U.S. patent number 3,811,795 [Application Number 05/322,956] was granted by the patent office on 1974-05-21 for high pressure fluid intensifier and method.
This patent grant is currently assigned to Flow Research, Inc.. Invention is credited to John H. Olsen.
United States Patent |
3,811,795 |
Olsen |
May 21, 1974 |
HIGH PRESSURE FLUID INTENSIFIER AND METHOD
Abstract
A pressure intensifying apparatus to deliver a very high
pressure stream of water through a nozzle. There is a single
working piston having two pressure surfaces of a relatively large
area, the working piston being connected to two high pressure
pistons each having a pressure surface of a relatively small area.
A control valve delivers a high pressure working fluid alternately
to opposite sides of the working piston to cause it to reciprocate
so that the pressure pistons alternately deliver water at high
pressure to the nozzle. In shifting between its two end positions,
the control valve passes through an intermediate position at which
a restricted flow passage is provided for the working fluid, this
restricted flow passage having an effective cross sectional area
relative to the effective area of the discharge nozzle such that
the back pressure of the restricted passage of the control valve
matches the back pressure exerted on the working fluid so that a
substantially constant back pressure is imposed on the high
pressure source of working fluid. Further, there is a valve
shifting mechanism comprising two shifting valves, each of which is
responsive not only to physical contact by the working piston, but
also to pressurization of its related working chamber to cause
rapid shifting of the control valve.
Inventors: |
Olsen; John H. (Vashon,
WA) |
Assignee: |
Flow Research, Inc. (Kent,
WA)
|
Family
ID: |
23257183 |
Appl.
No.: |
05/322,956 |
Filed: |
January 12, 1973 |
Current U.S.
Class: |
417/53; 91/313;
91/306; 417/397 |
Current CPC
Class: |
F03C
1/10 (20130101); F04B 9/10 (20130101); F15B
3/00 (20130101) |
Current International
Class: |
F04B
9/10 (20060101); F04B 9/00 (20060101); F03C
1/00 (20060101); F03C 1/10 (20060101); F04b
035/00 (); F01l 025/04 () |
Field of
Search: |
;91/305,306,307,313
;417/397,403,404,390,53 |
References Cited
[Referenced By]
U.S. Patent Documents
Primary Examiner: Croyle; Carlton R.
Assistant Examiner: Sher; Richard
Attorney, Agent or Firm: Graybeal, Barnard, Uhlir &
Hughes
Claims
1. A hydraulic power apparatus for operation, for example, as a
pressure intensifying system to deliver a high pressure flow of
output fluid, said apparatus comprising:
a. a working cylinder,
b. a working piston mounted in said working cylinder and separating
said working cylinder into first and second working chambers, said
working piston and cylinder being so arranged with inlet and outlet
passage means as to receive pressurized working fluid of at least a
predetermined pressure level alternately into said first and second
working chambers, to cause said piston to reciprocate between
opposite first and second end positions in said cylinder,
c. a control valve means to direct pressurized fluid from a working
fluid source alternately to said first and second working chambers
to cause the reciprocation of said working piston, said control
valve means having first and second positions to direct pressurized
fluid alternately to respectively said first and second working
chambers,
d. a first shifting valve means operatively connected to said
control valve means and having an actuating position where it
causes said control valve means to be positioned in its first
position to deliver pressurized working fluid to said first working
chamber, said first shifting valve having first actuating means to
move said first shifting valve means to its actuating position,
which first actuating means is responsive to the piston moving to
one of its end positions of travel and also responsive to working
fluid in one of said chambers being at said predetermined pressure
level, and
e. a second shifting valve means operatively connected to said
control valve means and having an actuating position where it
causes said control valve means to be positioned in its second
position to deliver pressurized working fluid to said second
working chamber, said second shifting valve means having second
actuating means to move said second shifting valve means to its
actuating position, which second actuating means is responsive to
said piston moving to the other of its end positions of travel and
also responsive to working fluid in the other of said chambers
2. The apparatus as recited in claim 1, wherein the actuating
member of each shifting valve means is positioned to be moved to
its acutating position by engagement of said piston at its
respective end position of travel, and also moved to its actuating
position by exposure to pressure
3. The apparatus as recited in claim 2, wherein said actuating
member extends into an end of its respective working chamber and is
urged by
4. The apparatus as recited in claim 1, wherein said control valve
means has first and second control chambers, which, when
pressurized and depressurized, move said control valve means
between its first and second positions, said first shifting valve
means being operatively connected to the first control chamber so
as to reduce pressure in said first control chamber to move the
control valve means to its first position, and said second shifting
valve means being operatively connected to the second control
chamber so as to reduce pressure in said second control chamber
to
5. The apparatus as recited in claim 4, wherein there are other
positioning means for said valve control means, which other
positioning means urges said control valve means to an intermediate
position between said first and second positions, where working
fluid at said predetermined pressure
6. The apparatus as recited in claim 5, wherein said other
positioning means comprises spring means which resiliently urges
said control valve
7. The apparatus as recited in claim 5, wherein said control valve
means at its intermediate position functions to connect both said
first and second
8. The apparatus as recited in claim 4, further comprising other
valve operating means having a first position wherein said first
control chamber is exposed to low pressure, and a second position
wherein said second control chamber is exposed to low pressure,
whereby said control valve means can be moved between its first and
second positions by operation of
9. A hydraulic power apparatus for operation, for example, as a
pressure intensifying system, to deliver a high pressure stream of
output fluid, said apparatus comprising:
a. a working cylinder,
b. a working piston mounted in said working cylinder and separating
said working cylinder into first and second working chambers, said
working piston and cylinder being so arranged as to receive
pressurized working fluid of at least a predetermined pressure
level alternately into said first and second working chambers, to
