Variable output vane pump

A variable output vane pump for use with an automotive automatic transmission. At least one pressure chamber is formed between the circular bore of a housing and the outer circumference of a ring by means of at least one seal pin which is fitted in the axial groove formed in the ring circumference and which is opened in the housing bore. The ring is urged by the action of a spring, which is fitted in the housing, in a direction opposite to the pressure to be built up by the pressure chamber. An elastic member is fitted in the bottom of the axial groove to urge the seal pin both toward the housing bore and at the same time toward one of the side walls of the axial groove. The urging direction of the elastic member and a tangential line, on which the seal pin is tangential to the housing bore, are angularly positioned at an acute angle with respect to each other.

BACKGROUND OF THE INVENTION 
1. Field of the Invention 
The present invention relates to a variable output vane pump and, more 
particularly, to a variable output vane pump which is suitable for use 
with an automotive automatic transmission. 
2. Description of the Prior Art 
A hydraulic pump for an automatic transmission according to the prior art 
generally uses an inner or outer gearing of the constant output type. 
Since, in this case, an output proportional to the r.p.m. of the engine 
can be attained, it becomes more than necessary for a high r.p.m. so that 
most of it is returned from a flow regulator to an oil tank. The energy 
loss at this portion raises a problem. In order to solve this problem, it 
is necessary to make the pump of the variable output type so that the 
theoretical output of the pump itself may be reduced for an r.p.m. higher 
than a predetermined value. 
The variable output pump thus far described is exemplified by a variable 
displacement pump, for example, which is disclosed in U.S. Pat. No. 
4,342,545. This pump is an improvement over the prior art, but a 
displacement control chamber is formed between a ring and the inner wall 
of a housing so that seal portions are formed at two positions, i.e., 
between the ring and a pivot pin providing a pivot for the ring and 
between the ring and the pump housing. The sealing performance of the vane 
pump is so influenced by an error in the shape of the circular bore of the 
housing and by an error in the shape of the outer circumference of the 
ring that it can be maintained at a satisfactory level in some cases but 
with the sliding resistance being increased to raise a difficulty in the 
responsiveness to the output. In other cases, on the other hand, the 
responsiveness to the output is improved but with the sealing performance 
being degraded, thus raising an inconsistency. This inconsistency is 
followed by a disadvantage that the housing and the ring have to be 
machined with high accuracy. 
There arises another disadvantage in that difficulty is encountered during 
machining because of the structure in which one half of the pivot pins for 
rocking the ring are fitted in the inner wall of the pump housing whereas 
the other half are fitted in semi-cylindrical grooves which are formed at 
the ring side. Moreover, since the suction and discharge of the pump are 
conducted in the axial direction through arcuate ports formed in the pump 
housing, there arises a defect in that the total length of the pump is 
enlarged in the axial direction. (In the case of a front-engine and 
front-drive automobile, it is necessary to make the axial length of the 
pump as small as possible.) On the other hand, the suction and discharge 
part arrangement is the so-called "one-side suction" which raises a 
problem in producing suction for a fast run. As to the sealing at the side 
opposite to the pivot of the ring, moreover, the center of the arc of the 
sealing face of the housing is offset from the center of rotation of the 
sealing portion, i.e., the center of the pivot. As a result, the spacing 
between the sealing portion of the ring and the sealing face of the 
housing is varied by the rocking motions of the ring which makes it 
difficult to ensure reliable sealing performance and durability so that 
the sliding resistance at the sealing portion is varied which raises the 
concern that the smooth sliding motions of the ring could be obstructed. 
Still moreover, the hydraulic force F.sub.R, which is generated in the 
ring by the pumping action exerted upon the ring, is applied as it is in 
the pivotting direction so that the contact load at the pivot portion 
becomes high thereby increasing the sliding resistance. There arises 
another problem that any excessive force F.sub.R blocks the smooth rocking 
motions of the ring. 
According to the disclosure of U.S. Pat. No. 3,656,869, two separate 
opposed pressure chambers are formed between a cam ring and the bore of a 
housing by means of a sealing pin. Moreover, fluid circuitry including an 
automatic pressure regulator valve is in fluid communication with each of 
the pressure zones so that the pressure balance or ratio of pressures 
across the cam ring may be controlled. The structure thus disclosed 
succeeds in improving the problems of the sealing structure of the prior 
art described above. However, the seal pin is merely fitted loosely in the 
transverse openings of the outer circumference of the cam ring and may 
fail to ensure a complete sealing effect. Moreover, when the cam ring is 
controlled by the differential pressure between the two opposed chambers, 
there arises a concern that the maximum eccentricity of the cam ring 
cannot alway be held at the start of the pump so that the pump may not 
operate as expected. 
