Note: Descriptions are shown in the official language in which they were submitted.
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Description
Multi-stage Centrifugal Compressor
Technical Field
This invention relates to compressors for
air or other gases and more particularly to centrifugal
or radial flow compressors in which rotating blades
are situated in a flow passage that increases in
diameter towards an outlet end.
Background Art
Compressors which employ rotating vanes or
blades may be divided into two broad categories on the
basis of the configuration of the air flow passage
Centrifugal or radial flow compressors, which constitute
the first category, have a flow passage which increases
in diameter in the direction of the air flow. Axial
flow compressors form the second category and have a
flow passage of constant or almost constant diameter.
Centrifugal compressors are basically simpler,
more compact and less costly than the axial flow form.
These characteristics are highly desirable in many
compressor usages such as in gas turbine engine and
; engine turbo chargers as well as others but heretofore
it has been necessary to tolerate a relatively low
isentropic efficiency in order to gain the benefits
of these advantages. The relatively low efficiency
typical of prior centrifugal compressors results from
the fact that such mechanisms are single staged devices
except where a lengthy and complex construction, in-
volving what amounts to a series of such compressorscoupled in tandem, is resorted to.
A single compressor stage consists of a set
of revolving compressor blades fo1lowed by at least
one set of diffuser blades which may be stationary or
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counter rotating. A set of diffuser blades in addition
to a set of compressor blades is required to form a
compressor stage since a sizable proportion of the
energy imparted to incoming air by the revolving set
of compressor blades is initially tangential energy
of motion of the air flow. To complete the compression
process the air flow must then pass through at least
one set of diffuser blades oriented at a different
angle than the compressor blades to convert the tan-
gential velocity energy into static pressure head.
The degree of compression accomplished ina rotary compressor is expressed by the pressure ratio
which is the ratio of pressure at the outlet to that
at the inlet. A high pressure ratio across a single
compressor stage requires a high loading on the
compressor blades, the blade loading being ~uantitative-
ly expressed by the diffusion factor. Where sizable
pressure ratios are to be achieved in a single stage
device, the diffusion factor must necessarily be high.
Isentropic efficiency, which is an inverse function
of the diffusion factor, is therefore necessarily low.
The high efficiency of many axial flow
compressors largely results from the common practice
employing a series of stages in the axial flow type
of compressor. An axial flow compressor may typically
have several sets of compressor blades alternated with
sets of diffuser blades along the axial flow path,
each set of compressor blades in conjunction with
a following set of diffuser blades constituting an
individual stage. In such a device, the overall
pressure ratio of the compressor is the product of
the lower pressure ratios of the individual stages of
the series. As each stage individually has a low
pressure ratio, each stage operates at high efficiency
and the efficiency of the compressor as a whole is also
high.
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Prior applications of the multiple staging
principle to centrifugal compressors have gained
efficiency at the cost of sacrificing much of the
inherent advantages of the centrifugal compressor
configuration. More particularly, most prior multi-
staged centrifugal compressors have as a practical
matter included a series of essentially separate
single stage centrifugal compressors connected to-
gether in tandem through bulky and complex air ducting
for channeling the outlet flow fxom one stage radiallyinward to the smaller diameter inlet of the next
subsequent stage. This results in a lengthy, comple~
and costly construction such as is found in axial
flow compressors.
Prior centrifugal compressors also exhibit
other problems aside from a relatively low efficiency
where a high pressure ratio is to be realized. De-
signing for a hi~h pressure ratio in a single staged
compressor results in extremely high tangential air
velocity behind the single long set of compressor
blades. This in turn dictates that a bulky and heavy
diffuser structure, including lengthy diffuser blades
and a voluminous diffusion chamber, be provided at
the outlet.
A further practical problem encountered in
prior centrifugal compressors arises from the fact
that different compressor usages require different
pressure ratios and flow capacities. If each com-
pressor model in a family of compressors of different
pressure ratio and capacities must be manufactured
with a large number of distinct parts usable only
in the one particular model, the cost of manufacture
of the line of compressors as a whole is increased.