cause said piston to reciprocate between opposite first and second
end positions in said cylinder,
c. a control valve means to direct pressurized working fluid from a
working fluid source alternately to said first and second working
chambers to cause the reciprocation of said working piston,
1. said control valve means having a first position where it
delivers pressurized working fluid to said first chamber, a second
position where it delivers pressurized working fluid to said second
chamber, and a third intermediate position where said first and
second working chambers are connected to a lower pressure area,
2. said control valve means having first and second control
chambers, which, when pressurized and depressurized, moves said
control valve means between its first and second positions,
3. said control valve means having other valve positioning means
arranged to urge said control valve means to its intermediate
position,
d. a first and second valve shifting means operatively connected to
the control chambers of the control valve means in a manner to
cause movement of the control valve means between its first and
second positions, the first valve shifting means having an
actuating position in which it is located by said piston moving to
one of its end positions of travel and also by action of said
pressurized working fluid at said predetermined level in one of
said chambers against said shifting valve means, said second
shifting valve means also having an actuating position at which it
is located by action of the piston moving to the other of its end
positions of travel and also by pressurized working fluid at said
predetermined level in the other of said working chambers acting
against
10. The apparatus as recited in claim 9, wherein said first
shifting valve means is operatively connected to the first control
chamber so that with said first shifting valve means in its
actuating position pressure is reduced in said first control
chamber so as to cause movement of the control valve means to its
first position, and said second shifting valve means being
operatively connected to said second control chamber so that with
said second shifting valve means in its actuating position pressure
is reduced in said second control chamber to move said control
valve means
11. The apparatus as recited in claim 10, wherein each of said
shifting valve means comprises an actuating member which is
positioned to be moved to its actuating position by engagement with
said piston at its related end position of travel and also
positioned to be exposed to pressure in
12. The apparatus as recited in claim 11, wherein said other
positioning means comprises spring means which resiliently urges
said control valve
13. The apparatus as recited in claim 9, further comprising other
valve operating means having a first position wherein said first
control chamber is exposed to low pressure, and a second position
wherein said second control chamber is exposed to low pressure,
whereby said control valve means can be moved between its first and
second positions by operation of said other shifting valve
means.
14. A hydraulic power apparatus for operation as a pressure
intensifying system to deliver a high pressure flow of output
fluid, said apparatus comprising:
a. a working cylinder,
b. a working piston mounted in said working cylinder and separating
said working cylinder into first and second working chambers, said
working piston and cylinder being so arranged with inlet and outlet
passage means as to receive pressurized working fluid of at least a
predetermined pressure level alternately into said first and second
working chambers, to cause said piston to reciprocate between
opposite first and second end positions in said cylinder,
c. first and second high pressure output piston means operatively
connected to said working piston so as to be moved along
reciprocating paths thereby to deliver a high pressure flow of
output fluid,
d. a control valve means to direct pressurized working fluid from a
working fluid source alternately to said first and second working
chambers to cause the reciprocation of said working piston,
1. said control valve means having a first position where it
delivers pressurized working fluid to said second chamber to cause
said working piston to move to its first end position, a second
position where it delivers pressurized fluid to said first working
chamber to cause said working piston to move to its second end
position, and a third intermediate position where said first and
second working chambers are connected to a lower pressure area,
2. said control valve having first and second control chambers,
which, when pressurized and depressurized, moves said control valve
means between its first and second positions,
3. said control valve means having other valve positioning means
arranged to urge said control valve means to its intermediate
position,
e. a first valve shifting means having an actuating position and
operatively connected to said first control chamber so as to
depressurize said first control chamber when in its actuating
position so as to initiate movement of said control valve means
from its first position toward its second position, said first
valve shifting means being responsive to move to its actuating
position both by said working piston moving to its first position
and by pressure in said first working chamber being at said
predetermined pressure level,
f. a second valve shifting means having an actuating position and
operatively connected to said second control chamber so as to
depressurize said second control chamber when in its actuating
position so as to initiate movement of said control valve means
from its second position toward its first position, said second
valve shifting means being responsive to move to its actuating
position both by said working piston moving to its second position
and by pressure in said second working chamber being at said
predetermined pressure level,
whereby when said working piston reaches one of its end positions
of travel, the working piston moves a proximate shifting valve to
its actuating position to cause movement of said control valve
means through its intermediate position, thereby causing a pressure
reduction in the previously pressurized working chamber so that the
other of said shifting valve means moves to its non-actuating
position and causes continued movement of said control valve means
to its opposite position, thereby effecting rapid response of said
control valve means between its first and
15. The apparatus as recited in claim 14, wherein said other valve
positioning means for said control valve means comprises means to
urge
16. The apparatus as recited in claim 14, wherein each of said
shifting valve means comprises an actuating member extending into a
related working chamber, whereby each of said actuating members is
responsive to pressure in its related chamber as well as to
engagement of said working piston.
