Note: Descriptions are shown in the official language in which they were submitted.
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VARIABLE MARINE JET PROPULSION
Technical Field
This invention relates to Marine Jet Propulsion Systems, and more particularly
to such systems of an improved
design, which are more efficient over a range of vessel speeds and loads.
Background Art
A marine jet propulsion system includes an inlet duct, a pumping means and a
nozzle. The inlet duct delivers water
from under the hull to the pumping means, which is driven by an engine. The
pumping means delivers the water
through the nozzle, which produces a water jet, thereby propelling the
watercraft through the body of water in
which the watercraft moves. In the prior art, a reversing bucket redirects the
jet flow back under the boat fully for
reverse thrust and partially for neutral thrust.
My US Patents 5,658,306, 5,679,035, and 5,683,276, which are incorporated by
reference, disclose systems and
methods for simultaneously optimizing the hydraulic efficiency of the inlet
duct and the pumping means. Such
increased hydraulic e~ciency has allowed a substantial increase in the design
system flow rate, which is well
understood in the propulsion field of art to improve propulsion efficiency at
low watercraft speeds. The increased
hydraulic efficiency of the system and the methods preserves propulsion
efficiency at higher watercraft speeds, so
that the systems operate more efficiently over a wide range of boat speeds and
accelerations.
From disclosures in my US Patents and through common knowledge in the
propulsion field of art, it is known that
larger mass flow rates and concomitantly lower nozzle velocities are more
efficient at lower watercraft speeds,
whereas lower mass flow rates and concomitantly higher nozzle velocities are
more efficient at higher watercraft
speeds. To achieve these ends, it is well understood in the art that a larger
nozzle area is useful at low watercraft
speeds, whereas a smaller nozzle area is most useful at higher watercraft
speeds. Such reduction of nozzle size with
watercraft speed was a natural consequence of the operation of the systems and
the methods disclosed in my US
Patents. However, a greater reduction of nozzle size with watercraft speed
would be desirable for increased
propulsion efficiency over a range of watercraft speeds.
When the watercraft is operating in a planing mode, the water jet obliquely
strikes the water surface behind the
watercraft, which results in turbulence on the water surface. Such turbulence
is dependent on the velocity of the
water jet relative to the water surface. When the velocity of the water jet
relative to the water surface is high, as is
common in the prior art, the water jet interacts with the water surface to
produce a high turbulent spray of water
behind the boat, which is commonly called a "rooster tail." The rooster tail
is commonly considered objectionable
for water skiing and wakeboarding behind the watercraft. Reducing the velocity
of the water jet relative to the
water surface eliminates the rooster tail, but still leaves a turbulent trail
of surface water in the wake of the
watercraft, which is still objectionable to wake boarders, who like to use
short ropes. A further reduction of the
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velocity of the water jet relative to the water surface would be desirable for
the further reduction of the turbulent
trail of surface water in the wake of the watercraft.
40 Another shortcoming of the prior art is the fact that the engine commonly
operates at substantially higher rpm than
would be most efficient, which results in greater fuel consumption, greater
engine wear, and more noise than would
result from operation at the engine's most efficient rpm. The operation of
such systems in the prior art has been
made more efficient by incorporating a two-speed transmission, but at higher
cost, weight and axial length.
45 Many marine jet propulsion systems of the prior art feature a direct
connection between the pump and the engine to
eliminate the cost and axial length of a transmission or clutch. In these
designs of the prior art, the neutral position
that could be provided by the transmission or clutch is approximated by
partially reversing the flow from the jet.
The operator cannot easily maintain the balance of this partial reversing,
especially given the sudden surge when
starting the engine, so that the watercraft moves unpredictably. A true
neutral control position would be desirable
50 to enhance the operator's control of the watercraft.
Trash management is another shortcoming of the marine propulsion systems of
the prior art. Many types of floating
debris can become lodged on the grate that covers the inlet of the system,
which restricts the flow of water into the
pump and reduces propulsion efficiency. There are three types of such debris:
solid objects, like rocks; fibrous
55 material, like rope, fishing line, grass, reeds, and the stems of aquatic
plants; and sheet material that can blind large
sections of the grate, like large kelp leaves and plastic bags. The fibrous
material is also well known to lodge on the
leading edges of pump and stator vanes, reducing pump efficiency. The rope is
particularly difficult to disentangle,
when it becomes wrapped around the impeller and the drive shaft. Some jet
boats carry hand rakes with right angle
bends in the handle to remove debris from the inlet grate, and some integrate
moveable grate sections to remove
60 such debris, but these methods are awkward and only partially effective.
Some commercial water jet propulsion
systems are equipped with a reversing transmission, which is used to back
flush both the pump vanes and the gate.
As a last resort, commercial systems and river boats are commonly equipped
with a clean-out hatch, which can be
removed to allow the operator to remove debris from the pump inlet by hand. It
would be desirable to reduce or
eliminate the need for the trash handling mechanisms and methods by providing
trash handling and back flushing
65 methods integral to the design of the marine jet propulsion system.
In the marine jet propulsion systems of the prior art, reverse thrust is
achieved by redirecting the water jet back
under the boat along hydraulic reaction surfaces. Such reaction surfaces are
commonly carried on a structure
known as a "bucket", which is mechanically moved into the jet stream by the
operator to get reverse thrust.
70 Buckets for large jets take up considerable space and add weight and cost
to the system. It would be desirable to
eliminate the need for the bucket by incorporating a method of producing
reverse thrust in the pump design.
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Disclosure of the Invention
Accordingly, it is an object of the invention to provide an improved marine
jet propulsion system, which combines
75 a variable pitch pump impeller and a variable nozzle under microcontroller
controls to create a continuously
variable power transmission, so that the engine is always operating close to
its most efficient rpm.
It is a further object of the invention to use full pitch on the variable
pitch pump impeller and maximum nozzle area
on the variable nozzle at low speeds, which both increases propulsion
efficiency and reduces the turbulent trail of
80 surface water in the wake of the watercraft.
It is a further object of the invention to reduce the variable-pitch impeller
pitch and the variable nozzle area with
increasing watercraft speeds, so that both the impeller pitch and the nozzle
area are minimum at the top boat speed,
which is well understood in the art to increase propulsion efficiency.
It is a further object of the invention to maintain the variable pitch
impeller pump close to its most efficient
operating conditions over both a wide range of shaft rpm and a wide range of
watercraft speeds, while
simultaneously achieving the objects and advantages stated above.
It is a further object of the invention to achieve these objects and
advantages in combination with a variable inlet
duct, that efficiently converts excess velocity at the duct entrance into
pressure at the pump inlet, as described in my
US Patent 5,683,276.
It is a further object of the invention to incorporate a novel pump design,
which allows the variable pitch to be
reduced to near zero, which results in no effective pumping action, which is
effectively a true neutral power
transmission.
It is a further object of the invention to provide a method of further varying
the pitch of the variable pitch impeller
pump to create a reverse pumping action, which provides grate and vane
cleaning by back flushing the system.
100
It is a further object of the invention to provide a vane design method, which
results in close tolerances between the
leading edges and the trailing edges of the vanes as they rotate through zero
pitch, which results in a scissoring
action between the leading edges and the trailing edges of the vanes, which
effectively cleans the leading edges of
the vanes. The scissoring action will also be seen to be effective in cutting
rope and fishing line that may be sucked
105 into the system, and it can be used to effectively chop up larger pieces
of debris in the pump inlet into smaller
pieces, which can escape through the grating or the nozzle.
It is a further object of the invention to provide for the further variation
of the variable pitch vanes to produce a
reverse pumping action through the system, which becomes an effective reverse
thrust when controlled in concert
110 with the variable inlet and the variable nozzle, thereby eliminating the
need for the reversing bucket.
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It is a further object of the invention to utilize the same nozzle vanes for
reverse steering as are used for forward
steering and nozzle flow regulation.
SUM1VIARY
115
These and other objects are met by providing an improved marine jet propulsion
system, which combines a novel
variable pitch spherical pump impeller and a variable steering nozzle to
create a continuously variable power
transmission, so that the engine is always operating close to its most
efficient rpm. Reducing the pitch on the
variable pitch spherical pump to near zero provides a neutral power
transmission. Further reducing the pitch results
120 in a scissoring action between the pump vanes, which cleans debris off the
leading edges of the vanes. Further
reducing the pitch results in reverse pitch and in reversing the pump flow,
which back flushes the system for trash
removal. Further reducing the pitch results in a reverse pumping action, which
is an effective reverse thrust,
particularly when used in concert with the variable steering nozzle and in
concert with the variable inlet duct, which
can act as a reverse nozzle. The swim platform and power trim function, which
are both common on recreational
125 boats of the prior art, can be used to reduce vortex formation and
cavitation in the reverse thrust mode.
