U.S. patent number 3,877,839 [Application Number 05/402,973] was granted by the patent office on 1975-04-15 for torque limiting means for variable displacement pumps.
Invention is credited to Richard Joseph Ifield.
United States Patent |
3,877,839 |
Ifield |
April 15, 1975 |
**Please see images for:
( Certificate of Correction ) ** |
Torque limiting means for variable displacement pumps
Abstract
The loading of rotor support journals, of variable displacement
pumps, which is proportional to the theoretical driving torque of
the pump, is converted to a fluid pressure which is proportional to
the theoretical pump driving torque. The fluid pressure is applied
to a plunger in opposition to an adjustable load, to operate a
servo valve to regulate the pump displacement. The loading may be
varied to provide variations in the pump driving torques to
correspond to the torques available from a prime mover under a
variety of power settings.
Inventors: |
Ifield; Richard Joseph (Sidney,
New South Wales, AU) |
Family
ID: |
3765433 |
Appl.
No.: |
05/402,973 |
Filed: |
October 3, 1973 |
Foreign Application Priority Data
Current U.S.
Class: |
417/217;
417/222.1 |
Current CPC
Class: |
F04B
1/324 (20130101); F04B 1/2071 (20130101); F04B
49/002 (20130101); F04B 49/08 (20130101); F02B
1/04 (20130101) |
Current International
Class: |
F04B
1/20 (20060101); F04B 1/32 (20060101); F04B
49/00 (20060101); F04B 1/12 (20060101); F04B
49/08 (20060101); F02B 1/00 (20060101); F02B
1/04 (20060101); F04b 049/00 () |
Field of
Search: |
;91/504-506
;417/217,218,222,15 |
References Cited
[Referenced By]
U.S. Patent Documents
Primary Examiner: Freeh; William L.
Attorney, Agent or Firm: McGlew and Tuttle
Claims
The claims defining the invention are as follows:
1. In a variable displacement piston pump including a driven rotor,
torque limiting means, for limiting the driving torque of the pump,
comprising, in combination, a hydrostatically balanced slipper
bearing supporting the pump rotor, pressure responsive means
operable to regulate a pump displacement, and means transmitting
the hydrostatic bearing pressure to said pressure responsive means
to effect operation of said pressure responsive means.
2. Torque limiting means according to claim 1, wherein said
pressure responsive means comprise a pilot valve device including a
moving element responsive to the bearing pressure in opposition to
a resilient loading, a servo motor controlling the pump discharge
and a servo control valve operated in consequence of the operation
of said pilot valve and controlling said servo motor.
3. Torque limiting means according to claim 2, in which the pump is
driven by a prime mover, and means controlling the torque as a
function of the speed of the prime mover including means applying a
fixed negative load to said moving element in combination with
means applying, to said movable element, a load varying as a
function of the prime mover speed and power control setting.
4. Torque limiting means according to claim 2, including a
stabilising device comprising a piston in a cylinder device
connected to the servo motor to produce a fluid pressure varying as
a function of the rate of change of the pump displacement and
acting on the pilot valve loading to enhance the stability of the
control system.
5. Torque limiting means according to claim 2, wherein operation of
said pilot valve overrides the effect of the other control means
acting on the servo valve so as to limit the pump driving torque by
variation of the pump displacement irrespective of the setting of
said other control means.
6. Torque limiting means according to claim 5, wherein said servo
valve is a direction control selector valve, having an axially
moveable spool such that, at one limit of the spool axial travel,
the servo valve controls the servo motor in response to the setting
of said other control means whereas, at the other limit of the
spool axial travel, the servo valve causes the servo motor to
reduce pump capacity; the spool being moved to said other limit by
a pressure change in consequence of operation of said pilot
valve.
7. Torque limiting means according to claim 6, wherein said servo
valve is a combined manual and automatic displacement and
directional control selector valve, comprising a manually operated,
ported sleeve and a rotary spool driven by the servo motor and
subject to axial displacement in consequence of operation of said
pilot valve.
8. Torque limiting means according to claim 5, wherein said pilot
valve comprises a bearing pressure responsive plunger, bearing
against a springloaded valve element, the springload being
mechanically adjustable to regulate the maximum torque value.
Description
FIELD AND BACKGROUND OF THE INVENTION
This invention relates to hydraulic pumps and motors of the kind
comprising a body, and a rotor surrounded by and mounted in an
external support, the body having an inlet and outlet for liquid to
be pumped and the rotor carrying a plurality of equiangularly
spaced pistons which are disposed parallel to or inclined to the
rotor rotational axis and are reciprocated as the rotor rotates in
order to pump liquid. Such a pump will be referred to hereinafter
as being of the variable displacement piston pump type.
