Power control for a hydrostatic transmission

The present invention is directed to a power control for an automatic transmission which consists of two valves, the first of which is a characteristic valve providing a pressure differential linearly proportional to an input speed signal provided by a prime mover driven fixed displacement pump, and a second valve which is modulated by the pressure differential generated by the first valve and by a pressure feedback proportional to the torque transmitted by the transmission. The power control provides a series of linear control curves proportional to input speed and is useful as either an automotive type control or an anti-stall control, or a combination of both.

FIELD OF THE INVENTION 
The field of the present invention is a hydraulic control circuit utilized 
with a variable displacement hydrostatic transmission and responsive to a 
hydraulic signal representing the input speed to the transmission to 
automatically modulate the transmission displacement. This is particularly 
useful in automotive type controls wherein the input speed signal is used 
to increase hydrostatic pump displacement, and/or reduce motor 
displacement, to increase the transmission output speed in proportion to 
input speed increases above a predetermined minimum rpm. The control is 
also useful in an anti-stall control where pump displacement is decreased, 
and/or motor displacement is increased, as prime mover rpm drops below a 
predetermined rpm indicating overloading of the prime mover. 
BACKGROUND OF THE INVENTION 
There have been various single valve controls for use in either an 
automotive type control or an anti-stall control for a prime mover driven 
hydrostatic transmission. The term automotive type control, when used 
relative to hydrostatic transmissions, refers to a displacement control 
which is responsive to pump input speed, and thus driving prime mover 
speed, to gradually increase the transmission ratio as prime mover speed 
increases above a predetermined rpm such as engine idle. In one previous 
automotive control taught in a paper numbered 851505 entitled, "An 
Automotive Type Hydrostatic Transmission Control", by Peter J. Strommel 
and Charles Sticklin, a single speed sensing valve, also referred to as a 
type DA Automotive Style Cartridge Valve, utilizes the .DELTA.P generated 
across a fixed orifice to position a valve spool in response to flow from 
an input speed pump. This valve also has output feedback applied against a 
lesser area to provide a multiplication gain in output relative to speed 
pump flow generated .DELTA.P. The output of the speed sensing valve is 
used to operate a swashplate servo mechanism in response to engine speed. 
The particular construction of the cartridge valve taught in this 851505 
article makes it impossible, or at least difficult, to manufacture and 
assembly the valve. Furthermore, since the flow is generated across a 
fixed orifice, the generated .DELTA.P is not linear, and thus does not 
provide a linear output. 
Another example of an automotive control is taught in U.S. Pat. No. 
3,986,358 issued to Helmuth Koffmann on Oct. 19, 1976 wherein the 
displacement control flow to both the variable displacement pump and a 
variable displacement motor in a transmission are controlled by generated 
input speed pump flow across a manually controlled (gas pedal) variable 
orifice. As the .DELTA.P across the variable orifice increases, a bleed 
valve from the control line is gradually closed to increase the pressure 
to the servo mechanisms. This particular control has no feedback, 
generates no gain in the power control, and furthermore does not provide a 
linear control relationship between speed pump flow and hydraulic unit 
displacement since there is no pressure control of the size of the 
variable orifice. 
It is also known in anti-stall controls to use a fixed displacement pump to 
provide a hydraulic speed signal which is sensed across an orifice to 
generate a .DELTA.P to destroke the hydrostatic transmission pump to 
reduce prime mover load. This may be done by applying the anti-stall 
signal to move the displacement control valve as taught in U.S. Pat. No. 
3,946,560 issued to McIntosh et al on Mar. 30, 1976. U.S. Pat. No. 
3,803,843, issued to Wyman et al, on Apr. 16, 1974, teaches using a fly 
weight valve to generate a speed signal which is applied against an 
accelerator adjusted spring force to position a govenor valve to modulate 
servo control pressure and also is applied against a spring and feedback 
biased pressure regulator valve which acts as a gain amplifier and 
provides charge supply to the govenor valve. In various control stages, 
this control provides both automotive type and anti-stall functions, but 
requires three valves and the flow generated .DELTA.P does not provide a 
linear control curve. 
