U.S. patent number 3,784,318 [Application Number 05/213,417] was granted by the patent office on 1974-01-08 for variable diffuser centrifugal pump.
This patent grant is currently assigned to General Electric Company. Invention is credited to Donald Y. Davis.
United States Patent |
3,784,318 |
Davis |
January 8, 1974 |
VARIABLE DIFFUSER CENTRIFUGAL PUMP
Abstract
A centrifugal pump is supplied with a discharge shutter valve
which provides a variable diffuser for the pump. The shutter valve
includes a hollow, slotted cylinder positioned for movement into a
variety of operative positions between the impeller and the
diffuser vane passages. Each operative position completely closes
one or more of the diffuser passages while leaving the remaining
passages completely open, thereby providing a pump with high head
rise over the complete operating range of the pump.
Inventors: |
Davis; Donald Y. (Cincinnati,
OH) |
Assignee: |
General Electric Company
(N/A)
|
Family
ID: |
22795048 |
Appl.
No.: |
05/213,417 |
Filed: |
December 29, 1971 |
Current U.S.
Class: |
415/158;
415/164 |
Current CPC
Class: |
F04D
29/468 (20130101); F04D 15/0038 (20130101); F05D
2270/64 (20130101) |
Current International
Class: |
F04D
15/00 (20060101); F04D 29/46 (20060101); F04d
015/00 (); F04d 029/46 () |
Field of
Search: |
;415/157,158,151,150,164,165 |
References Cited
[Referenced By]
U.S. Patent Documents
Foreign Patent Documents
|
|
|
|
|
|
|
706,107 |
|
Mar 1931 |
|
FR |
|
736,266 |
|
Jun 1943 |
|
DD |
|
133,892 |
|
Sep 1929 |
|
CH |
|
Primary Examiner: Raduazo; Henry F.
Attorney, Agent or Firm: Thomas J. Bird Jr. et al.
Claims
What I claim is:
1. In a centrifugal pump of the type including an impeller adapted
to propel fluid through a radial outlet to a diffuser which
includes a plurality of stationary diffuser vanes defining diffuser
passages therebetween, the improvement comprising:
a discharge shutter valve positioned between said impeller and said
diffuser vanes and capable of movement between an inoperative
position and at least one operative position, said valve including
meand for completely closing the inlets to individual diffuser
passages when in an operative position while leaving the inlets to
the remaining passages completely open.
2. The improved centrifugal pump recited in claim 1 wherein said
valve comprises a hollow cylinder.
3. The improved centrifugal pump recited in claim 2 wherein said
cylinder has a plurality of axially extending slots located
therein, said slots being separated by solid portions of said
cylinder.
4. The improved centrifugal pump recited in claim 3 wherein each of
said slots is positioned opposite the inlet of one of said diffuser
passages.
5. The improved centrifugal pump recited in claim 4 wherein at
least one of said slots is of a different axial length from the
remaining slots.
6. The improved centrifugal pump recited in claim 5 wherein said
valve includes actuating means for moving said cylinder between its
operative and inoperative positions.
7. The improved centrifugal pump recited in claim 6 wherein said
actuating means includes a piston extending from one end of said
cylinder, a chamber for receiving said piston, and means for
delivering servo fluid to opposite sides of said piston.
8. The improved centrifugal pump recited in claim 7 wherein said
slots extend from the end of said cylinder opposite the end which
includes said piston, and said slots extend toward said piston.
Description
BACKGROUND OF THE INVENTION
This invention relates generally to centrifugal pumps and, more
particularly, to a variable diffuser exit passageway for such a
pump.
Many fluid pumping applications require a very large flow range
with low pump temperature rise throughout the flow range. For
example, current and advanced gas turbine engines and variable
thrust liquid rocket engines require very large fuel flow ranges
with minimum fuel temperature rise from the pumping system. This
minimum fuel temperature rise requirement is rather easily attained
at high flow rates because of the high flows and the short dwell
times within the pumping system. The low pump temperature rise
requirement during low flow operation is not as easily attained. In
fact, fuel pump performance considerations show that the
requirement of a high fuel flow range (high flow turndown) is
inconsistent with the requirement of low fuel temperature rise
across the pump at low pump outputs.
As a result of the above consideration, traditional gas turbine
engine and variable thrust liquid rocket engine fuel systems
utilize combinations of several fuel pumping elements operating in
parallel to solve the flow range problem. The number of pumping
elements operated at any one time depends upon the required flow
output, with the remaining pumps being inactivated until needed.
This approach to obtaining a wide fuel flow range, while successful
in providing low fuel temperature rise at low pump flows, presents
problems of added size, complexity and control.