One form of piston pump to which the invention may be applied
comprises a rotary cylinder block assembly in which the pistons are
equally spaced around the cylinder block axis with the cylinder
bores parallel to the axis. The cylinder block is linked to the
driving mechanism in such a way that the driving mechanism and
cylinder block rotate at the same speed. As the shaft rotates, the
distance between any one piston and the porting surface changes
continually. Each piston moves away from the porting surface during
one half of the revolution and toward the porting surface during
the other half, the inlet port being in line as the piston moves
away, and the outlet port being in line as the piston moves
closer.
An object of the invention is to provide a variable displacement
piston pump, with torque limiting means in a convenient form, as to
ensure a simplified and cheaper production of the torque means,
which are easily adjusted on site, or controlled to suit any
available torque from a power supply unit.
The invention consists in torque limiting means for limiting the
driving torque of a variable displacement piston pump comprising a
hydrostatically balanced slipper bearing, supporting the pump
rotor, and pressure responsive means, responsive to the hydrostatic
bearing pressure, to regulate the pump's displacement.
When variable displacement piston pumps are employed for power
transmission purposes, it is usually necessary to provide a manual
control to regulate the maximum pump displacement and to provide
overriding devices, one to reduce the pump displacement if the
transmission pressure exceeds a predetermined value, in order to
protect the transmission components against overload, and another
to reduce the pump displacement if the driving torque exceeds a
predetermined value, in order to prevent overloading of the engine
or electric motor which drives the pump.
The theoretical pump driving torque is the product of the pump
displacement and the pressure difference, and the torque limiting
means are based on this fact. Hitherto the torque limiting means
have including a plunger subject to the pump delivery pressure to
obtain a servo valve to regulate the pump displacement through a
servo motor, the plunger movements being opposed by a spring
loading which is varied by a cam responding to the movements of the
servo motor, so that the spring loading is inversely proportional
to the pump displacement. Such torque limiting means are complex
and costly to produce and they can be adjusted correctly only on
special test rigs incorporating torque dynamometers. Such torque
limiters cannot be adjusted in a proportional manner to suit
varying available torques, such as the varying torques of petrol
engines operating over a range of throttle openings.
SUMMARY OF THE INVENTION
The invention is based on an appreciation that the rotor support
journal loading of variable displacement pumps is the product of
the pump displacement and the operating pressure, so the journal
loading is proportional to the theoretical driving torque.
According to the invention, the torque responsive journal loading
is converted to a fluid pressure, so that the fluid pressure is
proportional to the theoretical pump driving torque. This pressure
is applied to a plunger in opposition to an adjustable load to
operate a servo valve to regulate the pump displacement. The
invention includes novel means for varying the loading, to provide
variations in the pump driving torques to suit the torques
available from a prime mover under a variety of power settings.
The invention is described as applied to a particular form of axial
piston pump, which is controlled overcenter for forward and reverse
drive, but it can equally well be applied to other forms of
variable displacement pumps.
The invention is described with reference to a particular type of
hydro-statically balanced slipper bearing, but it can equally well
be applied to operate from any design of hydrostatically balanced
bearing, where the hydro-static pressure supports the torque
responsive loading. In order to illustrate the function of the
invention in relation to the other controls, a suitable complete
control system is described, but the invention is not limited to
the type and details of the control system or of servo motor
employed.
BRIEF DESCRIPTION OF THE DRAWINGS
An embodiment of the invention is described with reference to the
accompanying drawings.
FIG. 1 is a longitudinal section through a typical axial piston
pump of a kind to which the invention may be applied.
FIG. 2 is an end view of the stationary port face for the pump
shown in FIG. 1.
FIG. 3 is a diagrammatic representation of a pump control system
according to one embodiment of the invention.
FIG. 4 shows a preferred form of servo control valve.
FIG. 5 is a graphical representation of the performance of an
engine speed responsive torque limiter according to the
invention.
FIG. 6 shows a method for achieving the control characteristics
shown in FIG. 5.
FIGS. 7, 8 and 9 are sectional views taken along the lines 7--7,
8--8 and 9--9 respectively of FIG. 4.
Referring to FIG. 1, the pump cylinder barrel 1 contains nine axial
cylinders with pistons 2 having ball and socket joints 3 to flat
faced slippers, bearing against a variable angle oblique surface 4
and held in contact with the surface by the slipper retaining plate
5.
