U.S. patent application number 10/591025 was filed with the patent office on 2007-12-20 for axial flow pump and marine propulsion device.
Invention is credited to Donald E. Cornell, William M. Farrell.
Application Number | 20070292278 10/591025 |
Document ID | / |
Family ID | 34704011 |
Filed Date | 2007-12-20 |
United States Patent
Application |
20070292278 |
Kind Code |
A1 |
Cornell; Donald E. ; et
al. |
December 20, 2007 |
Axial Flow Pump and Marine Propulsion Device
Abstract
A mechanically reconfigurable marine propulsion device that
adapts to engine torque and/or vessel speed thereby providing
improved propulsive efficiency and performance. The axial flow
propulsion device has two or more stages each having an impeller
section and a stator section. Stator vanes and/or the pumping
chamber provide a flow diffusion that generates increased
hydrostatic pressure from ram pressure recovered from high velocity
working fluid which, due to reduced fluid velocity and increased
hydrostatic pressure, lowers cavitation events and frictional
losses within the propulsion device. Optionally, variable-pitch
vanes in the stator section control the amount of ram pressure
imparted to the working fluid. Also optionally, variable-pitch
inlet guide vanes control the whirl angle and/or mass flow rate of
incoming fluid independently of rotor or vessel speed. A set of
fixed or variable exit guide vanes aft of the pumping chamber
provides flow straightening and pressure maintenance at a discharge
nozzle. Other options include a dual flow concentric pumping
arrangement that improves performance at low to moderate vessel
speeds, an inlet diffuser that recovers ram pressure within an
intake duct, and/or a variable area throat in the discharge nozzle
that controls the water jet exit velocity according to vessel speed
in order to maintain propulsive efficiency. Advantageously, the
variable geometry propulsion device enables a shipmaster to achieve
improved performance and fuel efficiency over a wide range of
vessel speed, vessel loading, sea state conditions, power settings,
and/or engine set points in order to achieve higher vessel speeds
(e.g., 10 to 30 kts) and/or greater range (e.g., 25-40%) over
conventional marine pump jets (i.e., centrifugal or mixed-flow
waterjets), which is particularly useful for vessels utilizing 10
to 100 megawatt power plants.
Inventors: |
Cornell; Donald E.; (New
Port Richey, FL) ; Farrell; William M.; (Walton,
NY) |
Correspondence
Address: |
Lawrence Harbin;McIntyre Harbin & King
500 Ninth Street SE
Washington
DC
20003
US
|
Family ID: |
34704011 |
Appl. No.: |
10/591025 |
Filed: |
March 15, 2005 |
PCT Filed: |
March 15, 2005 |
PCT NO: |
PCT/US05/08474 |
371 Date: |
July 3, 2007 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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10801705 |
Mar 17, 2004 |
7108569 |
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10591025 |
Jul 3, 2007 |
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60455578 |
Mar 19, 2003 |
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60606905 |
Sep 3, 2004 |
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Current U.S.
Class: |
417/245 ; 417/53;
440/38 |
Current CPC
Class: |
B63H 2011/046 20130101;
B63H 2011/084 20130101; B63H 11/11 20130101; B63H 2011/081
20130101; B63H 11/103 20130101; Y02T 70/56 20130101; B63H 11/08
20130101; F04D 29/566 20130101; Y02T 70/50 20130101; F04D 3/00
20130101; F04D 15/0022 20130101 |
Class at
Publication: |
417/245 ;
417/053; 440/038 |
International
Class: |
B63H 11/00 20060101
B63H011/00; F04B 1/00 20060101 F04B001/00; F04B 49/00 20060101
F04B049/00 |
Claims
1. A multistage axial flow pumping device comprising: a housing, a
longitudinal chamber within the housing to convey a working fluid
from an inlet to an outlet, said chamber including at least two
rotor stages that each include a rotor section, said chamber
further including a stator section between respective rotor
sections, said rotor section including a multi-bladed drive wheel
positioned downstream of said inlet and operative to rotate around
an axis to thereby to pump working fluid from the inlet to the
outlet through a high pressure section defined within said chamber,
and said stator section being positioned downstream of a rotor
section and including plural stator vanes substantially fixed
relative to said housing and geometrically arranged to define a
flow path having a cross-sectional area between vanes that
increases from an entry point to an exit point of the stator
section.
2. The device of claim 1, wherein said stator vanes are variable in
pitch to adjust the angle of attack of the working fluid upon entry
into a following rotor section.
3. The device according to claim 1, wherein said outlet comprises a
convergent annular chamber at a discharge nozzle whereby to
increase velocity of liquid discharged from said outlet.
4. The device of claim 3, wherein said convergent annular chamber
includes a variable throat area positioned immediately upstream of
said outlet to vary the velocity of the liquid discharged from said
outlet.
5. The device of claim 4, further including an actuator to vary the
axial position of a nozzle plug whereby to vary effective area of
said outlet.
6. The device of claim 4, further including balancing pistons
operated by pressure differentials obtained by sensing discharge
nozzle jet velocity and vessel velocity respectively to obtain a
desire optimum discharge velocity relative to vehicle velocity.
7. The device of claim 6, further including a nozzle plug position
override effective to reposition said nozzle plug by overriding
automatic positioning of said balancing pistons with hydraulic
pressure.
8. The device of claim 1, further including at least one variable
inlet guide vane positioned downstream of said inlet and operative
to adjust the whirl angle of fluid entering said inlet.
9. The device of claim 1, further comprising a fixed set of inlet
guide vanes positioned downstream of said inlet to adjust inlet
whirl angle of liquid entering said inlet.
10. The device of claim 9, further comprising an inlet diffuser
serving as an inlet duct positioned upstream of said inlet guide
vanes.
11. The axial flow pumping device of claim 1, further comprising a
larger diameter booster section preceding the high pressure section
and operative to increase fluid pressure at said inlet.
12. The axial flow pumping device of claim 11, wherein said booster
section comprises multiple rotor stages, said booster section
further including a stator section between respective rotor
sections.
13. The axial flow pumping device of claim 11, wherein said booster
section includes variable discharge nozzles and an actuator to
control the discharge nozzles to maintain a given pressure at the
inlet of the high pressure section.
14. The axial flow pumping device of claim 11, wherein said booster
section further includes variable pitch stator vanes.
15. The axial flow pumping device of claim 11, further comprising a
common drive shaft for said pumping device and said booster
section.
16. The axial flow pumping device of claim 11, wherein said booster
section includes a set of inlet guide vanes to control the whirl
angle of incoming fluid.
17. The axial flow pumping device of claim 1, wherein said
longitudinal chamber has a cross-sectional area that initially
decreases in a downstream direction and that subsequently increases
upstream of said outlet to form an expansion region prior to
discharge of the working fluid whereby to convert ram pressure of
higher speed working fluid to static pressure in lower speed
working fluid in order to reduce internal friction within said
longitudinal chamber and to improve propulsive efficiency of said
pumping device.
18. The axial flow pumping device of claim 17, wherein said
longitudinal chamber is annular.
19. The axial flow pumping device of claim 18, wherein rotor and
stator vanes of successive stages in the chamber have decreasing
heights in the downstream direction and rotor blade drive wheels
have corresponding increasing diameters in the downstream
direction
20. The axial flow pumping device of claim 18, wherein rotor and
stator vanes of successive stages in the chamber have decreasing
heights and the housing diameter decreases in the downstream
direction to attain a decreasing effective area of the annular
chamber in the downstream direction.
21. The axial flow pumping device of claim 18, wherein rotor and
stator vanes of successive stages in the chamber have increasing
heights in the downstream direction and rotor blade drive wheels
have corresponding decreasing diameters in the downstream
direction.
22. The axial flow pumping device of claim 18, wherein rotor and
stator vanes of successive stages in the chamber have increasing
heights and the housing diameter increases in the downstream
direction to attain an increasing effective area of the annular
chamber in the downstream direction.
23. A method of conveying a working fluid through an axial flow
pumping device comprising: defining a flow path in the axial flow
pumping device to convey the working fluid from an inlet to an
outlet, providing multiple stages within said flow path that each
include a rotor section, providing at least one stator section
between at least two rotor sections, driving working fluid through
said flow path by rotating the rotor section, and lowering the
speed of working fluid in said path by providing increased flow
path areas between vanes of the stator section as the working fluid
travels in the downstream direction.
24. The method of claim 23 wherein said inlet includes variable
inlet guide vanes having controllable pitch and said method further
including controlling the whirl angle at said inlet by altering the
pitch of said inlet guide vanes.
25. The method of claim 23 wherein said outlet includes an exit
guide vane stage operative to straighten flow of said working fluid
and to increase static pressure prior to discharge.
26. The method of claim 23, further including boosting pressure of
said working fluid prior to conveying the working fluid from the
inlet to the outlet.
27. The method of claim 23, further comprising oppositely altering
ram and static pressure of the working fluid by varying the
cross-sectional area of the flow path in the downstream direction
whereby to reduce internal friction and improve propulsive
efficiency.
28. A method of controlling discharge velocity of water discharged
from an axial flow pumping device relative to water speed a vessel,
said method comprising: detecting discharge velocity of water
discharged from the axial flow device, detecting water speed of the
vessel, providing a discharge nozzle in said axial flow device
having a variable area throat, and utilizing said discharge
velocity and boat velocity to control the area of said throat
according to a desired set point based on the discharge velocity of
the water and the speed of the vessel.