Centrifugal pumps, while capable of providing the high output
requirements of gas turbine engine and variable thrust liquid
rocket engine systems, have historically been limited to relatively
narrow flow ranges because of the excessive temperature rise and
stability problems associated therewith at low percentages of rated
flow. Previous attempts at providing a centrifugal pump with a
relatively wide flow range have met with various degrees of
success. For example, variable speed drive centrifugal pumps can
achieve the necessary high flow turndowns but the reduced speed
results in substantial head dropoff from the pump. Pump inlet
throttling, such as by using the inlet vapor core approach, or pump
discharge throttling with a diffuser inlet shutter can each be
utilized to provide the reduced temperature rise at low flow. Each
of these approaches, however, also results in significant head rise
dropoff at high flow turndown ratios.
One further proposed approach consists of providing a pump with a
variable angle diffuser vane, which can successfully provide the
low temperature rise without significantly causing head rise
dropoff. The complexity of such a pump, however, in most cases far
outweighs the advantages gained therefrom.
SUMMARY OF THE INVENTION
It is an object of this invention, therefore, to provide a
relatively high flow range centrifugal pump having low temperature
rise throughout the range of operation. It is a further object of
this invention to provide such a pump without the necessity of
variable angle diffuser vanes.
Briefly stated, the above and other related objects are attained by
providing a centrifugal pump with a variable diffuser which
utilizes a slotted valve to selectively close off diffuser vane
entry passages as the fuel rate of the pump is decreased. In one
embodiment, the slotted valve takes the form of a slotted
cylindrical sleeve radially located between the pump impeller tip
and the pump diffuser entry passages. Means are provided for moving
the valve axially as a function of pump-delivered flow in such a
manner that selective diffuser vane entry passages are completely
closed off while the desired flow is delivered through the
remaining fully open diffuser vane entry passages. The net effect
of the slotted valve is to increase overall efficiency at low
delivered flows while maintaining near rated head rise of the pump
throughout its range of operation.
DESCRIPTION OF THE DRAWINGS
While the specification concludes with a series of claims which
particularly point out and distinctly claim the subject matter
which Applicant regards as his invention, a complete understanding
of the invention will be gained from the following description of a
preferred embodiment. This description is given in connection with
the accompanying drawings in which:
FIG. 1 is a generally sectional view, with portions deleted,
describing a centrifugal pump constructed in accordance with the
prior art;
FIG. 2 is an enlarged partial sectional view of a centrifugal pump,
similar to FIG. 1, describing a pump constructed in accordance with
the present invention;
FIG. 3 is a sectional view, with portions deleted, showing the
cross section of the pump of FIG. 2 with the slotted valve in one
of its operative positions;
FIG. 4 is a perspective view of the slotted valve of FIGS. 2 and 3;
and
FIG. 5 is a graphical plot showing the outputs of the pumps
illustrated in FIGS. 1 and 2.
DESCRIPTION OF A PREFERRED EMBODIMENT
Referring now to the drawings wherein like numerals correspond to
like elements throughout, attention is directed initially to FIG. 1
wherein a centrifugal pump constructed in accordance with the prior
art is designated generally by the numeral 10. The pump 10 is shown
to include an axial inlet 12, a rotating impeller wheel 14 which
includes a plurality of impeller vanes 16, and a casing 18 which
defines a radial outlet 20 surrounding the tips of the impeller
vanes 16. A diffuser 22 surrounds the radial outlet 20 and includes
a plurality of stationary diffuser vanes 24 at its inlet. Each pair
of the diffuser vanes 24 defines a diffuser entry passage 26 from
the radial outlet of the pump 10 to a toroidal-shaped collector 27.
The impeller wheel 14 extends from a rotatable shaft 28, which is
journaled for rotation in bearings 30.
As previously mentioned, the elements described above are typical
of prior art centrifugal pumps. For this reason, the structure of
the centrifugal pump 10 is shown somewhat schematically in FIG. 1.
As will become apparent from the following description, Applicant's
inventive concept may be applied to any type centrifugal pump and
the structure shown in FIG. 1 is merely meant to be illustrative
and not limiting in any manner.
Centrifugal pumps, such as that shown in FIG. 1, have been limited
to relatively narrow flow ranges because of the excessive
temperature rise and stability problems at low percentages of rated
flow. These problems may be substantially reduced or eliminated by
reducing low flow power loss. For this reason it is known to
provide the centrifugal pump 10 with a diffuser inlet shutter 32
which throttles the pump discharge by partially closing all of the
diffuser entry passages 26. In its simplest form, as shown in FIG.
1, the diffuser inlet shutter 32 takes the shape of a hollow
cylinder positioned between the radial outlet 20 of the pump 10 and
the diffuser vanes 24 as shown in FIG. 1. The cylinder 34 is
provided with some actuating mechanism, such as a piston 36 formed
integrally with one end of the cylinder 34. The piston 36 is
positioned within a chamber 38 which is supplied with actuating
fluid by means of ports 40 and 42, respectively.