SUMMARY OF THE INVENTION 
It is therefore an object of the present invention to provide a pump from 
which the defects of the prior art thus far described are eliminated and 
which is enabled to have its sealing performance improved to provide high 
durability and performance. 
Another object of the present invention is to provide a variable output 
vane pump which has such an elastic member that provides complete elastic 
restoration having a large deflection and high strength even within a 
limited space so as to can sufficiently satisfy the required function. 
A further object of the present invention is to provide a variable output 
vane pump which has smooth ring motions so that it can provide a high 
output responsiveness to changes in its r.p.m. to ensure the necessary 
output in response to the minimum input. 
Still further object of the present invention is to provide a variable 
output vane pump which can have a small axial length and provide suction 
from the two sides of a rotor so that it finds a suitable application as a 
hydraulic pump for an automotive automatic transmission. 
Yet another object of the present invention is to provide a variable output 
vane pump which has its parts reduced in number and its piping simplified. 
Still another object of the present invention is to provide a variable 
output vane pump which can be used in an automatic transmission, while 
replacing a constant output pump of the prior art, without any drastic 
change in the hydraulic system of the conventional automatic transmission. 
The several objects thus far described and the other objects to be 
described hereinafter can be achieved by a variable output vane pump 
comprising: a pump housing; a rotor supported rotatably within said 
housing; a plurality of vanes fitted in the outer circumference of said 
rotor in such a manner that they can be displaced radially into and out of 
said rotor; a ring supported pivotally by a pivot portion, which provides 
a connection between the circular bore of said housing and the outer 
circumference of said ring, and encloses said rotor and said vanes; and 
inlet and outlet ports formed in said housing, at least one pressure 
chamber for pivotting said ring being formed between the circular bore of 
said housing and the outer circumference of said ring by means of at least 
one seal pin which is fitted in an axial groove formed in the outer 
circumference of said ring and opened in said circular bore of said 
housing, said ring being urged by the action of a spring, which is fitted 
in said housing, in a direction opposite to that of the pressure, which is 
built up by said pressure chamber, namely, in a direction such that the 
axis of said ring leaves the axis of rotation of said rotor, wherein the 
improvement resides in that an elestic member is fitted in the bottom of 
the axial groove, which is formed in the outer circumference of said ring 
and opened in the circular bore of said housing, thereby to urge said seal 
pin both toward the circular bore of said housing and at the same time 
toward one of the side walls of the axial groove; and in that the urging 
direction of said elastic member and a tangential line, on which said seal 
pin is tangential to the circular bore of said housing, are angularly 
positioned at an acute angle with respect to each other. According to the 
construction thus far described, the pressure chamber ensure reliable 
seals despite the low sliding resistances of the seals, because their one 
or two ends are sealed dynamically by means of the seal pin, so that the 
sealing performance is improved to provide a high-performance pump which 
has a high durability and a low input power loss. At the start of the 
pump, moreover, the cam ring can always be held at the maximum eccentric 
position by the action of the spring. 
According to a preferred embodiment, each of said elastic members may 
preferably be made of a formed wire spring which is contoured to resemble 
the outer periphery of the bottom of each of said grooves and which has 
its inner end portion bent inward to have such a height as not to be 
folded. 
According to this construction, the formed wire spring is so devised that 
its contact height may become equal to the diameter thereof when it is 
depressed and that its total length may be elongated as much as possible. 
As a result, the formed wire spring does not require any large space in 
the radial direction of the ring but can reduce the size of the ring so 
that it can enjoy not only a large free length and a small contact height 
but also a complete elastic restoration having a sufficient deflection and 
a high strength even within the groove having a limited space. As a 
result, the seal pin can always be urged properly within the circular bore 
of the housing irrespective of the values of the relative displacements of 
the ring and the housing bore thereby to complete the sealing function. 
According to another aspect of this invention, there are two pressure 
chambers: a first pressure chamber having communication with said outlet 
port and enclosing the pivot portion of said ring for generating such an 
urging force as to urge said ring toward said pivot portion against the 
internal pressure, which is generated in said ring by a pumping action 
that, is weaker than said internal pressure; and a second pressure chamber 
disposed adjacent to the first-named pressure chamber through said seal 
pin and adapted to be supplied with the pump output pressure either 
directly or indirectly through a control valve for generating an urging 
force against the biasing force of said spring, the region extending 
between said circular bore of said housing and the outer circumference of 
said ring other than said pressure chambers being made to communicate with 
said inlet port. 