Description
BACKGROUND OF THE INVENTION
1. Field of the Invention
This invention relates to a high pressure fluid intensifier, such
an an intensifier arranged to deliver a stream of fluid at a very
high pressure, to accomplish a function such as cutting, drilling
or waterblast cleaning.
2. Description of the Prior Art
There are in the prior art various pressure intensifying systems
where a larger working piston is reciprocated to provide a high
pressure output through smaller high pressure pistons. One of the
problems associated with many such prior art devices is that of
obtaining a rapid reversal of the working piston so that there is
no significant interruption of the flow of high pressure output
fluid. Yet another problem in the prior art is that of alleviating
pressure surges in the working fluid while the working piston is
being reversed. Aside from these operational problems, there are,
especially when very high pressures are used, safety considerations
with respect to one of the hydraulic lines or other components
breaking or rupturing.
Typical of the prior art devices which show pressure intensifying
systems and various valve switching mechanisms adaptable for such
systems are the following U.S. Pats.: Atkinson, No. 153,296; West
et al., No. 2,000,805; Rethmeier, No. 2,942,584; Murray, No.
3,045,611; Pennther, No. 3,540,349; and Bowen, No. 3,565,191.
It is an object of the present invention to provide a high pressure
fluid intensifier having desirable operating features, particularly
with respect to the problems and considerations mentioned
above.
SUMMARY OF THE INVENTION
In the apparatus of the present invention, there is a working
cylinder in which a working piston is mounted for reciprocating
motion, the piston dividing the cylinder into first and second
working chambers. Two high pressure pistons are connected to the
working piston in a manner that reciprocation of the working piston
causes a flow of high pressure fluid to be produced alternately
from the two high pressure pistons. The high pressure flow is
directed through a discharge nozzle to produce a high velocity
stream of water. To cause reciprocation of the working piston,
there is a control valve having first and second positions to
deliver pressurized working fluid to, respectively, the first and
second working chambers.
According to one facet of the present invention, the control valve
has a third intermediate position through which it passes in moving
between its first and second positions. In the intermediate
position pressurized working fluid is directed through a pressure
reducing flow passage, which produces a back pressure substantially
balancing the back pressure resulting from transmitting power
through the working piston and the pressure intensifying pistons to
produce a high pressure fluid flow through the output nozzle. Thus,
during reversal of the working piston when the control valve passes
through its intermediate position, any substantial surge of back
pressure against the working fluid source is alleviated.
In accordance with another facet of the present invention, there
are two shifting valves to cause the control valve to move between
its first and second positions. Each shifting valve is made
responsive not only to movement of the piston to an end limit of
travel proximate the shifting valve, but is responsive also to
pressurized working fluid in its respective chamber being at a
predetermined level. Actuation of one or the other of the shifting
valves causes a pressure imbalance at the control valve to cause
rapid shifting of the control valve. A centering spring urges the
control valve toward its intermediate position. If there is a
pressure reduction in the working chamber that is pressurized at
any particular moment, for example by a rupture in one of the high
pressure output lines, the decrease in pressure causes deactivation
of its related shifting valve to move the control valve to its
intermediate position and stop further operation of the
apparatus.
Other more specific features of the present invention will become
apparent from the following detailed description.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a semi-schematic side elevational view of the overall
apparatus of the present invention;
FIG. 2 is a view partly in sections of the pressure intensifying
apparatus of the present invention;
FIG. 3 is a sectional view of one of the shifting valves of the
apparatus of FIG. 2;
FIGS. 4-6 are semi-schematic drawings illustrating the operating
sequence of the present invention;
FIGS. 7A-7E are a series of semi-schematic drawings showing the
sequence of operation of the control valve of the present
invention;
FIG. 8 is a graph illustrating the flow characteristics in the
control valve in the sequence of operation of FIGS. 7A-7E;
FIG. 9 is a graph illustrating the pressure characteristics of the
control valve in the sequence of operation of FIGS. 7A-7E;
FIG. 10 is a semi-schematic illustration of a second embodiment of
the control valve of the present invention;
FIG. 11 is a semi-schematic drawing of a third embodiment of the
control valve of the present invention.