29. The method of claim 28, further comprising providing an
actuator that drives the discharge nozzle, said detecting steps
includes detecting respective pressures associated with said
discharge velocity and water speed of the vessel, and said
utilizing step includes using the respective pressures to drive
said actuator to an equilibrium position that defines a desired
optimum throat area of the discharge nozzle.
30. A dual flow axial flow liquid pumping device comprising: a low
pressure section that conveys liquid along a downstream path in a
chamber having a first cross-sectional area, a high pressure
section in communication with the low pressure section that conveys
said liquid along a downstream path in a chamber having a second
cross-sectional area that is smaller than said first
cross-sectional area, and said first and second sections each
having at least one rotor-stator stage mounted on respective shafts
that drive the respective stages to stepwise pressurize the liquid
in the respective sections.
31. The dual flow axial flow liquid pumping device of claim 30,
wherein at least one of said high and low pressure sections
includes a set of stationary vanes operative to increase static
pressure between said at least one rotor-stator stage of a section
by converting ram pressure of the liquid to static pressure.
32. The dual flow axial flow liquid pumping device of claim 31,
wherein said low pressure and high pressure sections have
concentric annular chambers in communication with each other and
said low pressure section includes control nozzles that control
discharge of liquid around the annular chamber of said high
pressure section.
33. The dual flow axial flow liquid pumping device of claim 32,
wherein said respective shafts are common.
34. The dual flow axial flow liquid pumping device of claim 32,
wherein said respective shaft are geared together to provide
different rotational speeds of the respective shafts.
35. A dual flow pumping method that pumps water during water jet
propulsion, said method comprising: pressurizing the water at a
first lower pressure, conveying the water along a first downstream
path having a first cross-sectional area, utilizing the first lower
pressure of the water to establish a higher pressure in said water
along a second downstream path having second cross-sectional area
that is smaller than said first cross-sectional area, providing at
least one rotor-stator stage in each of said first and second
downstream paths in order to stepwise pressurize the water in each
path, and powering said rotor stator stages in each path to propel
a vessel by water jet propulsion.
36. An water jet propulsion device comprising: a propulsion pump;
and a diffuser preceding said propulsion pump, said diffuser
including an inlet, a larger outlet in communication with said
propulsion pump, multiple sections between said inlet and outlet
that have increasing cross-sectional area in the downstream
direction, and multiple groups of deflector vanes between said
sections that re-direct and convert ram pressure of the liquid to
static pressure between respective sections.
37. The propulsion device of claim 36, wherein said deflector vanes
are positioned between respective sections to re-direct said liquid
approximately 30.degree. from a nominal flow direction of a
preceding section.
Description
BACKGROUND
[0001] The present invention is directed to an axial flow pumping
device and method, as well as an application thereof to achieve
efficient marine propulsion.
[0002] In a multistage axial flow pump having two or more impeller
stages, a certain amount of energy is transferred from a power
plant to working fluid (e.g., water) at successive stages. Pressure
is essentially stepwise increased at the succeeding stages until
discharge of the working fluid through an outlet nozzle whereby to
generate thrust. Increased pressure inside the pump suppresses
damaging cavitation that may otherwise act upon the impellers. An
axial flow pump differs from a more voluminous centrifugal or
mixed-flow pump that is generally limited to single stage. Thus, an
axial flow pump, if properly designed, may have a higher power
density than conventional pumps. Apart from use in marine
propulsion, other applications of the present invention include
high volume pumps for fire or flood control, irrigation, and in
large cooling towers requiring extremely high volumes or
pressures.
[0003] In marine applications, design parameters of the propulsion
pump are ideally matched with engine torque and speed, i.e., the
engine power or performance curve. Most engines, however, have only
a single optimum operating speed that delivers peak horsepower and
another single operating speed for peak efficiency, which may not
optimally match the desired thrust and/or hull speed of the vessel.
Such desired hull speed or thrust generally varies with vessel
loading, sea state conditions, and/or temperature and density of
the ocean. Thus, certain inefficiencies inherently exist in
conventional engine-propulsor combinations during operation of a
vessel.
[0004] To compensate for such inefficiencies during the desired
operating condition, it has been known to vary the pitch rotor
blades in a pump's impeller section according to optimum torque,
speed, or fuel efficiency of the engine. It is also known, but not
necessarily applied to marine propulsion, to include fixed stator
vanes between impeller sections of a multi-stage pumping device to
counteract whirl or rotational velocity that rotor blades impart to
the fluid, such as that disclosed by U.S. Pat. Nos. 5,755,554 and
5,562,405 (both issued to Ryall). The stator vanes has the effect
of redirecting fluid flow to achieve a desired angle-of-attack of
fluid acting on rotor blades in the succeeding stage, but such
prior stator vane designs significantly increased internal
friction. Ryall, for example, provides a substantially constant
absolute velocity in flow passages between fixed stator blades. Due
to their geometric structure, prior stator vane designs endured
losses in efficiencies and generally operated, at best, around 65
to 72% propulsive efficiency.
[0005] The present invention, in part, takes advantage of the
relationship between static and ram pressure, that is, the fact
that the total pressure at within the pumping device (as well as at
the pump's intake and discharge) comprises the sum of hydrostatic
(static) of the pumped fluid plus the impact (ram) pressure
imparted to the fluid by the impellers. It is also known the extent
of internal frictional losses, due to barrier layer effects of the
fluid transgressing internal components of the pump, increases
exponentially with fluid speed. At the pump inlet, a diffuser may
be used to alter static pressure before water enters the impeller
section. Also, it was not heretofore known, among other things, to
alter the static pressure component of total pressure between rotor
stages (by impeller design or geometric shape of the pump housing);
to provide a low pressure booster for pumping mechanism; or to
provide variable pitch stator vanes or a variable inlet guide vanes
that mechanically reconfigure the pump to compensate for variations
or operating characteristics of the power plant, desired mission
profile, ship velocity or loading, water density or water
temperature changes.
SUMMARY OF THE INVENTION
[0006] Thus, one aspect of the invention comprises a multistage
axial flow pump that includes an outer housing, a substantially
annular chamber within the housing that conveys working fluid
(e.g., water) from an inlet to an outlet, multiple stages within
the chamber that may each include a rotor section and a stator
section. Stator vanes in the stator section may be fixed relative
to housing and have a geometrical shape to define a fluid flow path
having a cross-sectional area that increases as the fluid
transgresses the stator section in the downstream direction whereby
to increase static pressure. Optionally, variable internal geometry
may be provided to reconfigure the pump parameters to match desired
operating conditions of a vessel. For example, the stator vanes may
optionally have variable pitch and the chamber may optionally
include a variable nozzle having a discharge area (i.e., throat)
that is controlled to optimally match the water jet discharge speed
with the vessel speed. In yet a further aspect of the invention, a
set of variable inlet guide vanes controls inlet fluid flow by
changing the inlet area and/or whirl angle of incoming fluid. Such
variable geometry enables the propulsion device to match a wide
range of prime movers of different power.
[0007] The invention also includes a method implemented by the
pumping device. One aspect of the method includes providing an
actuator that varies the throat area of the discharge nozzle,
detecting respective pressures associated with the discharge
velocity and speed of the vessel, and using the respective
pressures to drive the actuator to an equilibrium position that
defines a desired optimum throat area of the discharge nozzle
according to the instantaneous speed of the vessel. Other aspects
are set forth in the claims.
[0008] In another aspect of the present invention, it was
recognized that if a trade-off is made between static and ram
pressure by increasing static pressure and reducing ram pressure
using a diffuser-type annular chamber between stages of a
multi-stage pumping device (with total pressure remaining
constant), inherent frictional losses can be significantly lowered
since friction exponentially decreases with fluid speed. This
results in a more efficient propulsion device. Thus, an aspect of
the invention also comprises a multi-stage pumping device, such as
that described in U.S. patent application Ser. No. 10/801,705,
having diffuser-like chamber between stages to control static and
ram pressures according to desired a relationship. Construction of
either the housing or the annular chamber may be varied where the
effective area of the chamber in a direction normal to fluid flow
is progressively increased or decreased. The height of the stator
and rotor blades is correspondingly varied according to the height
(clearance between the shaft wheel and housing) within the annular
chamber.
[0009] Other aspects of the invention are pointed out by the
appended claims.
BRIEF DESCRIPTION OF THE DRAWINGS
[0010] FIG. 1 is partial cut-away view of a multistage axial flow
propulsion or pumping device according to one embodiment of the
present invention that includes three rotor-stator stages, variable
geometry inlet guide vanes, a thrust reversing/steering mechanism,
variable-area discharge nozzle, and a mechanism to control nozzle
discharge area.
[0011] FIG. 2A shows an axial flow pumping device where, by virtue
of increasing drive wheel diameters in the downstream direction,
the effective area of an annular chamber decreases in the
rotor-stator sections, and then the effective area of the annular
chamber substantially increases prior to reaching the discharge
nozzle thereby converting ram pressure of the working fluid to
static pressure (e.g., by slowing down fluid speed and decreasing
internal frictional losses).
[0012] FIG. 2B shows an axial flow pumping device where, by virtue
of decreasing housing diameter but constant drive wheel diameter in
the downstream direction, the effective area of the annular chamber
decreases in the rotor-stator sections, and then the effective area
of the annual chamber substantially increases prior to reaching the
discharge nozzle thereby converting ram pressure in the working
fluid to static pressure (e.g., by slowing down fluid speed and
decreasing internal frictional losses).
