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BACKGROUND OF THE INVENTION
This invention relates generally to compressor apparatus and, more
particularly, to a method and apparatus for compressing a fluid in a
centrifugal compressor with relatively high efficiencies and over a
substantial operating range.
In a centrifugal compressor, it is desirable to convert the gas kinetic
energy leaving the impeller to potential energy or static pressure. This
is commonly accomplished by way of a diffuser which may be of either fixed
or adjustable geometry. The fixed geometry diffuser may be of the vaneless
type, or it may be of the fixed vane type. An adjustable geometry diffuser
may be either of the vaned or vaneless type and take the form of a
throttle ring as shown in U.S. Pat. No. 4,219,305 assigned to the assignee
of the present invention, a movable wall as shown in U.S. Pat. No.
4,527,949 assigned to the assignee of the present invention, or include
rotatable vanes as shown in U.S. Pat. No. 4,378,194 assigned to the
assignee of the present invention. Each of these various types of
diffusers have peculiar operating characteristics that tend to favor or
discourage their use under particular operating conditions.
Centrifugal chillers used in air conditioning systems are normally required
to operate continuously between full load and part load (e.g., 10 percent
capacity) conditions. At this 10% flow condition, the air conditioning
system still requires a relatively high pressure ratio (i.e., from 50-80%
of the full load pressure ratio) from the compressor. This requirement
puts an extreme demand on the stable operating range capability of the
centrifugal compressor. Therefore, to prevent early compressor surge
caused by impeller stall, centrifugal compressors are typically provided
with a variable inlet geometry device (i.e. inlet guide vanes). Rotatable
inlet guide vanes are able to reduce the flow incidence angle at the
impeller under part load conditions, thus enabling stable compressor
operation at much lower capacities.
In addition to the instability which may be introduced by the particular
impeller and its inlet design, the diffuser may also be cause for
instability under part load conditions. Of all types of diffusers, the
vaneless type generally provides the broadest operating range since it can
handle a wide variation of flow angles without triggering overall
compressor surge. If variable geometry, such as is discussed hereinabove,
is added to such a vaneless diffuser, further stability can be obtained,
but such features add substantially to the complexity and costs of a
system.
Typically associated with the broader operating range of a vaneless
diffuser is substantially lower efficiency levels because of the modest
pressure recovery in the diffuser. The vaned diffuser, on the other hand,
allows higher efficiencies but generally demonstrates a substantially
smaller stable operating range. To increase this operating range, some
type of variable diffuser geometry may be added to the vaned diffuser to
prevent surge when operating under off-design conditions so as to thereby
obtain relatively high efficiency over a broad operating range. But again,
such a structure is relatively expensive.
One type of fixed geometry diffuser that has demonstrated an exceptionally
higher efficiency level is that of the fixed vane or channel diffuser,
which may take the form of a vane island or wedge diffuser as shown in
U.S. Pat. No. 4,368,005, or a so-called pipe diffuser design as shown in
U.S. Pat. No. 3,333,762. The latter was developed for efficiency
improvement under transonic flow conditions occurring in high pressure
ratio gas turbine compressors. Like other vaned diffuser compressors as
discussed hereinabove, higher efficiencies are obtained, but they normally
introduce an associated narrow stable operating range, which for the gas
turbine compressor is not of concern, but when considered for centrifugal
chiller application is of significant concern as discussed hereinabove.
In one instance as shown in U.S. Pat. No. 4,302,150, a pipe diffuser was
used, supposedly to obtain higher efficiencies, with the associated narrow
operating range being broadened by the introduction of a so-called
vaneless diffuser space between the impeller outer periphery and the
entrance to the diffuser. However, the increased stability of such a
design is minimal and only occurs under full load operating conditions
(i.e., no inlet guide vanes). Further, the larger vaneless diffuser space
reduces the compressor lift capability under part load conditions.
Moreover, the introduction of a relatively large vaneless space tends to
move the peak efficiency closer to the surge point, an operating condition
that cannot be tolerated for safe compressor operation.
