Constant power variable volume pump

A pump control module is provided which may form an integral part of a variable volume pump or which may take the form of a pump attachment and which is operative to continually vary the volumetric output in inverse proportion to the operating pressure at the pump discharge so as to utilize a high percentage of the available horsepower at maximum output over a portion of the operating range of the pump.

FIELD OF THE INVENTION 
This invention relates to an improved hydraulic pump control to provide a 
simple, low cost method of inversely varying the maximum fluid volume 
output with the output pressure of a pressure compensated hydraulic pump 
over a portion of the operating range. Such a pump could find applications 
for example in portable airless paint pumps where it would be highly 
desirable to continually vary the maximum operating speed in inverse 
proportion to the operating pressure so as to utilize a high percentage of 
the available horsepower over a portion of the operating range. 
BACKGROUND OF THE INVENTION 
Controls for hydraulic pumps to vary the maximum flow with pressure are 
generally complex and because of physical space limitations are available 
only on larger pumps. In portable airless painting equipment the pumps are 
generally small and thus it is common practice to use a simple pressure 
compensated pump. It is then necessary to build different machines for 
high volume/low pressure and for low volume/high pressure applications. A 
hydraulic pump which would automatically and continuously vary the maximum 
flow in inverse proportion to the maximum pressure would allow the use of 
one machine to cover the full range of outputs from low flow/high pressure 
to high flow/low pressure without overloading the power source and at the 
same time to utilize a high percentage of the maximum available power over 
the full range. The subject invention covers a simple and inexpensive 
method of accomplishing this type of control in a pressure compensated 
hydraulic pump. 
The pressure compensated hydraulic piston pump is widely used in industry 
as a means of driving a wide range of hydraulic devices. The pressure 
compensated pump delivers a fixed maximum volume of fluid at pressures 
below the design level and then an abrupt cutoff of the flow as the design 
pressure level is reached. Such pumps are usually equipped with a variable 
pressure control which allows the cutoff pressure to be easily adjusted. 
They can also be equipped with a volumetric control, usually in the form 
of a hand wheel, a lever or an adjusting screw. The volumetric control 
allows the maximum displacement of the pump to be varied independently of 
the pressure control. 
One such type of pump is shown in FIG. 1 This pump is of the axial piston 
type and is representative of the prior art. 
The operation of the pump can be described as follows. The pump is driven 
by an external power source through shaft (1. A cylinder block 2 is 
connected to and rotates with the shaft. A series of pistons 3 in the 
cylinder block rotate with the block and rest at the right hand side 
against a tilted swash plate 4 which is located in the pump housing 5. 
As the cylinder block rotates the pistons move back and forth in it with 
their displacement controlled by the angle of the swash disc 4 to the axis 
of the drive shaft 1. Valve orifices are located in a plate 6 to allow 
entry of fluid into the pump on the backstroke of the pistons and out of 
the pump on the forward stroke of the pistons. The inlet and outlet to the 
pump (not shown) are located in the end cap 7. At a fixed rotational speed 
and fixed swash plate angle the pump will deliver a constant flow of fluid 
at varying pressures. A pressure compensating valve generally designated 8 
senses the hydraulic system pressure and as the pressure approaches the 
maximum operating pressure selected, opens conduit 9 to a flow of fluid 
under pressure. Conduit 9 is connected to cavities 10 and 11 through an 
opening generally designated 12 in a piston 13. The piston 13 is fixed in 
position in the end cap 7. A movable cylinder 14 slides over the piston 
and is in contact with the end of the swash plate 4. The introduction of 
fluid under pressure into conduits 9, 10, 11 and 12 by the action of the 
pressure compensating valve will exert a force on the cylinder 14 and when 
the pressure becomes high enough to overcome the force of spring 15 will 
move the swash plate to the neutral position and pumping will cease. 
An adjustment screw 16 is connected through a rod 17 to contact the inner 
end of cylinder 14. Movement of the adjusting screw inwards will cause the 
rod 17 to limit the travel of cylinder 14 to the left and thereby limit 
the maximum swash plate angle and hence flow. The screw 16 therefore 
provides an independent control of the maximum output of the pump. The 
adjusting screw function is commonly incorporated into a hand wheel or 
lever control. 
The theoretical operating characteristics of a pump of this type are 
generally described in FIG. 2. As the pressure increases to Pmax the 
maximum volume remains constant at Vmax and then abruptly drops to zero. 
The actual operating characteristics will vary slightly from theoretical 
with leakage and the sharpness of the cutoff of the compensator valve. 