The variable pitch spherical pump incorporates concentric spherical surfaces
on the impeller hub and on the pump
housing. The axes of rotation of the variable pitch impeller vanes are radii
of the concentric spherical surfaces, and
the inner and outer edges of the variable pitch impeller vanes are also
spherical surfaces, which fit closely to the
130 spherical surfaces of the impeller hub and of the pump housing,
respectively. This geometry allows the variable
pitch impeller vanes to rotate about the axes of rotation, while constantly
maintaining close fits between the inner
and outer edges of the vanes and the impeller hub and the pump housing,
respectively. The close fits are well
known in the pump design field of art to contribute to efficient pump
operation. In particular, this geometry allows
the vanes to rotate to near zero pitch required for effectively neutral power
transmission, while providing close fits
135 at the full pitch required in any application. It also allows the vanes to
rotate fully into reverse pitch, while
maintaining the close fits, which is well understood to result in a reverse
pumping action, which is useful for back
flushing trash and for providing reverse thrust.
In the forward thrust mode of operation, the variable nozzle is controlled to
maintain the most efficient head on the
140 variable pitch impeller pump for the current shaft rpm, as is described in
my US Patent 5,679,035. It is well
understood in the art that the most efficient head on the variable pitch
impeller pump is largely dependent on the
square of the shaft rpm. It is also well understood in the art that the most
efficient head on the variable pitch
impeller pump is only very slightly dependent on impeller pitch. Hence, the
pump will always be operating close to
peak efficiency, when the variable nozzle is controlled to maintain pump head
as a function of square of the shaft
145 rpm.
It is well understood in the art that efficiency is nearly constant over a
broad range of impeller pitch. The resulting
flow through the pump is well understood to be a function of the impeller
pitch. The shaft power demand of the
pump is well understood to be directly dependent on the product of pump head
and flow, when e~ciency is
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150 constant. From this, it is clear that varying the impeller pitch varies
the pump shaft power demand. It is further
clear that this variation of power demand occurs without significant loss of
efficiency, when most efficient pump
head is simultaneously maintained by varying nozzle area. It will also be
clear to those schooled in the art that
knowledge of instantaneous pump head and shaft rpm can be used to compute the
system flow by means of the
pump affinity constants, and hence the shaft power demand of the pump. It will
also be clear to those schooled in
155 the art that knowledge of actual system flow can be compared to the flow
indicated by pump head and shaft rpm to
monitor the efficiency of the pump operation, which can be used to alert the
operator of pump inefficiency, which is
probably due to debris on the inlet grate or on the pump vanes.
A microcontroller incorporates inputs from differential pressure transducers
to determine the head on the pump and
160 the flow through the system. The microcontroller gets an rpm input from an
engine tachometer. The control
program in the microcontroller incorporates a look-up table of the pump
efficiency as a function of shaft rpm. From
these inputs the control program determines the shaft power demand of the
pump. The control program also
incorporates a look-up table, which allows interpolation of the most efficient
power supplied at each shaft speed by
the engine, as is well understood in the art of industrial controller
programming. The control program compares the
165 calculated pump power demand to the power most efficiently supplied by the
engine at the input rpm, and adjusts
the pitch on the variable pitch impeller to adjust pump shaft power demand to
approximate the most efficient power
supply of the motor at the input rpm. Simultaneously, the variable steering
nozzle is adjusted to maintain the pump
at its most efficient operating head for the shaft rpm.
170 In an alternate embodiment, the pitch on the variable pitch impeller is
controlled by reference only to the throttle
position on the engine. The efficient power supplied by the motor is largely
dependent on the throttle position, and
the pump power demand is largely dependent on impeller pitch, so linking the
impeller pitch to the throttle position
approximately maintains e~cient engine operation. Simultaneously, the variable
steering nozzle is adjusted to
maintain the pump at its most efficient operating head for the shaft rpm.
175
In another alternate embodiment, the pitch on the variable pitch impeller is
adjusted based on an engine loading
output from a combustion microcontroller on the engine. It is well understood
that such combustion
microcontrollers commonly use a variety of sensors on the engine to control
fuel injection, ignition timing and
electric servo valve timing. Such combustion microcontrollers also commonly
output engine-loading signals to
180 automobile transmission microcontrollers, which incorporate engine
conditions into their shift point control
calculations. By these means, variations in elevation, humidity, fuel quality,
and other engine operating parameters
are incorporated in the most efficient shift point control decisions, so that
the engine operates most efficiently.
Similarly, this alternate embodiment adjusts the impeller pitch to operate the
engine most efficiently.
Simultaneously, the variable steering nozzle is adjusted to maintain the pump
at its most efficient operating head for
185 the shaft rpm. It will be clear from the following disclosure that several
fortunate consequences result from this
pump and nozzle design and from these control methods.
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When the watercraft is at the dock, the operator can manually control the
pitch on the variable pitch impeller to be
effectively zero, so that no pumping action results from the rotation of the
variable pitch impeller. This is a true
190 neutral position for starting the engine and for sitting at rest in the
water. The operator can also reverse the pitch to
clean the vanes and to back flush the system. By increasing the pitch, the
operator increases the flow through the jet
in a controllable way, either in forward or reverse, eliminating any starting
jerks or uncontrollable movement of the
watercraft. The same steering wheel or other steering control method is
effective in steering the boat in either
forward or reverse. When the operator has set the impeller at full forward
pitch and increases the engine rpm, the
195 microcontroller maintains efficient operation, as described above.
At low speeds, the power demanded to propel the boat at constant speed is low.
To match the power demanded by
the pump to the most efficient rpm of the engine, the microcontroller sets the
pump impeller pitch near maximum.
To maintain the pump close to its most efficient operating conditions, the
microcontroller opens the variable
200 steering nozzle to maximum. In addition to maintaining engine efficiency,
this control strategy has the fortunate
consequence of providing maximum flow at low speeds for maximum propulsion
efficiency. The flow through the
maximum nozzle opening also occurs at the lowest possible velocity. Thus,
motor efficiency, pump efficiency, and
flow rate efficiency are all close to optimum, and wake turbulence is
minimized.
205 When the system is under full acceleration, as in pulling up a water
skier, the control system will reduce the pump
impeller pitch to match the pump's shaft power demand to the engine's most
efficient power supply at the
instantaneous shaft rpm. The control system will also reduce the nozzle area
to maintain the most efficient head on
the pump for its current rpm.
210 When the boat reaches steady wakeboarding speed in the approximate range
of 15 to 20 mph, the impeller is close
to full pitch to reduce the engine rpm to the most efficient operating point.
The variable nozzle is close to being
fully open to maintain the most efficient pump head at the relatively low
shaft rpm. A further advantage is that the
variable inlet duct opening is near maximum due to the high flow, which
results in no losses from the conversion of
inlet entrance velocity to pressure at the pump inlet. This again has the
fortunate consequence of providing close to
215 maximum flow at this relatively low boat speed for maximum propulsion
efficiency, which also results in minimum
nozzle velocity through the large nozzle area and consequently in minimum wake
turbulence. The system rpm is
further reduced relative to systems of the prior art by this higher propulsion
efficiency, which requires less shaft
power and consequently lower shaft rpm to maintain the boat speed. Thus, motor
efficiency, pump efficiency, and
flow rate efficiency are all close to optimum, and wake turbulence is
minimized.
220
When the boat reaches steady water skiing speed at approximately 30 mph, the
recovery of pressure in the inlet duct
has increased, which will cause a slight reduction in nozzle area to maintain
the most efficient system flow and head
on the pump. The power required to maintain this higher boat speed is also
higher, so the engine must operate at a
higher rpm to supply the necessary power. The most efficient pump head rises
as the square of the shaft rpm.
225 Higher engine rpm causes the control system to reduce the impeller pitch,
which reduces the most efficient pump
flow. The nozzle area control function implicitly accounts for higher inlet
head at this boat speed, higher pump
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head at the higher shaft rpm, and the reduced flow resulting from reduced
impeller pitch. As a result of all these
factors, the nozzle area is reduced and the nozzle velocity relative to the
boat is increased. However, the nozzle
velocity relative to the water surface is reduced by the increased boat speed,
so that the velocity of the jet relative to
230 the water surface has only slightly increased. Wake turbulence is thereby
only slightly increased, and the use of
longer towropes at this higher boat speed makes wake turbulence less critical,
since it has more time to dissipate
before the skier reaches it.
