Note: Descriptions are shown in the official language in which they were submitted.
12~37~71
Hydraulic pum~
The present invention relates to a hydraulic pump, in
particular for steering power assistance.
In a preferred embodiment, the present invention is
directed to a hydraulic pump comprising the following features:
the hydraulic pump has a rotor which is driven at a varying
speed (n) and which with stationary pump portions ~orms at
least one displacement region to which lead inlet ports and
outlet ports; the inlet ports of each displacement region
communicate with a feed system and the outlet ports of each
displacement region communicate with a pressure chamber; the
pressure chamber and the feed system communicate with each
other by way of a flow control valve which bypasses an excess
delivery flow into a relief duct of the feed system and
outputs a controlled output flow (Q) to an external pump
outlet; the flow control valve includes a spool guided in a
valve bore and having a first higher-pressure spool face and a
second lower-pressure spool face, a valve spring and a metering
orifice at which a pressure drop of the controlled output flow
(Q) is tapped off and fed to the two spool faces o~ the spool;
the spool has an extension por~ion whose position reduces the
effective width of the orifice with increasing pump speed (n)
thus producing a generally falling leg in the output flow~pump
speed characteristic curve, characterised in that the orifice
is formed at an output flow bore duct extending generally
radially in the direction of the ~low control valve, the
output flow bore duct and the respective relief duct are at
an axial spacing (c) ~rom each other, as considered in the
direction of di~placement oE the spool, which is less than the
width tb) of an annular chamber which i5 formed at the spool
between the first spool face and a third spool face; and the
annular chamber communicates with the pressure chamber by wa~
a hollow space in the Sp~Ql,
*
3~7~l
Steering power assistanoe pumps are usually in the form of
vane-type pumps and are rigidly connected to the drive engine of
the motor vehicle in which the steering assistance system is used.
Accordingly the pump delivery flow increases with increasi~g
engine speed. However, there is generally no need for a strong
steering power assistance effect when the engine is rotating at
high speeds. For that reason, the systems generally use a flow
control valve for bypassing a part of the pump delivery flow
while the remaining regulated output flow is taken back to the
tank by way of the steering valve. When that happens, the hydraulic
fluid which is under what is referred to as dynamic pressure
experiences release of pressure which results in a waste of
energy, unless the power is used by the steering system. In a
practical situation, in the high range of engine speeds, such a
high level of power consumption does not occur because it is not
possible to produce sharp steering movements when travelling
quickly. Accordingly, in the high range of speeds of the pump
the system maintains a constant condition of power readiness
which is not required at that level and which thus results in an
unnecessary waste of ener~y.
In order to overcome that disadvantage, it is already known
for the flow control valve to be designed and arranged to provide
that the output flow-pump speed characteristic curve has a falling
leg (German published specifications DE-A-22 65 097 and DE-~-26 52 707).
The pressure inlet port to the flow control valve is arranged
~253'7'7~L
radially in that design, and likewise the relief passage,
while the output flow is arranged axiall~, in the direction
of movement of the spool of the flow control valve. The
projection portion on the spool is in the form of a valve
needle with needle head, the valve needle extending through
the axial outlet so as to form a metering orifice whose width
depends on the position of the needle head relative to the
axial outlet. A disadvantage with that arrangement is that
just very small changes in the position of the components
result in considerable variations in the cross-sectional area
of the axial outlet, through which the flow passes.
The object of the invention is to design a hydraulic pump
having a falling characteristic of output versus rotational
speed wherein a wide range of falling legs in said character-
istic can be produced by slight alterations at the flowcontrol valve. The invention therefore seeks to make it
possible to produce a characteristic which is suited to the
situation of use.
In accordance with the invention, the pump has a bypass
flow control valve wich includes a spool of special
construction.