[0013] FIG. 3 shows a partial cut-away view of an alternative
design of a multistage propulsion or pumping device according to
yet a further aspect of the present invention, which includes
variable pitch stator vanes in order to provide adaptive
power-torque conversion between an engine and propulsion
device.
[0014] FIG. 4 shows a second, segmented stator vane design that may
be incorporated in the propulsion device of FIG. 3, according to
yet another aspect of the present invention.
[0015] FIG. 5 shows a first stator vane design that may be
incorporated in the pumping or propulsion device of FIG. 3,
according to yet another aspect of the present invention.
[0016] FIG. 6 depicts actuator and control mechanisms that may be
incorporated in a multistage propulsion device to control variable
inlet guide vanes and/or the pitch of stator vanes according to yet
other aspects of the present invention.
[0017] FIG. 7A shows a front view of one embodiment of a control
ring and actuator that may be used to control the pitch of variable
pitch stator vanes of an axial flow pumping or propulsion device
according to yet a further aspect of the present invention.
[0018] FIG. 7B shows a side view of slip ring and control arm
mechanism to vary the pitch angle of stator vanes according to an
aspect of the present invention.
[0019] FIG. 7C is a plan view of a control arm of FIG. 7 that
controls the pitch angle of the stator vanes.
[0020] FIG. 8 shows a piston-cylinder drive and pressure balancing
mechanism to vary the discharge area of the multistage propulsion
device according to sensed water jet velocity and vessel speed
according to yet a further aspect of the present invention.
[0021] FIG. 9A is a rear perspective view of an exemplary rotor
blade the may be used with the illustrative pump or propulsion
device.
[0022] FIG. 9B is a side view of the rotor blade of FIG. 9A.
[0023] FIG. 9C is a rear view (viewed from a downstream position)
of the exemplary rotor blade of FIG. 9A.
[0024] FIG. 9D is a top view of the exemplary rotor blade of FIG.
9A.
[0025] FIG. 10 is a conceptual view of a series of rotor-stator
sections of an exemplary three-stage pumping or propulsion device
that optionally includes a set of inlet guide vanes and a set of
exit guide vanes.
[0026] FIG. 11A shows an improvement that comprises a low pressure
booster section preceding the inlet of the exemplary propulsion
device of FIG. 3 or FIG. 1, which enables increased mass flow rates
to enhance acceleration from zero to slow speeds and/or higher
propulsive efficiencies at low to moderate vessel speeds.
[0027] FIG. 11B shows variable discharge nozzles of the booster
section of FIG. 11A which variably open and close (manually or
under automatic control) according to vessel speed.
[0028] FIGS. 12A, 12B, and 12C illustrate a split housing design of
a multistage axial flow pump adapted to receive stator vane groups
having a desired fixed pitch in order to adapt a given pump design
to a wider range of power plants.
[0029] FIGS. 13A, 13B, and 13C show details of the stator vane
assembly used in the split housing design of FIGS. 12A, 12B, and
12C.
[0030] FIGS. 14A and 14B illustrates an inlet diffuser that may
precede a pumping device, such as an multistage axial flow pumping
device, in order to boost or increase hydrostatic pressure by
recovering a portion of ram pressure brought about by vessel
speed.
[0031] FIG. 15 shows a specific embodiment of an inlet diffuser of
FIGS. 14A and 14B having stepped sections and deflector vanes
therebetweeen to effectively recover or convert ram pressure in a
preceding section to static pressure in a succeeding section.
DESCRIPTION OF ILLUSTRATIVE EMBODIMENTS
[0032] FIG. 1 shows a first embodiment of a pump or propulsion
device 10 having a substantially cylindrical outer casing 12, an
inlet 14 through which a substantially incompressible working fluid
(e.g., water) enters, and an outlet 18 that discharges the working
fluid as an accelerated jet discharge 22. In marine, firefighting,
irrigation, or flood control applications, the working fluid is
water. Device 10 includes an internal annular chamber 19 extending
along and circumscribing an axis 13. Chamber 19 conveys working
fluid from inlet 14 to outlet 18 under power imparted by multiple
pumping stages each of which may comprise a rotor section and an
intervening stator section. Respective rotor sections of device 10
include a rotor blade 30, 32, or 34 attached to a corresponding
rotating wheel, such as wheel 45 centered on axis 13. Blade 30 is
coupled to and rotated with wheel 45. Multiple concatenated wheels
and an internal wall of casing 11 define the annular chamber 19
within the cylindrical housing of device 10. Although a cylindrical
housing is preferred, housing 12 may have a non-cylindrical
shape.
[0033] Concatenated wheels are driven in unison by drive shaft 20,
which may be coupled to any one of a number of conventional
engines. Mounting flange 24 couples device 10 to a fluid conduit
that supplies working fluid to device 10. Forward and aft sets of
thrust bearings 15 and 17 support the shaft along axis 13 within
casing or housing 12. Thrust bearings 15 and 17 also absorb or
counteract a relatively large opposing axial force between housing
12 and shaft 20 developed by multiple rotor sections during
operation of the device. Preferably, each of the rotor blades 30,
32, and 34 radially extends from axis 13 of an associated rotating
wheel to a given design height, width, thickness, and twist angle
so as to impart maximum energy to a working fluid.
[0034] Stator vanes 40, 42, and 44 lay in respective stator
sections following respective rotor sections but instead are
fixedly attached relative to wall 11 of the casing or housing 12,
rather than being attached to a rotating wheel. Vane design is
similar to the blade design of the rotors. Stator vanes 40, 42, and
44 serve to redirect and/or diffuse the flow of working fluid from
the rotor blades, e.g., rotor blades 30, 32, and 34, in the
preceding section. In operation, rotor blades impart energy to the
working fluid by accelerating fluid in a partial tangential and
partial axial direction relative to axis 13, thus increasing the
ram or impact pressure of the fluid as it enters the next stage.
The stationary vanes redirect the working fluid in an opposed
tangential direction, e.g., to counteract whirl imparted by the
preceding rotor section, as the fluid flows in annular chamber 19
along axis 13 towards outlet 18.
[0035] According to an important aspect of the invention, the
stator vanes are arranged to effectively reduce the velocity of the
working fluid but retaining total pressure therein by providing an
expanding area between vanes as fluid flows through the stator
section. In part, this is accomplished by providing, in embodiments
illustrated in FIGS. 4 and 5, an airfoil-shaped stators (with or
without a segmented flap portion) having a thicker leading edge
portion and a thinner trailing edge portion. Other geometric shapes
achieving the same or similar results may also be utilized. In one
practicable embodiment, the flow path area in a direction of fluid
flow through the stator section may increase, for example, from a
factor of about 1.15 to 1.5 (e.g., 23%), more or less. Such
expanding flow path area between stator vanes correspondingly
decreases the working fluid speed and simultaneously increases the
static pressure of the fluid prior to entry into the next rotor
stage. Fluid velocity decreases proportionately, more or less.
However, total pressure of the fluid, i.e., static pressure plus
impact or ram pressure imparted by the preceding rotor section,
remains relatively constant (except for minor frictional losses)
within the stator section. Thus, the geometric arrangement of the
stator vanes relative to fluid flow enables a speed reduction of
the working fluid without sacrificing total pressure thereby
reducing internal frictional and flow losses associated with higher
fluid speeds. The arrangement of the stator vanes also increases
static pressure of the fluid prior to the next stage thereby
providing a higher initial static pressure upon which the rotor
blades work in order to impart energy. Thus, the rotor blades in
effect deliver further impact or ram energy to the working fluid by
increasing pressure derived from the preceding stage. Successive
stepwise increases in static pressure provided by the stator
sections and successive recovery and supplementation of impact or
ram energy provided by successive rotor sections significantly
increase the final working fluid pressure at the discharge nozzle
and thus significantly increase the overall effectiveness of the
pump or propulsion device.
[0036] Preferably, device 10 has three or more stages although two
stages may also suffice. Each rotor section may or may not include
a subsequent stator section. Fluid enters the next or succeeding
stage at essentially the same total pressure of the fluid being
discharged from the preceding stage. The rotor sections impart
additional pressure to the fluid at each stage. Stepwise increases
in pressure is repeated as many times as necessary to attain the
desired design point pressure at region 21, which supplies
pressurized fluid to an annular discharge nozzle. The discharge
nozzle includes an axially variable plug 60 that controls the size
of the area of throat 28 between deflector 52 and plug 60.
Preferably, region 21 defines an annular nozzle that is convergent
to eject water at an increased velocity thereby generating
propulsive thrust. Thrust, which can be measured in pounds, equals
mass flow times velocity.
[0037] In a preferred embodiment, the size or area of throat 28 in
the annular discharge nozzle is variable and controllable, and may
be used to trim the water jet discharge velocity to maximize boat
velocity.
[0038] Inlet 16 of device 10 preferably includes a series of inlet
guide vanes 46 that serve to control, redirect, or throttle
incoming fluid flow and/or to change the angle of attack of
incoming fluid. This alters the load on the rotor blades in the
first stage of device 10. Due to differential cross-sectional areas
of inlet duct 26, the velocity of water at entry into the inlet
duct is lower than the velocity of the water entering the casing of
device 10. In the inlet duct, there is a transition section 23 from
larger to smaller area so that the difference is not so abrupt as
to cause losses from eddies thereby maintaining streamline fluid
flow. A principal embodiment of the invention does not require
inlet guide vanes 46 in the first stage although other embodiments
do. In a fixed inlet guide vane embodiment, the vanes direct water
flow into the first rotor-stator stage 30, 40 at a prescribed angle
and function as a flow director. In an embodiment utilizing
variable inlet guide vanes, i.e. variably controlled vanes actuated
by actuator ring 48 and actuator 47, the flow angle of water
entering the first rotor stage of blade 30 is variable. This not
only changes the incidence angle of the working fluid entering the
pump but also changes the amount of flow and therefore the inlet
guide vanes function as a throttling mechanism. Thus, guide vanes
46 provide mass flow throttling of the working fluid, and include
control linkage to rotate the vanes 46 about .+-.30.degree. from a
neutral position according to a desired mass flow rate.