Further increases in boat speed demand increased engine power, which the
engine can only supply at higher rpm.
235 The control system reduces impeller pitch to allow the engine higher rpm.
Reduced impeller pitch requires a
commensurate reduction in nozzle area. Pump head is rising as the square of
the rpm. Inlet head is rising as the
square of the boat speed. The increasing pump rpm, the reducing pitch, and the
higher inlet pressure are all factors,
which will result in the control system's reducing the nozzle area to maintain
peak pump efficiency. Hence, nozzle
area is reduced with increasing rapidity as boat speed increases as a natural
consequence of the system operation,
240 until minimum nozzle area is reached at the top design speed of the
system. The minimum nozzle area at top speed
is also ideal for reducing the system flow rate, hence improving propulsion
efficiency at the higher speed.
Brief Description of Drawings
FIG. 1 is a plan view of the bottom of a boat, which incorporates an Improved
Marine Jet
245 Propulsion System, showing the hull, inlet duct, pump housing, variable
nozzle, and the swim platform.
FIG. 2 is a midline vertical section view indicated on FIG. 1, showing the
internal details of the
improved marine jet propulsion system and the control system schematic.
FIG. 3 is an enlarged view of the area indicated on FIG. 2, which shows the
details of the
hydraulic control piston for the vane pitch.
250 FIG. 4 is an enlarged view of the area indicated on FIG. 2, which shows
the details of the impeller
hub and vane pitch operating mechanism.
FIG. 5 is a section view indicated on FIG. 2 Showing the vanes in the inlet
duct and the sliding
gate beneath the vanes.
FIG. 6 is the section view indicated on FIG. 2 showing the variable vane
operating mechanism of
255 the pump.
FIG. 7 is a rear section view of the boat indicated on FIG. 2 showing the
variable rectangular
nozzle under the swim platform.
FIG. 8 is an schematic overhead view of the variable steering nozzle showing
the various vane
positions that result from the actions of the hydraulic nozzle controls.
260 FIG. 9 is a schematic representation of the nozzle hydraulic system, which
shows the integration
of the steering function, the nozzle area reduction function, and the nozzle
pitch function.
FIG. 10 is a section view indicated on FIG. 1 showing the power trim
adjustment of the
propulsion system and the maximum declination, which is used in reverse mode.
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FIG. 11 is a graph on which shaft power is plotted against shaft rpm, showing
the relationships
265 between pump power demand and efficient engine power supply.
FIG. 12 is a flow chart for the microcontroller program used to control the
variable pump vane
pitch, the variable nozzle area and the variable inlet entrance area in all
embodiments of the invention.
FIG. 13 is a flow chart for three alternative microcontroller programs used to
control the variable
impeller vane angle for efficient engine operation in the forward mode of
operation.
270
-Table of Reference Numerals:
19watercraft 84 nozzle vanes
20marine jet propulsion 85 integral nozzle
system vane shafts
21engine 88 nozzle guard
22tachometer 89 rectangular discharge
opening
23combustion control computer91 hydraulic ram
29body of water 92 vane operating arm
30inlet duct 93 hydraulic nozzle
cylinders
31adjustable inlet slide 94 ball-ended couplings
32inlet entrance opening 95 transom
33inlet tunnel 100 gear reduction
34inlet hydraulic cylinder101 driven gear
35clevis pin 102 hydraulic cylinder
assembly
36inlet cylinder shaft 103 bell housing
37leading edge of inlet 104 fasteners
slide
38entrance angle 105 end piece
39upper surface of inlet 107 hydraulic fluid
tunnel passage
40grate structure 108 square post
41rectangular passages 109 piston
42grate vanes 110 roller thrust bearing
43middle vane 111 bearing plate
44water flow 115 hydraulic pump
45inlet entrance flow 121 hydraulic steering
line
46slide rails 122 hydraulic steering
line
47slide rail fasteners 123 hydraulic helm
48grate structure fasteners124 steering wheel
49rear exit opening of 125 balancing cylinder
inlet duct
50spherical pump 126 driven cylinder
51drive shaft 127 nozzle closing circuit
52impeller 128 balancing connection
53impeller hub 129 balancing connection
54return spring 130 driving cylinder
55spider 133 flow control module
56operating arms 134 trim control valve
57impeller vanes 140 microcontroller
58bearing holes 141 single handle control
59locking pin 142 engine throttle
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61impeller hub cone bolts 145 head differential pressure
transducer
62split spherical pump housing146 nozzle pitot tube pressure
63circumferential fasteners 147 inlet pitot tube pressure
64fasteners 149 flow differential pressure
transducer
65push rod 150 pump inlet pressure
70diffuser 151 speed pressure transducer
71stator vanes 152 speedometer pressure
72diffuser hub 154 tachometer input
73tapered roller bearing 155 engine load signal
74tapered roller bearing 157 operator preference input
75pump shaft 158 impeller vane control
module
76bearing collar 159 inlet control module
77mechanical seal 160 pump power demand curve
78fasteners 161 engine power supply curve
79water seal 163 horizontal line
80variable rectangular steering164 horizontal line
nozzle
81top plate 165 pump power demand curve
82bottom plate
83wing walls
Best Mode for Carrying out the Invention
In the accompanying FIGS. 1-13, there is shown an improved marine jet
propulsion system, generally referred to as
275 20, designed to achieve higher propulsion efficiency, greater
maneuverability, and better injury prevention features
than currently available marine propulsion systems.
The system 20 includes a variable water inlet duct 30 for admitting water into
the system 20, a variable-pitch
spherical pump 50 capable of receiving and pumping a relatively large amount
of incoming water, and an
280 adjustable, large variable rectangular discharge nozzle 80 capable of
forcibly exiting the water pumped by the pump
40 to propel the watercraft 19 through the body of water 29. A microcontroller
120 controls the variable inlet duct
30, the variable pitch spherical pump 40 and the variable discharge nozzle 80.
By simultaneously controlling the
variable inlet duct 30, the variable-pitch spherical pump 50, the large
variable rectangular discharge nozzle 80, the
propulsion e~ciency of the system 20 is greatly improved over marine jet
propulsion systems of the prior art.
285
The inlet duct 30 is designed so that hydraulic efficiency of the system 20 is
optimally maintained at all watercraft
19 velocities, as described in my LTS Patents. In this embodiment, the
entrance area of the entrance opening 32 is
varied by the action of the hydraulic cylinder 34 on the adjustable slide 31
to match the velocity of the water in the
entrance opening 32 to the velocity of the water passing under the watercraft
19.
290
As shown in FIGS. 1-5, the inlet duct 30 includes an adjustable slide 31
located over the entrance opening 32 of the
hydraulically efficient, elongated inlet tunnel 33 formed or attached to the
bottom of the watercraft 19. The
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hydraulic cylinder 34 moves the adjustable slide 31 to vary the effective area
of the entrance opening 32. The inlet
tunnel 33 is longitudinally aligned on the watercraft 19 with a front entrance
opening 32 and a rear exit opening 49.
295 The inlet tunnel 33 gently curves upward inside the watercraft 19 and has
a larger cross-sectional area at its exit
opening 49 than at its entrance opening 32, when the adjustable slide 31 is in
its forward position as shown in FIG.
2. The surrounding surface of the entrance opening 32 of the inlet tunnel 33
is gently curved from tangent to bottom
of the watercraft 19 so that turbulence is minimal at the entrance opening 32
of the inlet duct 30.
300 The grate structure 40 fits in the elongated inlet tunnel 33 and attaches
to the watercraft 19 with fasteners 48, so that
the conversion of excess entrance velocity at the entrance opening 32 into
pressure at the rear exit opening 49 takes
place largely in the rectangular passages 41 between the grate vanes 42. It is
well understood in the art of hydraulic
design that dividing the flow into such rectangular flow channels 41 reduces
turbulence losses the water flow 44,
which are larger than the frictional losses against the vane surfaces. The
control system 100 moves the adjustable
305 slide 31 by means of the hydraulic cylinder 34 to adjust the size of the
entrance opening 32 so that the velocity of
the incoming water therethrough matches the velocity water under the
watercraft 19 in the body of water 29 in
which the watercraft 19 moves. By controlling the relative velocity of the
incoming water through the entrance
opening 32 and by using a hydraulically e~cient inlet tunnel 33, which
gradually increases in cross-sectional area
between its entrance opening 32 to its exit opening 49, the dynamic head of
the incoming water may be efficiently
310 recovered at the pump 50.