Disclosure of Invention
The present invention is directed to over-
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coming one or more of the problems as set forth above.
In one aspect of the present invention there is
provided a centrifugal compressor having relatively
rotatable inner and outer elements defining an annular
flow path therebetween having a progressively diminishing
cross-sectional area from a flow inlet end to a flow
outlet end, the flow outlet end having a diameter greater
than the diameter of the flow inlet end in the flow path,
a portion of outward curvature in the flow path wherein
the flow path becomes progressively more radially directed
as the flow path approaches the outlet end, and blades
forming a plurality of compression-diffusion stages within
the portion of outward curvature of the flow path.
By providing a series of sets of compressor
blades alternated with sets of diffuser blades within a
radial flow path of progressively increasing diameter,
the efficiency of a centrifugal compressor is greatly
increased while preserving the compactness of prior
centrifugal compressors and while preserving much of
the structural simplicity as well. For a given
pressure ratio and flow capacity, gains in efficiency
may in fact exceed those realized in multiple stage
axial flow compressors having a similar number of
stages as centrifugal effects aid the blade action in
generating a pressure rise. Consequently, fewer
internal stages may be needed to achieve a given
pressure ratio with a given enerby input.
As diffusion is largely accomplished
internally in stages, a large diffuser is not
necessarily needed at the outlet of the compressor.
Thus the invention may be relatively compact in the
diametrical direction relative to prior centrifugal
compressors. In addition, the multi-staged
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construction enables the manufacture of a family of
centrifugal compressors od different pressure ratio
and/or flow capacity without requiring a large number
of different structual elements for each model.
In instances where the multi-stage centrifugal
compressor is a component of a gas turbine engine, an
engine turbocharger, or the like, the high efficiency
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of the multi-stage compressor significantly reduces
power losses in such apparatus as a whole.
Brief Description of Drawings
Figure 1 is a broken out side elevation
view of a first embodiment of a centrifugal compressor
constituting the air intake component oE a gas turbine
engine.
Figure 2 is an enlarged axial section view
of a portion of the air compressor of Figure 1.
Figure 3 is a broken out perspective view
of the impeller and stator portions of the air com-
pressor of Figures 1 and 2 further illustrating blading
structure within the compressor.
Figure 4 is a view taken along curved line
IV - IV of Figure 2 illustrating the configurations
and relative inclinations of the blades of successive
stages within the air compressor of the preceding
figures.
Figure 5 is a graph depicting input power
losses as a function of blade loading or diffusion
factor in the present invention and in typical prior
air compressors.
Figure 6 is an axial section view of a portion
of an air compressor basically similar to that of
Figure 2 but with modifications o~ the blading structure
to vary the air flow capacity.
Figures 7A to 7G are diagrammatic views
illustrating further modifications of the compressor
of Figure 1 which enable realization of any of a
series of different pressure ratios and/or flow capa-
cities utilizing much of the same basic structural
components.
Figure 8 is an axial section view of an
engine turbocharger having a compressor section in
accordance with an embodiment of the invention.
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Best Modes of Carrying Out the Invention
Referring initially to Figure 1 of the
drawings a radial flow or centrifugal compressor 11
has an impeller 12 disposed for rotation within an
annular stator member 13 respectively constituting
relatively rotatable inner and outer elements that
jointly define an annular flow path 14.
Impeller 12 is of progressively increasing
diameter from the air inlet end 16 of the flow path
10 14 towards the air outlet end 17. Stator member 13
has an inner diameter which also progressively increases
along the flow path 14 but at a lesser rate so that
the spacing of the impeller from the stator diminishes
towards the outlet end 17 of the flow path. The
diminishing spacing of the impeller 12 from stator
member 13 along flow path 14 compensates for the
progressively increa~sing diameter of the flow path
towards the outlet end which would otherwise cause
the flow path to have a progressively increasing
cross-sectional area. The decrease in spacing also
compensates for the air compression that occurs along
the flow path 14 and which progressively reduces
the volume occupied by unit mass of air as it travels
along the path.