[0039] At the discharge end of device 10, the axial position of
nozzle plug 60 is controllable to effectively open or restrict the
water jet throat area 28. When plug 60 is extended, as shown in
FIG. 1, the area of throat 28 is smaller thereby resulting in a
faster water discharge speed for a given mass flow rate. A
retracted nozzle plug 62, shown in phantom, opens the area of
throat 28 to a larger area and thus lowers water discharge speed
for a given constant mass flow rate. A plug position control
mechanism including pressure sensors, such as pitot tubes 66 and
68, provide balanced pressure settings in a piston drive head 64 to
attain optimum positioning of nozzle plug 60 in relation to speed,
loading, or other parameters of the vessel.
[0040] When deployed in marine applications, steering may be
accomplished by redirecting the water jet at the discharge nozzle.
In the embodiment of FIG. 1, the device 10 may include thrust
reversers on each side thereof in the form of a deflector 52
hydraulically actuated by cylinder 55 and control arm 56. When
driven to a reverse position to seal off the throat 28 by engaging
the head of plug 60, as shown by deflector 53 (shown in phantom),
fluid flow is redirected from region 22 and is forced in a
direction 58 (also shown in phantom). When corresponding deflectors
are provided at four quadrants of outlet 18, simultaneously
actuating the deflectors to a reverse position produces a reverse
thrust to reverse the direction of travel of the vessel. Respective
deflectors on left and right sides of the vessel may be
independently operated to provide steering. In addition, the
discharge region of device 10 may be mounted on a gimbal to effect
redirection of thrust to provide steering.
[0041] FIG. 2A shows an axial flow pumping device 200 where, in the
downstream direction, the effective area of the annular chamber
decreases in the rotor-stator sections 218, 220, and 222 due to
increasing diameters of drive wheels 238, 239, and 240. This has
the effect of decreasing the effective area of the annular chamber
216 in the downstream direction thereby increasing fluid flow
speed. However, upon reaching region 217 of the annular chamber,
the effective area abruptly increases thereby substantially
increasing the effective area of the annular chamber prior to
reaching the throat 218 of discharge nozzle 250. This abrupt
increase in area converts ram pressure to static pressure (slowing
down fluid speed and decreasing internal frictional losses) within
the propulsor. In the embodiment of FIG. 2A, it is also seen that
the height of the rotor blades and stator vanes decrease in the
downstream direction. Variable pitch stator or rotor vanes (either
or both being variable) may be employed to effect power delivery to
the fluid according the desire fluid speed within the respective
rotor-stator sections.
[0042] FIG. 2B shows an axial flow pumping device 230 where, in the
downstream direction, the effective area of the annular chamber 216
also decreases due to a decreasing diameter of housing 242, but the
diameters of drive wheels 238, 239, and 240 remain constant. Here,
it is seen that the heights of the respective rotors 230, 232, and
234, as well as the height of the respective stators 231, 233, and
235 decrease in the downstream direction. This construction also
effects and increase in fluid speed in the rotor-stator sections,
but at region 217, the effective area of annular chamber 216
abruptly increases thereby substantially slowing down fluid speed r
prior to reaching the throat 219 of discharge nozzle 250. Ram
pressure of the high-speed fluid is converted to a high static
pressure at lower fluid speed thereby reducing internal friction
and providing the desired thrust at nozzle 250. Similarly, variable
pitch stator or rotor vanes (either or both being variable) may be
employed to effect power delivery to the fluid according the desire
fluid speed within the respective rotor-stator sections.
[0043] In an alternative design, rather than providing an effective
area of annular chamber 216 that progressively decreases in the
downstream direction, the effective area of the annular chamber in
the downstream direction may increase (rather than decrease) while
transgressing the rotor-stator sections and then merge with a
region 217 in a less abrupt transition. Such alternative design may
be accomplished by varying the blade height or housing geometry, as
explained in the earlier embodiment. Thus, the invention embraces
various geometrical designs that take advantage of diffuser designs
to trade off static and ram pressures, while maintaining constant
total pressure, to improve efficiencies of a multi-stage axial flow
pumping device.
[0044] FIGS. 3-5 shows yet another embodiment of a pumping or
propulsion device in which, rather than providing "fixed pitched"
stator vanes 30, 32, 34 (FIG. 1), variable pitch stator vanes 80
and 82 are provided. Control arms 85, 89 control the effective
pitch of vanes 80, 82 upon tangential translation of linkages 83,
87. A control ring (not shown) actuates linkages 83, 87 when
rotated upon the outer casing 12, and such rotational control may
be implemented manually or under computer control in response
monitored operational parameters such as engine torque, vessel
speed, velocity of the water jet, vessel loading, fluid density, or
a combination thereof, in order to attain optimum performance or
efficiency. The stator vanes may also be segmented into a
stabilizer section 92 and a trailing section 90 that about a shaft
91, as depicted in FIG. 4. Shaft 91 is preferably integrally formed
with trailing section 90 of the stator vane. In an exemplary
embodiment, trailing section 90 is designed to pivot plus or minus
thirty degrees, more or less, about a neutral position. About
twelve to fourteen stator vanes 80 are circumferentially and evenly
spaced within the annular chamber 19, which extend radially from
axis 13 from about 3.0 inches to about 4.5 inches. A similar or
smaller number of rotor blades may be used on each wheel.
[0045] In the exemplary embodiment of FIG. 3, the outer radius of
the wheels, such as wheel 45, defines the inner surface of annular
chamber 19 at about 3.0 inches from axis 13 while the outer radius
of chamber 19 is about 4.50 inches from axis 13. Preferably, the
height of the rotor blades and stator vanes is about 1.5 inches and
the ratio of blade or vane height to its cord is about 1:1 or
higher. The ratio of blade height to drum radius in the exemplary
embodiment is preferably between 0.66 and higher, i.e., a blade
height of at least 2/3.sup.rd of the drum radius, or more. None of
these exemplary dimensions, however, constitutes a limitation of
the invention. This exemplary embodiment can be driven with a 1250
horsepower engine at a propulsive efficiency exceeding 84 to 86%.
When used to pump water in other applications, the exemplary
embodiment is designed to pump water over six hundred and fifty
feet vertically at a flow rate of about 8500 gallons/minute.
[0046] Instead of using a segmented vane structure, vanes 80, 82
may take on the form 86, as depicted in FIG. 5, which is
constructed much like a standard airfoil having a single section 95
that pivots about a shaft 93. Shaft 93 is preferably integrally
formed with vane 95.
[0047] The material of the vanes and stator may comprise any of a
variety of materials known in the art such as titanium, bronze, a
high carbon stainless steel, a composite material, or other
material that is preferably non-corrosive and/or adapted for marine
applications.
[0048] In addition, there is provided a "fixed pitch" exit guide
vane 84 (FIG. 3) that is fixedly attached to wall 11 of housing
12.
[0049] FIG. 6 illustrates one type of mechanism to vary the pitch
of stator vanes according to the variable pitch stator vane aspect
of the invention where an actuator 100 under manual or automated
control includes an actuator rod 102 that translates control arm
104 in direction 105 parallel to axis 13. A series of L-shaped
linkages 106, 108 and 110 interconnect control arm 104 with
respective pitch actuating turnbuckles 112, 114, and 116 to vary
the pitch of inlet guide vane 46 as well as the pitch of a series
of stator vanes, one of which is shown at 30. Turnbuckle 112
couples control ring 124 via flange 126, turnbuckle 114 couples
control ring 122 via flange 128, while turnbuckle 116 couples
control ring 120 via extension 129. The tumbuckles include a
threaded adjustment rod that may be adjusted to properly align the
pitch angle of the stator vanes and inlet guide vane relative to
each other. Upon translation of control arm 104 in an axial
direction, the trailing portion of variable pitch stator vane 30
(shown in cut-away view) changes pitch by pivot action of linkage
108 about pivot point 109. This action drives control ring 122
circumferentially around casing 12 via connecting flange 128.
Circumferential movement of control ring 122 turns the stator vane
30 via arm 140. As indicated above, a preferred embodiment may vary
the angular pitch of stator vane 30 (or trailing portion thereof)
by as much as plus or minus twenty or thirty degrees. A similar
action occurs with respect to control ring 120 to actuate the lever
130 to vary the pitch of the interconnected stator vanes underneath
casing 12 (not shown). Levers 131 and 132, which are ganged to
control ring 120 with other levers, similarly vary the pitch of
interconnected stator vanes.
[0050] As apparent from the illustrated actuating mechanism,
control of the stator vanes and the inlet guide vane 46 occur in
unison for simultaneous pitch angle changes. Pitch angle changes
alter the angle of attack of, and hence, the torque applied against
or energy delivered to the working fluid by the rotor blades of the
following section. Each rotor section thus stepwise increases the
energy imparted to the working fluid. Control of the inlet guide
vanes of ring 124 may, however, be separated from control of the
stator vanes of rings 120 and 122. As control arm 104 axially
translates, linkage 110 pivots about pivot point 111 to advance and
retract turnbuckle 112, which drives control ring 124 via flange
126. Control ring 126 couples the shaft of inlet guide vane via
actuating arm 150. Preferably, actuator 100 is controlled in a way
to attain peak power output or peak propulsive efficiency of the
pumping device as working fluid enters the inlet 16.
