Title:
Hydraulic swash plate pump
United States Patent 3873240
Abstract:
Swash plate pump comprising a rotatable swash plate the inclination of which is controlled by a hydraulic jack means. The jack means is in turn controlled by a slide valve which detects the variations in the delivery pressures of the pump and causes a high or a low pressure to actuate the jack means, whereby the inclination of the swash plate is modified and consequently the output of the pump varied. The slide valve is further acted upon by a spring.
US Patent References:
Pressure compensator for variable volume pumps
Parr et al. - May 1958 - 2835228
Variable displacement check valve pump
Roberts - February 1967 - 3304886
AXIAL PISTON PUMP
Russell - May 1970 - 3512178
CONSTANT TORQUE HYDRAULIC PUMP
Leduc - April 1971 - 3575534
/3588286.html
Skinner - June 1971 - 3588286
Pressure compensator for variable volume pumps
Parr et al. - May 1958 - 2835228
Variable displacement check valve pump
Roberts - February 1967 - 3304886
AXIAL PISTON PUMP
Russell - May 1970 - 3512178
CONSTANT TORQUE HYDRAULIC PUMP
Leduc - April 1971 - 3575534
/3588286.html
Skinner - June 1971 - 3588286
Inventors:
Leduc, Gerard (88 Saint-Die, FR)
Leduc, Michel (54 Azerailles, FR)
Leduc, Michel (54 Azerailles, FR)
Application Number:
05/370545
Publication Date:
03/25/1975
Filing Date:
06/15/1973
View Patent Images:
Export Citation:
Primary Class:
Other Classes:
417/270
International Classes:
F04B1/14; F04B1/18; F04B1/29; F04B49/08; F04B1/12; F04B49/00
Field of Search:
417/212,218,222,270 91/506 60/452
Primary Examiner:
Freeh, William L.
Assistant Examiner:
Lapointe, Gregory
Attorney, Agent or Firm:
Darby & Darby
Claims:
We claim
1. In a hydraulic swash plate pump whose driving torque is to be kept constant as a function of delivery pressure, said pump comprising:
2. A pump as in claim 1 wherein said jack moving means includes restricted passage means (24a, 25-24a, 25) for supplying the fluid to said chamber (26--26).
3. A pump as in claim 1 wherein said control means includes a slide valve communicating with and responsive to the pressure in said chamber (26--26) and further comprising spring means for exerting a force on said slide valve means in opposition to the force exerted thereon by the fluid in said chamber.
4. Pump according to claim 3 wherein said spring has a predetermined rigidity curve corresponding to desired variations in deliveries and pressures.
5. Pump according to claim 1 wherein said control means comprises a valve located inside the jack.
6. Pump according to claim 3 wherein the jack includes a piston having a surface communicating with said chamber, said piston of the jack controlling the swash plate having a double section of different sizes, said jack moving means including two chambers situated so that one communicates with the large section and the other communicates with the small section of said jack piston, a calibrated passage means providing communication between said two chambers, the chamber situated behind the large section receiving either the high pressure or the low pressure by means of the valve, so that, with equilibrium, the said valve being closed, these two chambers are at the same pressure, which is a function of the delivery pressure.
7. Pump according to claim 6, wherein the slide member of the valve projects into the chamber which communicates with the small section of the jack and is subjected to the pressure obtaining in this chamber, so that any variation of the delivery pressures causes a displacement of the slide member of the valve.
8. Pump according to claim 7, further comprising a fixed stop against which said spring bears said fixed stop comprising a partition separating the two chambers situated behind the jack.
9. Pump according to claim 7, further comprising a fixed stop against which said spring bears said fixed stop comprising a partition separating the two chambers situated behind the control jack of the swash plate, and second spring means bearing on the jack.
10. Pump according to claim 6, further comprising a partition separating said two chambers and wherein said calibrated passage is formed by an annular space provided between the small-section of the jack and an orifice formed in the partition separating the said chambers.
11. Pump according to claim 10, wherein the jack includes a rod having a shoulder extending through said orifice such that the said annular calibrated passage is of variable length.
12. Pump according to claim 3 wherein the pump includes pumping pistons having bearing studs thereon, said swash plate being formed with a plurality of passages on the path of the bearing studs of the pumping pistons in a path corresponding to their delivery travel, a central chamber formed at the back of the swash plate with which said passages communicate, said jack having a piston including a supporting stud for providing communication between said central chamber and a passage formed in the said piston and ending at the valve.