The horsepower required to deliver a volume of fluid at pressure is 
directly proportional to the flow times the pressure or 
EQU HP=KVP 
The horsepower required at mix flow will therefore vary linearly from 0 to 
a maximum value as the pressure increases as shown in FIG. 2. 
A hydraulic system driving such a pump can only utilize the maximum 
horsepower of the power source at the maximum pressure and flow condition 
and will utilize considerable less than the maximum horsepower at lower 
pressures. In many applications it would be highly desirable if the 
maximum output flow could be increased as the pressure drops below Pmax so 
as to utilize a greater portion of the available horsepower over a portion 
of the operating range. 
In FIG. 3 the dotted curve indicates the way in which the flow would vary 
for pressures below the maximum if the full horsepower of the source were 
utilized. 
SUMMARY OF THE INVENTION 
It is a principal purpose of this invention to provide a simple module 
which an be mounted on a standard production pressure compensated pump to 
provide improved horsepower utilization output over a portion of the 
pressure range. 
It is a further purpose of this invention to provide a flow control as 
described which does not materially affect the operating characteristics 
of the pressure compensator on the pump. 
Briefly, the present invention carries out the principals of the invention 
through provision of a swash plate control module which is incorporated as 
an integral part of a pressure compensated pump or which is in the form of 
a module which is adapted for installation on a conventional pressure 
compensated pump. The module incorporates a piston energized, spring 
biased plunger which is operated to control the maximum angle of the swash 
plate responsive to the pressure sensed at the discharge port of the pump. 
After a predetermined pressure has been reached as determined by 
precompression of the plunger retarding spring, the plunger will impart 
movement to the swash plate, reducing the stroke of the pistons in direct 
relation to output pressure. 
Other and further features and advantages of the present invention will 
become apparent to one skilled in the art upon consideration of this 
entire disclosure. The form of the invention, which will now be described 
in detail, illustrates the general principals of the invention, but it is 
to be understood that this detailed description is not to be taken as 
limiting the scope of the present invention. 
BRIEF DESCRIPTION OF THE DRAWINGS 
So that the manner in which the above recited features, advantages and 
objects of the present invention are attained and can be understood in 
detail, more particular description of the invention, briefly summarized 
above, may be had by reference to the embodiments thereof which are 
illustrated in the appended drawings, which drawings form a part of this 
specification. 
It is to be noted, however, that the appended drawings illustrate only 
typical embodiments of this invention and are therefore not to be 
considered limiting of its scope, for the invention may admit to other 
equally effective embodiments.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENT 
The invention relates to a simple method of improving the overall 
performance of a pressure compensated pumping system by automatically 
varying the maximum hydraulic fluid flow in inverse proportion to the 
operating pressure. Such a system would prove for much greater utilization 
of the available power over the operating range. One embodiment of the 
invention is shown in FIG. 4. From FIG. 4 it can be seen that an extension 
generally designated 18 has been added to the end plate 7' of the pump 
body. It will be obvious to one skilled in the art that such an extension 
would be added with little change to the structure of the pump. The cavity 
19 in the extension 18 houses a spring 20 which will be generally 
precompressed during assembly. A plunger 21 contacts the spring on surface 
22 and at the opposite end contacts the inside of the movable cylinder 14. 
A plug 23 is screwed into the end of extension 18 and forms a cylinder 24 
in which rides a free floating piston 25 sealed in the cylinder by an 
o-ring 26. A conduit 27 to the left of piston 25 connects the cavity 28 to 
some point at the output of the hydraulic pump to sense the output 
pressure. The piston 25, at its right hand end, contacts the plunger 21. 
The operation of the device can be described as follows. Under zero 
pressure the pump components will rest as shown in FIG. 4. When operating 
at maximum flow below the set pressure, the hydraulic system pressure 
acting on the left side of piston 25 will act against the combined force 
of spring 20 and spring 21. By proper selection of the precompression on 
spring 20 the piston 25 can be restrained from moving until a 
predetermined pressure level is reached. As the pressure is further 
increased the piston 25 will cause the spring 20 to compress and the 
plunger 21 will move to the right, pushing n cylinder 14', spring 21, and 
hence the swash plate 4 to reduce the maximum volumetric output of the 
pump. Correct selection of the spring rate of spring 20 will provide for 
the desired rate of change of maximum volume with pressure. 