SUMMARY OF THE INVENTION 
The present invention is directed to a power control for a hydrostatic 
transmission in which a pair of control valves, the first being a 
characteristic valve and the second being a modifier valve, are connected 
in series relationship to conduct a modified input speed signal generated 
by a fixed displacement pump to a displacement servo mechanism for 
controlling hydrostatic transmission speed ratio. The characteristic valve 
acts as a spring biased variable orifice to provide a .DELTA.P having a 
linear characteristic proportional to increasing input speed pump flow, 
and thus prime mover rpm. The .DELTA.P generated by the characteristic 
valve and control feedback are utilized to position the modifier valve 
which controls the flow to the servo mechanism. In the preferred form, the 
modifier valve feedback is compared to case pressure and provide a 
multiplication or amplification gain in the output signal. 
Such power control, when utilized in an automotive type control system, 
provides a rising linear curve characteristic for increasing pump 
displacement as prime mover speed increases so that the control system is 
responsive to prime mover rpm. By simply adjusting the spring preload of 
the characteristic valve, the predetermined pressure at which the speed 
signal flows to the displacement control servo can be modulated. 
Since the characteristic valve operates as a variable orifice with pressure 
feedback rather than a fixed orifice the .DELTA.P generated across the 
characteristic valve will be substantially linear rather than parabolic as 
would be normal for the flow generated .DELTA.P across a fixed orifice. 
This linear .DELTA.P signal is then applied across the modifier valve with 
output feedback to provide the desired gain and reference change to tank 
pressure. 
When such power control is utilized in the anti-stall system, the control 
also provides a relatively steep linear curve characteristic which is used 
for decreasing pump displacement as prime mover speed decreases so as to 
reduce the transmission load on the prime mover as a stall condition 
approaches. 
The term power control is utilized since the control, in both the 
anti-stall configuration and automotive configuration, controls 
transmission displacement relative to both a prime mover input speed and 
the torque absorbed by the transmission. Increased torque load on a 
hydrostatic transmission, and thus increased pressure in the system, 
generates swashplate moments which tend to destroke the variable 
displacement pump. When used as an anti-stall control, the control backs 
off pump displacement until the torque absorbed by the transmission, which 
is proportional to the power absorbed by the transmission, is equal to the 
torque available from the prime mover, which is proportional to the power 
available from the prime mover. When used in an automotive control, the 
control is both input speed responsive due to the input speed signal and 
torque responsive because of the destroking swashplate moments. Thus, the 
control regulates power of the transmission as a function of input speed, 
taking into account changing torque loads. 
It is also an object of the present invention to provide relatively simple 
characteristic and modifier valves which are easy to assemble, inexpensive 
to produce and which are adjustable so as to control the point that the 
control pressure modulates hydraulic unit displacement. 
A further object of the present invention is to provide power control for a 
hydrostatic transmission which has a pair of hydraulic units at least one 
of which being of variable displacement, and the variable displacement 
hydraulic unit has a hydraulic servo mechanism for varying the 
displacement thereof. The hydrostatic transmission furthermore has a fixed 
displacement speed signal pump driven at a speed proportional to the input 
speed of the transmission and provides a hydraulic input speed signal. The 
power control comprises first and second valves in series relationship 
between the speed signal pump and the servo mechansim. The first valve is 
a variable orifice valve which is spring biased toward a closed position 
and biased by the input speed signal toward an open position to establish 
a pressure drop having a generally linear relationship to the input speed 
signal. The second valve is spring biased toward a closed position 
relative to communicating the output of the first valve to the servo 
mechanism and has first and second opposed pilots connected across the 
first valve so that the input speed signal induced pressure drop biases 
the second valve toward an open position. The second valve also has an 
output feedback proportional to pressure in the servo and connected to a 
third pilot biasing the second valve toward the closed position. 
It is also an object of the present invention to provide a valve for use in 
a hydraulic circuit control which has a housing providing an elongated 
housing bore and an axial moveable valve spool slidable within the housing 
bore. The valve spool has an internal bore with a central member. The 
valve housing bore is sealed at each end to form with opposite ends of the 
spool first and second housing pilot chambers. A pair of axially fixed 
pistons extend from the housing ends so as to be received one in each end 
of the internal bore and forming with the central member first and second 
pilot chambers. The first and second housing pilot chambers are in fluid 
communication with a pair of pressure signals generating a first .DELTA.P 
across the valve spool. The output of the valve is modulated by a pair of 
lands located on the valvespool. A radial opening through the valve spool 
between the pair of lands communicates one of the inner pilot chambers 
with the valve output to provide output feedback.