As shown in FIGS. 1, 2, and 5, the grate structure 32 includes a plurality of
spaced apart, longitudinally aligned
elongated grate vanes 42. The middle grate vane 43 is vertically truncated to
allow passage for the shaft 36 of the
hydraulic cylinder 34, which passes through the watercraft 19. The shaft 36 is
attached to the adjustable slide 31
315 with the clevis pin 35, so that the action of the cylinder 34 moves the
adjustable slide 31 in response to the
microcontroller 120. The adjustable slide 31 is held in place by the slide
rails 46, which are attached to the grate
structure 40 with the fasteners 47. The leading edge 37 of the adjustable
slide 31 bends downward so that the
effective entrance angle of the leading edge 37 is approximately parallel to
the upper inlet surface 39 of the inlet
tunnel 33, so that the velocity of entrance flow 45, which is parallel to the
upper surface 39 will approximately
320 match the entrance angle of the leading edge 37, which is well understood
in the art of hydraulic design to provide
efficient separation of the inlet flow from the water under the boat.
When the watercraft 19 is stationary or at low speed, water enters the inlet
entrance opening 32 via the suction
created by the pump 50. During this stage, the adjustable slide 31 is in its
rearmost position as shown by the ghost
325 line position 46 in FIG. 2., so that the entrance opening 32 is wide open
and the grate vanes 42 act as diffusers to
reduce entrance swirl. As the watercraft's speed increases, water enters the
entrance opening 32 by the forward
movement of the watercraft 19 through the body of water 29 and by the suction
of the pump 40. The
microcontroller control system 120 adjusts the position of the slide 31
through the cylinder 34 and shaft 36 so that
the velocity of the water entering the inlet opening 32 matches the velocity
of the water under the watercraft 19. At
330 the top design speed of the system 20, the slide 31 is in the forward
position shown in FIG. 2.
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As the velocity of the incoming water at the entrance opening 32 relative to
the velocity of the incoming water at
the exit opening 49 in the inlet tunnel 33 increases, the controller 120
progressively moves the slide 31 forward. It
can be seen that this has two effects --first, it reduces the effective area
of the entrance opening 32 of the inlet tunnel
335 33; and second, it increases the effective length of the inlet duct 30. It
can be seen that the changes both in cross-
sectional area and change in flow direction within the inlet tmmel 18 are
always gradual, which are design
requirements well known in the art for the efficient recovery of pressure head
in the turbines and venturi flow
meters. It can also be seen that the increasing effective length of the inlet
tunnel 33 with decreasing effective area of
the entrance opening 32 maintains a nearly constant rate of change in area
over the inlet tunnel's range of operation.
340 The total dynamic head of the incoming water can then be efficiently
recovered at the pump 50.
Disposed adjacent to the exit opening 49 of the inlet tunnel 33 is the pump
50, which is coupled via a drive shaft 51
and gear reduction 100 to an engine 21. In the embodiment shown, the pump 50
is contained in a split spherical
pump housing 62, which is attached to the grate structure 40 with the
fasteners 64. The pump 50 is axially aligned
345 with the inlet duct exit opening 49, so that the drive shaft 44 extends
forward there from and connects to the
transmission 100. In the embodiment shown, the pump 50 includes a spherical
impeller 52, which rotates to forcibly
deliver the incoming water from the exit opening 49 to the discharge nozzle 80
located on the opposite side of the
pump 50. In the preferred embodiment, the pump 50 is designed to be used with
a 300 horsepower engine so that
the mass flow equals approximately 2200 lbs/sec and the pump head is
approximately 70 feet at full power with 18-
350 degree discharge angle on the variable pump vanes. The pump 50 uses a 16-
inch spherical impeller 46, which
matches the size of the diffuser 70, which is disposed over the aft position
of the pump 50 to recover the vortex
velocity produced by the pump 50 as useful propulsive momentum, as is common
in the art of pump design. The
stator vanes 71 of the diffuser 70 support the diffuser hub 72, which contains
the tapered roller bearings 73 and 74.
355 In the assembly of the pump 70, the bearings 73 and 74 are first mounted
on the pump shaft 75, which is inserted
into the diffuser hub 72. The bearing collar 75 is bolted to the hub 72 to
carry the thrust of the bearing 73 and to
provide mounting surfaces for the mechanical seal 77. The spherical impeller
hub 53 is bolted to the pump shaft 75.
The return spring 54 and the spider 55, engaging the operating arms 56, are
held in place by a press, while the vanes
are inserted radially through the bearing holes 58 and the operating arms 56.
Each vane is rotated to align with the
360 pin holes in the operating arms 56 and the locking pins 59 are inserted to
lock the assembly together. When the
press is removed, the spider 55 is held in place by the operating arms 56,
which restrains the return spring 54. The
hub cone 60 is fitted over the spider 55 and drawn up against the impeller hub
53 by the progressive sequential
tightening of the bolts 61. This process compresses the spring 54 and results
in the impeller vanes 57 being held in
the full pitch position by the spring 54 against the hub cone 60.
365
The split spherical pump housing 62 is assembled around the impeller 52 and
pinned together circumferentially
with the fasteners 63. The diffuser70 is attached to the pump housing 62 with
the fasteners 78. The splined drive
shaft 51 is assembled into the internally splined gear 101 trapping the water
seal 79. Matching the internal spline in
the pump hub cone to the splined shaft 51, the assembled pump 50 and diffuser
70 slide onto the splined shaft 51
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370 and are attached to the grate structure 40 with the fasteners 64. Internal
to the splined shaft 51 is the pushrod 65,
which acts on the spider 55.
A vane adjustment means is connected to the pump impeller 52 for controlling
pitch of the pump vanes 57, and,
hence, the most~efficient flow rate of the pump 50. As shown in FIGS. 2, 3, 4
and 5, the vane adjustment means
375 includes the hydraulic cylinder assembly 102 internal to the driven gear
101. The hydraulic cylinder assembly 102
is solidly mounted to the bell housing 103 using the fasteners 104. The end
piece 105 of the hydraulic cylinder
assembly 102 incorporates a hydraulic fluid passage 107 and a square post 108,
which fits a square hole in the
piston 109 to prevent the rotation of the piston 109. The piston 109 acts on
the roller thrust bearing 110. On the
other side of the roller thrust bearing 110, the bearing plate 111 engages the
internal spline in the driven gear 101,
380 so that the bearing plate 111 rotates with the driven gear 101 and the
splined drive shaft 51.
It can be seen that when hydraulic fluid is forcibly introduced through the
fluid passage 107, the piston 109 is
driven against the bearing 110, which acts on the bearing plate 111 and the
push rod 65, which is driven through the
rotating drive shaft 51 to act on the spider 55 to compress the spring 54 in
the impeller hub 53 and move the
385 operating arms 56, which reduce the pitch of the impeller vanes 57.
Located aft position of the pump's diffuser 70 is the variable rectangular
steering nozzle 80. The nozzle 80 is
formed between a top plate 81 and a bottom plate 82, which are held parallel
by their attachment to the two wing
walls 83. The nozzle vanes 84 have integral shafts 85. The shafts 85 are born
by bearing holes in the top plate 81
390 and bottom plate 82. The nozzle vanes 84 are formed so that their top and
bottom edges fit closely to the top plate
81 and bottom plate 82, respectively. The axes of the vane shafts 85 are held
perpendicular to the plates 81 and 82,
so that the rotation of the shafts 85 results in the movement of the vanes 84
between the plates 81 and 82, while
maintaining close fits between the edges of the vanes 84 and the plates 81 and
82. As a result of this geometry,
there is formed a rectangular discharge opening 88, which is bounded by the
plates 81 and 82 and the vanes 84.
395
FIG. 8 shows top views of the nozzle 80, which shows how the angle of the
nozzle vanes 84 can be controlled both
to provide steering control and to reduce nozzle area. Each of the steering
vanes is positioned by a hydraulic ram
91, which operates on the respective vane operating arm 92. The hydraulic
nozzle cylinders 93 are mounted inside
the transom 95, so that only the rams 91 penetrate the transom 95. FIG. 8A
shows the nozzle vanes 84 in the wide-
400 open straight position. FIG. 8B shows the nozzle vanes 84 in the full low
speed turn position turn position. Figure
8C shows the nozzle vanes 84 in the high-speed flow reduction position.
FIG. 9 is a schematic of a hydraulic system for controlling the nozzle vanes
84 for steering, flow reduction, and
nozzle azimuth simultaneously. The azimuth movement is commonly used in
planing watercraft as a power trim to
405 adjust the planing angle of the boat, as is well understood in the art. As
will be seen in the discussion of FIG. 10
below, the adjustment of the nozzle discharge angle in the vertical plane is
also useful for reducing vortexing in the
reverse mode.