According to this construction, the ring is controlled by at least two 
pressure chambers which are defined by means of the improved seal pins, 
wherein the first one effects hydraulic balance with the internal pressure 
acting upon the pivot portion thereby to drop that pressure to a 
remarkably low level. Thus, since the stress exerted upon the ring is 
remarkably reduced so that the weight of the ring can be reduced and since 
the force for urging the pivot portion is weak, the control stability of 
the ring pivotting motions can be markedly improved to enhance the output 
responsiveness so as to produce prompt responses to changes in the r.p.m. 
of the pump thereby providing a pump having a reduced input power loss and 
a high performance. Since no excessive pressure is applied to the pivot 
portion, moreover, this pivot portion can be so simplified as to employ a 
spherical member such as a ball in the pivot portion so that a better 
control stability of the ring pivotting motions can be attained. 
Still moreover, since the suction and discharge can be performed in the 
radial directions as is different from the prior art in which they are 
performed from only one side, the axial length of the pump can be reduced 
to provide a two-side suction pump in which the suction is effected from 
both sides of the rotor. 
In a further preferred embodiment, the aforementioned pump can use a ball, 
a ball having a spherical face, or a spherical roller in the pivot 
portion. As a result, not only the friction at the pivot portion can be 
reduced to smoothen the ring motions but also the pivot portion can be 
machined with ease and at a low cost. 
According to a further preferred embodiment, said control valve includes: a 
spring fitted in a spring chamber communicating with said outlet port; a 
spool having its one end biased by said spring and its hollow portion 
formed on its outer circumference with at least two lands and in its 
inside with an orifice; a tank port having communication with said pump 
inlet port; a first port having communication with said pump outlet port 
and said hollow portion; and a second port adapted to communicate with the 
second-named pressure chamber and to be opened between said two lands and 
arranged at such a position as to communicate with said tank port, when 
the pressure difference between the upstream and downstream sides of said 
orifice is equal to or lower than a predetermined level, and with the 
first-named port when said pressure difference exceeds said predetermined 
level. 
According to the construction thus far described, the throttle valve and 
its accompanying piping can be omitted so that the pump as a whole can be 
made compact and manufactured at a low cost. Since the control valve can 
be attached to or built into the pump without any difficulty, moreover, it 
is possible to provide a variable output vane pump which can have its 
parts interchanged with those of the product of the prior art.

DESCRIPTION OF THE PREFERRED EMBODIMENTS 
The present invention will be described in the following in connection with 
the embodiments thereof with reference to the accompanying drawings. FIG. 
1 is a sectional view showing an embodiment of the present invention, and 
FIG. 2 is a section taken along line II--II of FIG. 1. Indicated generally 
at reference numeral 1 is a variable output vane pump which has a housing 
including a housing (A) 10, a housing (B) 30 and their accessories, unless 
specified. The variable vane pump 1 is mounted to a transmission body (not 
shown) through mounting holes 11 by means of bolts (not shown). The 
housing (A) 10 and the housing (B) 30 are joined by means of bolts (not 
shown) through tapped holes 12 which are formed in the housing (A) 10. The 
pump has its inlet port 13 communicating with an oil tank by way of a 
passage of the transmission body. On the other hand, the outlet port 14 of 
the pump is made to communicate with an actuator by way of another passage 
(not shown) of the transmission. The pump housing (A) 10 is formed therein 
with a generally circular bore 20 which extends in the axial direction. 
This bore 20 is formed in its upper portion with a semicircular groove 21 
in which a pivot ball 22 is fitted. In the bore 20, there is fitted a ring 
50 having its upper portion formed with a semicircular groove 51, in which 
the pivot ball 22 is fitted. As a result, the ring 50 is supported in the 
bore 20 such that it can be rotated in the bore in the radial direction 
while using the pivot ball 22 as its pivot. The openings of the ring 50 at 
both sides in the axial direction are blinded by the opposed inner walls 
of the housings (A) 10 and (B) 30 to define pump chambers 83 in 
cooperation with both a rotor 70 fitted in the ring 50 and vanes 71. The 
ball 22 is used as the pivot member in the embodiment under consideration 
but may be exemplified by a round roller having a round face or a pin 
having a shape similar to that of FIG. 8 in which case, such roller or pin 
may simultaneously be used as a sealing pin as disclosed in FIG. 1 of U.S. 