The cylinder rotor is loaded against its port face 6 by a number of
springs 7, reacting through the drive shaft 8 to the thrust bearing
9. A suitable prime mover, such as an electric motor or an internal
combustion engine, is arranged to be coupled to drive shaft 8.
The oblique surface 4 is on a part cylindrical element 10, bearing
in the housing end cover 11 and the angle of the oblique surface is
adjustable by movements of the servo motor piston 12 through a gear
tooth rack 13 engaging teeth formed on the part cylindrical element
10. Movement of the servo motor piston is capable of rotating the
cylindrical element 10 to an oblique angle of say 20.degree. in
either direction, thereby varying the strokes of the pistons and
the direction of flow through the portings.
When the element 10 is in the position shown and the pump is driven
clockwise, as viewed from the end of the drive shaft 8, the
delivery will be from port A shown in FIG. 2, which will be
referred to as the forward high pressure port. When the element 10
is tilted in the opposite direction, the delivery will be through
port B, which will be referred to as the reverse high pressure
port.
The drive shaft 8 is provided with a high degree of flexibility, so
that the cylinder barrel 1 is free to align itself against the flat
port face surface 6 and the cylinder barrel is located by four part
spherical slipper pads. The journal loading at the cylinder barrel
is applied at the slipper 14, when the oblique surface 4 is tilted
in the direction shown. When the oblique surface is tilted in the
opposite direction, the journal loading would be at the slipper 15.
There are no theoretical sideways loadings at the cylinder barrel,
but a slipper 16 as shown in the partial section of FIG. 2, is
placed at each side for location purposes only.
The slippers 14 and 15 are hydro-statically balanced against the
journal loadings. They may for example be connected through the
connection points marked C and D to the highest porting pressure
through restricted orifices. The sealed area at the spherical faces
of the slippers is sufficiently great for the requirements at
maximum pump displacement. The required hydro-static balance
pressure becomes progressively less than the porting pressure as
the pump displacement is reduced. Under these conditions the
slipper operates with a small clearance at its spherical surface,
the resulting leakage providing an appropriate pressure drop at the
restricted orifice.
The inside of the pump housing is maintained at the same pressure
as that in the low pressure port, so the journal loading is
proportional to the product of the porting pressure difference and
the tangent of the angle of the oblique surface. The pump driving
torque is also theoretically proportional to the product of these
two variables, so the hydro-static pressure at the slipper is
proportional to the theoretical driving torque. According to the
invention, this hydro-static pressure at the loaded slipper bearing
is employed as a signal for the torque limiting device.
The two ends of the servo cylinder are shown with connection points
marked E and F and the various connection points identified by the
letters A,B,C,D,E, and F in FIGS. 1 and 2 are similarly identified
in the control system diagram FIG. 3.
Referring to FIG. 3, a low pressure pump 17, draws from the
reservoir 18 and its delivery pressure is limited by the setting of
the low pressure relief valve 19. It delivers directly to the
inside of the variable displacement piston pump housing, from which
it flows to whichever is the low pressure port through one of the
non return valves 20 or 21. Under clockwise forward conditions port
A is at high pressure and port B is charged from the low pressure
pump through the non return valve 20.
Ducts from the pump ports A and B have non return valves 22 and 23,
the downstream sides of which are inter-connected, supplying the
selected highest pressure through the restrictors 24 and 25 to the
hydro-statically balanced slippers 14 and 15, respectively, through
the connection points C and D as described with reference to FIG.
1.
The selected highest pressure from the non return valves 22 and 23
is supplied to the center gallery of a manually operated reversing
valve 26, which also contains a small gallery and a central
drilling to one end, which is ducted to the reservoir 18.
Two galleries 28 and 29 formed in the housing 27 of the manually
operated reversing valve 26 are ducted through a servo valve 30 to
the cylinders of the transmission pump servo motor 12 via the
connection points E and F, respectively. The manually operated
reversing valve 26 is shown in the position to supply the highest
pressure to the servo motor connection point E, with the connection
point F ducted through the gallery 29 to the reservoir. This causes
the servo piston to move to the position of maximum pump
displacement as shown at FIG. 1. Movement of the reversing valve 26
outwards reverses the pressures at the connection points E and F
and causes the servo piston 12 to move to the other extreme
position.