DESCRIPTION OF THE PREFERRED EMBODIMENTS
With reference to FIG. 1, there is shown an electric motor 10 which
drives a hydraulic pump 14, which in turn supplies working fluid to
a pressure intensifier unit 16. The intensifier 16 draws fluid
(i.e., water) from a suitable source, such as a reservoir 18, and
discharges the water at a very high pressure through an output,
which as shown herein is a tube 20 with a small area exit nozzle
22. This results in a discharge of a fluid jet stream of a small
diameter (e.g., 0.03 inches) and a very high velocity (e.g., 1,200
feet per second or greater).
In describing the present invention in detail, first the physical
components of the pressure intensifying unit 16 will be described
with reference to FIGS. 2 and 3. Second, the overall operation of
the total apparatus will be described with reference to the
sequential schematic drawings of FIGS. 4 through 6. Thereafter, the
precise manner in which the control valve 38 functions to
accomplish the proper operation of this apparatus will be described
in more detail with reference to the sequential illustrations of
FIGS. 7A through 7E, and the two graphs of FIGS. 8 and 9, with two
modified valves being shown in FIGS. 10 and 11.
The physical components of the pressure intensifying unit 16 are
illustrated in FIG. 2. For clarity of illustration, the various
fluid lines and passages built into or attached to the unit 16 are
not illustrated in FIG. 2, but rather are shown schematically in
FIGS. 4 through 6. With reference to FIG. 2, the pressure
intensifying unit 16 comprises a main housing 24, comprising a main
cylinder 26, right and left end bell members 28 and 30,
respectively, mounted to the ends of the cylinder 26, and right and
left high pressure cylinders 32 and 34, respectively, threaded into
respective bells 28 and 30. Connected to the housing 24 is a
manifold block 36 on which is mounted a flow control valve 38.
Mounted for reciprocating motion within the housing 24 is a unitary
piston assembly 40. This assembly comprises a larger diameter
central working piston 42 mounted within the main cylinder 26, and
right and left high pressure pistons 44 and 46, respectively,
extending oppositely from the center working piston 42. The working
piston 42 divides the interior of the main cylinder 26 into right
and left working chambers 48 and 50 respectively. The high pressure
piston 44 reciprocates in the right high pressure chamber 52
defined by the right cylinder 32, while the left high pressure
piston 46 reciprocates in the left high pressure chamber 54 defined
by the other cylinder 34.
The aforementioned control valve 38 comprises a valve housing 56
defining a transfer chamber 58, in which is slide mounted a valve
spool 60. In the housing 56 is a centrally located high pressure
fluid inlet port 62, right and left transfer ports 64 and 66,
respectively, on opposite sides of the inlet port 62, and right and
left low pressure outlet ports 68 and 70, respectively, positioned
outside of the two transfer ports 64 and 66.
The valve spool 60 comprises right and left lands or pistons 72 and
74, respectively, and right and left outermost end closure members
76 and 78, respectively. Outside the two closure members 76 and 78
are respective right and left centering springs 80 and 82,
respectively, which urge the spool 60 to its center position in the
housing 56; each of the springs 80 and 82 has a stop collar 83
engaging a stop shoulder 83a to prevent either spring 80 or 82
urging the valve spool 60 beyond its center position.
The center port 62 is connected to a high pressure line from the
pump 14, while the ports 68 and 70 are connected to the low
pressure return line of the pump 14. The right transfer port 64
connects to the right working chamber 48, and the left transfer
port 66 connects to the left working chamber 50. The groove or
chamber 84a located between the two piston elements 72 and 74 is a
high pressure fluid transfer chamber, and functions to direct high
pressure fluid from the port 62 to either the right transfer port
64 or the left transfer port 66 when in, respectively, a right or
left hand position. The right piston 72 and the right closure
piston 76 define a groove or chamber 84b which is a low pressure
transfer chamber that functions to connect the transfer port 64
with the low pressure outlet port 68 when the spool 60 is in its
left hand position. In like manner, the left transfer piston 74 and
left closure piston 78 define therebetween a groove or chamber 84c
which functions to connect the left transfer port 66 with the low
pressure outlet port 70 when the spool 60 is in its right hand
position.
To move the spool 60 of the valve 38 between its right and left
positions, there are two shifting valves 85 and 86, respectively,
located in, respectively, the right and left bell sections 28 and
30 of the housing 24. For convenience of illustration only the
right shifting valve 85 is shown in section in FIG. 2 (the valve 85
and 86 being substantially identical). Each of the valves 85 and 86
comprises a sleeve 88 in which is slidably mounted a plug 90. There
is an inlet port 92 and an outlet or venting port 94, the port 92
being closed from the venting port 94 when the valve is in the
closed position shown in FIG. 2. The plug 90 is urged to its closed
position by a compression spring 96. To open the valve 85 or 86
there is provided an actuating pin 98 which butts against the plug
90 and extends through the housing to project into the end of its
respective working chamber 48 or 50. A locating stop 100 on the pin
98 properly positions the plug 90 in its closing position with the
pin 98 extending into the chamber 48 or 50. Through holes 101 in
the plug 90 permit flow from the inlet port 92 to the venting port
94 when the plug 90 is pushed by pin 98 against the urging of the
spring 96 to its open position shown in FIG. 3. The inlet port 92
of the right shifting valve 85 is connected through an end opening
102 in the housing 56 of the valve 38 to a right control chamber
104 at the right end of the spool 60 of the valve 38, while the
inlet port 92 of the left shifting valve 86 is similarly connected
to the left control chamber 105 through opening 103 in the left of
the valve 38.