Pat. No. 4,342,545 to seal one end of a pressure chamber which will be 
described in detail hereinafter. The ring 50 is formed at its lower 
portion with a protrusion 52 which is urged to the left, as viewed in FIG. 
1, by the action of a spring 53 fitted in a hole 15 of the pump housing so 
that the ring 50 is offset to the left with respect to the rotor 70. The 
housing bore 20 is arranged, as better seen from FIG. 2, with a suction 
chamber 16, which has communication with the pump inlet port 13, and a 
discharge chamber 17 which has communication with the pump outlet port 14. 
The suction chamber 16 and the discharge chamber 7 are arranged in the 
radial directions, i.e., in the directions perpendicular to the rotor 
shaft 80. Moreover, the suction chamber 16 has communication with arcuate 
inlet ports 35 and 36, which are disposed at both sides of the pump 
chambers 83, across the two sides of the ring 50, whereas the discharge 
chamber 17 also has communication with arcuate ports 37 and 38, which are 
disposed at both sides of the pump chamber 83, across the two sides of the 
ring 50. This ring 50 has its outer circumference formed with three 
protrusions 54', 55' and 56' which extend outward and which in turn are 
formed with axially extending grooves 54, 55 and 56. In these grooves, 
there are fitted both seal pins 60, 61 and 62, which are in respective 
sliding contacts with the bore 20, and three elastic members 63 each of 
which underlies the seal pins 60, 61 and 62. The space defined by the bore 
20 of the pump housing and by the outer circumference 57 of the ring is 
divided into three sealed compartments by means of the three seal pins. 
Specifically: the compartment between the seal pins 60 and 62 defines the 
suction chamber 16; the compartment between the seal pins 61 and 62 
defines the discharge chamber 17; and the compartment between the seal 
pins 60 and 61 defines a control chamber 18. The ring 50 is formed at its 
lefthand side with a protrusion 58 which has its leading end 59 providing 
a stopper, which is brought into contact with the pump housing bore 20, in 
case it is urged to the left by the action of the spring 53. In the ring 
50, there are fitted the rotor 70, the vanes 71 which can move freely into 
and out of the rotor 70, a guide ring 72 which is made operative to urge 
the vanes 71 to contact with the inner circumference of the ring 50 even 
when the pump is stopped. The rotor 70 has its radially inside portion 
splined to the shaft 80 so that it is supported rotatably within the 
housing 10 and 30 by the action of the shaft 80. Reference numeral 82 
indicates a bearing. The shaft 80 may be formed into the shape of a sleeve 
shaft, which extends usually from the engine transmission as well known in 
the art, although it is schematically shown in the drawing. 
The sealing operations by the seal pins 60, 61 and 62 will be described 
with reference to FIGS. 3 to 8 (of which FIG. 3 is an enlarged view 
showing the portion around the protrusion 56' located at the righthand 
upper portion of FIG. 1). The seal pin 62 (as shown in FIG. 8) has an 
axial length substantially similar to that of the ring 50 and is urged at 
all times in the direction of arrow F.sub.S (as shown in FIG. 3) by the 
action of the elastic member 63 such as a leaf spring 63A (as shown in 
FIG. 6) or a synthetic rubber 63B (as shown in FIG. 5). The seal pin 62 is 
always in contact at and urged by the elastic member 63 toward two points 
a and b (as shown in FIG. 3) because the contact angle against the bore 20 
of the housing 10 is acute. When the discharge pressure of the pump rises, 
this high pressure oil exerts its high pressure upon the side of the seal 
pin 62 to urge the seal pin in the direction F.sub.H of FIG. 4 showing the 
operating state thereby to ensure seals at the two points a and b. More 
particularly, as shown in FIG. 3, the tangential line C, on which the seal 
pin 62 contact with the housing bore 20, and the urging direction F.sub.S 
of the elastic member 63 have to be angularly positioned at an acute angle 
.alpha. smaller than 90 degrees. The ring 50 is formed at its two sides 
with grooves 124, each of which has its one end 125 communicating with the 
pump suction chambers 35 and 36 and its other end 126 blinded, thereby to 
prevent the pressure of the oil in the pump chambers 83 from leaking out 
into the second pressure chamber 18. 