As the pressure level is increased the compensator valve will open to admit 
pressurized fluid into conduit 9' and hence inside cylinder 14'. If the 
diameter of piston 25 is smaller than that of the inside diameter of 
cylinder 14' the net effect of this pressure introduction will be to 
increase the loading on springs 20 and 21 and the swash plate will 
immediately be moved to the right and the fluid output decreased. This 
movement will occur as soon as pressure is sensed inside cylinder 14'. 
Under normal operation of a pressure compensated pump the pressure inside 
cylinder 14' must rise to a level sufficient to compress spring 21 before 
the flow reduction starts. The operation of the device with piston 25 
smaller in area than cylinder 14 will result in considerable deterioration 
of the sharp pressure cutoff characteristics which are highly desirable in 
this type of pump. 
If the diameter of piston 25 is made equal to that of cylinder 14' the 
introduction of pressure into cylinder 14' will increase the load on 
spring 20 by acting on the right hand face of cylinder 14' and attempting 
to compress spring 20. At the same time the fluid pressure will decrease 
the load on spring 20 by acting in the reverse direction on piston 25. If 
piston 25 and cylinder 14' are equal in area, the opposing forces will be 
equal and there will be no tendency for piston 25 to move to the right as 
the compensator valve first opens and admits pressure into cylinder 14'. 
When the pressure in cylinder 14' reaches a sufficient level to compress 
spring 21 then normal unloading will take place with no adverse effect 
from the presence of the maximum volume control assembly. 
The use of a piston 25 equal in diameter to that of cylinder 14' can result 
in very high forces on spring 20 by virtue of the full discharge pressure 
acting in cavity 28. The spring forces required can be substantially 
reduced using a stepped piston as shown in FIG. 4A. 
The forward part 25' of the piston in FIG. 4A is of the same size as 
cylinder 14' and hence will provide the advantages described above. The 
opposite end of the piston 31 is reduced in diameter and extends through 
bore 29 to the atmosphere. An O-ring 30, seals the piston extension 31 in 
bore 29 and conduit 27' connects the discharge pressure to cavity 3. In 
this configuration the discharge pressure acts only on an area equal to 
the difference between the cross sectional area of the right hand portion 
(25') of the piston and cross sectional area of the piston extension 31. 
The loading on spring 20' can thereby be significantly reduced without 
affecting the performance of the device. 
The performance characteristics of a pressure compensated pump with the 
improvements shown can be seen in FIG. 5. For purposes of explanation it 
will be assumed that the precompression of spring 20 is adjusted so that 
volumetric control will start at 50% of the maximum pressure and the 
spring rate is set so that the volumetric flow is reduced in half between 
one half and full pressure. It is further assumed that the variation of 
volume with pressure is linear. 
It can be seen from FIG. 5 that the invention provides control of volume 
inversely proportional to pressure over 1/2 the operating range. The 
horsepower requirements shown in the dotted line vary only 12% over a 50% 
change in pressure and volume. 
An improvement in the invention can result if a nonlinear spring or 
combination of springs is used in place of spring 20 in FIG. 4. One such 
configuration is shown in FIG. 6 spring 20' is precompressed so as to 
start movement at 50% of maximum pressure. The spring rate of spring 20 is 
selected to prove a 33% reduction in flow at 75% of maximum pressure. 
Spring 29, inside spring 20 is shorter than spring 20 and is not contacted 
by the head 30 until spring 20 is compressed to its position at 75% of 
maximum pressure. The spring constant of spring 29 is set so that the 
combination of springs 20' and 29 give a flow variation of from 67% to 50% 
between 75% and 100% of maximum pressure. The characteristics of this pump 
are shown graphically in FIG. 7. In this configuration flow versus 
pressure characteristics between 50% and full pressure more closely 
approximate the theoretically required hyperbolic relationship and the 
horsepower variation over a 50% change in pressure and volume is less than 
5%. This is a significant improvement over the characteristics of FIG. 5 
with very little additional mechanical complexity. 
It will be obvious to one skilled in the art and familiar with commercially 
available pressure compensated axial piston pumps that the proposed 
invention can be incorporated with little change to the structure of the 
hydraulic pump. It will also be found that the selection of springs to 
provide the required control can be easily determined. 
In view of the foregoing, it is respectfully submitted that a constant 
power variable volume pump mechanism has been provided herewith which 
accomplishes all of the features and objects hereinabove set forth 
together with other features which are inherent in the valve mechanism 
itself. It will be understood that certain combinations and 
subcombinations of this invention are of utility and may be employed 
without reference to other features and subcombinations. This is 
contemplated by and is within the scope of the present invention.