In addition to the design considerations for the diffuser as discussed
hereinabove, the impeller design features can also be chosen so as to
generally optimize efficiency and operating range. While it is generally
understood that impeller efficiency peaks when its blade exit angle
.beta..sub.2 approaches 45 degrees (as measured from the tangent
direction), there is also a general understanding that, to a point, the
operating range of a centrifugal compressor increases as the impeller
blade exit angle .beta..sub.2 decreases. For a given ratio between the
impeller inlet relative velocity and the impeller exit relative velocity,
reducing the impeller blade exit angle .beta..sub.2 (i.e., increasing the
backsweep) will reduce the absolute flow exit angle .beta..sub.2 leaving
the impeller. If this angle .alpha..sub.2 decreases too far, however, the
radial pressure gradients near the impeller periphery tend to cause flow
separation, and the operating range thus becomes narrower. Therefore, in
centrifugal refrigeration impeller practice, the impeller absolute flow
exit .alpha..sub.2 angle is normally chosen to be within the range of 20
and 40 degrees. Further, heretofore, it was generally understood that to
reduce the impeller flow exit angle .alpha..sub.2 below 20 degrees would
inherently lead to flow separation and a narrowed operating range. The use
of impellers with such flow exit angles have thus been avoided.
It is, therefore, an object of the present invention to provide an improved
centrifugal compressor method and apparatus.
Another object of the present invention is the provision for a centrifugal
compressor which demonstrates high efficiency and a broad stable operating
range.
Yet another object of the present invention is the provision in a
centrifugal compressor for obtaining higher efficiencies without any
substantial loss in operating range.
Still another object of the present invention is the provision in a
centrifugal compressor for a diffuser apparatus which is effective in use
and economical to manufacture and operate.
Still another object of the present invention is the provision for a
centrifugal compressor which is economical to manufacture and effective in
use.
These objects and other features and advantages become more readily
apparent upon reference to the following description when taken in
conjunction with the appended drawings.
SUMMARY OF THE INVENTION
Briefly, in accordance with one aspect of the invention, a fixed vane or
channel type diffuser is provided with a relatively few number of channels
so as to thereby maximize the "wedge angle" therebetween. The associated
impeller is, in turn, so designed that its flow exit angle is relatively
small. The combination of the relatively large wedge angle with the
relatively small flow exit angle allows for a relatively large angle of
incidence without causing flow separation and degradation of the operating
range.
By another aspect of the invention, the diffuser comprises a series of
conical channels having center lines which extend substantially
tangentially to the outer periphery of the impeller. The channel structure
itself brings about increased efficiencies, and the tangential orientation
of the channels to the impeller further enhances the efficiency
characteristics of the system.
In accordance with another aspect of the invention, the impeller is so
designed that its absolute flow exit angle .alpha..sub.2 is maintained
below 20 degrees. This is accomplished in one form by the use of backswept
vanes. Flow separation that might otherwise occur is then prevented by
maintaining the associated wedge angle .alpha..sub.2 between the adjacent
diffuser channels above 15 degrees. In this way, both high efficiency and
a broad stable operating range is obtained.
By yet another aspect of the invention, the vaneless space, between the
outer periphery of the impeller and the leading edge circle defined by the
leading edges of the wedges, is limited in radial depth to thereby reduce
the likelihood of flow separation in the vaneless space. In particular,
the radial dimension is limited so as not to exceed the throat diameter of
the channels.
In the drawings as hereinafter described, a preferred embodiment is
depicted; however, various other modifications and alternate constructions
can be made thereto without departing from the true spirit and scope of
the invention.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a graphic illustration of a performance map for a fixed speed
centrifugal compressor with variable inlet guide vane geometry as compared
with that for the fixed diffuser geometry of the present invention.
FIG. 2 is a partial, axial cross sectional view of a centrifugal compressor
having the present invention incorporated therein.
FIG. 3 is a radial view of the diffuser and impeller portions thereof.
FIGS. 4 and 5 are radial views of the impeller of the present invention
showing the effect of backsweep on the absolute flow exit angle
.alpha..sub.2.
FIGS. 6 and 7 are axial cross sections of the blades showing the effect of
impeller back sweep on the height .beta..sub.2 of the impeller blades at
discharge.
FIGS. 8, and 9 show the flexibility of the present invention in
accommodating various flow rates without diffuser leading edge separation.
THE DESCRIPTION OF THE PREFERRED EMBODIMENT
Referring now to FIG. 1, there is shown a plurality of performance map
curves representative of various configurations of centrifugal compressors
with different inlet guide vane positions as compared with the fixed
diffuser geometry of the present invention. In order to understand the
significance of the present invention, it is desirable to consider some of
the performance characteristics of existing systems.