BRIEF DESCRIPTION OF PREFERRED EMBODIMENTS 
The present invention is directed to a power control as would be utilized 
in a hydrostatic transmission such as the transmission 10 of FIG. 1. The 
transmission consists of a pair of hydraulic units such as a pump 12 
driven by a prime mover 14 and motor 16 driving a load 18. The pump 12 and 
motor 16 are interconnected by a pair of hydraulic lines 20 and 22 so as 
to form a closed hydraulic main loop. The pump 12 is of the reversible 
variable displacement type as diagramatically shown to include a 
swashplate 24 operated by a spring centered servo mechanism 26 including a 
pair of piston/cylinder arrangements 28 and 28'. The position of the 
swashplate 24 determines the amount and direction of flow from the pump 12 
to the motor 16 so as to provide a reversible drive of the load 18. While 
the motor 16 is shown to be fixed displacement, it could also be a 
variable displacement hydraulic unit in staged control relationship with 
the pump 12 so that the motor goes from maximum displacement to minimum 
displacement after the pump 12 has reached maximum displacement in either 
direction. The control of the present invention could also provide control 
fluid to a servo mechanism for the motor if of the variable displacement 
type but for purposes of simplicity, the controls only shown to control 
pump displacement. Such variable displacement hydrostatic transmissions 
utilizing various types of hydraulic units are well known and need not be 
described in greater detail relative to understanding the concepts of the 
present invention. 
As is also quite typical in hydrostatic transmissions a fixed displacement 
charge pump 30, which as used in the present invention is also an input 
speed pump, is used to provide both a source of control fluid and make-up 
fluid for the main loop. The pump 30, since it is also to be utilized to 
provide a speed signal is driven at a speed equal to, or directly 
proportional to, the rpm of the prime mover 14. The flow from the pump 30 
proceeds through line 32, line 34 including a characteristic valve 36 to 
be described in detail later, and line 38 to a charge relief valve 40 and 
a pair of check valves 42. Thus the pump 30 provides make-up fluid to 
either main loop line 20 or 22, depending on which side is the low 
pressure side of the main loop, as is well known in the prior art. 
Characteristic valve 36 acts as a variable orifice. As fluid flows through 
an orifice a pressure drop is generated proportional to the volume of flow 
through the orifice. Since the pump 30 of fixed displacement and driven 
proportional to prime mover rpm, the pressure drop, or .DELTA.P, across 
the orifice will also be proportional to engine rpm. However, the 
relationship between flow through a fixed orifice and .DELTA.P is a 
parabolic relationship while the relationship between the flow through a 
properly designed variable orifice and its .DELTA.P could be made 
substantially linear. The characteristic valve 36 of the present invention 
is so designed to provide this linear relationship. 
The details of the characteristic valve 36 are shown in FIG. 2. The 
characteristic valve has a housing 44 containing a stepped bore 34 which 
is represented by the line 34 in the schematic of in FIG. 1. The large 
diameter portion of the stepped bore 34 is connected to the line 32 while 
the small diameter portion of the bore 34 communicates with a pair of 
passages 38 and 38' represented by lines 38 and 38' in FIG. 1, the purpose 
of the latter being described in detail below. Located at the step in the 
bore 34 is a tapered poppet 46 biased toward the step by a spring 48 
located in spring chamber 50. The bias of spring 48 on the poppet 46 can 
be adjusted or varied by the typical threaded adjustment means 52. While 
the spring 48 biases the poppet 46 toward the closed position, the 
pressures of lines 32 and 38 bias the poppet toward the open position. The 
pressure in line 32 applied at the larger diameter portion of thestepped 
bore 34 acts on the outer portion of the poppet 46 to provide an opening 
bias on the characteristic valve as represented by line 54 in FIG. 1. The 
pressure in line 38 is supplied against the inner diameter portion of the 
poppet 46 as represented by line 56 in FIG. 1 and also biases the 
characteristic valve 36 toward an open position. Aiding the spring bias 
and moving the poppet 46 toward the closed position is a restricted pilot 
pressure applied through a restrictive orifice 60 in the poppet 
communicating the pressure of line 38 with the spring chamber 50, and thus 
the backside of the poppet 46. This is represented by line 58 in FIG. 1. 