[0051] Thus, according to the structure of FIG. 6, the rotors are
fixed pitch while the stators are variable pitch. The
pitch-changing mechanism is simple in design, construction, and
maintenance.
[0052] FIG. 7A depicts one of the control or actuator rings, i.e.,
actuator ring 120, in greater detail. As apparent, upon actuation
of hydraulic or electrical actuator 104, angled link 106 rotates
about a pivot point 107 to effect a vertical excursion of
turnbuckle 116 which, in turn, circumferentially rotates actuator
ring 120 around casing 12 to alter the pitch angle of the stator
vanes, e.g., stator vanes mechanically coupled with control arms
130, 131, and 132 of FIG. 6. FIG. 7B shows stator vane control arm
130 in operative relation with actuator ring 120 and shaft 135 of a
variable pitch stator vane. There, a slot in guide block 121
enables the actuator ring 120 to circumferentially rotate when
actuated by turnbuckle 116 (FIG. 6) that, in turn, sweeps the end
of control arm 130 through slot 134 via locking pin 133 extending
through hole 138 of control arm 130. This action effects rotation
of shaft 135, which is interlocked with control arm 130 via inset
137, as further illustrated in FIG. 7C. Bushing 136 confines shaft
135 to an axial position and seals water pressure inside casing
11.
[0053] FIG. 8 illustrates yet an addition aspect of the invention,
which is designed to optimally match boat speed with water jet
speed when deployed in marine applications. The apparatus and
method may be used to automatically or manually control the throat
of the discharge nozzle by altering the axial position of nozzle
plug 60 to attain optimum propulsive efficiency according to boat
speed and water jet speed. In determining how such control is to be
implemented, sea level static thrust=W/g*V. The net thrust of a
vessel underway, however, is characterized by: Thrust
T=W/g*(V.sub.j-V.sub.b) (1) where thrust T=mass flow rate in weight
of working fluid (i.e., water) per unit volume per second,
g=gravitational acceleration constant (e.g., expressed as 32
ft/sec.sup.2), velocity Vj=exit velocity of the fluid jet at the
discharge nozzle, and velocity Vb=exit velocity of the vessel
relative to the water. The exit velocity exerts a dynamic pressure
P.sub.d equal to 1/2 the density Rho of the working fluid times the
velocity squared divided by two times the acceleration of gravity,
or P.sub.d=(Rho*V.sup.2)/2g (2)
[0054] It is known that dynamic pressure P.sub.d at the discharge
nozzle is directly proportional to the velocity squared V.sup.2 of
the fluid. Propulsive efficiency (Np) equals the useful thrust
output divided by the combination of useful thrust output and
losses (e.g., frictional losses). So, if Vb represents the velocity
of boat and Vj represents the velocity of the water jet at the
discharge nozzle, then the Absolute (or effective) Discharge
Velocity Va equals Vj-Vb. Therefore, propulsive efficiency
Np=((W/g)*Va*Vb)/{(W/2g)*(V.sub.j.sup.2-V.sub.b.sup.2)} (3)
[0055] Simplifying the expression of Np, then Np=2/(1+Vj/Vb)
(4)
[0056] Therefore, it is seen that the propulsive efficiency Np is
indirectly proportional to the ratio of the water jet and boat
velocities. Propulsive efficiency Np is also proportional to the
ratio of the dynamic pressures generated by the jet and boat
velocities, i.e., Np=Pd(jet)/Pd (boat). Using equation (4) above,
the propulsive efficiency Np is 67% for a hull design speed of 30
knots at a water jet speed of 60 knots.
[0057] FIG. 8 illustrates one type of mechanical arrangement to
capture these relationships and control nozzle discharge area, or
the speed of the water jet in relation to boat speed. The fluid
discharge area is defined by throat 208, which is confined by head
200 of the nozzle plug and the internal walls of casing 11 at the
throat area. Nozzle head 200 axially moves in a direction indicated
by line 207 to alter the effective area of throat 208, which
extends within an annular path of chamber 19. A first pitot sensor
210 senses pressure of the working fluid in throat 208 while a
second pitot tube 212 senses pressure of the water in the hull ship
stream that is exerted by boat speed. Pitot tube 212 extends
downwardly below water level 215 and opens to the direction of
travel of the vessel. A line 211 communicates sensed pressure of
pitot tube 210 with nozzle head retraction chamber 204. Flex line
213 communicates pressure sensed by pitot tube 212 with nozzle head
extension chamber 203. In chambers 203 and 204, which are
preferably cylindrical in construction, forces acting upon opposing
sides of preferably cylindrical piston 202 are measured by pressure
times the area of respective surfaces 203a and 204a. In a circular
piston, a circle defines area 203a whereas concentric circles
define area 204a. Piston 202, however, may be non-circular. Thus,
the respective velocities of water sensed by the pitot tubes 210
and 212 are translated to opposing forces acting on opposing sides
of piston 202, which is mechanically coupled to or integrally
formed with nozzle head 200.
[0058] A balance in the opposing forces is achieved when the
individual products of pressure and area equalize, which drives
piston 202, and consequently nozzle head 200, to an equilibrium
position (e.g., from position indicated by phantom nozzle 201)
thereby providing a mechanism and method to optimize water jet
speed for a given boat speed, assuming the operator has knowledge
of characteristics of the boat, e.g., optimum hull speed. In
mechanical construction, the diameter d of neck 205 defines the
areas of respective surface 203a and 204a, which due to their
respective areas automatically effects equilibrium at the
appropriate nozzle head position. In the exemplary device, the area
of surface 203a is 1.88 times the area of surface 104a.
[0059] To automatically control or override the pressure-driven
equilibrium position of nozzle head 200, automated computer control
may be implemented to actuate servos according to sensed pressure
at pitot tubes 210 and 212, or conventional transducers and
amplifiers may be deployed to produce appropriate control signals
to drive a servo or actuator. Instead of using pitot static
pressure, the axial position of nozzle 200 in larger propulsion
devices may be electrically or hydraulically actuated. In addition,
a pressure regulator may be interposed on either or both lines 211,
213 (or elsewhere) to alter the equilibrium position of or control
piston 202.
[0060] FIGS. 9A through 9D show an exemplary rotor blade design
that may be used with the illustrated pumping or propulsion device.
Stator vanes may have a similar blade construction, but
incorporating a shaft as shown by FIGS. 4 and 5. The illustrative
rotor blade of FIGS. 9A and 9B includes a base 300 having a curved
head 304 to support blade 302. According to an aspect of the
invention, particularly in connection with the stator vane design,
blade 302 has a thin or sharp trailing edge 306 so that an area of
the flow path that is normal to fluid flow expands as fluid travels
from leading edge 308 to trailing edge 306 of blade 302.
Preferably, blade 302, head 304, and base 300 are integrally formed
of non-corrosive material, such as stainless or high carbon steel,
bronze, or other materials known in the art. In relation to the
central rotor axis 13 (FIG. 1), the height of the exemplary blade
at equally spaced points A-F from head 304 to the outer tip 310
(FIG. 9C) are 5.93136, 6.67501, 7.41866, 8.162308, 8.905955, and
9.6436023 inches. FIG. 9C shows the relative twist of the exemplary
blade and FIG. 9D shows the cross-sectional geometry of the blade
from its leading edge 308 to its trailing edge 306. As known in the
art, increasing the radius ratio (i.e., the ratio of blade height
to tip radius) decreases blade efficiency. Such losses stem from
differential pressures between the root and tip of blade 302, which
result from an increased tip velocity of the blade relative to the
working fluid.
[0061] FIG. 10 illustrates exemplary blades and vanes of a
three-stage device where working fluid travels in direction 416
through the device upon rotation of the rotor blades in direction
417. As depicted, the three stages comprise rotor-stator section
stages 402-403, 404-405, and 406-407. Only a couple of blades or
vanes are shown in each section, which is conceptually represented
by cross-cuts at a mean blade or vane height. To simplify the
illustration, blade or vane twist is not shown in the
illustration.
[0062] The illustration of FIG. 10 includes an optional,
variable-pitch inlet guide vane stage 408, as well as an optional,
fixed-pitch exit guide vane stage 410 that straightens the flow of
the working fluid prior to discharge. In a preferred structure, it
is desired to obtain at each stage a ratio of inlet velocity
V.sub.1 to exit velocity V.sub.2 of about 1.15 to 1.50 where
V.sub.1/V.sub.2=1.15 to 1.0 (5)
[0063] Due to a decreasing area of the flow path between the inlet
guide vanes 412, 414 and 418, which define the respective flow
paths, the velocity of the working fluid for a given mass flow rate
increases as it flows through section 408. As seen, the
cross-sectional area between inlet guide vanes 412 and 414
decreases in downstream direction 416 because the vane geometry
provides a wider width W1 at its section entry and a narrower width
W2 at its section outlet. The cross-sectional area of the flow path
between vanes is measured by width multiplied by vane height,
assuming the guide vanes have the same twist angle and constant
height throughout the section. As measured in a plane normal to
flow path 400, the area of the flow path between the vanes
decreases in the downstream direction. According to an aspect of
the invention, the flow path area between the inlet guide vanes can
be altered by changing the pitch angle of the inlet guide vanes, as
shown by exemplary vane 418, for example.