In operation, fluid is supplied to the centrifugal pump 10 at the
inlet 12 in any known manner, while the shaft 28 and impeller wheel
14 are driven by some suitable driving mechanism (not shown). The
fluid is acted upon by the impeller vanes 16 and is delivered
through the radial outlet 20 at a much higher pressure and velocity
than its inlet pressure and velocity. The fluid then passes through
the diffuser entry passages 26 of the diffuser 22, where its
velocity pressure is converted to static pressure. From the
diffuser 22 the fluid is directed into collector 27. The fluid may
then be directed to a suitable control device (not shown) or to
some other component, depending upon the use of the pump 10.
In cases where the maximum output of the pump 10 is desired,
sufficient servo fluid is delivered to the rod end of the piston
36, through the port 42, to position the cylinder 34 completely
outside of the diffuser inlet passages 26.
Being a fluid dynamic device dependent upon fluid velocity and
velocity vectors for adequate flow, pressure and pressure
stability, the centrifugal pump 10 can only be designed to produce
optimum effiency at a single value of output flow. This optimum or
best efficiency point is usually obtained at or near the maximum
flow point in order to satisfy the required pressure rise with
minimum impeller size. This necessity of sizing the pump for high
flow results in poor efficiency at low flows since the pump is
grossly oversized in terms of fluid volume during low flows. In
addition, the blade and diffuser shapes are mismatched for low flow
operation. At these low flows, excessive power loss to the fuel can
result in large fuel temperature rise, and numerous forms of pump
and pump/system related pressure instabilities can occur.
For this reason, whenever the particular application of the pump 10
calls for low flow outputs, servo fluid is delivered to the head
end of the piston 36 through the port 40 and the piston 36 is moved
partially into the radial outlet 20. By thus throttling the pump
discharge, the low flow power losses can be substantially reduced
and the fluid temperature rise associated with the power losses can
be minimized. Unfortunately, by partially blocking each of the
diffuser entry passages 26, the effectiveness of the diffuser 22 is
decreased, thereby resulting in significant head rise dropoff at
high flow turndown ratios. This drawback to the use of diffuser
inlet shutter is shown graphically by the dotted line curves of
FIG. 5. As shown therein, as the percent of rated flow decreases
the effect of partially closing the diffuser inlet passages 26
reduces shaft power somewhat, but the effect on the Q-H curves is
also significant with the percent of rated head rise dropping
significantly near the lower end of the rated flow spectrum.
Referring now to FIGS. 2 through 4, Applicant's inventive concept
for diminishing the fluid temperature rise while maintaining near
rated head rise throughout the range of operation of a pump 10'
will be described. The basic construction of the pump 10' remains
unchanged from that previously described. That is, the pump 10'
still includes an inlet 12, an impeller wheel 14, a diffuser 22, a
plurality of diffuser vanes 24 which define diffuser inlet passages
26, and a collector 27. The primary difference between the pump 10
described in connection with FIG. 1 and the improved pump 10',
shown in FIGS. 2 through 4, lies in the fact that the diffuser
inlet shutter 32 of the pump 10 is replaced with a variable
diffuser valve 50.
The variable diffuser valve 50, as shown most clearly in FIG. 4,
consists of a cylinder 52 having a plurality of various length
slots 54, 56 and 58, located therein. The cylinder 52 is further
shown to include a piston member 60, similar to the piston 36 of
FIG. 1, which may be formed integrally with the cylinder 52 or
joined thereto in any known manner.
Each of the slots 54 - 58 begins in the end of the cylinder 52
opposite the piston 60 and extends axially toward the piston 60.
The slots 54 - 58 are of approximately equal width but are of
varying axial lengths, with the slots 54 being the shortest and the
slots 58 the longest, as shown in FIG. 4.
The slots 54 - 58 are spaced around the periphery of the cylinder
52 and positioned in such a manner that each of the slots lies
between one of the diffuser inlet passages 26 and the tips of the
impeller vanes 16. This arrangement is shown in the perspective
view of FIG. 4 wherein the eight slots are positioned as if there
were twelve equally spaced slots, i.e. the centerlines of the slots
lie 30.degree. apart. Since there are twelve diffuser passages 26
and only eight of the slots 54 - 58, there are four solid portions
of the cylinder 52, generally designated by the numeral 59, which
block four of the diffuser passages 26 as soon as the valve 50 is
moved into a first operative position within the radial outlet 20.
While the cylinder 52 is shown to include a total of eight slots of
three various lengths, the total number of slots and the sizing
thereof will, of course, depend upon the number of diffuser
passages and the design of the centrifugal pump, as will become
apparent from the following description of the operation of the
variable diffuser centrifugal pump 10'.