In describing the operations of this apparatus, for convenience the
overall operation will first be described and then the means for
starting the pumping action. In FIG. 4, the valve 38 is in its left
hand position, so that high pressure fluid from the high pressure
line 106a of the pump 14 is directed into the left working chamber
50, while the right working chamber 48 is connected through a low
pressure return line 106b to the fluid reservoir 107 of the pump
14. This causes the working piston 42 to move to the right, as seen
in FIG. 4, which in turn causes the right high pressure piston 44
to force output fluid from the high pressure chamber 52 through a
check valve 108 and out the discharge nozzle 22. At the same time,
output fluid is being drawn from the source 18 through a check
valve 109 into the left high pressure chamber 54.
The pumping pressure of the pump 14 is sufficiently large, relative
to the force of the return springs 96 and the cross sectional area
of the actuating pins 98 of the shifting valves 85 and 86, that
when one of the working chambers 48 or 50 is pressurized by the
pump 14, the resulting pressure on its related actuating pin 98 is
sufficient to force the pin 98 against the urging of its related
spring 96 to open the valve 85 or 86. A portion of the high
pressure working fluid is directed from the pump 14 to the valve
control chambers 104 and 105 through respective restricted flow
orifices 110 and 112, and also to the high pressure inlet ports 92
of the shifting valves 85 and 86. The venting port 94 of each
shifting valve 85 or 86 is connected to the pump reservoir 107.
Thus, when one of the valves 85 or 86 is closed, its related valve
control chamber 104 or 105, respectively, is pressurized, but when
one of the valves 85 or 86 is open, its respective valve control
chamber 104 or 105 is depressurized.
With reference to FIG. 4, since the working chamber 50 is
pressurized from the pump 14, the shifting valve 86 is open so that
the left valve control chamber 105 is depressurized. Since the
working chamber 48 is connected to the low pressure reservoir 107,
the actuating pin 98 protrudes into the chamber 48, so that the
shifting valve 85 is closed, with the right valve control chamber
104 being pressurized. The difference in pressure in the two valve
control chambers 104 and 105 holds the spool element 60 of the
valve 38 in its left hand position against the urging of the spring
82 as seen in FIG. 4.
As the working piston 42 continues to move to the right, it
approaches its end limit of travel as shown in FIG. 5. Near its end
limit of travel, the piston 42 engages the actuating pin 98 to push
the pin 98 and the plug 90 of the shifting valve 85 to its open
position to depressurize the right valve control chamber 104. With
both the control chambers 104 and 105 depressurized, the left
centering spring 82 pushes the spool element 60 toward its center
position, as shown in FIG. 5.
When the spool element 60 reaches its center position, two things
occur. First, both of the working chambers 48 and 50 become
connected to the low pressure pump reservoir 107 through the low
pressure return line 106b so as to be depressurized. Second, the
high pressure supply line 106a from the pump 14 becomes connected
through a restricted flow passage, indicated schematically at 114,
to permit limited flow from the pump 14 back to the pump reservoir
107. The particular manner in which this is accomplished and how
this alleviates pressure surges in the high pressure supply line
will be described later herein in a more detailed description of
the functioning of the valve 38.
The immediate effect of depressurizing both of the working chambers
48 and 50 is to permit the spring 96 to move the left shifting
valve 86 outwardly toward the left working chamber 50 to close the
left shifting valve 86 and thus cause immediate pressurization of
the left valve control chamber 105. This immediately causes the
spool element 60 to continue movement through its center or
intermediate position to its right position, shown in FIG. 6, where
high pressure working fluid is delivered to the working chamber 48,
with the left working chamber 50 being connected to the low
pressure line 106b leading to the pump reservoir 107.
With the right working chamber 48 now pressurized, the working
piston 42 moves to the left so that the left high pressure piston
46 forces output fluid from the left high pressure chamber 54
through a check valve 116 and out the output nozzle 22. At the same
time, additional output fluid is being drawn into the right high
pressure chamber 52 through a check valve 118. With the right
working chamber 48 pressurized, as the working piston 42 moves away
from physical engagement with the right actuating pin 98, the
pressure of the working fluid in the chamber 48 keeps the right
actuating pin 98 in its retracted position to maintain the right
shifting valve 85 in its open position and maintain the
depressurization of the right valve control chamber 104 so that the
spool element 60 remains in its right position, as shown in FIG. 6
because of the pressurization of the left valve control chamber 105
due to the left shifting valve 86 being closed.