13. Pump according to claim 1 wherein said pump includes delivery pistons and further comprising means for selecting the highest delivery pressure including an assembly of interconnected passages equipped with non-return valves, said passages being located in the casing of the pump body downstream of the delivery pistons.
14. Pump according to claim 13, further comprising additional non-return valves at the outlet of each delivery pistons, said non-return selection valves being located downstream of said additional non-return valves which are located at the outlet of the bore of each delivery piston.
15. Pump according to claim 13, further comprising additional non-return valves at the outlet of each of said delivery pistons, said selection non-return valves being positioned upstream of the additional non-return valves.
16. Pump according to claim 13 further comprising selection non-return valves for all the delivery systems of the pump.
17. Pump according to claim 13 wherein said selection non-return valve provides a single delivery by flow.
18. Pump according to claim 1, further comprising first means for applying the highest selected delivery pressure to the valve positioned inside the jack by means, said first means comprising a passage drilled in the casing of the pump body and a passage drilled in the jack.
19. Pump according to claim 18, further comprising a joint between two passages, said joint including a ring which is centered with friction on the jack and is held immobile to rotation, said ring having a passage providing communication between the said two passages.
20. Pump according to claim 12 wherein the high pressure selected at the passage ending at the valve communicates with a further passage connecting two passage assemblies supplying the said high pressure to the bearings of the said shaft in two diametrically opposite zones.
21. Pump according to claim 14 further comprising selection non-return valves for all the delivery systems of the pump.
22. Pump according to claim 14 further comprising a selection non-return valves for providing a single delivery by flow.
23. Pump according to claim 13 wherein means for selecting the high pressure communicates with a passage connecting two passage assemblies supplying the said high pressure to the bearings of the said shaft in two diametrically opposite zones.
1. In a hydraulic swash plate pump whose driving torque is to be kept constant as a function of delivery pressure, said pump comprising:
2. A pump as in claim 1 wherein said jack moving means includes restricted passage means (24a, 25-24a, 25) for supplying the fluid to said chamber (26--26).
3. A pump as in claim 1 wherein said control means includes a slide valve communicating with and responsive to the pressure in said chamber (26--26) and further comprising spring means for exerting a force on said slide valve means in opposition to the force exerted thereon by the fluid in said chamber.
4. Pump according to claim 3 wherein said spring has a predetermined rigidity curve corresponding to desired variations in deliveries and pressures.
5. Pump according to claim 1 wherein said control means comprises a valve located inside the jack.
6. Pump according to claim 3 wherein the jack includes a piston having a surface communicating with said chamber, said piston of the jack controlling the swash plate having a double section of different sizes, said jack moving means including two chambers situated so that one communicates with the large section and the other communicates with the small section of said jack piston, a calibrated passage means providing communication between said two chambers, the chamber situated behind the large section receiving either the high pressure or the low pressure by means of the valve, so that, with equilibrium, the said valve being closed, these two chambers are at the same pressure, which is a function of the delivery pressure.
7. Pump according to claim 6, wherein the slide member of the valve projects into the chamber which communicates with the small section of the jack and is subjected to the pressure obtaining in this chamber, so that any variation of the delivery pressures causes a displacement of the slide member of the valve.
8. Pump according to claim 7, further comprising a fixed stop against which said spring bears said fixed stop comprising a partition separating the two chambers situated behind the jack.
9. Pump according to claim 7, further comprising a fixed stop against which said spring bears said fixed stop comprising a partition separating the two chambers situated behind the control jack of the swash plate, and second spring means bearing on the jack.
10. Pump according to claim 6, further comprising a partition separating said two chambers and wherein said calibrated passage is formed by an annular space provided between the small-section of the jack and an orifice formed in the partition separating the said chambers.
11. Pump according to claim 10, wherein the jack includes a rod having a shoulder extending through said orifice such that the said annular calibrated passage is of variable length.
12. Pump according to claim 3 wherein the pump includes pumping pistons having bearing studs thereon, said swash plate being formed with a plurality of passages on the path of the bearing studs of the pumping pistons in a path corresponding to their delivery travel, a central chamber formed at the back of the swash plate with which said passages communicate, said jack having a piston including a supporting stud for providing communication between said central chamber and a passage formed in the said piston and ending at the valve.
13. Pump according to claim 1 wherein said pump includes delivery pistons and further comprising means for selecting the highest delivery pressure including an assembly of interconnected passages equipped with non-return valves, said passages being located in the casing of the pump body downstream of the delivery pistons.