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The double acting steering cylinders 93 penetrate the transom 95 with ball-
ended fittings or rubber grommets, as is
410 common in the art, and are connected to the vane operating arms 92 with
ball-ended couplings 94, as is common in
the art. The hydraulic steering lines 121 and 122 are connected to a hydraulic
helm 123, which is driven by the
steering wheel 124, as is common in the art. The steering cylinders 93 are
series connected for reverse action, so
that the cylinders 93 move equal distances in opposite directions in response
to fluid delivered from the hydraulic
helm. This steering action can be seen to result in the common rotation of the
vane shafts 85, until the steering
415 vanes 84 reach the position shown in FIG. 8(B).
The balancing cylinder 125 of FIG. 9 is composed of three hydraulic cylinders
in tandem. The driven cylinder 126
is single ended, and is so constructed that the area of the piston is twice
the area of the shaft. The nozzle closing
circuit 127 is connected on the closed end of the driven cylinder 126, so that
the fluid displacement is proportionate
420 to the area piston. The balancing circuit 128 is connected on the shaft
side of the piston, so that the fluid
displacement in the circuit 128 is equal to half of the displacement of the
piston in the circuit 127 and opposite in
direction. The balancing circuit connection 129 is made to tandem cylinder in
the driven cylinder 130, so that the
displacement is in the balancing circuit 129 is also equal to half of the
displacement of the piston in the circuit 127
and opposite in direction. It can be seen that the result of this arrangement
is that the steering cylinders 93 move in
425 the same direction and by the same amount in response to the movement of
the common shaft of the tandem
cylinders 126 and 130. It can also be seen that no net fluid displacement
occurs in the hydraulic circuit 121 and 122
from the hydraulic steering helin 123 to the steering cylinders 93. Hence, the
displacement of the common shaft of
the tandem cylinders 126 and 127 has the effect of increasing and decreasing
the angle between the nozzle vanes
84, and this movement may cause the vanes 84 to reach the positions shown in
FIG. 8(C). It should also be noted
430 that the action of these driven cylinders 126 and 127 is independent of
the action of fluid flows from the steering
helm at 121 and 122 and may occur simultaneously, so that the system allows
simultaneous nozzle area control and
steering.
Referring ftwther to FIG. 9, the driving hydraulic cylinder 133 controls the
displacement of the driven cylinders 126
435 and 130. The driving cylinder 133 moves in response to two hydraulic power
sources. The flow control valve 134,
which is also shown in FIG.2, responds to commands from the microcontroller
150. It is a 4-way valve that
controls the motion of the driving cylinder 133, as is common in the art.
Through this means the microcontroller
150 acts to adjust the effective nozzle area of the rectangular discharge
opening 89 in order to maintain the efficient
operation of the spherical pump 50, as will be detailed below. The second
hydraulic power source is a trim control
440 valve 135, which is controlled by the operator to adjust the azimuth of
the nozzle 80 for power trim, as shown in
FIG. 10. The hydraulic circuit from the trim control valve 135 is connected in
series with the trim cylinder 136. It
can be seen that the effect of this circuit is to displace the trim cylinder
136 and the driving hydraulic cylinder 133
in the same direction. As a result of the motion of the driving cylinder 133,
the steering cylinders 93 are also
displaced in the same direction. This common motion can be seen to reduce the
effect of power trim adjustment on
445 the position of the nozzle vanes 84.
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FIGS. l0A and lOB are side elevation section views indicated on FIG. 1,
showing the action of the optional trim
cylinder 136 on the nozzle 80. The piston area of the trim cylinder 136 is so
chosen relative to the piston area of
the steering cylinders 93 and the driving cylinder 133 that the angular
displacement of the connection points 136
450 and 137 is approximately equal. As a result, the action of the trim
control valve 135 of FIG. 9 is to extend both the
trim cylinder ram and the steering cylinder rams by a proportion that
minimizes the effect of the trim movement on
the steering and nozzle area control functions. The extreme down position of
the trim range shown in FIG. l OB is
useful in increasing submergence of the nozzle 80, which acts as a water inlet
in the reverse mode. The nozzle
guard 88 serves both to prevent vortex cavitation and to prevent human limbs
and other objects from approaching
455 possible pinch points in the nozzle 80 mechanisms.
Propulsion system efficiency is the product of four efficiency components:
inlet duct, pump, nozzle, and engine.
The nozzle has relatively small losses, which can be ignored without
significant loss of system efficiency. The inlet
duct recovery efficiency is maximized independently by maintaining the duct
entrance velocity to approximate the
460 velocity of the water under the boat, as detailed in my US Patents. Pump
efficiency is maximized independently by
adjusting the nozzle area to maintain the most efficient head on the pump for
the current shaft rpm, as detailed in
my US Patents. In this disclosure engine efficiency is maintained by
incorporating a variable pitch spherical pump
50 in the propulsion system 20 design, which provides continuously variable
power demand to track the most
efficient power supply of the engine. It is worth noting that the head nozzle
control method and the inlet duct
465 control methods from my US Patents work well in concert with the variable
pitch pump. It is also worth noting that
this propulsion system 20 simultaneously maximizes all of the four e~ciency
components (inlet duct, pump,
nozzle, and engine) over a wide range of boat speeds and accelerations. As a
result, the design flow of the system
20 can be increased with a smaller efficiency penalty, which allows the use of
a higher mass flow rate for better
propulsion e~ciency, as is well understood in the art. The relevant principles
and their interrelation are discussed
470 in more detail below.
FIG. 2 shows the schematic diagram connections of the microcontroller 140. The
operator uses the single handle
control 141 to control both propulsion direction and the throttle 142 for the
engine 21, as is common in recreational
boats. The single handle control 141 incorporates a throttle dead band, so
that the throttle is set at idle from about
475 10 degrees forward of the straight-up or neutral position to about 10
degrees back of the neutral position. In the
prior art these forward 10 degree and reverse 10 degree travels operate a
gear, which shifts the transmission into
forward and reverse, respectively, and further travel of the handle out of the
throttle dead band increases the engine
throttle 142, as is well known in the art. In the present embodiment, the
single handle control 141 has the same
appearance and function to the operator, but the integral gear shifting
mechanism is omitted and replaced with a
480 shaft encoder 143, which provides the angular position of the single
handle control 141 to the microcontroller 140.
As will be more fully explained in the discussion of FIGS. 12 and 13 below,
the microcontroller 140 is programmed
to position the vane actuator piston 109 through the hydraulic control module
157, so that the vane angle follows
the position of the single handle control over the throttle dead band.
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485 Another input to the microcontroller 140 shown in FIG. 2 is the head
differential pressure transducer 145, which
provides the difference between the pitot tube pressure 146 after the pump 50
and the pitot tube pressure 147 at the
inlet of the pump 50. This difference is well understood in the hydraulic art
to be the commonly accepted measure
of the head h on the pump 50.
490 Another input to the microcontroller 140 shown in FIG. 2 is the flow
differential pressure transducer 149, which
provides the difference between the inlet pitot tube pressure 147 and the
inlet static pressure 150. For purposes of
the calculations discussed below it is well known that the differential
pressure on the transducer 149 is equal to the
flow velocity V squared divided by twice the acceleration of gravity g, or
V~2/2g. It is also well understood that
the volume flow rate Q is the product of the velocity V and the cross section
flow area (Q=VA) and that mass flow
495 rate q is the product of volume flow rate and the density of the fluid w,
so that q=Qw.
Another input to the microcontroller 140 shown in FIG. 2 is the speed pressure
transducer 151, which provides the
speedometer pitot tube pressure 152 from the boat speedometer pitot tube. For
purposes of the calculations
discussed below it is well known that this pressure is approximate to the
speed of the water craft divided by twice
500 the acceleration of gravity, so the discussion in the previous paragraph
also applies here.
Another input to the microcontroller 140 shown in FIG. 2 is the engine
tachometer 154. This tachometer input 154
is commonly a pulse train that is read with a timed counter integral to the
microcontroller, as is well known in the
art.
505
Another input to the microcontroller 140 shown in FIG. 2 is the engine load
signal 155, which is output by the
engine combustion microcontroller. This interface is well known in the
automotive art.
Another input to the microcontroller 140 shown in FIG. 2 is the operator
preference input 157, which is a variable
510 resistance or optical encoder to indicate the operator preference for
performance or economy operation.