The communication between line 38 and the spring chamber 50 could also be 
through a separate passageway provided in the characteristic valve housing 
44 and such passageway would also include a restrictive orifice such as 
60. It is noted that under stable conditions when there would be 
substantially no flow through the orifice 60, and thus spring chamber 50 
would be at the same pressure as line 38, the net effect of the pressure 
of line 38 is to bias the poppet toward the closed position since the area 
of the poppet valve at spring chamber 50 is larger than the small diameter 
portion of the stepped passage 34. The net effect of the spring 48 bias 
and the three pilot biases on the poppet is to modulate the position of 
the poppet 46 relative to the stepped portion of the bore 34 to provide a 
small annular gap therebetween. This annular gap forms the variable flow 
orifice between line 32, and thus the speed signal pump 30, and line 38. 
When the prime mover is driving pump 30 there will always be some flow 
through the small annular gap generating a .DELTA.P representative of 
prime mover rpm between line 32 at the input speed signal pressure and 
line 38 at the characteristic valve output pressure, which could also be 
referred to as the conditioned signal. Due to the charge relief valve 40, 
the conditioned signal is normally at a preset pressure, which when 
combined with the characteristic valve flow generated .DELTA.P, 
establishes the pressure in line 32 or the input speed signal pressure. 
Furthermore, since this small annular gap is a variable orifice with the 
appropriate pilot pressures, the .DELTA.P across the characteristic valve 
36 is linearly proportional to prime mover rpm. 
Downstream, in a supply relationship sense, from the characteristic valve 
36 is the second valve in the power control, that is a modifier valve 62 
which is in flow communication with the outlet side of the characteristic 
valve by means of the previously mentioned line 38'. An orifice 64 is 
located in line 38' to restrict the flow through the modifier valve 62 
since only a limited amount of flow will eventually be necessary to 
operate the servo mechanism 26. Also connected to the modifier valve is a 
tank line 66 in fluid communication with a tank or reservoir 68 at 
atmospheric pressure. The outlet of the modifier valve 62 is connected to 
a line 70 which provides a control signal which is applied to one of the 
servo cylinder arrangements 28 and 28' of the servo mechanism 26. 
Modulation of the modifier valve 62 selectively connects the output line 
70 with conditioned pressure as applied from line 38 and 38' through 
orifice 64 or with tank through line 66. 
The modifier valve 62 is selectively positioned by a modifier valve spring 
72 and four pilots schematically shown in FIG. 1 and shown in 
constructional detail in FIG. 3. The modifier valve has a housing 74 with 
elongated bore 76 extending the length thereof. Located within the bore 76 
is an axially moveable valve spool 78. Located one at each end of the 
elongated housing bore 76 are plugs 80 and 82 which form with the 
respective ends of the valve spool 78 housing pilot chambers 84 and 86 
which are respectively connected to lines 32' and 38" shown in FIG. 1. 
The valve spool 78 has an internal bore 88 with a central plug 90 fixed to 
the valve spool 78. The central plug has two machine flats which form 
axially extending passageways 92 and 94, the purpose of which will be 
explained below. Axially positioned relative to the valve housing 74 are a 
pair of dumbbell shaped pistons 96 and 98 which extend into the open ends 
of the valve spool internal bore 88 so as to form inner bore pilot 
chambers 100 and 102. Centrally located on the valve spool are a pair of 
lands 104 and 104' which modulate the fluid communication between the 
modifier valve outlet 70 and either the input of line 38' or drain 68. 
Located between the pair of lands 104 and 104' is a radial passage 106 
which connects with the previously mentioned axial passage 94 to provide 
fluid communication between the outlet 70 and the inner bore chamber 102. 
This fluid communication, schematically shown as line 70' and orifice 106 
in the schematic of FIG. 1, provides a feedback between the valve output 
of line 70 and the inner bore chamber 102 closed by piston 98. Similarly, 
inner bore chamber 100 is in fluid communication with drain through axial 
passage 92 and a radial passage 108 formed in the valve spool outboard of 
land 104'. This is schematically shown in FIG. 1 as line 66' connecting 
with line 66 and tank 68. 