In the embodiment of the invention depicted
in Figure 1, compressor 11 constitutes the air intake
component of a gas turbine engine 18 and certain
structural features of this particular compressor 11
are specialized for this context. For example, the
impeller 12 is supported on and driven by a forward
extension of the main shaft 15' of the gas turbine
engine 18 and the inner stator member 13 is secured
to an outer stator member 19 which is itself secured
to the main housing 21 of the gas turbine engine and
supported thereby.
Aside from the air intake section defined
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by the compressor 11, the gas turbine engine 18 may
be of a known design such as that described in prior
United States Patent 4,030,288 and therefore will
not be further described except for certain components
which directly coact with elements of the compressor.
It should be understood that usage of a compressor
embodying the invention is not limited to the context
of gas turbine engines. When the invention is employed
in other contexts or for other purposes, the impeller
12 may be journaled within the stator members 13 and
19 by suitable bearing structures known to the art
and may be driven by any of a variety of known external
motors. Similarly, the stator may be provided with
appropriate support means of any of various known forms.
The inner and outer stator members 13 and
19 jointly form an annular diffusion chamber 22 which
receives air from the outlet end 17 of the compressor
flow path 14. In order to minimize the size of the
compressor in the radial direction, outer stator
member 19 is shaped to situate most of the volume
of diffusion chamber 22 adjacent the smaller diameter
forward portion of inner stator member 13. This is
a practical configuration in that the lengthy, radially
extending diffuser vanes required at the outlet end
of the flow path in many conventional single stage
centrifugal compressors are not necessarily required
in the present invention.
Diffusion chamber 22 is communicated with a
compressed air outlet tubulation 23 which in the
present example supplies the compressed air to the
combustor 24 of the gas turbine engine 18 through
a heat exchanger module 26 which transfers heat from
the exhaust of the engine to the incoming compressed
air. In instances where the compressor 11 is used
for purposes other than in a gas turbine engine, the
outlet tubulation 23 may be replaced with a hose or
other conduit means suitable for connection with the
compressed air utilizing device.
Compression of air within the flow path
14 is accomplished by blading means 27 depicted on
a larger scale in Figure 2, which form a plurality
of internal compression-diffusion stages 30a to 30f
of low blade loading or diffusion factor within the
flow path 14.
Referring now to Figures 2 and 3 in con-
junction, a plurality of spaced apart sets of compressorblades 28 extend radially from impeller 12 into the
flow path 14, there being six such sets 29a, 29b,
29c, 29d, 29e, 29f of compressor blades, proceeding
from the air inlet end 16 to the air outlet end 17,
in this example. The individual compressor blades
28 of each set 29a to 29f are equiangularly spaced
around the rotational axis of the impeller 12 and owing
to the progressively diminishing thickness of the
flow path 14, the blades of each successive set extend
progressively smaller distances from the impeller~
A plurality of spaced apart stationary sets
of diffuser blades 31 extend into flow path 14 from
the inner stator member 13, there being seven sets
32a, 32b, 32c, 32d, 32e/ 32f, 32g of diffuser blades
31 in this example. The sets 32a to 32g of diffuser
blades are alternated with the sets 29a to 29f of
compressor blades 28 except that the two final sets
of diffuser blades 32f and 32g are both behind the
final set 29f of compressor blades. Individual blades
31 of each set 32a to 32g of diffuser blades are also
equiangularly spaced apart with respect to the rota-
tional axis of the compressor and the blades of each
successive set 32a to 32g e~tend progressively shorter
distances from the stator member to accommodate to
the progressively diminishing thickness of the flow
path 14.
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Each set 29a to 29f of compressor blades
in conjunction with the following set of diffuser
blades 31 constitutes one of the plurality of
compression-diffusion stages 30a to 30f situated
in the flow path 14. Thus in the present example
compressor blade set 29a and diffuser blade set 31a
form a first compression-diffusion stage 30a and
compressor blade set 29b in conjunction with diffuser
blade set 32b form a second compression-diffusion
stage 30b, there being six such stages in this example.