An annular chamber formed at an extension portion on
the spool has a certain width of overlap with respect to a
metering orifi~e bore and relief ducts means of which there
are pre~erably two. The annular chamber is fed by a hollow
chamber in the extension portion on the spool. In the annular
chamber, the ~low or output is divided up to provide the
useful flow and the bypassed excess delivery flow. The latter
is relieved over a short distance back into the inlet of the
pump so that low levels of flow losses occur. The fact that
the Elow is divided in the annular chamber as indicated above
has the Eurther advantage that the pulse ~orces applied are in
mutual opposition to each other, thufi providing substantial
compensation in respect of the ~low forces actin~ on the spool.
S3~7~
,
In accordance with a development of the invention, the valve
output duct has a second feed means formed by an annular gap
between the extension portion on the valve spool and the wall of
the valve bore in the region of movement of the extension portion.
The feed cross-section of the second feed means is smaller than
the normal cross-section of the opening between the annular
chamber in the extension position and the valve output duct.
In accordan oe with further features of the invention, the
extension portion on the valve spool may be of varying gecmetrical
configurations in order to influence the configuration of the output
pump speed characteristic.
The invention is described hereinafter with reference to the
drawings in which:
Figure 1 is a partly broken-away view in vertical longitudinal
section through a rotary vane pump,
Figure 2 is a partly broken-away view in horizontal longitudinal
section taken along line II-II in Figure 1,
Figure 3 is a view in cross-section taken along line III-III
in Figure 1,
Figure 4 is a diagrammatic view on an enlarged scale o~ a
detail from Figures 1 and 2, and
Figures 5 through 8 show spool shapes and associated diagrams
in respect of the output useful flow in relation to the pump
speed.
The rotary vane pump comprises a main casing portion 1 and
a casing cover portion 2 which enclose an internal cavity or chamber
la which is sealed in relation to pressure fluid. Disposed in
the internal chamber la an~ fixed with respect to the casing are
a pressure plate 4 and a cam rin~ 5. The pressure plate 4 and
the cam ring 5 are prevented from rotating by means of pins 6.
Disposed within the cam rin~ S and between the casin~ cover
~25~3~7~1l
~ 5
portion 2 and the pressure plate 4 is a rotor 7 which, as shown
in Figure 3, has a plurality of xadial guide slots~ Vanes 8 are
radially slidably mounted within the guide slots. The rotor 7 is
arranged to be driven by way of a shaft 9 which is mounted in a
mounting bore in the Q sing cover portion 2~ The rotor 7 is of
a cylindrical configuration while the cam ring 5 is of an
approximately oval internal configuration, with the minor axis
thereof approximately corresponding to the diameter of the rotor
while the major axis defines the distance by which the vanes 8
can extend from the rotor 7. In that way, defined between the
Qm ring 5 and the rotor 7 are two sickle-shaped displacement
regions 11 and 12 which are sukdivided by the vanes 8 into a
plurality of cell spaces or chambers. At the suction side of the
system, the cell spaces or chambers increase in size while they
decrease in size at the pressure side.
Hydraulic fluid is supplied from a tank 14 (see Figure 3) and
a distributor region 16 by way of two bores 17 (see Figure 2)
which slope slightly dcwnwardly, elbow-bent inlet duct portions
18 and intake ports 20 into the respective displacement regions
of the pump. The elbow-bent inlet duct portions 18 each have
a radial limb portion which ~nmunicates with a relief duct
19 (see Figures 2 and 4).