As shown in FIGS. 3 and 4, the gaps .delta..sub.1 and .delta..sub.2 between 
the outer circumference of the ring and the bore of the housing are varied 
slightly by the rocking motions of the ring around the pivot. It is 
necessary to ensure reliable seals even for those variations. Thanks to 
the construction thus far described, the seal pin never fails to contact 
with the outer circumference of the ring and the bore of the housing at 
those two points a and b by the action of the elastic member 63. The 
spring force of the elastic member in this instance may be such a 
remarkably low load as to merely raise the pin so that it can suppress the 
sliding resistance at the sealing points a and b at a low level while the 
ring is rocking. Under increased pressure, on the other hand, the 
hydraulic pressure is so exerted by the angle .alpha. as to ensure 
automatically the contacts at the points a and b thereby to ensure the 
sealing performance. Thus, the levels of "the order of the pump discharge 
pressure"&gt;"the control pressure"&gt;"the pump suction pressure" are 
maintained at all times so that the contact point b is never released. 
As a result, in spite of the low sliding resistance of the seals, the 
sealing effects are ensured to provide a pump which has its sealing 
performance improved to provide a high performance having a high 
durability and a low input power loss. 
The operations of the pump shown in FIG. 1 will be described in the 
following. The rotor 70, the vanes 71 and the guide ring 72 are rotated 
clockwise. The vanes are brought radially outward by the centrifugal force 
generated during the rotations and by the guide of the guide ring so that 
they slide while having their leading ends contacting with the ring bore 
at all times. As a result, the pump chambers 83 are supplied below the 
line X--X of FIG. 2 with the working oil by way of the inlet port 13 
communicating with the tank, the suction chamber 16 and the arcuate intake 
ports 35 and 36 at both sides of the ring, thus effecting the sucking 
operations. The oil thus sucked is discharged above the line X--X to the 
outside by way of the arcuate discharge ports 37 and 38 at the two sides 
of the ring, the discharge chamber 17 and the outlet port 14. 
In the state of FIG. 1, the central axis of the ring 50 is offset to the 
most eccentric position with respect to the center, i.e., the axis of 
rotation of the rotor 70 is urged by the urging force of the spring 53 so 
that the discharge or output of the pump is changed in proportion to the 
r.p.m. of the same. If, now, the control chamber 18 defined by the seal 
pins 60 and 61, the housing bore 20 and the ring outer circumference 57 
are supplied from the outside with the control pressure by way of a port 
84, there is established in the ring a force which will urge the ring to 
the right, as viewed in FIG. 1, against the urging force of the spring 53. 
However, the ring is always subjected to the urging force which is 
directed upward, as viewed in FIG. 1, namely, toward the pivot ball so 
that it is rocked counter-clockwise around the pivot ball. As a result, 
the eccentricities of the ring and the rotor are so reduced that the 
discharge of the pump is accordingly dropped, as is well known in the art. 
In other words, the theoretical discharge of the pump can be varied by 
varying the pressure of the control chamber. 
The first pressure chamber 17, i.e., the discharge chamber encloses the 
pivot portion 22 and is subjected to the pump discharge pressure to 
generate an urging force FRO which is directed against the internal 
pressure FRT generated in the ring 50 by the pumping action but weaker 
than the same. Thanks to this construction, the force for the ring 50 to 
urge the pivot ball 22 is remarkably weakened by the pressure of the 
discharge chamber 17 so that the stress exerted upon the ring 50 can be 
dropped to a markedly low level. Moreover, the control stability of the 
pivotal motions of the ring is improved to an outstanding level thereby 
enhancing the discharge responsiveness. On the other hand, the pivot 
portion can be manufactured at a low cost without any high accuracy. Still 
moreover, the suction can be effected at both sides of the rotor 70 so 
that the pump can have its axial length reduced. 
The elastic member 63 is exemplified by the synthetic rubber 63b shown in 
FIG. 5 and by the leaf spring 63A shown in FIG. 6. In order to ensure the 
actual function, however, the elastic member 63 is required to have a 
considerable deflection and a complete elastic restoration for the 
repeated stresses applied. However, the synthetic rubber 63B is deficient 
in the deflection and restoration characteristics whereas the leaf spring 
63A is difficult to design and so weak in strength that it is liable to 
flex so that it cannot be used because of its deficient elastic 
restoration properties. Thus, both the synthetic rubber 63B and the leaf 
spring 63A may sometimes be short of the functions required as the elastic 
member. 
In order to satisfy the required function, as shown in FIG. 7, it is 
sufficient to use one or more coil springs 100, for example. In this 
instance, however, it is necessary that the groove 56 of the ring 50 be 
machined with a hole 101 in which the coil spring 100 is to be fitted. 