Referring now to Figure 4, the individual
compressor blades 28 of each set 29a to 29g are in-
clined relative to the rotational axis 18' of the
impeller to impart an increment of flow velocity to
intercepted air as the blades turn in the direction
indicated by arrows 33 in the drawing. The compressor
blades 28 of each successive set 29a to 29g have a
progressively increasing angulation relative to axis
18' to accommodate to the progressive increase of
free stream velocity which occurs along the flow
path. The blades 31 of the successive sets 32a to
32g of diffuser blades have an opposite angulation
relative to axis 18', which also becomes progressively
greater for each successive set of diffuser blades,
in order to convert tangential velocity energy imparted
to air by the preceding set of compressor blades into
static pressure head energy.
Thus, with reference to Figures 1 and 2,
the compression achieved by compressor 11 as a whole
is accomplished in six distinct compression-diffusion
stages 30a to 30f along the flow path 14. The pressure
ratio of each individual stage 30a to 30f may there-
fore be low relative to a conventional centrifugal
compressor having a single long set of compressor
blades followed by a single long set of diffuser
blades, designed to accomplish the same degree of
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compression. Since each component stage 30a to 30f
of compressor 11 operates at a low pressure ratio
and therefore a high level of efficiency, the aggregate
efficiency of the several stages in combination is
itself high in comparison with conventional single
staged devices.
In order to fully realize the gains in
efficiency inherent in the multiple stage construction,
each compression-diffusion stage 30a to 30f is designed
to have a free stream flow velocity which is below
supersonic throughout the region of blading means
27 and to have a diffusion factor below about 0.55
at each stage. As is known in the art, the diffusion
factor of a single compression-diffusion stage may
be selected, within limits, by an appropriate fixing
of the shape, angulation and number of compressor
blades and diffuser blades in relation to the configura-
tion of the flow path and the rotational velocity
of the compressor blades, More particularly, in a
compressor stage wherein the free stream air velocity
is subsonic throughout the region of the blading as
is the case in the compressors of the present invention,
diffusion factor (D.F.2 is given by the expression:
(D.F.) = 1 - V2 + ~ Ve
V- 2~ Vl
where: Vl = inlet flow velocity relative to blade row
V2 = outlet flow velocity relative to blade row
Ve = tangential flow velocity relative to
blade row
~ = blade row solidity (proportion of open
flow space to total cross-sectional area
of flow path in blade region)
The benefit of establishing a diffusion
factor below about 0.55 at each of the several
compression-diffusion stages 30a to 30f may be seen
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by referring to Figure 5 which is a graph depicting
measured input energy losses, that is energy which
does not become available as pressure energy at the
outlet of the compressor, as a function of diffusion
factor for three different types of rotary compressor
all of which achieve the same overall pressure ratio
or degree of compression. Rectangles 34 designate
measured losses for a conventional single stage
centrifugal compressor which necessarily must have
a relatively high diffusion factor to accomplish
the desired degree of compression in the single stage.
Circles 36 indicate the relatively low measured losses
in a conventional multiple stage axial flow compressor
in which the diffusion factor for each individual
stage may be much lower and therefore more efficient.
Triangles 37 indicate the measured losses in a multiple
stage centrifugal compressor embodying the present
invention. It should be observed that the compression
is accomplished in the present invention with a
diffusion factor 37 per stage which is substantially
lower than that 36 of the lengthier and more complex
axial flow compressor. The reason for this greater
efficiency of the present invention as indicated by
triangle 37 is believed to be that centrifugal force
supplements the direct effect of the blading in
achievin~ compression. This effect does not occur
in the non-radial flow path of an axial flow compressor.
The significance of a diffusion factor
value of about 0.55 as an upper limit for the individual
stages of the present invention is also evident in
Figure 5. It may be seen that there is not a linear
relationship between power loss and diffusion factor.
Instead, as the diffusion factor is increased from
a very low value, losses rise at a relatively moderate
rate, indicated by lines 38, until a value of about
0.55 is reached. Thereafter losses increase much
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more sharply with increasing diffusion factor as
indicated by lines 39. Efficiency is an inverse
function of power losses and thus it may be seen
that efficiency drops off relatively sharply after
the diffusion factor value of about 0.55 is passed.