The discharge of hydraulic fluid takes place by way oE outlet
ports 33 (see Figure 1) through the pressure plate 4 on the
rear side thereof into a pressure chamber 35. At a flow control
valve 40, the pump delivery flow is divided into a controlled
output flcw, which flows by way of a bore 38, to
an ~ernalpump outlet 37 (see Figure 2), and an excess delivery
flow which is b~passed throughthe relief ducts 19. l~he boxe 38
represents a valve or output duct and at the same time a
Fart of a metering oriice 36 through which the output
~S~3~7~
passes, with the pressure drop thereof being tapped off. The valve
output flow passes by way of an inclined discharge duct 39 (see
Figure 1) to the pump outlet 37 (see Figure 2~. From there, a
communication goes to a control chamber 47 of the flow control valve
40 by way of a damping throttle or restrictor means 48. The
flow control valve 40 has a spool 41 which is guided in a valve
bore S5 and which is urged towards the pressure plate 4 by
the force of a spring 42 and which is possibly caused to bear against the
pressure plate 4. The spool 41 has first and second spool faces
53 and 54 as well as shoulders or lands 43 and 44 between which
extends an annular groove 45. The land 43 is narrower than the
r~lief ducts 19 (Figure 2~ which meet the annular groove 45. From
the annular groove 45, a duct 46 which extends partly radially
and partly axially passes through the spool 41 into the control
chamber 47 and the duct 46 is controlled by a ball valve which
responds when a given permissible pressure in the control chamber
47 is exceeded, and thus relieves the control chamber 47 so that
the spool 41 acts as a controlled pressure relief valve, as is
known. Whether operating as a flow control valve or as a pressure
relief valve, when it responds the valve 40 assumes the position
shown in Figure 4.
The spool 41 has an extension portion 49 in which a hollow
space or cavity 50 is disposed. It canmunicates by way of a
series of bores 51 with an annular chamber 52 of a width b. The
annu~rchamber 52 is de~ined by the first spool face 53 and a third
spool face 56 which, co-operating with the relief ducts 19 and
the valve output duct 38, operate as control edges so
that the valve 40 represents a dual-edge control device. The
radial bore 38 and the radial relief duct 19 are shown in Figure
4 in the same axial plane of the valve 40 whereas in actual fact
they are disposed in di~ferent axial planes which for example
inclucle ~n an~le of 90 relative to each other. Projected on to the
~3~'7~
axial sectional plane shown in Figure 4, there is a spacing a
between the axes of the relief duct 19 and the bore 38, with a
land thickness as indicated at c. The ducts 19 and 38 only need
to be in a general radial direction with respect to the valve 40,
the important consideration being that a land width c is formed.
It will be seen that the spacing b is greater than the distance
c, that is to say in a given position of the spool 41 the annular
chamber 52 can connect the bore 38 to the relief duct 19. The
diameter of the extension portion 49 is denoted by d2 while the
valve bore 55, in the region of the extension portion 49, is
of a diameter indicated at dl.
Operation of the pump is as follows:
The rotor 7 is driven b~ way of the shaft 9 and the vanes 8
pass through the displacement regions 11 and 12 so that fluid
is fed by way of the fluid outlet system 33, 35, 50, 38 and 39
to the external or service pump outlet 37 and fluid is sucked
in by way of the external Eump inlet 16 and the fluid inlet
system 17, 18 and 20. If the flow of fluid through the bore
38 exoe eds the desired value, the pressure drop in the orifice
36 at the bore 38 is sufficiently hi~h to overcome the force
of the valve spring 42, in other words the pressure force act-
ing on the face 53 is greater than the pressure force acting on
the face 54 plus the spring force 42. A part oE ~he delivery
pump flow is now bypassed by way of the relief duct 19 while
the valve output flow continues to be taken off by way of the
bore 38. The effective cross-sectional area thereof decreases
as a result of the control edge 56 moving in the closing
direction, that is to say, the measuring orifice 36 is reduced
in size and the pressure drop in the output flow increases.
30Figure ~ shows a diagr~m in respect of the controlled out-
put 10w Q in rel~tion ~o the sEeed of rotation of the pump n when
S37t71
dl~ d2. As long as the annular chamber 52 only communicates with the
bore 38, the output flGw rises linearly with the speed of rotation
n of the pump. Thereafter a progressively increasing part of the
pump flow is bypassed until finally the control edge 56 campletely
S shuts off the output flow. By virtue of the dimensions a, b, c, d3,
d4 being of suitable magnitudes, it is possible to influence the
configuration of the falling leg of the characteristic curve,
that is to say the value n at which the output flow Q goes to
zero can be defined.