When the working piston 42 moves toward its left end position to
engage the left actuating pin 98, a similar shifting sequence
occurs, as described with reference to FIG. 5, to reverse the fluid
flow in the working chambers 48 and 50 and cause the working piston
42 to begin movement back to the right.
To initiate operation of the apparatus, there is provided a
starting valve, this being shown schematically at 120. The valve
120 has an up position where the right valve control chamber 104 is
directly connected to the reservoir 107, a down position where the
left valve control chamber 105 is connected to the reservoir 107,
and an intermediate position where the valve 120 provides no
operative connection to the chambers 104 and 105. In the three
illustrations of FIGS. 4 through 6, the starting valve 120 is shown
in its center position where it has no effect on the operation of
the apparatus. To describe the operation of the starting valve 120,
let it be assumed that the pump 14 has been turned off and the
entire system has become depressurized, with the spool element 60
of the valve 38 returning to its center position by action of the
centering springs 80 and 82. Further, let it be assumed that the
working piston 42 is in some intermediate position, as shown in
FIG. 4.
When the pump 14 is started, with the valve spool element 60 in its
center position, neither of the working chambers 48 or 50 becomes
pressurized. However, both of the valve control chambers 104 and
105 are pressurized, since both of the shifting valves 85 and 86
remain closed. These circumstances contribute to the safety of the
operation in that the intensifying unit 16 will not inadvertently
start pumping when the hydraulic pump is started, provided that
piston 42 is not at that moment holding either valve 85 or 86 open.
It should be noted that when the motor 10 is turned off to stop the
pumping action, there is a gradual decline of hydraulic pressure
due to the inertia of the motor and pump. If the motor is turned
off and the hydraulic pressure has declined below the level
required to hold pin 98 in against spring 96 while the working
piston 42 is in contact with one of the pins 98 of one or the other
of the shifting valves 85 or 86, this will cause the control valve
38 to remain in either its right or left position to pressurize the
working chamber 48 or 50 at which the piston 42 is depressing the
pin 98. This in turn causes the piston 42 to move out of engagement
with that pin 98 allowing the valve 85 or 86 to close and cause the
control valve 38 to return to its center position where neither of
the working chambers 48 and 50 is pressurized. Thus, piston 42 will
stop in a position remote from pins 98, and when the motor 10 is
restarted, the pumping action will not start. By pushing the valve
120 to its down position, the left control chamber 105 becomes
exposed to the low pressure pump reservoir 107 so that the high
pressure in the right valve control chamber 104 causes the valve
element 60 to move to the left (i.e., the position shown in FIG.
4.) to cause the piston 42 to move to the right. As soon as the
working piston 42 reaches its full right position to engage the
right actuating pin 98, the normal shifting sequence goes into
effect as described above herein. By moving the starting valve 120
to its up position to cause the spool element 60 to move to the
right, the working piston 42 can be caused to move to the left. So
if there is some reason that the normal shifting sequence of the
apparatus does not function as described above, for example by
reason of excess air in the high pressure lines, then the manual
starting valve 120 can be used to move the working piston 42 back
and forth to clear the hydraulic lines so that the normal shifting
sequence becomes operative, with the working piston 42 then
reciprocating with automatic shifting of the valve 38 as described
above.
In the event that there is a break in one of the high pressure
output lines to the nozzle 22, there will be an immediate drop in
pressure in the output chamber 52 or 54 that is at the time
pressurized, and a corresponding drop in pressure in the related
working chamber 48 or 50, which happens to be pressurized at that
particular instant. When there is such a pressure drop in the
pressurized working chamber 48 or 50, its related actuating pin 98
is moved outwardly by spring 96 to its valve closing position to
close its related shifting valve 85 or 86 to pressurize its related
valve control chamber 104 and 105 (both chamber 104 and 105 then
being pressurized) so that the valve element 60 returns to its
center position by action of the centering springs 80 and 82 to
vent both working chambers 48 and 50 to low pressure and halt
movement of the working piston 42. Thus, in the event of any break
in the high pressure lines, the system immediately shuts itself
off.
To describe in more detail the operation of the control valve 38,
reference is now made to FIGS. 7A through 7E.
In FIG. 7A, the spool element 60 is shown in its full left position
(shown schematically in FIG. 4), where the right piston 72 is
positioned between the high pressure port 62 and the right transfer
port 64 so as to block any flow therebetween, while the left piston
74 is positioned so as to block any flow from the left transfer
port 66 into the low pressure outlet port 70. In this position,
there is free flow from the high pressure port 62 to the left
transfer port 66 through the center high pressure chamber 84a to
pressurize the working chamber 50 as described above. Also, the
right transfer port 64 communicates with the low pressure port 68
through the right low transfer chamber 84b so that the right
working chamber 48 is depressurized. As described above, the spool
element 60 remains in this position until working piston 42 reaches
its extreme right end of travel to cause shifting of the valve
spool element 60 to the right.