Returning to Figure 1, the high efficiency
of the compressor 11 in turn increases efficiency of
the gas turbine engine 18 itself as power losses
in the compressor section of the engine are reduced.
As compared with a gas turbine engine utilizing an
axial flow compressor configuration for the purpose
of realizing somewhat comparable efficiencies, the
engine 18 of this example is much more compact and the
compressor section is simpler and less costly.
While the compressor 11 described above
is provided with six internal compression-diffusion
stages 30a to 30f, varying numbers of stages may be
provided by changing the number of sets of compressor
blades 28 and diffuser blades 31. Moreover, the
construction readily lends itself to manufacture of
a family of compressors of different pressure ratio
and/or flow capacities by varying only the number and
disposition of the sets of blades 28 and 31 within
the flow path 14 while otherwise utilizing identical
components for the several compressor models. Refer-
ring to Figure 6 for example, a compressor 11' having
a lower pressure ratio but a smaller air mass flow
rate and therefore a smaller driving power requirement
may be produced simply by removing the first set 29a
of compressor blades and the first set 32a of diffuser
blades, shown in phantom in Figure 6, while otherwise
utilizing components, such as impeller 12 inner stator
member 13 and outer stator member 19 identical to
those of the previously described embodiment. In
general, the elimination of compression-diffusion
stage blading means 27 from the air inlet 16 region
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of the flow path 14 has an effect of reducing both
air mass flow and pressure ratio while the elimination
of stages of blades from the region nearest the air
outlet end 17 has the predominate effect of reducing
pressure ratio. Adding of stages at the inlet end
increases mass flow and pressure ratio while additional
stages near the outlet end predominately raise pressure
ratio.
Thus while a limited number of specific
blading modifications will be described with reference
to Figures 7A to 7G and specific parameters will
be given, such examples are not exhaustive of the
possible modification. In accordance with the above
discussed relationships, other modifications may be
made to provide other mass flows and pressure ratios.
Figures 7A to 7C diagramatically illustrate
how a series of compressors lla, llb, llc respectively
of different pressure ratio and/or air flow capacity
may be configured by simply varying the numbers of
sets of blades in the air flow path while otherwise
utilizing identical components. Where the compressors
are embodied in gas turbine engines as previously
described, this enables production of a family of
engines 18a, 18b, 18c of different output power rating
and fuel consumption requirements simply by varying
the blading in the compressor section.
While the gas turbine engines 18a, 18b and
18c may be of known construction apart from the
compressors lla, llb, llc, the coaction of the com-
pressor sections with the other portions of the enginesmay best be understood by briefly reviewing certain
basic structure of such engines. Referring specifi-
cally to Figure 7A, for example, such engines 18a
have a fuel burning combustor 24a receiving compressed
air from compressor lla through heat exchanger 26a.
Output gasses from the combustor 24a drive a gasifier
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turbine 42a that turns the impeller 12a of the com-
pressor lla. Nozzle vanes 43a direct the gas flow
from combustor 24a and gasifier turbine 42a to a
power turbine 44a which turns the engine output shaft
46a, the exhaust gas from the power turbine being
discharged through the heat exchanger 26a to preheat
the compressed air which is delivered to the combustor.
The modified compressor lla of Figure 7A
is similar to that previously described with reference
to Figure 2 except that the first two compression-
diffusion stages 30a, 30b have been eliminated by
removing the first two sets 29a and 29b of compressor
blades 28 and the first two sets 32a and 32b of
diffuser blades 31. As a result of this simple modi-
fication, the compressor lla of Figure 7A has a lowerair flow and a lower pressure ratio of about 4.5.
The output power rating of the gas turbine engine
18a is then typically about 1200 horsepower realized
with a fuel efficiency of less than about 0.4 brake
specific fuel consumption (BSFC).