Centrifugal compressors with vaned diffusers (such as diffusers using
airfoil vanes, single thickness vanes, vane islands or conical pipes) have
higher efficiencies than compressors with vaneless diffusers and are
therefore very attractive, but they also have a smaller stable operating
range and therefore need expensive and complicated variable diffuser
geometry devices and control schemes to prevent surge under off-design
conditions. Considering the definition of the stable operating range as:
##EQU1##
wherein choke mass flow=the maximum flow when the flow reaches sonic
velocity at the throat (represented by curve 1)
surge mass flow=minimum or surge flow representing the lowest stable
operating condition in the compressor (represented by the curves C or D)
It can be stated that well designed centrifugal compressors of intermediate
pressure ratio (i.e. 2.5 to 1 to 5 to 1) with vaneless diffusers can have
a stable operating range of 30%, whereas a centrifugal compressor of
similar pressure ratio with some type of vaned diffuser is limited at best
to a 20% stable operating range.
Many centrifugal compressor applications require part load characteristics,
wherein the head or pressure ratio drops less fast than the flow rate.
Curve A in FIG. 1, for example, represents a typical load line of a water
cooled chiller. In practice, even better part load head capability is
required for water cooled chillers since variations from the typical load
line A are not uncommon. Curve B in FIG. 1, for example, is a typical load
line of a water cooled chiller under variable capacity,
constant-temperature-lift operating conditions.
Vaned diffuser centrifugal compressors with only variable inlet geometry,
part-load control devices are not capable of providing the required head
under off-design conditions. The limited range at full load also results
in limited range under part load conditions. The end result is a steep
surge line on the compressor performance map such as shown at line C in
FIG. 1.
In contrast the performance map of a centrifugal compressor constructed in
accordance with the present invention is shown in the curve D of FIG. 1.
It will be recognized that, in addition to the high efficiency (i.e. in
access of 85%) a very wide stable operating range (i.e. in access of 35%)
is demonstrated. This surge line which exceeds the most severe load line
condition demands (i.e. constant temperature lift water-cooled chiller
operation), is obtained with fixed diffuser geometry and with only one
variable geometry mechanism, i.e. the variable inlet guide vanes. The
specific structure of a centrifugal compressor incorporating the present
invention will now be described.
Referring now to FIGS. 2 and 3, the invention is shown generally at 10 as
comprising a particular configuration of a pipe diffuser 11 combined with
an impeller 12, as installed in an otherwise conventional centrifugal
compressor having a volute structure 13, suction housing 14, blade ring
assembly 16, inlet guide vanes 17, and shroud 18. The impeller 12 is
mounted on a drive shaft 19, along with a nose piece 21. When the assembly
is rotated at high speed, it draws refrigerant into the suction housing
14, past the inlet guide vanes 17, and into the passage 22 where it is
compressed by the impeller 12. It then passes through the diffuser 11,
which functions to change to kinetic energy to pressure energy. The
diffused refrigerant then passes into the cavity 23 of the volute 13, and
then on to the cooler (not shown).
Referring now to FIG. 3, the impeller wheel 12 is shown in greater detail
to include a hub 24, an integrally connected and radially extending disc
26, and a plurality of blades 27. It will be seen that the blades 27 are
arranged in a so called backswept configuration which is a significant
feature of one aspect of the present invention as will be more fully
described hereinafter.
The pipe diffuser 11 is shown in its installed position in FIG. 2, and in
combination with the impeller 12 only in FIG. 3. It comprises a single
annular casting which is secured near its radially outer portion to the
volute structure 13 by a plurality of bolts 28. A plurality of
circumferentially spaced, generally radially extending, tapered channels
31 are formed in the diffuser 11, with their center lines 32 being tangent
to a common circle indicated generally at 30 and commonly referred to as
the tangency circle, which coincides with the periphery of the impeller
12.