Then, not only the machining and assembly become difficult but also the 
height L.sub.1 of the recess becomes partially increased so that the ring 
50 has its thickness reduced thereby raising a problem of strength. This 
problem is accompanied by a problem that the size of the ring 50 has to be 
enlarged. This problem is solved by the improved formed wire springs which 
are shown in FIGS. 9-11 and in FIG. 16. The embodiment of FIG. 9 is 
directed to a formed wire spring 102 which has its respective outer sides 
103 formed into a generally rectangular shape resembling generally the 
outer peripheries of the rectangular bottoms of the aforementioned grooves 
54, 55 and 56 and which has its inner end portion 104 bent inward to have 
such a height h (as shown in FIG. 10), i.e., a free length as not to be 
folded. The formed wire spring 102 of FIG. 9 is made of a solid metal wire 
having a generally circular section but may be made hollow or may have a 
square section. Moreover, the formed wire spring 102 is formed into a 
generally angular G-shaped top plan view, as seen in FIG. 11. On the other 
hand, the formed wire spring 105 shown in FIG. 16 has its end portion 106 
bent twice inward generally at a right angle to enlarge the free length 
and to urge the seal pin at two points in a stable manner. 
Thanks to the construction thus far described, the formed wire springs 102 
and 105 can have their contact heights reduced to the same level as the 
diameter thereof when they are depressed from the above as in the 
direction F of FIG. 9. Since the outer sides 103 are shaped to resemble 
the outer peripheries of the bottoms of the grooves 54, 55 and 56, 
moreover, the total length of the springs can be enlarged as much as 
possible so that the free length, i.e., the height h can be accordingly 
enlarged. As a result, as shown in FIGS. 12 to 15, the formed wire springs 
102 and 105 can retain sufficient deflection and complete elastic 
restoration having high strength, even if they are fitted in grooves 
having limited space, so that they can depress properly the seal pin 62 to 
the housing bore 20 at all times, no matter how the ring 50 and the 
housing bore 20 might be displaced relative to each other, thus completing 
their sealing functions. On the other hand, the grooves can be made 
shallow as indicated at L.sub.2 (as shown in FIG. 12) so that the ring can 
have its size reduced and its strength improved. On the other hand, the 
formed wire springs 102 and 105 may be driven in the direction C (as shown 
in FIG. 15) when they are assembled, after the seal pin 62 has been 
inserted, so that their assemblies can be remarkably simplified. Moreover, 
the springs are formed of wire so that they are easy to design and 
inexpensive to manufacture. 
FIG. 17 shows an embodiment of the control valve for controlling the pump 
of the present invention. The parts corresponding to those of FIG. 1 are 
indicated by corresponding reference characters. 
The outlet port 14 of the pump is connected with the actuator A of the 
automatic transmission by way of a discharge passage 110 so that the 
working oil is fed from the pump to the actuator A. From the discharge 
passage 110, on the other hand, there is branched a passage 111 which 
leads to a sequence valve 120. From the sequence valve 120, there leads a 
passage 112 which communicates with the control chamber 18 of the pump by 
way of the port 84. The sequence valve 120 is equipped with a spool 121 
and a spring 122, the former being urged downward, as seen in FIG. 17, by 
the latter. 
In accordance with the rise in the r.p.m. of the pump, the pressure of the 
valve chamber 123 of the sequence valve 120 is boosted so that the spool 
121 is shifted upward in the drawing. In response to the shift of the 
spool 121, the communication between the passages 111 and 112 is 
established to introduce the pressure liquid into the control chamber 18 
thereby increasing the pressure in the control chamber 18. Because of this 
pressure in the control chamber, the ring 50 is rocked counter-clockwise 
around the pivot ball 22 against the urging force of the spring 53 so that 
the discharge of the pump is reduced. When the output pressure of the pump 
is dropped, on the contrary, the spool 121 is shifted downward to block 
the communication between the passages 111 and 112 so that the discharge 
of the pump is accordingly increased. The operations thus far described 
are automatically conducted so that the maximum pressure in the circuit 
can be maintained at a constant level even if the r.p.m. of the pump is 
varied. 
FIG. 18 shows another embodiment of the control valve. This embodiment 
contemplates to maintaining the discharge of the variable vane pump at a 
constant level for a pump r.p.m. higher than a predetermined value. The 
discharge passage 110 has communication with the outlet port 14 of the 
pump as in the embodiment of FIG. 17 and further with the actuator A by 
way of an orifice 130 and a passage 131. From the discharge passage 110, 
there is branched a branch passage 132 which is made to communicate with a 
valve chamber 141 formed at one end of the spool 144 of a constant 
pressure difference valve 140. From the passage 131, there is branched a 
branch passage 133 which has communication with a valve chamber 142 formed 
at the other end of the spool 144. In the valve chamber 142, there is 
fitted a spring 143 which urges the spool 144 to the right of the drawing. 