[0064] As known in the art, total or absolute pressure of the
working fluid in an axial flow device includes two components,
i.e., a ram or impact pressure component and a static pressure
component. The rotor blades impart ram or impact pressure to the
fluid. Static pressure is ambient. Assuming total or absolute
pressure remains constant throughout the inlet guide stage, an
increase in fluid flow speed after passage through the inlet guide
stage 408 necessarily decreases the static pressure component of
the working fluid if total pressure is to remain the same. Thus,
the variable inlet guide vanes enable altering of pressure and
whirl angle of the fluid before entering the first rotor stage.
This provides an additional level of control of the performance of
the pumping or propulsion device.
[0065] In stages 402-410, however, the area of the flow passage
between rotor blades and stator vane increases from an entry point
to an exit point of each section thereby decreasing the speed of
the work fluid as it flows through the pumping or propulsion
device. In the succeeding stages 402-410, the width W1 at the entry
point between rotor blades 422 and 424 is less than the width W2 at
the exit point of these blades--resulting in expanding flow path
area when blade height is constant in the direction of axis 13.
Likewise, the width W1 at the entry point between stator vanes 426
and 428 is less than the width W2 at the exit point of these
vanes--resulting in expanding flow path area when vane height
remains constant in the direction of axis 13. A similar decrease in
working fluid velocity occurs in stages 404-410. Given a constant
overall mass flow rate through the pumping or propulsion device, it
is seen that the velocity of the working fluid decreases at each
section. The decreased velocity over the succeeding stages also
lowers internal frictional and eddy (non-laminar) flow losses
(which exponentially increases with speed) that are typically
encountered in axial flow devices, thus further improving
efficiency.
[0066] Advantageously, the difference in magnitude of W1 and W2,
and consequently the relative entry and exit speeds as well as the
extent of whirl of the working fluid when passing the stator
section, may be changed by altering the pitch angle of the stator
vanes 426 and 428, as indicated by variable pitch stator blade 430.
Changing the angle of attack of the fluid prior to the rotor stage,
i.e., changing the amount of whirl, alters the load placed on the
engine, or energy imparted to the fluid. Thus, this aspect of the
invention substantially improves the overall operating efficiency
at various operating set points of the vessel, or at various engine
speeds, torque or power. Although W1 and W2 designate entry and
exit point width of each section shown in FIG. 10, these lengths
may differ between or among or within the stages or sections
without departing from the scope of the invention. Blade or vane
twist may also differ among stages, sections, or even within a
stage or section. In addition, concentric cylinders substantially,
i.e., the internal wall of the outer casing and the exterior
surface of the rotor blade wheel, define the illustrated annual
chamber of the pumping or propulsion device but other geometries
may also be employed to define a suitable flow path.
[0067] The exit guide vanes 440, 442, and 442 serve to straighten
fluid flow at the discharge nozzle. Their pitch angle may be fixed
or variable. A mechanism similar to that use to vary the stator
vanes may be employed to vary the pitch angle of the exit guide
vanes. This provides an additional layer of control.
[0068] As apparent, the invention allows control of thrust either
by controlling mass flow via inlet guide vane position, by altering
the pitch of the stator vanes (in the variable pitch embodiment of
the invention) and thus the pressure imparted to the fluid by each
rotor section, by altering the discharge nozzle area or jet
velocity to optimally match boat and water jet speed, or any
combination thereof, for any given horsepower, torque, or drive
speed applied to a multistage axial flow pump or propulsion device.
Since it is desired to operate most turbine or piston engines
(diesel or gasoline) at an maximum power, at maximum fuel
efficiency, at an optimum constant engine speed for best hull speed
or sea state condition, or on an optimum performance curve, inlet
guide vane throttling (to control mass flow) and/or discharge jet
velocity may advantageously be adjusted at the will or desire of
the shipmaster to meet any varied performance characteristics of
the vessel. The inlet guide vanes may be configured to rotate plus
or minus thirty degrees, more or less, from a neutral position.
This way, mass flow is positively controlled independent of the
speed of the vessel.
[0069] For a long haul, the shipmaster may desire to operate on a
best speed-range curve to travel the known distance in the shortest
time. In other situations, the shipmaster may desire to travel the
farthest distance given the amount of fuel onboard. In yet other
situations, the shipmaster may desire to travel at the highest
speed given the available horsepower, sea state condition, loading
of the vessel, and/or design speed of the hull. The present
invention meets all of these demands.
Dual-Flow Propulsion System
[0070] FIGS. 11A and 11B show an alternative embodiment of the
invention that comprises a dual-flow propulsion system and method
where a multistage high pressure section 500 and a multistage low
pressure "booster" section 502, with a certain sacrifice in
weight-power or weight-volume density, cooperatively provide even a
greater range of efficient vessel operability, including efficient
operation at low to moderate vessel speeds. As indicated herein, a
single-flow multistage propulsor having only a high pressure
section provides greater efficiencies, typically 20-30% greater,
over prior art propulsors at higher speeds, e.g., above fifty to
sixty knots. In combination with a low pressure booster section,
however, similar efficiencies and improved performance can also be
achieved at lower speeds, e.g., between twenty and fifty knots.
[0071] Booster section 502 enables higher flow rates to attain a
higher propulsive achieve efficiency and greater acceleration with
minimal impeller cavitation. In a typical installation, both
sections of the dual-flow system are mounted inside the hull of a
vessel above to water line for easy access and maintenance. A
diffuser duct is typically used to channel water from the bottom of
the hull, and such a diffuser duct may be used to channel water to
the inlet of low pressure section 502. As vessel speed increases, a
series of nozzles disposed about a periphery 504 of the booster
section may be partially or completely closed, either gradually or
stepwise, using a conventional mechanism, in order to control water
flow towards the inlet guide vanes of high pressure section 500,
explained above. In one embodiment, the extent of nozzle opening
(or closure) may be control by rotating slip ring 515 clockwise or
counter-clockwise, as illustrated in FIG. 11B. Conventional
pressure sensors and servo controllers may also be employed to
control the nozzles 510 to maintain a constant pressure (e.g.,
equal to a pressure differential between two stages of the high
pressure section 500), the highest pressure possible, or other
regulated pressure at the input stage (inlet guide vane) of the
high pressure section 500 (depending on design parameters and
desired performance). Generally, nozzle opening (or closure) in the
booster section is controlled to minimize cavitation in the high
pressure section by controlling (e.g., maximizing) the first stage
inlet pressure, particularly when a high torque is applied to the
propulsor at a low or zero vessel speed, and/or to remove any
residual drag induced by the larger low pressure section at
moderate to higher vessel speeds, e.g., by discharging excess water
flow not taken by the high pressure section. To achieve the stated
goals, nozzle control may be effected according to speed or desired
mass flow of fluid in either of the respective high or low pressure
sections, rotational speed(s) of the drive shaft(s) in the
respective high and low pressure sections, relation between vessel
speed and mass flow within the propulsor, or in accordance with the
state of other parameters of the vessel or propulsor.
[0072] At lower vessel speeds, nozzles 510 remain open to provide
additional thrust around the periphery of the housing of high
pressure section 500. Total thrust produced by the embodiment of
FIG. 11A is the combination of the respective thrusts of the high
and low pressure sections 500 and 502. In the illustrated
embodiment, booster section 502 includes a set of inlet guide vanes
511 (only one shown) that is fixed internally of housing 512 to
control the angle of attack, e.g., swirl angle, of incoming fluid
as it impinges upon a first stage of the booster section. As shown,
booster section 502 includes two rotor sections 530, 531 (only one
rotor blade 520, 521 is shown in each section) and one peripheral
stator section (only one stator vane 523 is shown) fixed to the
inside of housing 512. The low pressure section is not limited to
the structure shown, but may have any number of stages, e.g., any
number rotor or stator sections. In certain cases, it may also be
desirable to provide variable pitch mechanism for booster section
stator vanes like that provided above for variable pitch stator
vanes of high pressure section 500. A series of peripherally
mounted struts 525 maintain the relative positioning of the high
and low pressure sections 500 and 502. Each section 500 or 502 is
preferably driven by a common drive shaft but may also be
separately driven by respective engines. If driven by a common
rotor shaft, an optional gearing mechanism may be interposed to
provide different rotational speeds for the high and low pressure
sections. Relative speeds of the high and low pressure section are,
in part, dictated by the blade-vane parameters including their
angle of attack relative to incoming fluid, such as depicted in
exemplary designs listed in Tables 1-6 set forth in the appendix.
Such design parameters, though, are not intended to limit the
invention. In the tables, "Ww" indicates water flow, "dp" means
differential pressure, "HP" means high pressure (unless associated
with "horse power"), "LP" means low pressure, and "so" indicates
speed. Radius ratio means the ratio of the blade/vane height to the
radius of the tip of rotation. Tip diameter ratio refers to the tip
diameter of the booster section to the tip diameter of the high
pressure section. Other terms are self-explanatory.