A second circle, located just outside the tangency circle, is referred to
as the leading edge circle and is indicated at 33 in FIG. 3. The leading
edge circle, by definition, passes through the leading edges of each of
the wedge shaped islands 34 between the channels 31. The radial space
between the periphery of the impeller 12 and the leading edge circle 33 is
a vaneless/semi vaneless space 25 whose radial depth is limited in
accordance with the present invention in order to broaden the operating
range of the system. That is, the applicant has found that, in order to
prevent flow separation in the vaneless space 25, this radial dimension
should be less than the throat diameter of the tapered channels 31. This
vaneless/semi-vaneless space 25, which, for purposes of simplicity will be
referred to as a "vaneless" space is more fully described in U.S. patent
application Ser. No. 605,619 filed on Oct. 30, 1990, assigned to the
assignee of the present invention, and incorporated herein by reference.
As will be seen in FIG. 3, each of the tapered channels 31 has three
serially connected sections, all concentric with the axis 32, as indicated
at 35, 36 and 37. The first section 35, which includes the "throat"
mentioned above, is cylindrical in form, (i.e. with a constant diameter)
and is angled in such a manner that a projection thereof would cross
projections of similar sections on either circumferential side thereof. A
second section indicated at 36 has a slightly flared axial profile with
the walls 38 being angled outwardly at a angle with the axis 32. An angle
that has been found to be suitable is 2.degree.. The third section 37 has
an axial profile which is flared even more with the walls 39 being angled
at an angle which is on the order of 4.degree.. Such a profile of
increasing area toward the outer ends of the channel 31 is representative
of the degree of diffusion which is caused in the diffuser 11 and is
quantified by the equation
##EQU2##
wherein the area at the exit of the channel is taken normal to the axis at
the location identified at A in FIG. 3.
It was seen in FIG. 3 that the formation of the tapered channels 31 results
in the tapered sections or wedges 34 therebetween. It will also be evident
that the more tapered channels 31 that are formed in the diffuser, the
smaller will be the angle .gamma. of the wedges 34. The particular
diffuser 11 shown in FIG. 3 has 16 tapered channels formed therein, such
that the angle .gamma. is then equal to 221/2.degree.. This relatively
large wedge angle tends to prevent flow separation that might otherwise
occur because of variations in impeller discharge flow angle .beta..sub.2.
As will be seen in the subsequent discussion of the impeller design and
performance, it is desirable to provide for relatively tangential flow.
This, in turn, tends to reduce the change in .beta..sub.2 with mass flow
rate variations. In general, it is therefore desirable to have a
relatively large wedge angle .gamma. to accommodate variations in
incidence. The number of tapered channels 31, however, must be
sufficiently high so as to accommodate the flow volume from the impeller.
The applicant has therefore determined that one can obtain high efficiency
performance over a broad operating range, as is desirable for the present
invention, by a pipe diffuser having a wedge angle, .gamma. as low as
15.degree. (i.e. 24 tapered channels). We will return to the issue of
leading edge separation after a discussion of the impeller design and
characteristics.
Referring now to FIGS. 4 and 5, there are shown impellers 2 and 43 having
different degrees of backsweep. The impeller 42 has blades 44 with a
60.degree. backsweep (i.e. an impeller discharge blade angle .beta..sub.2
of 30.degree.), and the impeller 43 has blades 46 with a 30.degree.
backsweep (i.e. an impeller discharge blade angle .beta..sub.2 of
60.degree.). The absolute tangential component of the flow leaving the
impeller, V.sub.2.sup..theta. can be obtained by the equation
V2.theta.=W2.theta.+U2
where
W2=the tangential component of the relative velocity and
U2=the propeller tip speed
For impellers with backsweep, the direction of the tangential component of
the relative velocity, W2.sup..theta., is opposite to the tip speed
direction. For such impellers, V2.sup..theta. becomes less than U2 and is
reduced further by higher impeller backsweep angles. However, since the
impeller tip speed U2 is several times larger than the total relative
velocity at the impeller discharge W2, the relative change in
V2.sup..theta. due to impeller backsweep is much less than the relative
change in radial velocity V2R caused by impeller backsweep. Because the
increased backsweep reduces the absolute radial velocity V2R to a much
larger extent than the absolute tangential velocity V2.sup..theta.,
another effect of increased impeller discharge blade angle backsweep with
constant shroud stream surface diffusion is a reduction in the absolute
flow angle .alpha..sub.2 leaving the impeller. It will therefore be seen
in FIG. 4 that for a 60.degree. backsweep, the impeller absolute flow exit
angle .alpha..sub.2 is 12.degree., and for a backsweep of 30.degree. as
shown in FIG. 5, the impeller absolute flow exit angle .alpha. .sub.2
=20.degree..