14. Pump according to claim 13, further comprising additional non-return valves at the outlet of each delivery pistons, said non-return selection valves being located downstream of said additional non-return valves which are located at the outlet of the bore of each delivery piston.
15. Pump according to claim 13, further comprising additional non-return valves at the outlet of each of said delivery pistons, said selection non-return valves being positioned upstream of the additional non-return valves.
16. Pump according to claim 13 further comprising selection non-return valves for all the delivery systems of the pump.
17. Pump according to claim 13 wherein said selection non-return valve provides a single delivery by flow.
18. Pump according to claim 1, further comprising first means for applying the highest selected delivery pressure to the valve positioned inside the jack by means, said first means comprising a passage drilled in the casing of the pump body and a passage drilled in the jack.
19. Pump according to claim 18, further comprising a joint between two passages, said joint including a ring which is centered with friction on the jack and is held immobile to rotation, said ring having a passage providing communication between the said two passages.
20. Pump according to claim 12 wherein the high pressure selected at the passage ending at the valve communicates with a further passage connecting two passage assemblies supplying the said high pressure to the bearings of the said shaft in two diametrically opposite zones.
21. Pump according to claim 14 further comprising selection non-return valves for all the delivery systems of the pump.
22. Pump according to claim 14 further comprising a selection non-return valves for providing a single delivery by flow.
23. Pump according to claim 13 wherein means for selecting the high pressure communicates with a passage connecting two passage assemblies supplying the said high pressure to the bearings of the said shaft in two diametrically opposite zones.
Description:
BRIEF SUMMARY OF THE INVENTION
The present invention relates to improvements in hydraulic pumps of variable cubic capacity and particularly to pumps of the type such as described in U.S. Pat. No. 3,575,534 and having as its title "Constant torque hydraulic pump." The pump described in this patent is a pump having a rotatable swash plate, of which the inclination is modified by means of a hydraulic jack, of which the supply is controlled by a valve detecting the increases in pressures in the circuit supplied by the pump.
The present invention concerns a pump operating by the same general principle, but comprising a distribution system which is considerably improved as compared with that described in U.S. Pat. No. 3,575,534 of Feb. 5, 1969.
Actually, pumps providing separate rates of flow are being used to an ever-increasing extent: for example, with a pump having six pistons, it is possible to split up the flow of the pump into two separate flows, each supplied by three pistons, or even with a pump having nine pistons, into three flows each supplied by three pistons. It is quite obvious that, in these cases, it is necessary to have a valve capable of forming the sum of the pressures obtaining in the three circuits so as to act as a function of the torque necessary for driving the pump.
The distribution system described in U.S. Pat. No. 3,575,534 would make it necessary, when the pump is one having several flows, to have a valve body having as many stages as there are flows or deliveries; it would only be possible to obtain a satisfactory functioning of such a system at the cost of complex, delicate and troublesome means.
Another object of the invention is to permit the manufacture of the pumps to be standardised, a single pump head being capable of being used with different bodies having three, or six, or nine pistons.
A last object of the invention is to permit the supply of the pump to be achieved by an axial bore in the pump body, this considerably improving the volumetric yield thereof, but is not possible with a pump such as that which is described in U.S. Pat. No. 3,575,534.
BRIEF DESCRIPTION OF THE DRAWINGS
In the accompanying drawings, for facilitating the understanding of the invention and given by way of example:
FIG. 1 is a longitudinal section of a pump with 6 pistons, in accordance with the invention;
FIG. 2 is a partial view, as a longitudinal section along a plane perpendicular to the sectional plane of FIG. 1, in which a single piston is represented and this piston is offset by one-twelth of a revolution, so that the FIGURE may be more clear;
FIG. 3 is a partial view illustrating a modified form of the valve in FIG. 1;
FIG. 4 is a top plan view of the swash plate of FIG. 1;
FIG. 5 is a partial view of FIG. 1, illustrating the swash plate in the position which it occupies when the delivery is zero;
FIG. 6 is a sectional view along A--A of FIG. 5;
FIG. 7 is a partial view, as a section B--B of FIG. 6;
FIG. 8 is a partial view, as a section C--C of FIG. 6;
FIG. 9 is a longitudinal sectional view along the line A--A of FIG. 10 of a 12-piston pump, grouped into four deliveries, comprising the improvements according to the present addition;
FIG. 10 is a sectional view along the line B--B of FIG. 9;
FIG. 11 is a partial view as a longitudinal section of a pump similar to that of FIG. 9, comprising a modification;
FIG. 12 is a sectional view along the line DD of FIG. 11.