The microcontroller 140 has several control outputs, through which it controls
the movement of the nozzle vanes
84, the pump impeller vanes 57, and the adjustable inlet slide 31. The
operation of the flow control module 133 has
been discussed in relation to FIG. 9 above. The balancing cylinder 125 of FIG.
9 has internal positional feedback to
515 the microcontroller 140, as is well known in the art. The inlet control
module 159 uses hydraulic power and
incorporates positional feedback. The vane hydraulic control module 158 also
uses hydraulic power and
incorporates positional feedback.
In the preferred embodiment, the program for microcontroller 140 is a PICmicro
D Microcontroller, which is
520 available from Microchip Technology. Programs for these devices are
developed using the Microchip's C
programming environment. This development system is capable of incorporating a
wide range of mathematical
functions in the control program. The following paragraphs provide background
on the functions to be incorporated
in the control program.
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525 The Basis of the Control Relationships
The relationships for controlling the inlet duct and nozzle are developed in
detail in my said US Patents, and will be
reviewed in the discussion of FIG. 12 below. The technical basis for
controlling the impeller vane 57 pitch to
maintain the engine 21 at its most efficient operation follows.
530
FIG. 11 is a graph of shaft power versus shaft rpm, showing the relationship
between pump shaft power demand
and a typical engine's most efficient power supply. In a typical water jet
propulsion system design, the gearing
between the pump and engine is chosen so that pump power demand curve 160
intersects the engine power supply
curve 161 at the highest allowable engine rpm, which is taken to be 5000 rpm
in FIG. 11. When the pump is
535 maintained at its most efficient head and flow, the pump power demand
curve 160 is approximately a cubic curve as
shown in FIG. 11, as is well known in the art of pump design, and particularly
in the area of pump affinity
relationships. The difference between the most efficient power supply curve
161 and the pump power demand
curve 160 is unfortunately greatest in the most frequent operating range,
which falls between the horizontal lines
163 and 164. Hence, the engine 21 is operating furthest from its most
efficient operating rpm most of the time.
540 When the curve 160 represents the variable pitch spherical pump 50 with
the vanes 57 at about 18 degrees beta-2, as
vane pitch is commonly designated in the pump art, the full-pitch power curve
165 represents the power demand
curve of the same pump 50 with the vanes 57 set at full pitch of about 40
degrees beta-2. It is well understood in
the art of pump design that this range of efficient operation is common to
variable pitch propeller pumps. The
spherical pump 50 has the additional efficiency advantage of having close fits
between the tips of the vanes 57 and
545 the housing 62, even at pitches greater than 40 degrees beta-2. It is
clear that the full-pitch pump power demand
curve 165 much more closely matches the engine's most efficient power supply
curve 161 in the most frequent
operating range between the horizontal lines 163 and 164. Such reduction of
engine rpm is widely used in the
automotive power transmission art to increase fuel efficiency and engine 21
life. At the bottom end of operating
range on the line 163, the engine rpm is reduced from about 3,000 to about
1,900. At the top end of the operating
550 range along the line 164, the engine rpm is reduced from about 4,000 to
about 2,600. There is a continuous range of
efficient pump 50 power demand curves between the curves 165 and 160, which
result from the continuous
variation in vane 57 pitch possible in the spherical pump 50. One of these
intermediate curves can be seen to be the
most efficient for each possible engine rpm between 3,000 and 5,000 in FIG.
11.
555 The pump power demand curves 160 and 165 and the range of efficient curves
in between are based on the
assumption that the pump is maintained at its most efficient head and flow for
every shaft rpm and for every vane
pitch. Following my said US Patents, this function is approximated by a
control function based on the pump
affinity relationship: head (h) equals an affinity constant multiplied by the
square of the pump rpm (N) or h = kN~2.
This nozzle control function and method are detailed in my US Patent
5,679,035, which is incorporated here by
560 reference. It is well understood in the art of pump design that the
affinity relationship between pump head and shaft
rpm holds true for variable-vane pumps over a wide range of vane settings. It
is also well understood in the art that
the pump affinity constant is only approximate, because the pump efficiency is
reduced at higher shaft rpm. This
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efficiency deviation from the affinity relationship generally does not cause
significant losses in employment of the
nozzle control function, because the pump efficiency does not drop
significantly so long as the operating head and
565 flow are close to the most efficient operating point. However, factoring
in an efficiency correction factor based on
shaft rpm can increase the accuracy of the head affinity control relationship.
In practice, the efficiency reduction in
the pump with higher shaft rpm can be largely captured in the head affinity
constant, so that the control relationship
is still: head equals a constant (corrected for efficiency reduction with
increasing rpm) multiplied by the square of
the pump shaft rpm (h =1cN~2). This efficiency correction is also useful in
the pump shaft power demand
570 calculation, which is discussed below.
The curve 167 in FIG. 11 represents the power demand curve that can be
achieved at somewhat reduced pump
efficiency by either further increasing vane pitch or by reducing the nozzle
area below that required to maintain the
pump at its most efficient operating head and flow. In the preferred
embodiment this occurs at low watercraft
575 speeds and low engine rpm. For example, the nozzle 80 of the preferred
embodiment is designed for a rectangular
discharge opening 89 of 10" by 10", which is sufficient to maintain the pump
at its most efficient operating point on
the full-pitch curve 165 at a watercraft speed of 20 mph, where the inlet duct
30 is recovering about 12 feet of total
dynamic head at the pump inlet. However, at zero watercraft speed, and in the
absence of the 12 feet of recovered
head at the nozzle in addition to the pump head, the maximum nozzle area
restricts full-pitch pump flow. Hence,
580 pump head and shaft power demand are increased as is well understood in
the pump art. This results in a zero-
watercraft-speed, full-pitch power demand curve 166. The curve 166 can be
shifted down and to the right by
reducing the impeller pitch, which reduces the efficient flow and the
corresponding efficient nozzle area to
approach the maximum effective nozzle area of the variable rectangular
steering nozzle 80. Hence reducing the
pitch of the impeller vanes 57 reduces the pump shaft power demand in this
range, just as it does between the
585 curves 166 and 165. The control area between the curves 167 and 166 is
used to get more thrust at low engine rpm
at low boat speeds.
FIGS. 12, 13A, 13B, and 13C are flow diagrams for the microcontroller 140
program. The "d" values in FIGS. 12
and 13 are control dead band factors to prevent hunting, as is common in the
art. These are discussed in the
590 Operation of the Invention below.
Operation
The operation of the invention is controlled by the microcontroller 140 using
the control program diagrammed in
FIGs. 12 and 13. The physical components are shown in FIG. 2. The control loop
of FIG. 12 begins with reading
the position P of the shaft encoder 143 on the single handle control 141. If P
is in the throttle dead band range, the
595 microcontroller 140 increments the vane hydraulic control module 158 to
set the pitch of the impeller vanes 57 is
set to follow P. This has the effect of giving the operator direct control
over the forward or reverse flow through the
pump 50. The concentric spherical surfaces of the split pump housing 62, the
impeller vanes 57 and the spherical
surface of impeller hub 53 allow the impeller vanes 57 to rotate through 90
degrees or more, while maintaining
close fits between the vanes 57 and both housing 62 and the hub 53. In the
preferred embodiment, the impeller
600 vanes rotate to about plus 40 degrees beta-2 for full forward pitch and
through zero to minus 20 degrees beta-2 for
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full reverse pitch. The vane hydraulic control module 158 positions the piston
109 by controlling the flow of fluid
through the hydraulic fluid passage 107. The piston 109 acts through the
thrust bearing 110, the bearing plate 111,
the pushrod 65, and the spider 55 to move the operating arms 56, which rotate
the impeller vanes 57. If P is out of
the dead band in reverse, the microcontroller 140 holds full reverse pitch on
the vanes 57.
605
If P is greater than idle in the Forward Mode, the program of FIG. 12 branches
to adjust the nozzle according to the
pump head affinity relationship. The program of FIG. 12 then adjusts the inlet
slide to match entrance velocity to
boat speed, and passes control to 13A, 13B, or 13C for setting the pitch of
the pump impeller vanes 57.
610 When the operator moves the single handle control 141 out of the dead band
range in the forward direction, the
microcontroller program branches to the "Forward Mode" as shown in FIG. 12. In
accordance with the discussion
of FIG. 12 above, the control program adjusts the nozzle vanes 84 between the
positions shown in FIGS. 8A and 8C
to maintain the most efficient head on the spherical pump 50, according to the
pump affinity relationship h = kN~2.