The servo valve 30 contains a spool valve 31 having an enlarged
piston 32 at one end, sliding in a cylinder 33 and loaded by a
spring 34 in the opposite direction to the position shown. The
spring chamber and the remote end of the spool valve 31 are ducted
to the reservoir 18. The delivery port of the low pressure pump 17
is ducted to the cylinder 33 through the restrictor 35 to move the
spool valve 31 against the loading of the spring 34 to the position
shown, where it does not influence the operation of the servo motor
in response to the movements of the manually operated reversing
valve 27.
The galleries 36 and 37 of the servo valve 30 are ducted to the
transmission pump ports A and B respectively. If the cylinder 33 is
exhausted, the spool valve 31 will move in the opposite direction
to that shown, to isolate the galleries ducted to the manually
operated reversing valve 27 and to provide communication from the
pump ports A and B to the connection points F and E, respectively,
of the servo motor cylinder.
Port A is the forward drive high pressure port, so movement of the
spool valve 31 causes the servo motor piston 12 to reduce the
displacement of the pump under forward drive conditions, but it
cannot cause movements beyond zero pump displacement, because this
would cause port B to become the high pressure port. The operation
is similar when the manually operated valve 26 is moved to the
reverse position.
Under overrun conditions, the pump operates as a motor and, for
forward setting of the reversing valve, port B becomes the high
pressure port. Movement of the spool valve 31 then causes the servo
piston to increase the pump displacement, as desired under overrun
conditions.
For counter clockwise rotation of the pump, it is necessary only to
reverse the connections from ports A and B to the galleries 36 and
37.
In most applications of variable displacement transmission pumps,
it is necessary to provide means to stall the pump off stroke at a
predetermined maximum pressure to prevent overloading of the pump,
and also at a predetermined maximum pump driving torque, to prevent
overloading of the power plant driving the pump.
A pilot valve device 38 contains two small plungers 39 and 40
bearing against valves 41 and 42 which are loaded against their
seatings by adjustable springs 43 and 44 to close galleries which
are ducted to the cylinder 33 of the servo valve. The pilot valve
discharge chambers are ducted to the reservoir.
A duct supplies whichever is the higher pressure from ports A and B
to the end of the plunger 39, so that if this pressure is excessive
the valve 41 opens to vent the cylinder 33, causing the pump to
stall off stroke.
The pressures at the slippers 14 and 15 are ducted through non
return valves 45 and 46 to the end of the plunger 40, so this
plunger is responsive to the hydro-static pressure at whichever
slipper is supporting the journal loading. The force at the plunger
40 is therefore responsive to the pump driving torque. If this
pressure is excessive, the valve opens to vent the cylinder 33, in
the servo valve 42, so limiting the pumps driving torque by
variation of the pump's displacement.
Pump servo systems operated by the high pressure from the pump
delivery ports are prone to instability problems, particularly
under high gain conditions, i.e. at small pump displacement. The
invention includes a stabilising device which is a function of the
rate of change of pump displacement. A typical embodiment of this
part of the invention in shown in FIG. 3.
A piston 47 is shown with a cam roller 48, bearing against a cam 49
formed on the servo motor piston. For preference the cam form is
such that the movements of the piston 47, relative to the servo
motor 12, is an inverse function of the pump displacement. The
outer end of the cylinder for piston 47 communicates through a
central drilling to the inside of the pump housing and is therefore
subject to the pressure from the low pressure pump 17 to hold the
cam roller in contact with the cam. The annular cylinder chamber
communicates with the spring chamber of the pilot valve 38 and is
ducted to the reservoir through the restricted orifice 51. The
pressure difference across the restricted orifice 51, resulting
from movements of the piston 47, acts on the sealed pistons 52 and
53 to vary the loading at the valves 41 and 42.
When one of the valves 41 or 42 opens to limit the pump pressure or
the pump's driving torque, the pressure difference at orifice 51
provides a force to close the valve. A reduction in pressure at the
plungers 39 or 40, allowing the associated valve to close, results
in a pressure difference at orifice 51, providing a force to open
the valve. Thus the stabilizing device provides a leading term, to
stabilise the automatic control system, from the rate of change of
the pump displacement.
In many applications of variable displacement pumps, it is desired
to vary the maximum pump displacement by a manual control, with
maximum pressure and maximum torque overriding devices. FIG. 4
shows a valve to replace valves 27 and 30 of FIG. 3 and which
provides for manual displacement control.