DETAILED DESCRIPTION
Referring to FIG. 1, it is seen that the pump is in two parts: the pump body 1, comprising the pumping pistons 3, and the pump head 2, comprising the driving shaft 4 which, by means of a pin 5, supports the swash plate 6. When the shaft 4 is driven in rotation by a motor (not shown) it drives the swash plate 6, which alternately pushes back the pistons 3, which are countersupported by springs 7. As the swash plate 6 is pivoted on a pin 5, its inclination can be modified, the travel of the pistons 3 decreasing as a function of the decrease in the angle of the plate 6.
The modifications in the inclination of the plate 6 are caused by a hydraulic jack formed by a piston 8 sliding in a bore 8a drilled axially in the shaft 4 and possibly provided with a sleeve or liner 8b, so that during operation, the plate receives the thrust of the pistons 3 on one side and thrust of the piston 8 on the other side.
When there is an odd number of pistons 3, for example five pistons, the thrust exerted by the front face of the plate is alternately caused by two or three pistons, and the result thereof is a high frequency pulsation of the pressure obtaining at the rear of the piston 8, which pulsation cancels out the frictional effects of the various movable elements of the system. By way of example, with a pump having five pistons, the thrust experienced by the front face of the swash plate oscillates by 2.5 ± 20%, and this, when the shaft 4 is turning at 1,500 r.p.m., corresponds to a pulsation at a frequency of 250 cycles per second.
When there is an even number of pistons 3, for example, six pistons, and when the pump is a double-acting pump, the swash plate 6 also experiences high frequency pressure pulsations when the two deliveries are not at the same pressure. When the two deliveries are at the same pressure, or when the pump has a single delivery, there is also a high frequency pulsation which arises from the fact that, because of compressibility phenomena which are shown at the pressures under consideration, the travel of the pistons with a rise in pressure is not symmetrical with that with a fall in pressure, relatively to the line of maximum slope of the plate.
Referring to FIGS. 2 and 4, it is seen that the swash plate 6 comprises a central blind hole 9 which communicates by way of a passage 10 with a crescent-shaped passage 11, over which the studs 12 of the pistons 3 pass at the time of rotation of the plate 6. When the studs 12 pass over the crescent-shaped passage 11, the pistons 3 are in their suction stroke and the liquid surrounding the plate 6 in the chamber 13 is drawn into the bores in which the said pistons 3 are sliding.
A plurality of bores 14 have been placed along the path traversed by the studs 12 in the part corresponding to the delivery, each bore being provided with a non-return valve 15 and communicating with a central chamber 16 which is drilled in the rear face of the swash plate 6 (FIG. 2).
As shown in FIG. 4, the bores 14 are spaced from one another at a distance such that any one stud of a group of studs corresponding to a rate of flow always has its central orifice communicating with a bore.
The result of this arrangement is that when the pump is a pump having several rates of flow, the pressure obtaining in the chamber 16 is always the strongest of the pressures obtaining in the circuits connected to the said pump.
By referring to FIGS. 5 and 6, it will be seen that the pin 5 on which the swash plate 6 is pivoted comprises along its longitudinal axis a duct 34 which communicates with the said chamber 16 through a duct 35 and a passage 35a, drilled in the body of the swash plate 6. As is shown in FIGS. 7 and 8, there are passages 36 and 37 radiating from the two ends of the duct 34 towards the bearings 38 carried by the supporting shaft 4, in which the pin 5 is journalled. The passages 36 are directed towards the face of the plate 6, against which the pistons 3 are adapted to bear, and the passages 37 are directed towards the face of the plate 6, against which the piston 8 is bearing. The purpose of this lack of symmetry is to balance at least partially the torsional force to which the pin 5 is subjected, because the force of the pistons 3 on the plate 6 is unsymmetrical.
Referring to FIG. 1, it is seen that the chamber 13 is supplied with liquid originating from a reservoir through a duct 17 in the pump body 1 along the axis of this latter, this being favourable to the supply of the blind hole 9 and consequently of the crescent-shaped passage 11.
This arrangement in which the liquid arrives through a duct opening perpendicularly of the centre of the chamber 13 assists the supply to the crescent-shaped passage 11 by a centrifugal action of the hydraulic liquid, and cancels out the distrubing effect which is due to the movement of the pistons.
The piston 8 bears against the rear face of the swash plate 6 by means of a stud 18 which covers the chamber 16.