Figure 7B illustrates a gas turbine engine
18b of substantially greater output power rating
but which may be structurally identical to that of
Figure 7A except for another modification of the
blading structure within the compressor llb. Compressor
llb is similar to the compressor 11 of Figure 2 ex-
cept that the first and final sets 29a and 29g of
compressor blades of Figure 2 and the first and final
two sets 32a, 32f~ 32g of diffuser blades 31 have
been eliminated. The pressure ratio achieved by the
compressor llb of Figure 7B remains approximately the
same as that of Figure 7A but the volume of air
passing through the compressor llb of Figure 7B and
on to the combustor 24b is increased to the extent
that the power output of the turbine engine 18b is
now about 2,000 horsepower.
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Figure 7C depicts another modification,
confined to the blading means 27c of the compressor,
by which a similar basic gas turbine engine 18c in-
cluding similar impeller 12c and stator member 13c
elements in the compressor may be used to produce
an engine of still higher rated output power. The
compressor llc of gas turbine engine 18c is identical
to that of the first described embodiment of Figure
2 except that the final set 29f of compressor blades
of Figure 2 have been removed and the final two sets
32f and 32g of diffuser blades are now situated more
forwardly in the flow passage and configured for
that changed location. This makes the pressure ratio
of the compressor llc of Figure 7C about 6.5 and
provides an increase of volumetric air flow relative
to the Figure 7B embodiment. The rated power output
of the gas turbine engine 18c of Figure 7C is typically
about 3500 horsepower.
If the modifications of the compressor
blading arrangements are accompanied by modifications
of other components as well, the family of gas turbine
engines may be extended to still higher output power
ratings, examples of which are depicted in Figures
7D, 7E, and 7F. Referring initially to Figure 7D,
by forming the impeller 12d and inner stator member
13d to be relatively elongated at the front end 16d,
additional compression-diffusion stages, such as stage
30g, may be provided at the air inlet end of the
compressor lld to further increase rated power output
of the engine 18d. Thus the compressor lld of Figure
7D has an additional set of compressor blades 29g
followed by an additional set 32h of diffuser blades
at the front end of the air flow path 14d. The final
two sets of compressor blades 29e and 29f of the
embodiment of Figure 2 and the intermediate set of
diffuser blades 32e have been removed. The final
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two sets of diffuser blades 32f and 32g are again
situated more forwardly in the flow passage and have
configurations appropriate to that portion of the
passage. With these modifications, the pressure
ratio of the modified compressor lld of Figure 7D
remains at about 6.5 but air mass flow is sizably
increased causing the rated power output of the gas
turbine engine 18d to be increased to about 5,000
horsepower.
By making somewhat more extensive modifi-
cations, still greater power output ratings may be
obtained. For example as depicted in Figure 7E an
auxiliary compressor section 47e may be added between
the primary compressor lle and the heat exchanger
26e. The auxiliary compressor section 47e may for
example have two spaced apart sets 48e and 49e of
compressor blades on an auxiliary impeller 50e each
being followed by a set, 51e and 52e respectively
of diffuser blades. An annular air duct 53e is
provided to receive the output flow from the primary
compressor section lle and to return the flow radially
inward for delivery to the air inlet end of the
auxiliary compressor section 47e. Primary compressor
section lle is itself identical to the compressor
lld of the previous Figure 7D. To best realize the
advantages of the compressor modification of Figure
7E, other elements of the gas turbine engine 18e are
modified to the extent of providing an additional
gasifier turbine stage 54e to drive the impeller 50e
of the auxiliary compressor stage 47e. The modifi-
cations depicted in Figure 7E produce an overall
compressor pressure ratio of about 12 and raise the
rated power output of the gas turbine engine 18e to
about 5500 horsepower.
Figure 7F illustrates still a further
modification of the gas turbine engine 18f in which
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the structure remains similar to that described above
with reference to Figure 7E except that in the embodi-
ment of Figure 7F the annular air duct 53f which
communicates the primary compressor section llf with
the auxiliary compressor section 47f includes an
intercooler or heat exchanger 55f which acts to cool
the compressed air in passage between the two com-
pressor sections. Intercooling reduces the amount
of power required to drive a compressor and this
power reduction is reflected in an increased power
output at the output shaft 46f of the yas turbine
engine 18f. By this further modification, the gas
turbine engine 18f is made to deliver about 6500
horsepower.