A manually operated control lever 54 rotates a ported sleeve 55 in
a housing 56. A rotary valve element 57 inside the ported sleeve
has a gear tooth quadrant 58 driven by the gear 59 attached to the
element 10 of FIG. 1, or by rack teeth formed on the servo motor
piston 12. In either case, the gear tooth chamber communicates with
the inside of the pump housing and is subject to the pressure from
the low pressure pump 17. This pressure acts on an annular area of
the gear driven spool 57 and it is applied to a cylinder 60 on the
opposite side, through a restricted orifice 61. Normally the piston
in the cylinder 60 is moved to the position shown by the pressure
acting on the difference in areas, but if the cylinder 60 is
vented, the gear driven rotary spool moves to the left as viewed on
the drawing. This corresponds to the cylinder 33 of FIG. 3, the
venting valves may be similar to those described for the pilot
valve 38 of FIG. 3, and the connection K is identified in FIG.
3.
Directional control selector rotary valves of this type are well
known. As shown in FIGS. 4, 7, 8 and 9 they usually have galleries
for each different pressure with portings through the sleeve in
diametrically opposite pairs in order to avoid side loadings on the
valves.
The ports marked A and B are ducted to the transmission pump ports
similarly identified. The ports marked E and F are ducted to the
ends of the servo motor cylinder similarly identified. The port
marked G is ducted to the selected highest pressure identified
similarly in FIG. 3. The port marked H is connected to the
reservoir, 18, which is permanently ported to the end of the
cylinder containing the valve 57 and, through a central drilling,
to an annular chamber having an outer diameter equal to the outside
diameter of the sleeve 55.
In its normal operating position as shown, rotation of the valve
sleeve provides communication of the servo cylinder ports E and F
with the selected highest pressure port G or the reservoir port H.
The resulting movement of the servo motor piston 12 closes the
ports G and H, so that normally the valve operates as a position
responsive control. Under these conditions the ports A and B are
closed.
If either of the overriding valves 41 or 42 in FIG. 3 is opened it
vents the cylinder 60 causing the spool valve 57 to slide axially
to the left. This closes the ports H and G from communication with
the valve flutes and opens the ports A and B. Port A then
communicates with one servo cylinder of the servo motor through
port F, and port B communicates with the other servo cylinder
through port E. Thus this valve operates in the same manner as the
servo valve 30 described for FIG. 3.
The spring which loads the valve 42 for torque limitation is shown
with an adjusting screw. Obviously this screw may be replaced by a
push rod operated from a camshaft coupled to the controls of a
prime mover, so that the pump driving torque would be a function of
the setting of the prime mover power controls.
A preferred method of controlling the pump driving torque as a
function of the prime mover power control settings is to cause the
controlling torque to increase as a controllable function of prime
mover speed until the torque required line intersects the available
torque line from the prime mover. This is exampled by the curves in
FIG. 5. The curves a1, to a5 represent prime mover torque outputs
for given throttle openings or given fuel pump discharges for
diesel engines. The curves b define a controlled torque area
increasing with the square of engine speed, with a negative
constant, so that there is zero torque loading for engine speeds
below 15 percent and the torque loading increases to 100 at 40
percent of maximum engine speed. The actual controlled engine
conditions are at the intersection of the a curves and the b
curves.
The curve b.sub.1 represents a minimum control setting and the
curve b.sub.2 represents a maximum control setting for the torque
control, and any other curve may be provided between these limits,
for different prime mover power requirements.
A method of achieving these desired control characteristics is
shown in FIG. 6 which illustrates the torque limiting device, the
low pressure pump 17, the reservoir 18 and the low pressure relief
valve 19.
The lettered connection points to the torque limiter are J, the
torque responsive hydro-static pressure from the slipper bearing,
K, the pilot pressure identified at the servo valve device 30 in
FIG. 3, or the device illustrated in FIG. 4, and L, stabilising the
pressure from the cylinder 50 in FIG. 3.
A variable orifice 62 is fitted in the discharge passage from the
low pressure pump 17. The pressure drop at this orifice is applied
to the piston 63, against the loading of the spring 64, to close
the valve. The spring 64 provides the desired negative constant.
The pump 17 is driven by the prime mover, so the pressure at J,
which is proportional to the pump driving torque, will be K.sub.1
N.sup.2 -K.sub.2, where N is the engine speed, K.sub.2 is the
negative constant and K.sub.1 is varied with the setting of the
prime mover power controls, to provide controlled torque lines such
as b.sub.1 and b.sub.2 in FIG. 5. The arrangement shown in FIG. 6
can replace the equivalent arrangement 38 shown in FIG. 3.
* * * * *