The basic operation of the pump 10' is identical to that of the
pump 10 with fluid being delivered through an inlet 12 to the
impeller wheel 14 and being directed radially outwardly while being
acted upon by the impeller vanes 16. The fluid is then delivered
through the radial outlet 20 to the diffuser entry passages 26,
where its velocity pressure is converted to static pressure, and
then it passes into the collector 27.
When the pump 10' is operating near its rated maximum output, the
variable diffuser valve 50 is held in an inoperative position such
that all diffuser entry passages 26 are fully open to receive fluid
from the radial outlet 20 and to deliver the same to the collector
27. As the desired output flow of the pump is decreased, the valve
50 is moved axially into a first operative position between the
tips of the impeller vanes 14 and the diffuser vanes 24. The
diffuser valve 50 is stopped in this first position in which each
of the slots 54, 56 and 58 lies between the tips of the impeller
vanes 14 and one of the diffuser inlet passages 26. In this manner,
because there are eight slots positioned opposite the twelve
diffuser vane passages 26, eight of the passages 26 will remain
fully open while the remaining four diffuser passages 26 will be
completely closed by the solid portion 59 of the cylinder 52. In
this manner, the eight fully opened diffuser inlet passages 26
operate as if the pump 10' were delivering its total rated flow,
while the remaining four passages have absolutely no effect on the
pump output. That is, the eight diffuser inlet passages 26 are
still operating near their peak design efficiency in contrast to
the structure of FIG. 1 which partially closes down each of the
diffuser inlet passages causing throttling and resulting loss of
diffuser pressure recovery.
When the desired output flow from the variable diffuser centrifugal
pump 10' further decreases, the diffuser valve 50 is moved further
to the left (as shown in FIG. 2) to a second position in which the
slots 54 are no longer positioned between the impeller tips and the
diffuser vanes 24. That is, the diffuser valve 50 is moved to a
second operative position in which that portion of the cylinder 52
which includes the slots 54 lies within a groove 62 (FIG. 2) formed
in the casing 18. In this position, only the slots 56 and 58 are
positioned between the impeller wheel 14 and the diffuser vanes 24,
and thus only four of the diffuser vane passages 26 are fully open
while the remaining eight are fully closed. The four fully open
passages 26 continue to operate near their design efficiency, but
the required shaft power of the pump is reduced significantly. FIG.
3 is a sectional view of the pump with the diffuser valve 50
located in this second position. As clearly shown in this view, the
slots 56 and 58 deliver fluid from the impeller vanes 16 to four of
the diffuser vane passages 26, while the remaining eight passages
are completely blocked.
Similarly, if the desired output of the variable diffuser
centrifugal pump 10' is further decreased, the diffuser valve 50 is
moved axially to a third operative position in which only the slots
58 lie between the impeller wheel 14 and the diffuser vanes 24. In
this position, only two of the diffuser vane passages 26 receive
fluid from the impeller wheel 14, while the remaining ten passages
are completely blocked.
The output of a pump constructed as shown in FIGS. 2 through 4 is
plotted as solid lines in FIG. 5. As shown therein, the net effect
of the variable diffuser is to increase overall efficiency at low
delivered flows while maintaining near rated head rise. That is,
the shaft power curves for the fully open and three alternative
positions of the diffuser valve 50 are identical to those for the
centrifugal pump 10. However, the head rise curves show that the
percent of rated head rise actually increases as the number of open
diffuser inlet passages 26 decreases, as opposed to the dotted head
rise curves which show significant decreases in head rise for the
pump 10 of FIG. 1 at low rated flow levels. The increase in head
rise is caused by the increasing pressure as the number of diffuser
inlet passages is decreased.
As described above, Applicant has provided an improved centrifugal
pump which has a number of basic advantages. The primary advantage
of this pump is the high head rise over the full operating range of
the pump. This phenomenon, when combined with the further advantage
of low fluid temperature rise due to low power loss at low flow
rates, provides a centrifugal pump capable of usage in many areas
heretofore thought impossible. Other advantages become readily
apparent when one compares the simplicity of the described design
with the complexity of previously known variable diffuser
geometry-type pumps. Furthermore, the pump described above provides
the potential of maintaining stable pump operation at lower
delivered flow rates than pump 10.
Various changes could be made in the structure shown in FIGS. 2
through 4 without departing from the broader aspects of Applicant's
invention. For example, the means for causing axial movement of the
diffuser valve 50 could take many forms other than the piston 60
described above. For example, the diffuser valve 50 could be
positioned mechanically by some linkage arrangement. Likewise, the
shape and number of the slots within the cylinder 52 could vary
substantially from those shown while still providing the basic
function of completely closing a certain number of inlet diffuser
passages while leaving the remaining passages completely open.
These and other similar changes are meant to be covered by the
appended claims.
* * * * *