Referring now to Figure 7G the power output
and therefore the fuel consumption rate of any of
the gas turbine engines described above may be adjusted
downwardly as desired by disposing a set of air flow
reducing stator vanes 56g in the inlet end of the
air flow path 14g in front of the initial set 29g
of compressor blades 28g. Stator blades 56g are
angled relative to the air flow path 14g in order to
constrict the air flow path and thereby reduce air
mass flow to any desired extent.
As previously pointed out, the invention
is not limited to compressors which function as an
air intake component of gas turbine engines, but
may also advantageously be utilized in free standing
compressors for supplying compressed air to various
pneumatic systems or to other mechanisms which include
a compressor as one component. Figure 8 illustrates
an example of the latter category in which a compressor
llh embodying the invention constitutes an air intake
component of a turbocharger 57 for an internal com-
bustion engine 58.
A turbocharger 57 increases the fuel
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efficiency of the engine 58 by boosting intake mani-
fold pressure and uses energy recovered from the
exhaust gas of the engine for this purpose. More
specifically, the turbocharger includes a turbine 59
driven by the engine exhaust flow and which drives the
compressor llh that supplies compressed air to the
engine 58 intake manifold. Centrifugal compressors,
preferably in combination with a centripetal turbine,
are advantageous in turbochargers in view of the
basic compactness and structural simplicity of such
compressors but if a conventional single stage
centrifugal compressor is used, the adiabatic effi-
ciency of the turbocharger is undesirably limited.
This adversely affects the power output of the
associated engine 58 per unit of fuel consumed. Very
high efficiency togetner with simplicity and compact-
ness in both the axial and radial direction can be
realized by embodying a multi-stage radial flow
compressor llh in accordance with present invention
in a turbocharger 57.
The compressor llh and turbine 59 are secured
to opposite ends of a housing 61 which journals a
drive shaft 62 that defines the rotational axis of
both the compressor and turbine.
Compressor llh has an annular outer stator
member l9h secured to the front end of housing 61
in coaxial relationship with the drive shaft 62 and
which defines a broad air intake passage 64. Stator
member l9h also forms a volute or annular collection
chamber 66 which is communicated with intake mani-
fold 67 of engine 58, the collection chamber being
coaxial with intake passage 64 and being of greater
diameter. A rotatable impeller 12h is supported on
the forward end of drive shaft 62 within stator member
l9h and in conjunction with an inner stator member
13h forms an annular air flow path 14h leadlng from
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air intake passage 64 to collection chamber 66.
Impeller 12h and inner stator member 13h have configur-
ations which cause the air flow path 14h to be of
progressively increasing diameter in the direction
of air flow while being of progressively diminishing
thickness towards the air outlet end.
Multi-staged blading means 27h of the type
previously described is situated within the flow
path 14h to provide a plurality of sub-sonic internal
lQ compression-diffusion stages 30j, 30k, 30L each having
a diffusion factor below about 0.55. In this example,
the blading means 27h includes three spaced apart
sets of compressor blades 28h secured to impeller 12h
and alternated with three spaced apart sets of diffuser
blades 31h secured to stator member 13h. Thus three
compression-diffusion stages 30k, 30j, 30L are provided
in this embodiment each being defined by a set of
compressor blades 28h and the immediately following
set of diffuser blades 31h.
While the multi-staged compressor llh is
advantageous in a turbocharger employing any of a
variety of different types of turbine 59, very high
efficiency is best realized by using a centripetal
turbine 59 which is also of a multi-staged construction.
The turbine 59 of this example has an annular
stator 76 secured to the back end of housing 61 and
forming an exhuast gas outlet passage 77. A turbine
rotor 78 is secured to the back end of drive shaft
62 in coaxial relationship with the shaft and in
conjunction with an annular inner stator member 79
forms a gas flow path 81 which is of progressively
less diameter but progressively increasing thickness
from a gas inlet end 82 to a gas discharge end 83.