Figure 5 shows a spool 41 on which the extension portion 49
is of a diameter d2 which is smaller than the diameter dl of the
valve bore 55. That arrangement forms an annular gap between the
extension portion 49 and the valve bore 55, through which a flow
can pass between the pressure chamber 35 and the bore 38,
irrespective of the position of the spool 41. Therefore the flow
f' . Q does not fall back to zero when the edge 56 closes off the bore
38. It will be appreciated that the amQunt of fluid flowing
through the annular gap depends on the pressure drop which occurs,
as is indicated in the associated diagram by broken and dash-dotted
lines.
Figure 6 shows a spool 41 on which the extension portion 49
is of a slightly conical or tapering con~iguration. The annular chamber
52 therefore extends so-to-speak as far as the front edge 57
of the spool 41. If accordingly the spool 41 is moved against the
force of its valve spring 42, the width of opening of the annular
gap between the extension portion 49 and the valve bore 55 is
reduced, with the rate of reduction rising greatly as the edge
57 approaches the valve bore 55 so that the proportion of the output
flow which flows by way o the annular gap between the e~tension
portion 49 and the valve bore 55 greatly decreases. Depending on
the amount by which the inside diameter dl of the valve bore 55 is
:12,~3 771
larger than the outside diameter d2 of the extension portion 49
however, there is still a certain proportion of the output flow
flowing, as shown in the associated characteristic curve.
Figure 7 shows a spool 41 with an extension portion 49 which
S is a composite from the shapes of the extension portions shown in
Figures 5 and 6, comprising therefore a cylindrical region 58
and a conical or tapering region 59. When the face 53 comes into
communication with the relief duct 19, a progressively increasin~
proportion of the delivery pump flow is bypassed, but a given
proportion of the output flow can flow through the annular gap
between the tapering region 59 and the valve bore 55, into the
bore 38, until, when the spool 41 is in a given position, the
cylindrical region 58 passes into the valve bore 55. Depending on
the amount by which the outside diameter d2 of the cylindrical
r~gion 58 is smaller than the inside diameter dl of the valve
bore, the arrangement then provides a residual output flow of
larger or smaller magnitude, as indicated in the ~ssociated output
flow-pump speed characteristic curve.
Figure 8 shows an embodiment of the spool 41 with an extension
portion 49 having a spherical surface. That shape approximates
to the configuration shown in Figure 7 and accordingly it provides
a similar output flow-pump speed characteristic curve.
In tes-ts which were carried out, the dimensions a, b, c, dl, d2,
d3 and d4 were varied so that the confiquration of the characteristic
curves illustrated could be still further influenced. Tn the
embodiment shown in Figure 4, a was 10.3 mm, d4 was 5.5 mm, d3
was 3.1 to 6.0 mm and b was 7.7 to 10.7 mm.
With d3, the output flow Q also increased, that is to say the
m~ximum output flow was only attained at n = 1700 l/min, instead
of at n = 1000 l/min, and it was accordingly hiqher. Frcm there the
~ZS3t7~71
.
flow Q fell to zero at about n = 6-8000 l/min, with the higher
values being attained at higher pressures. With relatively small
values of b, the falling leg of the characteristic curve fell
away more sharply than with relatively larger values of b.
In the embodiment shown in Figure 5, the width of the gap
defined by dl minus d2 was varied from 0~21 to 0.71 mm. me
greater the width of the gap, the shallower was the angle of
inclination of the falling leg of the characteristic curve,
relative to the zero line. At higher pressures, it was also
possible to achieve a constant output flow, irrespective of the
pump speed n. The dimension b was varied between 7.7 and 8.7 mm,
higher output flow values being achieved at higher values in
respect of b, that is to say the falling leg of the characteristic
curve fell away less severely or remained constant.
~t can thus be seen that the characteristic curve af the controlled
output flow Q can be influenced in such a way that, after the valve
responds, at a given pump speed, the magnitude of the controlled
output flow Q is reduced as follows:
a) with a rising pump speed n, slowly falling away to zero;
b) with a rising pump speed n, slowly falling away to a minimum
value, and
c) with a rising pump speed n, initially constant and then
falling away to zero or a minimum value.
In all cases the flow forces acting on the spool 41 are partially
directed in opposition-to each other, thus providing a substantial
compensation effect.