Normally, neither the impeller 42 shown in FIG. 4 or impeller 43 shown in
FIG. 5 would be acceptable for operation where a broad operating range is
desired since the radial pressure gradients at the impeller periphery
would tend to cause flow separation. However, when used with the pipe
diffuser of the present invention, these lower absolute flow exit angles
.alpha..sub.2 are not only possible but, as discovered by the applicant,
allow one to obtain higher efficiency over a relatively broad operating
range.
It will be recognized that in comparing the impellers of FIG. 4 and 5, an
increase in the impeller backsweep reduces the blade to blade normal
distance n2 of the discharge normal flow area as shown in FIGS. 6 and 7.
That is, the high backsweep impeller of FIG. 4 with its attendant reduced
blade to blade normal distance n2 requires a greater impeller discharge
blade height b.sub.2 than the impeller discharge blade height b.sub.2 as
shown in FIG. 7, which is associated with the lower backsweep impeller 43
of FIG. 5. If we assume that we want to maintain the relative velocity
radio W.sub.2 /W.sub.1, where W.sub.2 is the relative impeller discharge
velocity and W.sub.1 is the relative impeller inlet shroud velocity, then
an increase in impeller backsweep angle will therefore result in an
increase in the impeller tip blade height b.sub.2. This relatively wider
tip impeller tends to provide stability at low flow conditions since it
results in smaller absolute impeller discharge flow angles .alpha..sub.2
which therefore will show smaller angle variations at reduced flow.
Consequently, incidence effects will be less to thereby promote stability.
In summary, there are three features in the diffuser and impeller
structures of the present invention which contribute to the high
efficiency, broad operating range characteristics of the present
invention. First, the number of tapered channels 31 is limited such that
the wedge shaped islands 34 therebetween have a relatively large wedge
angle .gamma. such that the occurrence of flow separation at the tips are
minimized. Secondly, the vaneless space 25 between the outer periphery 30
of the impeller 31 and the leading edge circle 33 is limited in its radial
depth such that the occurrence of flow instabilities are prevented. In
this regard, the combination of the small vaneless space 25 together with
the solidity of the wedges 34, create pressure fields inside the vaneless
space with the gradients being more parallel with the direction of flow
rather than creating radial gradients which would tend to cause flow
separation. Finally, the use of an impeller with high backsweep, and
therefore one with the wide tip impeller, a very shallow discharge flow
angle, and relatively small absolute angle variations, reduces the
sensitivity of the downstream component (i.e. diffuser) to variations in
flow rate and thus increases the stable operating range of the compressor.
These results are illustrated in FIGS. 8 and 9.
In both FIGS. 8 and 9, the pipe diffuser 11 and the impeller 12 are
identical to that in FIG. 3, that is with a 60.degree. backsweep in the
impeller, with a vaneless space whose radial depth is less than the
diameter of the tapered channel throat, and with a wedge angle of
221/2.degree.. When the flow is at the full design flow level, the
absolute flow exit angle .alpha..sub.2 in the flow direction is parallel
to the center line of each of the tapered channels 31 of the diffuser 11.
This is shown by the arrows in FIG. 8. It will be seen that the two
intermediate arrows represent the direction of refrigerant flow as it
engages the wedge 34 on its pressure and suction side. It will thus be
understood from this illustration that no flow separation will occur at
the tip of the wedge 34. The absolute flow exit angle .alpha..sub.2 is
12.degree. at this flow level.
Referring now to FIG. 9, the amount of flow is substantially reduced such
that the absolute flow exit angle .alpha..sub.2 is reduced to 2.degree..
Here, the flow direction is parallel to the suction side, and there will
of course be no flow separation. The two intermediate arrows again
represent the direction of flow that will engage the wedge 34 on its
suction side 48. Again, it will be seen that the angles are such that flow
separation at the tip of the wedge 34 will not occur.
While the present invention has been disclosed with particular reference to
a preferred embodiment, the concepts of this invention are readily
adaptable to other embodiments, and those skilled in the art may vary the
structure thereof without departing from the true spirit of the present
invention.
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