The particular structure of the modifier valve 62 described above provides 
an inexpensive construction which is simple and easy to manufacture. It 
also provides a modifier valve wherein the output has a multiplication or 
amplification in gain relative to a primary input signal consisting of a 
first .DELTA.P applied at the pilot chambers 84 and 86, particularly since 
inner bore chamber 100 is referenced to tank 68. For example, if the 
diameter of the housing bore 76 is twice the diameter of the internal bore 
88, the primary input signal .DELTA.P is applied against an area four 
times the area of the inner bore minus the area of the inner bore 88 while 
the output pressure feedback of chamber 102 is applied against the area of 
the internal bore 88. This results in an output multiplication gain of 
3:1. 
The position of the valve spool 78 is modulated by a combination of forces 
consisting of the spring force of spring 72 located in chamber 86, the 
first .DELTA.P applied at the housing pilot chambers 84 and 86 and a 
second .DELTA.P at inner pilot chambers 100 and 102. Since the first 
.DELTA.P is the pressure drop across the characteristic valve 36 applied 
through lines 32' and 38", the valve spool 78 position is linearly 
proportional to the prime mover 14 input speed once the force of spring 72 
is overcome. For example, the spring 72 has a sufficient force biasing the 
valve spool 78 to the left to require a minimum first .DELTA.P of 30 psi 
to start to move the valve spool 78 toward the right to permit flow from 
line 38' past land 104 to the valve outlet 70. Once the outlet 70 is so 
pressurized, there is a pressure feedback signal applied through orifice 
106 to the inner bore chamber 102 tending to move the valve spool toward 
the left in addition to the spring force of spring 72. This valve 
construction provides an output pressure in line 70 which is greater than 
the control first .DELTA.P by a multiplication factor, such as 3:1 in the 
example presented above. 
The modulated control signal at line 70 is then applied to one of the servo 
cylinders 28 or 28' as selected by a solenoid operated 
forward-neutral-reverse (F-N-R) valve 110 schematically shown in FIG. 1 
and controlled by an electrical input signal from switch 112. The other of 
the servo cylinders 28 or 28' is connected to tank 68 through the F-N-R 
valve 110 and line 114. The F-N-R valve 110 merely selects the direction 
of operation while the amount of swashplate displacement is controlled by 
the power control consisting of the characteristic valve 36 and the 
modifier valve 62 as modulated in response to pump 30 rpm. Since line 70 
is connected to the stroking servo cylinder 28 or 28' through the F-N-R 
valve 110, the feedback pressure in the pilot line 70' of modifier valve 
62 is proportional to the main loop pressure moments destroking pump 12 
and sensed by the servo mechanism 26. This makes the power control of the 
present invention both input speed and torque responsive. 
An additional feature that may be provided is an inching valve which 
consists of a variable orifice 116 controlled by a foot pedal 118 and 
located in line 120 extending between a modifier valve output line 70 and 
tank 68. The inching valve 116 is normally closed, but when it is decided 
to manually override the automotive control to provide slow speed 
operation, the inching valve 116 is gradually opened to bleed pressure 
from line 70 to tank. 
By automotive type control it is meant that the transmission ratio is 
responsive to prime mover input speed as set by the engine throttle whose 
position is determined by the vehicle operator. Prime mover rpm causes the 
speed signal pump 30 to provide flow across characteristic valve 36 to 
establish the first .DELTA.for the modifier valve 62. Operation of the 
automotive control can best be understood by viewing the graph of FIG. 4 
teaching the relationship between pump displacement versus pump speed. 
Below a given prime mover rpm, such as engine idle, it is desirable to have 
zero pump displacement so that there is zero flow in the transmission main 
loop and no transmission load is applied on the prime mover 14. Thus, as 
indicated in FIG. 4 there is an offset to prevent pump displacement until 
a predetermined engine rpm is reached. This offset is established by the 
aforementioned modifier valve spring 72 which prevents movement of the 
modifier valve spool 78 to the right until a predetermined first .DELTA.P 
is established. For example, the spring 72 provides a sufficient force to 
prevent movement of the valve spool 78 until a first .DELTA.P of 30 psi is 
established across the characteristic valve 36 representing the desired 
minimum prime mover speed. As prime mover speed increases above this 
predetermined rpm, further speed increases increase the .DELTA.P across 
the characteristic valve 36 above 30 psi, such .DELTA.P having a linear 
relationship with engine speed as described above, to provide a linear 
increase in control flow from line 38 through the modifier valve 62 to the 
servo mechanism 26. This linear characteristic is shown by the rising 
lines 122, 122' and 122" of FIG. 4 which increase until maximum pump 
displacement is reached as indicated by line 124. 