This action maintains the pump 50 at its most efficient head for the current
shaft rpm. As also shown in FIG. 12,
615 the control program sets the adjustable inlet slide a lto match the
velocity of the inlet entrance flow 45 to the
velocity of the water under the boat. This maintains the most efficient
possible recovery of total dynamic head at
the inlet of the pump 50.
Control is then passed to one of three methods to match the shaft power demand
of the pump 50 to the most
620 efficient power supplied by the engine 21. FIG. 13A sets the pitch of the
impeller vanes 57 based on pump shaft
power demand calculated from measured head and flow on the pump 50.
Alternatively, FIG. 13B sets the pitch of
the impeller vanes 57 based on throttle position as measured by the shaft
position encoder 143 on the single handle
control 141. Alternatively, FIG. 13C sets the pitch of the impeller vanes 57
based on feedback from the combustion
controller 23 on the engine 21. It will be appreciated by those skilled in the
art that each of these alternative
625 methods accomplishes the same function: they all adjust the pitch on the
impeller vanes 57, so that the shaft power
demand of the pump 50 approximates the most efficient power supplied by the
engine 21 at its current rpm.
This program of FIG. 12 has the following consequences. When the handle 141 is
at the forward end of the dead
band range, the impeller vanes 57 are at full forward pitch, which provides
maximum forward thrust. When the
630 handle 141 is in the middle of the dead band range, the vane 57 pitch is
about zero, which provides no pumping
action and therefore a true neutral. When the handle 141 is at the back end of
the dead band range, the vane pitch is
in the maximum negative position, which provides reverse thrust and back
flushing of trash. As the handle moves
out of the dead band range in either direction, it increases the engine
throttle 142, which increases thrust, as is
common with single handle controls on recreational boats. From this it is
clear that the action of the propulsion
635 system 20 in response to the position of the control handle 141 is
identical to the action of propulsion systems of the
prior art.
In the "Forward Mode" of FIG. 12, and referring to FIG. 2 and FIG. 9, the
microcontroller 140 reads the head
differential pressure sensor 145 to input the pump head h and the engine
tachometer input 154 to input engine rpm
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640 N. If the measured head is higher than the pump affinity value kN~2 plus a
small dead band factor d to prevent
hunting, the inicrocontroller 140 uses the flow control module 133, which
positions the driving cylinder 130 and
consequently the driven cylinder 126, which forces fluid into the hydraulic
circuits 128 and 129, while removing an
equal amount of fluid from the hydraulic circuit 127. Following the
explanation of FIG. 9 above, this results in a
balanced fluid flow to the steering cylinders 93, so that the steering rams 91
are equally retracted, acting through
645 ball-ended couplings on the nozzle vane operating arms to increase the
distance between the nozzle vanes 84, thus
increasing the effective nozzle area 89. This has the effect of increasing the
water flow 44 and consequently
reducing the head h on the pump 50. If the measured head h is lower than the
pump affinity value kN~2 minus a
small dead band factor d to prevent hunting, the microcontroller 140 uses the
flow control module 133, which
positions the driving cylinder 130 and consequently the driven cylinder 126,
which removes fluid from the
650 hydraulic circuits 128 and 129, while it forces an equal amount of fluid
from the hydraulic circuit 127. Following
the explanation of FIG. 9 above, this results in a balanced fluid flow to the
steering cylinders 93, so that the steering
rams 91 are equally extended, acting through ball-ended couplings on the
nozzle vane operating arms to reduce the
distance between the nozzle vanes 84, thus reducing the effective nozzle area
89. This has the effect of reducing the
water flow 44 and consequently increasing the head h on the pump 50. If the
head h is within the dead band range,
655 no nozzle control action is taken.
The next sequence in the control loop of FIG. 12 is setting the inlet slide
131. Referring to FIG. 2, the
microcontroller 140 reads the position of the inlet slide cylinder 34 from the
inlet control module 159 and computes
the effective inlet entrance area. It reads the flow differential pressure
transducer 149 and computes the system
660 flow 44 as in the description of the flow differential pressure input 149
above. The microcontroller 140 then
computes the entrance velocity through the inlet entrance opening from V =
system flow 44 divided by the effective
inlet entrance area. The microcontroller 140 reads the boat speed pressure
transducer 151 and compares watercraft
speed S. If V > S +d, the microcontroller 140 outputs to the inlet control
module 159, which actuates the inlet
cylinder 34 to move the adjustable inlet slide 31 back, which increases the
effective entrance area and reduces the
665 entrance velocity V. If V < S -d, the microcontroller 140 outputs to the
inlet control module 159, which actuates the
inlet cylinder 34 to move the adjustable inlet slide 31 forward, which reduces
the effective entrance area and
increases the entrance velocity V. This control function meets the requirement
of my said US Patents that inlet duct
efficiency requires that the flow velocity through the inlet entrance
approximate the velocity of the water under the
hull, which is indicated by the boat speedometer 152. From the discussion
above, it is clear that the
670 microcontroller 140 can also be programmed to calculate the system flow
from positional feedback from the vane
control module 157 on the angle of the impeller vanes 57, which would allow
the elimination of the flow pressure
transducer 149 input in the control loop.
At the end of FIG. 12 the microcontroller program control passes to FIG. 13A,
13B, or 13C to match the pump SO
675 power demand to the power most efficiently supplied by the engine 21 by
varying the pitch of the impeller vanes
57.
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The control scheme in FIG. 13A first computes the hydraulic power produced by
the pump 50, which is the product
of pump head h and system mass flow rate q. The pump shaft power demand is the
hydraulic power divided by the
680 hydraulic efficiency e, so that the control equation for efficient engine
operation could be written P = hq/e or Pe =
hq. The latter formulation is most convenient, because both the e~cient engine
power P and the hydraulic
efficiency a are dependent on shaft rpm. Hence, a table of Pe values, which is
entered with the rpm N and the boat
speed S can be highly accurate. In the preferred embodiment, the boat speed S
factor is only useful for boat speeds
of less than 20 mph, where the power demand curve falls between the curves 166
and 165 of FIG. 11. The pump
685 head h is constantly available from the nozzle control loop of FIG.12. The
system mass flow rate q calculation is
discussed above in the description of the flow differential pressure
transducer 149 input. It will be apparent to one
skilled in the control systems art that other methods of inputting the mass
flow rate could be used, including
mechanical, acoustic and optical flow sensing devices. The microcontrolIer 140
sets the impeller vane 57 pitch by
outputting to the impeller vane control module 158. Control passes back to the
top of FIG. 12.
690
Alternatively, in FIG. 13B the control loop uses a table of vane pitch targets
T, which is entered with control
position C. The T values include adjustments for pump efficiency variations
and other factors based on test results.
The control position C is a measure of the engine throttle 142 setting, which
has an associated most efficient
operating rpm. This rpm is implicitly included in the table of values for T,
which is entered with C. It will be
695 obvious to those skilled in the art that the watercraft speed S could be
incorporated in the table of values for T to
improve performance, as discussed above. The microcontroller program then sets
the vane pitch to T. This method
presumes that the engine is operating at peak performance. The operator
preference input 157 may be used to
reduce the shaft power demand when the engine is out of tune or laboring. It
may also be used to choose between
Iow-speed performance and fuel economy, as is common in automotive power
transmission. This preference factor
700 is O in the control equations of FIG. 13C. This operation can be similar
to trimming the propeller pitch in an
airplane. After the impeller vane 57 pitch is adjusted, control passes back to
the top of FIG. 12. This method
requires no flow input to and no power calculations by the microcontroller
I40. It is particularly useful for legacy
diesel engines at sea level. The microcontroller 140 sets the impeller vane 57
pitch by outputting to the impeller
vane control module 158, as described above. Control passes back to the top of
FIG. 12.
705
FIG. 13C uses the output from the combustion microcontroller on the engine 21
as the best measure of the power
most efficiently supplied by the engine at the current shaft rpm, which is
shown as 23 on FIG. 2. This control
method is well understood in the automotive field of art, as it is widely used
to determine shift points in automatic
transmission controllers and to control continuously variable transmissions.
In effect, the microcontroller 140 is
7I0 programmed to act as a slave to the engine 21 combustion microcontroller,
which dynamically determines the most
efficient power demand for the motor based on a complex set of environmental
and combustion variables, as is well
understood in the art. In each control cycle the microcontroller 140
incrementally increases, decreases, or leaves
unchanged the impeller pitch, based solely on the input from the combustion
microcontroller. This method requires
no pump head input, no rpm input, no system flow input and no positional
feedback for controlling the impeller
715 vane pitch to match shaft power demand to the most efficient power
supplied by the motor. After the impeller vane
57 pitch is adjusted, control passes back to the top of FIG. 12.