An outlet port 145 to be throttled by the land 146 of the spool 144 is 
made to communicate with the control chamber 18 of the pump by way of the 
port 84. 
Thanks to the construction thus far described, if the diameter of the 
orifice 130 and the strength of the spring 143 of the constant pressure 
difference valve are selected at suitable levels, the spool 144 is urged 
to the right of the drawing by the action of the spring 143, while the oil 
is flowing through the orifice 130 at a flow rate not higher than a 
regulated value, to block the communication between the branch passage 132 
and the passage 112 so that no oil flows into the control chamber 18. At 
this time, the pump has its discharge varying in proportion to the r.p.m. 
thereof. When the pump r.p.m. is increased so that the flow rate of the 
oil passing through the orifice 130 is accordingly increased, the pressure 
difference between the upstream and downstream of the orifice is augmented 
in proportion to the oil flow rate. Moreover, this pressure difference is 
exerted upon the two sides of the piston 144 by way of the passages 132 
and 133. As a result, when the pressure difference exceeds a predetermined 
level, the spool is disposed to the left of the drawing against the action 
of the spring. This results in establishment of the communication between 
the passages 132 and 112 to feed the control chamber of the pump 18 with 
the pressure oil thereby to reduce the discharge of the pump. The pressure 
difference, at which the communication between the passages 132 and 141 is 
started, is subjected to the force of the spring 143. The spool 144 
operates to maintain the pressure difference between the upstream and 
downstream ends of the orifice always at the value which is regulated by 
the spring 143. 
If the pressure difference is increased, more specifically, the spool is 
displaced to the left of the drawing to provide communication between the 
passages 132 and 12 thereby to reduce the discharge of the pump. If the 
pressure difference decreases, on the contrary, the communication between 
the passages 132 and 112 is blocked to increase the discharge of the pump. 
Thanks to the operations conducted automatically, the pump discharge can 
be maintained at a constant level for the rpm's higher than the 
predetermined value. 
FIG. 21 shows another embodiment in which the constant flow rate is 
maintained for the rpm's higher than the predetermined level. The parts 
corresponding to those of FIG. 1 are indicated at reference numerals to 
which capital letters A, B, C and D are attached. According to this 
embodiment, the pump outlet port 14 is made to communicate with the 
actuator A by way of the discharge passage 110, the orifice 130 and the 
passage 131. From the discharge passage 110, there is branched a branch 
passage 137 from which the discharge pressure is introduced through the 
control port 84 to the control chamber 18A defined by seal pins 60A and 
60B. From the passage 131, there is branched a branch passage 138 from 
which the discharge pressure having passed through the orifice is 
introduced through the control port 85 to a control chamber 18B defined by 
seal pins 60C and 60D. 
The oil thus discharged out of the pump is introduced through the orifice 
130 to the actuator. When the oil just passes through the orifice 130, 
there is established between the upstream and downstream ends of the 
orifice 130 a pressure difference which corresponds to the flow rate of 
the oil passing through the orifice. Since oil at these respective 
pressure is introduced to the control chambers 18A and 18B, the ring 50A 
is going to rotate to the right of the drawing by the hydraulic pressure 
which is built up by that pressure difference. However, the ring 50A is 
equipped at its one end 52 with the spring 53 which in turn urges the ring 
50A to the left of the drawing. As a result, while the hydraulic pressure 
is weaker than that spring force, the ring 50A is held in the state shown 
in FIG. 21 to discharge the oil in proportion to its r.p.m. If the r.p.m. 
is further increased so that the flow rate of the oil passing through the 
orifice is accordingly increased, however, the pressure difference between 
the upstream and downstream ends of the orifice is proportionally raised. 
As a result, the hydraulic pressure acting upon the ring is strengthened. 
If this hydraulic force overcomes the spring force, the ring 50A is 
rotated to the right of the drawing to have its eccentricity reduced to 
drop its discharge. In these ways, the pump flow rate is automatically 
controlled to a constant value by the hydraulic force of the control 
chamber and by the force of the spring so that the pressure difference 
between the upstream and downstream ends of the orifice may become 
constant for the rpm's higher than the predetermined level. Here, the 
optimum control can be selected by varying the orifice diameter, the seal 
section (or the control chamber range) by the seal pins and the spring 
force. 
The pump control by a control valve 150, which has its orifice 130 of FIG. 
18 built into the constant pressure difference valve 140, is shown in FIG. 