The slope of the lines 122 through 122" are determined by the spring 
characteristic of spring 48 of the characteristic valve 36. When the power 
control of the present invention is used for an automotive type control, 
the spring 48 has a relatively soft spring characteristic (when compared 
to an anti-stall type control described below) in order to provide a 
gradual slope to the lines 122 through 122". It is also noted that 
adjustment of the characteristic valve spring 48 by the threaded 
adjustment mechanism 52 can also modify the amount of offset established 
by the fixed spring 72 of the modifier valve 62. Since the spring 48 
determines the slope of the linear curve established by the characteristic 
valve 36, the amount of engine rpm necessary to establish a sufficient 
first .DELTA.P to overcome the force of spring 72, 30 psi in the example 
given, the point of control start-up such as point 126 in FIG. 4 can be 
modified. Thus, adjustable spring 48 in conjunction with fixed spring 72 
establish the amount of offset. 
The three lines 122, 122' and 122" represent three pressure curves for the 
transmission which occur due to increased torque creating higher pressures 
in the main loop which tend to destroke the variable displacement pump 12. 
These three linear curve represent main loop pressures of 1000, 3000 and 
5000 psi respectively. Transmission torque is proportional to pump 
displacement times main loop pressure while transmission power is 
proportional to transmission torque times pump speed. Thus, as torque 
loads on the transmission vary, and therefor the prime mover load varies, 
the power control must balance main loop pressures, pump displacement, and 
pump speed to prime mover input torque. For example, if the prime mover 
and transmission are driving a load in stable condition at maximum pump 
displacement and 1000 psi main loop pressure, represented by point 128 in 
the graph of FIG. 4, and then the load 18 increases so as to increase main 
loop pressure to 3000 psi causing destroking pump moments, pump 
displacement follows the vertical dotted line 130 until the 3000 psi curve 
122' is reached as at point 132. It is noted that points 128 and 132 are 
at the same pump speed, and therefore constant prime mover speed. Line 130 
represents a torque control portion of the control curve under increasing 
load torque. Since the main loop pressure increase between point 128 and 
point 132 is proportionally greater than the decrease in pump 
displacement, transmission operation at point 132 would represent an 
increased torque load on the prime mover 14. This would cause a decrease 
in prime mover speed following the curve 122' until point 134 is reached 
representing a balance between prime mover rpm and thus pump speed, main 
loop pressure, and pump displacement so the actual power consumption at 
point 134 is approximately equal to the power consumption of point 128. 
The control curve between point 132 and 134 along line 122' upon an 
increase transmission loading is actually an anti-stall portion of the 
control curve. The torque control portion and the anti-stall portion 
actually occur simultaneously and form a gradual smooth curve joining 
point 128 and point 134 rather than the two linear portions as depicted to 
graphically represent the understanding of the power control function. 
The power control of the present invention can also be utilized as an 
anti-stall control for a hydrostatic transmission. This is represented in 
the schematic of FIG. 5 and the graph of FIG. 6. Where similar parts are 
utilized, the numbering of the schematic of FIG. 5 is identical to the 
numbering of the schematic of FIG. 1. Again, the power control utilizes an 
input speed pump 30 to establish a first .DELTA.P across the 
characteristic valve 36 to modulate the position of the modifier valve 62 
to provide a control signal at line 70. Construction of the characteristic 
valve 36 and the modifier valve 62 is identical to the construction shown 
in FIGS. 2 and 3 respectively and explained in detail above relative to 
the automotive type control. However, in the anti-stall type control the 
prime mover 14 is generally operated at a fixed preselected speed and pump 
12 displacement is controlled by a displacement control valve 136 
replacing the F-N-R valve 110 of the automotive type control. The 
displacement control valve is modulated by operator input which may be of 
the mechanical, hydraulic or electrical type. Since prime mover rpm is 
generally constant, a constant first .DELTA.P across the characteristic 
valve 36 is established generating a fixed pressure at line 70 which is 
then modulated by the displacement control valve 136 to control the servo 
mechanism 26 establishing the angle of swashplate 24 and thus pump 12 
displacement. 