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Note that the method of FIG. 13C requires neither flow measurement nor vane
positional feedback, because it
incorporates control feedback from the engine combustion control computer.
When the direct flow measurement
720 means is not required in the "Set Vane Angle" control sequence, as in FIG.
13C or when flow is estimated by vane
57 pitch, it can be compared with the positional feedback from the vane angle
to monitor the operating efficiency of
the marine jet propulsion system 20. If the calculated flow is lower than that
indicated by the vane position, the
likely cause is debris on the inlet vanes 42, pump impeller vanes 57, and/or
stator vanes 71. The microcontroller
140 can be programmed to alert the operator by some alarm means, such as a
light or a horn.
725
With any of these combinations, the microcontroller 140 is programmed to
adjust the pitch of the impeller vanes 57
through the hydraulic control module 158, so that the pump 50 shaft power
demand is made to approximate the
most efficient power supplied by the engine 21 at the current shaft rpm.
730 The functional advantages of this program of operation are described more
fully below.
When the operator switches the ignition on, the microcontroller 140 outputs to
the vane control module 158 to set
the impeller vane 57 pitch to the position indicated by the shaft encoder 143
on the single handle control 141, which
is generally zero pitch for neutral pump flow. The operator then starts the
engine 21, which idles at about 1,000
735 rprn. In response to the movement of the single handle control 141 in the
+/- 10-degree dead band range, the
microcontroller 140 adjusts the impeller vane 57 angle to continuously vary
the forward, neutral, and reverse thrust
of the marine jet propulsion system 20, as detailed in FIG. 12 and the
associated discussion above. Moving the
handle through the straight up position results in a scissoring action between
the pump vanes, which cleans debris
off the leading edges of the vanes. Moving the handle further back results in
reverse pitch and in reversing the
740 pump flow, which back flushes the system for trash removal. Such reverse
pump flow also produces an effective
reverse thrust. Moving the handle 141 back further increases engine rpm and
consequently the magnitude of the
reverse thrust, just as is common in propeller driven boats.
This operation provides smooth, quick shifting from forward to reverse thrust
for low speed maneuvering, because
745 there is no change of shaft direction in transitioning from forward to
reverse or from reverse to forward. The swim
platform and power trim function, which are both popular on recreational boats
of the prior art, may be used to
reduce vortex formation and cavitation in the reverse thrust mode, as shown in
FIG. 10. The operator independently
controls power trim, just as in stern-drive and outboard propulsion systems of
the prior art.
750 The steering wheel controls the action of the nozzle vanes 84 through the
range of motion shown in FIG. 8A and 8B
through the hydraulic helm and hydraulic circuits shown in FIG. 9 and
described above. Turning the wheel 124 of
FIG. 9 to the right results in the left position of the nozzle vanes 84 in
FIG. 8B. The resulting directional change of
momentum of the system flow creates a reaction steering force to the right
along the transom 95, so that when the
wheel is turned to the left as in FIG. 8B, the transom 95 is driven to the
right. The reaction force resulting from
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755 reverse system flow is in the same direction as with forward flow, so that
the transom 95 always moves to the right
when the wheel 124 is turned to the left. Such reaction forces are well
understood in the hydraulic art.
In the forward thrust mode of operation, the variable nozzle is controlled to
maintain the most efficient head on the
variable pitch impeller pump for the current shaft rpm, as is described in my
the US Patent 5,679,035 and as further
760 described above. It is well understood in the art that efficiency is
fairly constant over a broad range of impeller
pitch. The resulting flow through the pump is well understood to be a function
of the impeller pitch. At low
cruising speeds, the power demanded to propel the boat at constant speed is
low. To match the power demanded by
the pump to the most efficient rpm of the engine, the microcontroller sets the
pump impeller pitch near maximum,
which is the state in which it is passed from the low speed maneuvering mode
to the forward mode. To maintain
765 the pump close to its most efficient operating conditions, the
microcontroller opens the variable steering nozzle to
maximum. In addition to maintaining engine efficiency, this control strategy
has the fortunate consequence of
providing maximum flow at low speeds for maximum propulsion e~ciency. The flow
through the maximum
nozzle opening also occurs at the lowest possible velocity. Thus, motor
efficiency, pump efficiency, and flow rate
efficiency are all close to optimum, and wake turbulence is minimized.
770
When the system 20 is under full acceleration, as in pulling up a water skier,
the microcontroller 140 will reduce the
pitch on the impeller vanes 57 to match the pump's shaft power demand to the
engine's 21 most efficient power
supply at the instantaneous shaft rpm. The control system will also reduce the
nozzle area 89 to maintain the most
efficient head on the pump 50 for its current rpm.
775
When the boat reaches steady wakeboarding speed in the approximate range of 15
to 20 mph, the impeller vanes 57
are close to full pitch to reduce the engine rpm to the most efficient
operating point along the line 163 of FIG. 11.
The variable nozzle 80 is close to being fully open to maintain the most
efficient pump head at the relatively low
shaft rpm. A further advantage is that the variable inlet duct opening is near
maximum due to the high flow, which
780 results in no losses from the conversion of inlet entrance velocity to
pressure at the pump inlet. This also again has
the fortunate consequence of providing close to maximum flow 44 at this
relatively low boat speed for maximum
propulsion efficiency, which also results in minimum nozzle velocity through
the large nozzle effective nozzle area
89 and consequently in minimum wake turbulence. The system 20 rpm is further
reduced relative to systems of the
prior art by this higher propulsion efficiency, which requires less shaft
power and consequently lower shaft rpm to
785 maintain the boat speed. Thus, engine 21 efficiency, inlet duct 30
efficiency, pump 50 efficiency, and flow rate 32
efficiency are all close to optimum, and wake turbulence is minimized.
When the boat reaches steady water skiing speed at approximately 30 mph, the
recovery of pressure in the inlet duct
30 has increased, so the microcontroller 140 has made a slight reduction in
effective nozzle area 89 to maintain the
790 most efficient system flow 32 and head on the pump 50. The power required
to maintain this higher boat speed is
also higher, so the engine must operate at a higher rpm to supply the
necessary power. The most efficient pump 50
head rises as the square of the shaft rpm. Higher engine 21 rpm causes the
microcontroller 140 to reduce the pitch
on the impeller vanes 57, which reduces the most efficient pump flow 32. The
nozzle area 89 head-affinity control
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function implicitly accounts for higher inlet head at this boat speed, higher
pump 50 head at the higher shaft rpm,
795 and the reduced flow 32 resulting from reduced pitch on the impeller vanes
57. As a result of all these factors, the
nozzle area 89 is reduced and the nozzle velocity relative to the boat
velocity is increased. However, the nozzle
velocity relative to the water 29 surface is reduced by the increased boat
speed, so that the velocity of the jet relative
to the water surface has only slightly increased. Wake turbulence is thereby
only slightly increased, and the use of
longer towropes at this higher boat speed makes wake turbulence less critical,
since it has more time to dissipate
800 before the skier reaches it.
Further increases in boat speed demand increased engine 21 power, which the
engine 21 can only supply at higher
rpm. The microcontroller 140 reduces the pitch on the impeller vanes 57 to
maintain efficient engine 21 operation
at the higher rpm. Reduced pitch on the impeller vanes 57 requires a
commensurate reduction in nozzle area 89.
805 Pump 50 head is rising as the square of the engine 21 rpm. Inlet 30 head
is rising as the square of the boat speed.
The increasing pump 50 rpm, the reducing vane 57 pitch, and the higher inlet
30 pressure are all factors, which will
result in the microcontroller 140 reducing the nozzle area 89 to maintain peak
pump 50 efficiency. Hence, nozzle
area 89 is reduced with increasing rapidity as boat speed increases as a
natural consequence of the microcontroller
140 operation, until minimum nozzle area 89 is reached at the top design speed
of the system 20. The minimum
810 nozzle area 89 at top speed is also ideal for reducing the system flow
rate 32, hence improving propulsion efficiency
at the higher speed.
In compliance with the statute, the invention, described herein, has been
described in language more or less specific
as to structural features. It should be understood, however, the invention is
not limited to the specific features
815 shown, since the means and construction shown comprised only the preferred
embodiments for putting the
invention into effect. The invention is, therefore, claimed in any of its
forms or modifications within the legitimate
and valid scope of the amended claims, appropriately interpreted in accordance
with the doctrine of equivalents.