Split-Casing Construction
[0073] FIGS. 12A-12C show yet a further embodiment of the invention
that may be particularly useful for smaller power plants in the
range of 200 to 2000 horse power range. Conventional axial flow
pumps are generally designed for an engine having a given
horsepower or a fixed engine set point, and thus can only be
efficiently operated within a small range (e.g., plus or minus
5-10%) of power input. To accommodate different engines, the
manufacturer had to provide a separate pump for each engine size or
power, which became relatively costly in terms of tooling costs,
service, replacement parts, etc. According to this additional
aspect of the invention, multiple stator vane ensembles are
provided to adapt a given pump to wider power input ranges-much
like that provided by variable pitch stators and inlet guide vanes
described above. It is envisioned that a given multistage pump
design, for example, can be adapted to input power that varies by a
factor of two to three, more or less. Using a factor two, a
standard 400 horsepower pump may easily be reconfigured for a
larger engine up to 800 horse power, or reconfigured for a smaller
engine of at least 200 horsepower. Such flexibility aids the
manufacturer as well as a vessel owner desiring an engine
upgrade.
[0074] To implement such embodiment, FIG. 12A shows a split-casing
design of an exemplary pump having an inlet guide section and
mating top and bottom sections 604 and 606 that house the
multistage sections that include both rotor and stator sections.
FIG. 12B show further details of a mating section 604 or 606
(preferable but not necessarily identical) having channels or slots
608, 610, and 612 to receive top and bottom stator ring ensembles
620 and 622 shown in FIG. 12C. The split-casing design also enables
convenient access and maintenance of the pump at sea without the
necessity port maintenance.
[0075] FIG. 13A, from a view looking downstream of fluid flow,
shows a partial cut-a-way section of stator ring assembly mounted
in the interior wall of casing half 604. As shown, the stator ring
assembly comprises an outer shroud 621 and an inner shroud 622 that
having a series of fixed-pitch stator vanes 623-627 sandwiched
therebetween. The rotor section (not shown here) is fixed to a
shaft that rotates within the housing 604, 606. As a matter of
reference, vane 625 has a leading edge 625' and a trailing edge
625,'' and may be geometrically shaped like the blades or vanes
illustrated in FIGS. 9A-9D. The outer shroud 621 is geometrically
shaped to be inserted, by circumferentially sliding action, into
slots 608-612 (FIG. 12B) of casing 604. FIG. 13B shows a plan view
of ring ensemble 620 inserted in slot 610 of casing half 604. Also
shown are a series of bolt holes 631-634 used to fasten together
the split casings 604 and 606. FIG. 13C shows a stator vane 625,
for example, sandwiched between inner and outer shrouds 621 and 622
of the stator vane ensemble 620. When set into the ensemble, the
pitch angle of vane 625 is adjusted to configure the pump to
receive a given input power. The stator vanes may be affixed to the
inner and outer shrouds by a furnace brazing process. By adjusting
the pitch angle of the vanes in different sets of ring ensembles, a
given pump design may be adapted to handle multiple horsepower
settings within a given, relatively wide range defined by the
combination of the stator and impeller sections.
[0076] Tables in the appendix proposed design parameters of
exemplary propulsors having both a low-pressure booster section and
a high-pressure primary section.
Inlet Diffuser for Ram Pressure Recovery
[0077] FIGS. 14A, 14B, and 15 illustrate an inlet diffuser or duct
design that also takes advantage of the stator vane aspects of the
present invention. This aspect of the invention may be applied to
axial or mixed-flow pump jets, or to centrifugal pumps with proper
ducting. The purpose of a diffuser is to convert the maximum amount
of ram pressure resulting from vessel speed to hydrostatic pressure
as the fluid arrives at the pump jet inlet. In the case where inlet
ram pressure is taken directly from the upstream direction (rather
than from the bottom of a hull), inlet pressure attributable to
vessel speed in psi (pounds per square inch) equals (1/2 rho
V.sup.2)/(2g*144), where rho is the density of water (64.2
lbs/ft.sup.3), g is the gravity acceleration constant, and V is the
velocity of the vessel in feet per second. When the propulsor inlet
lies below the water's surface, additional static pressure is
added. At one hundred and two miles per hour, for example, ram
pressure equals 78.0 psi, which represents the initial pressure
available to the pump jet propulsion system.
[0078] A diffuser that might be adapted to an inlet duct might have
a diffuser ratio of 0.666 (i.e., ratio of inlet to outlet
cross-sectional area of the diffuser), which would recover 51.9 psi
of 78.0 psi of ram pressure. Such a diffuser having a maximum
allowable divergent angle between 6.degree. and 11.degree., i.e.,
the angle from a central axis of diverging walls of the diffuser,
may be unduly long in the axial direction, as indicated by the
length of duct 702 extending behind vessel 706 of FIG. 14A. The
diffuser housing may have a rectangular, square, or circular
cross-sectional shape in a downstream direction.
[0079] To explain further, FIG. 14A shows an inlet duct 702 mounted
near the transom 704 on a planing hull of vessel 706. In the
downstream direction, the cross-sectional area of duct 702
increases from an opening 708 to the pump intake 710. In an
exemplary construction, the cross-sectional area is ten square
inches looking into the opening 708, as shown in FIG. 14B, where
the opening 708 measures one inch in height and ten inches across.
In a planing hull of a high speed vessel, it is desirable to
provide an intake that is a rather narrow in the vertical direction
at the aft edge of the planing surface. In a displacement hull, the
intake may be positioned practically at any location below the
water surface. The initial cross-sectional area at the opening 708
expands to about one hundred square inches at the pump intake
710--accounting for an area consumed by the pump draft shaft 712.
Advantageously, this provides an inlet to outlet ratio of about
10:1.
[0080] According to further refined aspect of the present
invention, the diffuser duct may be staggered in multiple, stepped
stages, as shown in FIG. 15, in order to recover more of the ram
pressure of the incoming fluid in a shorter downstream direction.
In the illustrative ram recovery apparatus of FIG. 15, diffuser 720
has a ram inlet 722 having a first inlet area, and an outlet 724
having a significantly larger discharge area. Outlet 724 discharges
working fluid into a multistage propulsion device, such as the pump
described herein. In a stepwise manner, working fluid is
re-directed, e.g., approximately 30.degree. (more or less), and the
cross-sectional area in the diffuser is stepwise increased in the
downstream direction. A series of deflector vane groups 726, 727,
and 728 are provided to re-direct the working fluid at each section
of the diffuser. Although re-direction is illustrated in the upward
direction, it may occur in any direction. This aspect of the
invention, though, is not limited to the specific illustration or
redirection angle, but include all constructions or method that
involve re-direction of working fluid to shorten the
forward-to-aft-distance of the diffuser. Shortening such distance
is important when the diffuser inlet lies at the aft end of a
planing hull, as shown in FIG. 14A, to permit positioning of the
propulsor unit near the transom. In the case where incoming fluid
is taken from the bottom of hull, a stepped diffuser may still
provide benefits due to for-to-aft flow velocity of water, although
not to the same extent as a direct ram intake.
[0081] In the illustrated diffuser, each set of vane groups 726,
727, and 728 may recover as much as 1/cos .theta. of the ram
pressure (assuming .theta. is about 30.degree.) preceding the
respective group. The deflection angle, however, may range between
a few degrees to less than ninety degrees. In certain
circumstances, it may also be practicable to deflect flow up to
180.degree. provided the net gain in efficiency exceeds duct
losses. A diffuser having four sets of deflector groups, for
example, would yield a recovery ratio of 1.8:1, or a total ram
pressure recovery 0.77 instead of 0.666 using a longer diffuser
without deflectors. Additional diffuser sections having additional
stator vane groups may be incorporated to further increase ram
recovery of a multi-section diffuser. Similar to the fixed stator
vanes of the multistage propulsor described herein, the vanes of
deflector groups 726, 727, and 728 are designed to increase the
cross-sectional area of fluid flow in the downstream direction in
order to convert ram pressure to static pressure. In one
embodiment, this is accomplished by providing deflector vanes
having an airfoil shape, like that illustrated by an airfoil of
FIG. 10. The individual vanes in groups 726-728, however, need not
have a twist angle (like those of blades illustrated in FIGS.
9A-9C) since they are not mounted for rotation on a shaft but,
instead, may be specifically configured according to diffuser duct
geometry. If, according to the teachings herein, there is provided
a deflector vane and duct design that yields a 90% recovery ratio,
ram recovery amounts to 70.2 psi (pounds per square inch) instead
of 51.9 psi of a conventional design having 66.6% ram recovery,
thereby yielding an additional 18.3 psi in static pressure prior to
fluid entry into the propulsor pump. This additional pressure may
be used to lessen the amount of horsepower required to maintain a
given vessel speed, or to provide increased vessel speed using the
same horse power. In the example given herein, a 90% ram recovery
equates to an additional 6.7% in fuel savings over and above that
provided by the multistage propulsor design and/or the dual flow
propulsor design, described above (less frictional losses induced
by the deflector vanes and stepped diffuser). Such savings become
extremely desirable in large, high-speed ocean vessels powered by
100,000 horsepower or more, such as modern day fast ferries and
cargo ships.
[0082] Embodiments of the invention may be used to propel
displacement or hydroplaning hulls, or in hydrofoil or submarine
applications. The invention may also be deployed in water or fluid
pumping applications to pump the greatest amount of water or other
fluid at the highest pressure for a given horsepower input, or to
control the amount of water or other fluid delivered by a pumping
station. Thus, the invention embraces all such modifications and
adaptations that may come to those skilled in the art in view of
the teachings herein.