In FIG. 7B, the valve element 60 is shown moving from its extreme
left hand position through a position where it is just about to
enter its intermediate position. It will be noted that the
laterally outermost portion 122 of the circumferential surfaces of
each of the pistons 72 and 74 is substantially cylindrical so that
it fits against the inner cylindrical surface of the housing 56.
However, the laterally inward circumferential surface portion 124
of each of the pistons 72 and 74 (i.e., those surface portions
closer to the center of the spool element 60) are tapered very
moderately inwardly toward the middle of the spool element 60. For
purposes of illustration, this taper is shown to be at a somewhat
larger angle than normally used, this taper ordinarily being in the
order of one degree from the longitudinal axis of the spool element
60. It can be seen that in the position shown in FIG. 7B the
tapered surface 124 of the left piston 74 forms with the inner edge
126 of the left transfer port 66 a restricted circumferential flow
passage 128.
As the piston continues to move from the position of FIG. 7B toward
the position of FIG. 7C, where the spool element 60 is centered, a
passage opens at 130 from the left transfer port 66 through the
left low pressure transfer chamber 84c to the low pressure outlet
port 70. Because of the very shallow taper of the surface 124, as
soon as the spool element 60 moves a very short distance from the
position of FIG. 7B, the passageway 130 has a much larger cross
sectional area than the passage of 128, so that there is a large
pressure drop from the inlet port 62 across the passageway 128, and
the pressure in the transfer port 66 almost immediately drops to
the pump reservoir pressure that exists in the outlet port 70. As
the spool 60 continues to move to the right from position 7B to
that of 7C, the left flow passage 128 becomes more restricted,
while the right flow passage 128 defined by the piston 72 with the
housing 56 becoming less restricted. Since the tapered surfaces 124
are both uniform, the rate of decrease of the cross section of the
left restricted passage 128 is substantially the same as the rate
of increase of the cross sectional area of the right restricted
flow passage 128, so that the total flow rate through both passages
128 is constant. This is illustrated in the graph of FIG. 8, where
the flow through the left restricted flow passage 128 is indicated
at a and the right restricted flow passage 128 is indicated at b,
with the combined flow through both passages 128 being shown by the
dotted line at c.
In the graph of FIG. 9, the pressure in the left transfer port 66
is indicated at a, while the pressure in the right transfer port 64
is indicated at b, in the travel of the spool element 60 from the
position of FIG. 7A to that of FIG. 7E.
When the spool element 60 is traveling through its intermediate
position (i.e., from the position of FIG. 7B, through the position
of FIG. 7C to the position of FIG. 7D), both the right and left
transfer ports 64 and 66 are at the low pump reservoir pressure
since the flow passages 130 have a substantially larger cross
sectional flow area than the passages 128 (in the order of perhaps
one hundred times as great when the spool element 60 is in the
position of FIG. 7C).
As described previously herein, as soon as the pressurized working
chamber 48 or 50 becomes depressurized, there is an immediate
pressure imbalance in the two valve control chambers 104 and 105,
which causes the spool element 60 to continue movement through its
center postion to its other end position. When the spool element
reaches the position shown in FIG. 7D, the spool element 60 is now
moving from its intermediate position to its right position. At
this point, the left restricted flow passage 128 is being
completely closed off, while the right restricted flow passage 128
has reached its maximum effective cross sectional flow area.
Simultaneously, the flow path from the right transfer port 64 into
the right low pressure outlet port 68 is being abruptly closed off
by the right piston 72 so that there is an abrupt rise in the
transfer port 64 from pump reservoir pressure to high pressure. As
the spool element 60 continues movement to the right to the extreme
right position of FIG. 7E, there is substantially unrestricted flow
from the high pressure port 62 to the transfer port 64 to cause
pressurization of the right working chamber 48.
It is important to note that the total cross sectional area of the
two restricted flow passages 128 remains substantially constant as
the spool element 60 is moving through its intermediate phase from
the position of FIG. 7B to that of 7D. These two restricted flow
passages 128 are, in effect, the same restricted flow passages
indicated at 114 in the schematic drawings of FIGS. 4-6. The
effective combined cross sectional area of the two flow passages
128 is so dimensioned that the back pressure exerted at the
passages 128 is substantially the same as the pressure existing in
either of the working chambers 48 and 50 when pressurized with the
control valve 38 in either its right or left position. The effect
of this is that when the valve element 60 is moving from its right
position through its intermediate position to its left position or
vice versa, the back pressure exerted on the high pressure line
from the pump 14 remains substantially constant, with the only
pressure surges experienced being those caused by inertial forces
of reversing the working piston 42.