As long as prime mover input torque is equal to or greater than 
transmission output torque demanded by load 18, the variable displacement 
pump 12 is at a selected displacement, represented by line 138 in graph of 
FIG. 6, as established by the operator input to the displacement control 
valve 136. However, if the torque demanded by load 18 increases above the 
torque available from the prime mover 14, swash plate moments first reduce 
the pump displacement such as shown by line 142, and then the load causes 
the prime mover speed to droop generating a lower speed signal. This 
reduces the first .DELTA.P across the characteristic valve 36 since the 
flow from the fixed displacement pump 30 has now been decreased. Since the 
first .DELTA.P is now lower, the valve spool 78 of the modifier valve 62 
is biased toward the left reducing the flow from line 38' to outlet 70 of 
the modifier valve 62 and thus the available control pressure for the 
displacement control valve 136. This causes a reduction in the 
displacement of pump 12 which follows the linear control curves 140, 140' 
or 140" of FIG. 6 dependent upon the pressure drop across the transmission 
main loop such as to point 144. Since it is desirable to have a relatively 
accurate anti-stall control point, that is as a point wherein pump 
displacement decreases upon prime mover speed droop, the curves 140 and 
140" of FIG. 5 are relatively steep when compared to the control curves 
122 to 122" of FIG. 4 of the automotive control. The steepness of the 
anti-stall control curves is obtained by utilizing a relatively stiff 
spring 48 in the characteristic valve 36. Due to this steepness in the 
anti-stall curves, an equivalent reduction in pump displacement occurs 
over a much smaller engine speed droop, and thus greater accuracy is 
obtained, for the FIG. 6 pure anti-stall control as compared to the engine 
speed droop for the FIG. 4 automotive type control, even though both 
controls operate in a similar manner. The anti-stall control backs off 
pump displacement until the transmission torque absorbed, which is 
proportional to transmission power absorbed, is equal to the prime mover 
torque available, which is proportional to the prime mover power 
available. 
It is noted that in the automotive type control it is desired to increase 
pump displacement as prime mover speed increases above a minimum rpm. 
Therefore, in the graph of FIG. 4 the offset is shown at zero pump 
displacement since the modifier valve spring 72 and the adjustment of the 
characteristic valve spring 48 are established to preselect the start 
point for increasing pump displacement, such as point 126. However, in the 
anti-stall control of FIG. 5 and the graph of FIG. 6, prime mover speed is 
generally fixed and the power control is designed to decrease pump 
displacement when prime mover rpm droops below the predetermined speed. 
Thus, in FIG. 6 the offset is shown as preselected at maximum pump 
displacement even though the offset is still established by the strength 
of spring 72 of the modifier valve 62 and the adjustment of the 
characteristic valve spring 48. 
An optional feature of the anti-stall control is an override valve 142 
which may be added to defeat the anti-stall function. The solenoid 
operated override valve 142 normally connects line 66 to tank 68 so that 
modulation of the modifier valve 62 to proportionally controls the 
pressure at output line 70 between the pressure in line 38' and 
atmospheric or tank pressure 68. When the solenoid override valve 142 is 
energized the valve is moved so as to connect line 38 through line 38'" to 
line 66 so that the output pressure at line 70 is always equal to line 38 
pressure regardless of the modulation of the modifier valve 62. Since line 
38 pressure is established by the charge relief valve 40, line 70 pressure 
is substantially constant when the operation of the override valve 142 is 
selected. 
As can be seen from the above, a power control has been established which 
can provide both an automotive type control function and an anti-stall 
control function. Furthermore, the primary elements of the power control 
comprise two valves, a first characteristic valve and a second modifier 
valve, which are both relatively simple and inexpensive in construction as 
easy to assemble. Since the characteristic valve acts as a variable 
orifice responsive to the .DELTA.P generated across the variable orifice 
and the feedback thereof, the output of the power control is linear in 
nature providing pump displacement which has a relatively linear 
relationship to prime mover speed. 
It can thus be seen that the present invention, as described above, meets 
the objectives of providing an inexpensive, easy to assemble, two valve 
power control which can be used in either an automotive type control or an 
anti-stall control for a hydrostatic transmission and which provides a 
linear control function. The preferred embodiments of the power control as 
specifically described above are illustrative of the concepts of the 
present invention, but not intended to limit the scope thereof.