19. The control valve 150 is constructed to include: a spring 153, which 
is mounted in a spring chamber 158 having communication with an outlet 
port 157 connected with the actuator A; a spool 154 which is formed on its 
outer circumference with at least two lands 154' and 154" and therein with 
a hollow portion 159 having an orifice 155 and which has its one end urged 
by the action of the spring 153; a tank port 152 communicating with the 
pump inlet port; a first port 151 communicating with the pump outlet port 
14 and the hollow portion 159 by way of the passage 110; and a second port 
156 which is made to communicate with the control chamber 18 through the 
port 84 by way of the passage 112, which is opened between the two lands 
154' and 154" and which is so arranged to communicate with the tank port 
152, when the pressure difference between the orifice 155 is not higher 
than a predetermined level, and with the first port 151 when the same 
pressure difference exceeds the predetermined level. The orifice 155 need 
not be positioned adjacent to the spring 153, as in the embodiment shown 
in FIG. 19, but may be formed either inside of the spool 154 or adjacent 
to the first port 151 at the opposite side. 
Thanks to the construction thus far described, if the diameter of the 
orifice 155 and the strength of the spring 153 of the constant pressure 
difference valve are selected at suitable values, the spool 154 is urged 
downward of the drawing by the action of the spring 153, while the oil is 
flowing through the orifice 155 at a flow rate not higher than a regulated 
value, to block the communication between the passage 112 and the 
discharge passage 110 so that no oil flows through the control chamber 18. 
At this time, the pump has its discharge varied in proportion to its 
r.p.m. If this r.p.m. is augmented to increase the flow rate of the oil 
passing through the orifice 155, the pressure difference between the 
upstream and downstream ends of the orifice is augmented in proportion to 
the oil flow rate. On the other hand, this pressure difference is exerted 
directly upon the upper and lower sides of the spool 154. If this pressure 
difference exceeds the predetermined level, the spool 154 is displaced 
upward of the drawing against the force of the spring 153. This results in 
establishment of the communication between the passages 110 and 112 to 
feed the pressure oil to the control chamber 18 so that the discharge of 
the pump is decreased. The pressure difference, at which the communication 
between the passages 110 and 112 is started, is governed by the force of 
the spring 153. The spool 154 operates to maintain the pressure difference 
between the upstream and downstream ends of the orifice always at such a 
level as is regulated by the spring 153. 
If the pressure difference is increased, more specifically, the spool 154 
is displaced upward of the drawing to provide communication between the 
passages 110 and 112 thereby decreasing the discharge of the pump. If the 
pressure difference is dropped, on the other hand, the communication 
between the passages 110 and 112 is blocked to resultantly augment the 
pump discharge. Thanks to the operations conducted automatically, the 
discharge of the pump can be maintained at a constant level for its r.p.m. 
not lower than the predetermined value. 
As has been described hereinbefore, as different from FIG. 18, the present 
example can have its valve number and its piping reduced to less than one 
half to simplify the piping so that the system as a whole can be made 
compact and manufactured at a low cost. 
Moreover, the control valve 150 can be mounted in or formed integrally with 
the housing 10 of the pump 1 because it can have its size reduced. With 
reference to FIG. 20, more specifically, the control valve 150 shown in 
FIG. 19 is made integral with the outlet port 14 of the housing 10, and 
the respective ports 152, 151, 156 and 157 of the control valve 150 are 
made to communicate, respectively, with the pump inlet port 13 through a 
housing hole (not shown), with the pump outlet port 14 in a direct manner 
through the port, and with the pump control chamber 18 through the hole 
112' formed in the housing by way of the passages which are formed in the 
housing. It will be apparent to those skilled in the art that effects 
similar to those of FIG. 19 can be attained by attaching the control valve 
150 to the pump 1 in such a way that the control valve 150 is formed as a 
cartridge valve which is screwed in the hole formed in the outlet port 14. 
Thanks to the construction thus far described, the system can be made so 
remarkably compact as to solve the problem that the system has been 
difficult to use because the automatic transmission to be mounted in the 
limited engine room requires two or more control valves in accordance with 
the prior art. Moreover, the hydraulic system has to be markedly modified 
so as to use the variable pump in the hydraulic system of the automatic 
transmission using the constant output pump of the prior art. Such 
modification raises considerable difficulty in practice. According to the 
present invention, it is possible to provide an automatic transmission 
using the variable output vane pump merely by replacing only the pump 
without any outstanding change in the hydraulic system. This means that no 
other changes are required in the hydraulic system of the prior art.