Stator 76 also forms an annular volute or
gas receiving chamber 84 which is communicated with
the inlet end 82 of gas flow path 81 and which is
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also communicated with the exhaust gas manifold 86
of engine 58. To cause the exhaust gas flow to drive
the turbocharger 57, three spaced apart sets 87a,
87b, 87c of rotor vanes are secured to rotor 78 and
extend into the flow path 81, the rotor vanes being
angled with respect to the direction of gas flow.
To maximize the reaction forces of the gas flow on
the rotor vanes 87a, 87b and 87c, one of three sets
88a, 88b and 88c of stator vanes precedes each set
of rotor vanes 87a, 87b and 87c respectively along
the gas flow path. As the pressure drop at each
individual set of rotor vanes 87a, 87b and 87c is
substantially lower than the total pressure drop
through the turbine 59 as a whole, each set of vanes
operates at a relatively high efficiency in comparison
with a single stage centripetal turbine having a
single long set of rotor vanes.
The above described turbocharger 57 con-
struction enables the impeller 12h and rotor 78 to be
situated on the same shaft 62 to turn at the same
speed and in most cases the two elements need not have
any large disimilarity in diameters. With the
rotational speeds and diameters of both the impeller
12h and rotor 78 closely matched, centrifugal stresses
are also closely balanced at a high but tolerable
level to optimize air and gas throughput in relation
to the size and weight of the turbocharger.
Industrial Applicability
In the operation of the embodiment of the
invention depicted in Figures 1 to 3, impeller 12
of the compressor 11 is turned by the gas turbine
engine main shaft 18. The resulting rotary motion
of the several sets 29a to 29f of compressor blades
causes air to be drawn into inlet end 16 and to be
forced along flow path 14 to the diffuser chamber 22
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from which it is transmitted to the fuel combustor
24 of the engine 18 through tubulation 23 and heat
exchanger module 26.
Air is compressed in stages during passage
through flow path 14 as each set of compressor blades
29a to 29f imparts additional energy to the air flow.
At each successive set 29a to 29f of compressor blades
the added energy appears in part as a rise of static
pressure, in part as tangential velocity energy of
motion and to some extent as heat. The set 32a to 32g
of diffuser blades 31 situated behind each set 29a
to 29f of compressor blades converts a substantial
portion of the velocity energy into additional static
pressure. This process of compression followed by
diffusion is repeated at each successive compression-
diffusion stage 30a to 30f and since the pressure ratio
at each successive stage issubstantially less than
the pressure ratio of the compressor as a whole, each
individual stage operates at high efficiency and the
overall compression process is therefore highly
efficient.
Where the compressor 11 is an air intake
component of a gas turbine engine 18 as in this
example, the gains in ef~iciency in the operation of
the compressor translate into increased efficiency
of the gas turbine engine itself. To the extent
that power losses in the compressor 11 are reduced,
the deliverable power output of the gas turbine engine
18 is increased. Moreover the compressor 11 is very
compact in both the axial and radial direction enabling
the gas turbine engine 18 as a whole to also exhibit
a very desirable degree of compactness.
Significant aspects of the operation of the
compressors lla to llg of the gas turbine engines
35 18a to 18g of Figures 7A to 7G are essentially similar
except insofar as different pressure ratios and air
1~4'~
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mass flows and therefore different output power ratings
for the gas turbine engines are realized in the manner
hereinbefore described.
In the operation of the turbocharger 57 of
Figure 8, the exhaust gasses from engine 58 drive
turbine 59 which in turn drives the compressor llh
through drive shaft 62. The blading means 27h of the
compressor llh draws air into flow path 14h and
delivers such air to the intake manifold 67 of the
engine 58 through diffusion chamber 66. Again, the
multiple staged blading means 27h of the compressor llh
enables the compression and diffusion process to be
accomplished in stages each of which individually
exhibits a small pressure ratio and low diffusion
factor thereby providing for high efficiency in the
operation of the compressor llh and thus in the
operation of the turbocharger 57 as a whole.
Other aspects, objects and advantages of this
invention can be obtained from a study of the drawings,
the disclosure and the appended claims.
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