present
compressor may
provide similar efficiency but an improved pressure ratio. Alternately, by
utilizing an
impeller exit radius less than that of the conventional centrifugal stage, the
present
compressor may provide a similar pressure ratio but a greater efficiency.
Intermediate
results are also available. Thus, the result may be a compressor with less
mass than a
comparable centrifugal compressor, which may result in a lower cost and/or
lighter
weight compressor which may be lighter and easier to install engine, due to a
smaller
engine diameter
[00391 The above description is meant to be exemplary only, and one skilled in
the art
will recognize that changes may be made to the embodiments described without
departing from the scope of the inventions disclosed. For example, although a
gas
turbine turboshaft engine is illustrated and described, the hybrid compressor
may be used
with any suitable applications, such as other gas turbine engines, such as
turboprop,
turbofan, APU, etc., or in other compressor applications. The hybrid
compressor 20 may
be used to compress gases in other situations, such as compressors for
supplying or
receiving compressed air to/from pneumatic systems, or even be used as a
suction
device. Application may also be suitable in any other air-breathing device
which
includes a rotor, such a APU load compressors, turbochargers, radial turbines
etc. The
vanes 60 may be mounted to any suitable support, whether on the radial outer
or inner
side of the gas path, and the vanes may be fixed or variably positionable.
Vanes 60 may
be mounted to the shroud housing 50, or any other suitable support, in any
suitable
manner. The rotor disks 70, 72 may be integrated in a monolithic element or be
an
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assembly of multiple elements, and can be made of any suitable material(s).
Still other
modifications will be apparent to those skilled in the art, in light of a
review of this
disclosure, and such modifications are intended to fall within the appended
claims.
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