APPENDIX
[0083] TABLE-US-00001 TABLE 1 Density Sea Water 84.2 Radius ratio
LP 0.5 Density Fresh Water 62.4 Radius ratio HP 0.866 One HP (Ft
Lbs/Sec) 550 Tip Dia Ratiio 1.76 Tip Dia Booster/Tip Dia Hi
Pressure g Accel Of Gravity 32.17 Tip Dia HP 9 inches LP Inlet flow
coeff 0.98 Tip Dia LP 15.75 inches HP Inlet flow coeff 0.99 HP
blade height 1.503 LP Noz dp psi @ 40 LP blade height 3.9375 RPM@
2500 HP inlet area 19.47399 LP Pitch Dia 11.8125 inches LP inlet
area 85.23738 LP Pitch Sp (in/sec) 1548.256 Num stgs LP 2 Ww
(Total) 1100865 in cu/sec Num stgs HP 4 Ww (Total) Lbs/sec 40900.21
Ram recovery Coeff 0.8 Ww net 40082.2 Eff HP 0.9 HP Pitch Dia 7.497
Eff LP 0.92 HP Pitch Speed 981.3573 HP req'd LP 6064.054 (horse
power) Ww(HP) Total 101309.5 in cu/sec HP req'd HP 3126.958 (horse
power) Ww(HP) Lbs/sec 3726.292 net Total HP req'd 9191-012 (horse
power) LP Flow (net) 37173.91 lbs/sec TOTAL THRUST 107.474 LBS LP
Noz Vel 75.97737 ft/sec LP Thrust 87785.34 Lbs Hp Noz dp 200 psi Hp
Noz vel 169.8906 ft/sec Hp Thrust 19678.64 Lbs
[0084] TABLE-US-00002 TABLE 2 Data Density Sea Water 64.2 Radius
ratio LP@ 0.5 Density Fresh Water 62.4 Radius ratio HP @ 0.666 One
HP (Ft Lbs/sec) 550 Tip Dia Ratio @ 1.5 g Accel of Gravity 32.17
Tip Dia Hi Pressure 9 inches LP Inlet flow coeff 0.98 Tip Dia LP
13.5 inches HP Inlet flow coeff 0.99 HP blade height 1.503 LP Noz
dp psi @ 40 LP blade height 3.375 RPM @ 2500 HP inlet area 19.47399
LP Pitch Dia 10.125 inches LP inlet area 62.62338 LP Pitch Sp
(in/sec) 1325.363 Num stgs LP 2 Ww (Total) 594219.7 in cu/sec Num
stgs HP 4 Ww (Total) Lbs/sec 22076.91 Ram recovery Coeff 0.8 Ww net
21635.37 Eff HP 0.9 HP Pitch Dia 7.497 Eff LP 0.92 HP Pitch Speed
981.3573 HP req'd LP 2993.474 (horse power) Ww(HP) Total 101309.5
in cu/sec HP req'd HP 3126.958 (horse power) Ww(HP) Lbs/sec
3726.292 net Total HP req'd 6120.432 (horse power) LP Flow (net)
18350.62 lbs/sec TOTAL THRUST 63018.14 LBS LP Noz Vel 75.97737
ft/sec LP Thrust 43339.51 Lbs Hp Noz dp 200 psi Hp Noz vel 169.8906
ft/sec Hp Thrust 19678.64 Lbs
[0085] TABLE-US-00003 TABLE 3 Data Density Sea Water Radius ratio
LP@ 0.5 Density Fresh Water Radius ratio HP @ 0.666 One HP (Ft
Lbs/Sec) Tip Dia Ratio @ 1.76 g Accel of Gravity Tip Dia Hi
Pressure 9 inches LP Inlet flow coeff Tip Dia LP 15.75 inches HP
Inlet flow coeff HP blade height 1.503 LP Noz dp psi @ LP blade
height 3.9375 RPM@ 1500 HP inlet area 19.47399 LP Pitch Dia 11.8125
inches LP inlet area 85.23738 LP Pitch Sp (In/sec) 927.7538 Num
stgs LP 2 Ww (Total) 660519.2 in cu/sec Num stgs HP 4 Ww (Total)
lbs/sec 24540.12 Ram recovery Coeff 0.8 Ww net 24049.32 Eff HP 0.9
HP Pitch Dia 7.497 Eff LP 0.92 HP Pitch Speed 588.8144 HP req'd LP
3638.432 (horse power) Ww(HP) Total 60785.73 HP req'd HP 1876.175
(horse power) Ww{HP) Lbs/sec 2235.775 net Total HP req'd 5514.807
(horse power) LP Flow (net) 22304.35 lbs/sec TOTAL THRUST 64484.39
LBS LP Noz Vel 75.97737 ft/sec LP Thrust 52677.21 Lbs Hp Noz dp 200
psi Hp Noz vel 169.8906 ft/sec Hp Thrust 11807.18 Lbs
[0086] TABLE-US-00004 TABLE 4 Data Density Sea Water 64.2 Radius
ratio LP@ 0.66 Density Fresh Water 62.4 Radius ratio HP @ 0.666 One
HP (Ft Lbs/Sec) 550 Tip Dia Ratio @ 1.75 g Accel of Gravity 32.17
Tip Dia Hi Pressure 9 inches LP Inlet flow coeff 0.98 Tip Dia HP
15.75 inches HP Inlet flow coeff 0.99 HP blade height 1.503 LP Noz
dp psi @ 40 LP blade height 2.63025 RPM @ 1500 HP inlet area
19.47399 LP Pitch Dia 13.11975 inches LP inlet area 59.63908 LP
Pitch Sp (in/sec) 1030.425 Num Stgs LP 2 Ww (Total) 570103.6 in
cu/sec Num Stgs HP 4 Ww (Total) Lbs/sec 21180.93 Ram recovery Coeff
0.8 Ww net 20757.31 Eff HP 0.9 HP Pitch Dia 7.497 Eff LP 0.92 HP
Pitch Speed 588.8144 HP req'd LP 3090.459 (horse power) Ww(HP}
Total 60785.73 in cu/sec HP req'd HP 1876.175 (horse power) Ww{HP)
Lbs/sec 2235.775 net Total HP req'd 4966.634 (horse power) LP Flow
(net) 18945.16 lbs/sec TOTAL THRUST 56550.84 LBS LP Noz Vel
75.97737 ft/sec LP Thrust 44743.65 lbs Hp Noz dp 200 psi Hp Noz vel
169.8906 ft/sec Hp Thrust 11807.18 lbs
[0087] TABLE-US-00005 TABLE 5 Data Density Sea Water 64.2 Radius
ratio LP@ 0.5 Density Fresh Water 62.4 Radius ratio HP @ 0.666 One
HP (Ft Lbs/Sec) 550 Tip Dia Ratio @ 1.75 g Accel of Gravity 32.17
Tip Dia Hi Pressure 9 inches LP Inlet flow coeff 0.98 Tip Dia HP
15.75 inches HP Inlet flow coeff 0.99 HP blade height 1.503 LP Noz
dp psi @ 50 LP blade height 3.9375 RPM@ 2500 HP inlet area 19.47399
LP Pitch Dia 11.8125 inches LP Inlet area 85.23738 LP Pitch Sp
(in/sec) 1546.256 Num stgs LP 2 Ww (Total) 1100865 in cu/sec Num
stgs HP 4 WW (Total) Lbs/sec, 40900.21 Ram recovery Coeff 0.8 Ww
net 40082.2 Eff HP 0.9 HP Pitch Dia 7.497 Eff LP 0.92 HP Pitch
Speed 981.3573 HP req'd LP 7580.067 (horse power) Ww(HP) Total
101309.5 in cu/sec HP req'd HP 3908.698 (horse power) Ww(HP)
Lbs/sec LP 3726.292 net Total HP req'd 11488.76 (horse power) Flow
(net) 37173.91 lbs/sec TOTAL THRUST 120159.6 lbs LP Noz Vel
84.94528 ft/sec LP Thrust 98158.18 Lbs Hp Noz dp 250 Hp Noz vel
189.9434 ft/sec Hp Thrust 22001.39 Lbs
[0088] TABLE-US-00006 TABLE 6 Data Density Sea Water 84.2 Radius
Ratio Booster r 0.5 Density Fresh Water 62.4 Radius Ratio Hi Press
0.666 One HP (Ft Lbs/sec) 550 Tip Dia Ratio (TDR) 1.5 Tip Dia
Booster/Tip Dia Hi Pressure g Accel of Gravity 32.17 Tip Dia Hi
Pressure 9 inches LP Inlet flow coeff 0.98 Tip Dia Booster 13.5
inches HP Inlet flow coeff 0.99 HP blade height 1.503 LP Noz dp psi
40 LP blade height 3.375 RPM 2500 HP inlet area 19.47399 LP Pitch
Dia 10.125 inches LP inlet area 62.62338 LP Pitch Sp (in/sec)
1325.363 Num stgs LP 2 Ww (Total) 594219.7 in cu/sec Num stgs HP 4
Ww (total) lbs/sec 22076.91 Ram recovery Coeff 0.8 Ww net 21635.37
Eff HP 0.9 HP Pitch Dia 7.497 Eff LP 0.92 HP Pitch Speed 981.3673
HP req'd LP 2993.474 (horse power) Ww(HP) Total 101309.5 in cu/sec
HP req'd HP 3126.958 (horse power) Ww(HP) Lbs/sec 3726.292 net
TOTAL THRUST 63018.14 lbs LP Flow (net) 18350.62 lbs/sec LP Noz Vel
75.97737 ft/sec LP Thrust 43339.51 lbs/sec Hp Noz dp 200 Hp Noz vel
169.8908 ft/sec Hp Thrust 19678.64 lbs/sec
* * * * *