The pressure balance is accomplished by properly selecting the
effective total cross sectional area of the passages 128 relative
to the effective cross sectional flow area of the output nozzle 22
and also relative to the effective pressure areas of the working
piston 42 and the high pressure pistons 44 and 46, and to the
frictional forces acting on the piston assembly. To explain this
relationship let us first assume these frictional forces are
negligable and represent only a small correction to the
relationship. It should be pointed out that the ratio of the
pressure in either of the high pressure chambers 52 or 54 to the
pressure in the working chamber 50 is inversely proportional to the
area of the piston 44 (which is the square of the radius of the
piston 44 times .pi.) and proportional to the working area of the
low pressure piston 42 (which is the square of the radius of the
piston 42 times .pi. minus the cross sectional area of the piston
44). The pressure drop through the output nozzle 22 is proportional
to the square of the average velocity of fluid flow from the nozzle
22 times the density of the output fluid (i.e., water). Likewise,
the pressure drop across the flow passages 128 is proportional to
the square of the average flow velocity of the fluid through the
passages 128 times the density of the working fluid from the pump
14. Thus, the effective cross sectional flow area of the passages
128, or of the passage shown schematically at 114, should be
proportional to the effective cross sectional flow area of the
nozzle 22 times the working area of the low pressure piston 42,
divided by the pressure area of either of the pressure pistons 44
or 46, multiplied by the square root of the ratio of the pressure
area of the working piston 42 to the pressure area of the high
pressure piston 44 and 46, multiplied by the square root of the
ratio of the density of the working fluid to the density of the
output fluid. Expressed mathematically, this relationship can be
stated as follows:
C.sub.v A.sub.v = (C.sub.n A.sub.n A.sub.w /A.sub.p) (.sqroot.
A.sub.w /A.sub.p) .sqroot.D.sub.w /D.sub.o
A.sub.w = the effective pressure area of the working piston 42
A.sub.p = the effective pressure area of the pressure piston 44 or
46
A.sub.n = the effective cross sectional flow area of the nozzle
22
A.sub.v = the effective flow area of the restricted flow passages
128 of the control valve
C.sub.n = orifice discharge coefficient of nozzle
C.sub.v = orifice discharge coefficient of restricted flow passages
128 of the control valve
D.sub.w = the density of the working fluid
D.sub.o = the density of the output fluid
To give a numerical example, let it be assumed that the densities
of the working fluid and output fluid are approximately equal, that
the discharge coefficients are equal, and that the effective
working area of the piston 42 is approximately four times the
working area of either of the pressure pistons 44 or 46. The total
flow of working fluid into either of the low pressure chambers 48
or 50 would be four times the flow from either of the high pressure
chambers 52 and 54. According to the above formula, the total
effective cross sectional flow area of the restricted valve
passageway (either the passages 128 or the schematic passage 114),
should be eight times the effective cross sectional flow area of
the nozzle 22. In practice the passages 128 would be chosen
slightly smaller to correct for the additional pressure required
for overcoming friction.
In FIG. 10, there is shown a second embodiment of the control valve
of the present invention, in which components corresponding to the
valve 38 of the first embodiment will be given like numerical
designations, with an a suffix distinguishing those of the second
embodiment. Thus there is a housing 56a in which is mounted a valve
spool 60a, comprising right and left pistons 72a and 74a,
respectively, and right and left outermost end closure members 76a
and 78a. However, instead of forming the inner circumferential
surface portions of the pistons 72a and 74a with a tapered surface,
these surface portions of the piston element 72a and 74a are
stepped radially inwardly as at 124a. Thus, when the valve element
60a passes through its intermediate position (corresponding to the
movement of the valve element of the first embodiment as it moves
from the position of FIGS. 7B to the position of FIG. 7D), there
are two restricted flow passages 128a, through which there are
substantially equal flows. As in the first embodiment, the combined
effective cross sectional flow area of the two passages of 128a is
such as to approximate the back pressure exerted from the output
nozzle back through the system to the pressurized working
fluid.
A third embodiment of the control valve of the present invention is
illustrated in FIG. 11. In this valve, there is a housing 140 in
which there is a movable valve element 142 comprising a center
piston 144 and two side pistons 146 and 147, respectively, and two
end closure pistons 148 and 149, respectively. There is a center
high pressure inlet port 150, right and left transfer ports 152 and
154, respectively, on opposite sides of the port 150, and right and
left low pressure ports 156 and 158, respectively, located outside
the ports 152 and 154.
These components are arranged so that when the valve element 142 is
in the right hand position, the right transfer port 152
communicates with the low pressure port 156 to permit outflow of
working fluid from its related working chamber, while high pressure
fluid is being directed from the port 150 through the transfer port
154 to pressurize the other working chamber. With the valve element
142 in its left hand position, the opposite situation occurs, with
high pressure fluid being directed into the port 152 and the port
154 being connected to low pressure.
The circumferential surface of the center piston 144 is stepped
radially inwardly at 160 at both of its circumferential portions
laterally outward from the middle of the center piston 144. These
two stepped surfaces 160 form with the housing 140 restricted flow
passages 162 which perform substantially the same function as the
aforementioned flow passages 128 and 128a of the first and second
embodiments. For this reason, no detailed description will be
provided of the operation of this valve of the third embodiment,
since it is apparent from the description of the operation of the
prior two embodiments of the control valve.
* * * * *