This stud 18 comprises an annular recess 19 facing the plate 6 and an annular recess 20 facing the spherical head of the piston 8. These two recesses are supplied with oil under pressure coming from the chamber 16 through a conduit 21 comprising a calibrated passage 22.
The piston 8 is extended at its rear end by a rod 24 which slides in a bore 25 which discharges into a chamber 26 closed by a plug 27.
The rod 24 comprises a cylindrical shoulder 24a which, with the bore 25, defines an annular calibrated passage of variable length.
The piston 8 and the rod 24, together with the bore 8a, define an annular chamber 23. The said piston 8 carries a valve which receives the high pressure and the low pressure and which causes one or the other to communicate with the said chamber 23 for actuating the piston 8 in one direction or the other, so as to modify the inclination of the swash plate 6 and hence to adapt the rate of flow delivered by the pump.
In the example illustrated, this valve is formed by a slide valve 29 which slides in a bore 28 formed in the rod 24 and the piston 8, the said valve being counter-supported by a spring 30. In its middle portion, the slide valve 29 comprises an annular chamber 29a, into which discharges a conduit 31 communicating with the chamber 23; with the displacement of the slide valve 29, this chamber is capable of being brought into communication either with the chamber 16 at high pressure, by means of the conduit 32, or with the chamber 13 at low pressure, by means of the conduit 33.
When the shaft 4 is stationary, the parts in the position which is shown in FIG. 1; on the other hand, FIG. 2 illustrates the parts in an equilibrium position.
The swash plate 6 receives on one face a thrust which is caused by the pistons 3, which thrust is balanced by that of the piston 8. If the outlet pressures (in the case of a pump with 2 x 3 pistons delivering two independent rates of flow) are called P 1 and P 2 , and if s is the total section of the piston 8, the balancing pressure p in the chambers 23 and 26 is a function of P 1 and P 2 , in the form:
p s = P 1 S 1 + P 2 S 2 ,
s 1 and S 2 being the active means surfaces of the piston groups 3, that is to say, half the total surface of the assembly of pistons connected to the same flow.
In the particular case where the pump has two identical rates of flow, that is to say: S 1 = S 2 = S, then there is obtained
p = S/S (P 1 + P 2 ).
This pressure p acts on the slide valve 29 in opposition with the spring 30; in the balanced state, these two forces are equal.
The purpose of the non-return devices 15 will be noted, this being to select the higher of the two pressures P 1 and P 2 for supplying the valve and thus to permit that a displacement of the piston 8 is always able to correspond to a displacement of the said valve.
When one of the pressures P 1 and P 2 increases, the thrust on the plate 6 likewise increases, and this has the effect of increasing the balancing pressure p in the chambers 23 and 26. The effect of the increase in the pressure in the chamber 26 is to drive in the slide valve 29, which brings the conduits 32 and 31 into communication, in such a way that the high pressure obtaining in the chamber 16 is directed into the chamber 23. The piston 8 then starts to be displaced towards the right at a speed determined by the speed at which the slide valve is driven in, this movement of said valve being continued as long as there is lack of equilibrium between the pressure in the chamber 26 and the action of the spring 30.
As soon as the movement of the piston starts, the volume of the chamber 26 increases. The delivery which is necessary for making good this increase in volume is proportional to the speed of the jack. This delivery or flow comes from the chamber 23, while undergoing a pressure drop through the calibrated orifice 24a, this drop in pressure being proportional to the said flow, that is to say, to the speed of the jack. The signalling pressure in the chamber 26, which was p in the state of equilibrium and which was changed to, shall we say, p', because of the increase in P 1 or P 2 , is thus lowered in proportion as the slide valve moves inwardly and the jack is accelerated. However, all this has taken place over approximately a few hundredths of millimetres. This fall in the pressure in the chamber 26 reduces the lack of balance in the slide valve 29, and this finally immobilises it. This immobilisation is produced substantially when the pressure in the chamber 26 has once again become p, since the movement of the slide valve 29 and the bringing of the piston 8 to speed are produced over a few hundredths of a millimetre and the calibration of the spring 30 is thus not substantially affected. The initial pressure drop in the calibration 24a is then almost exactly p'-p, that is to say, the initial speed of the jack is substantially proportional to the initial spacing between p' and the pressure corresponding to the tension of the spring 30.
However, with the valve open, the movement of the piston 8 is continued. The spring 30 becomes increasingly compressed and continuously moves the slide valve 29 towards the neutral position. It should be shown that, at any moment, the speed of the jack is substantially proportional to the residual variation between p' and the pressure corresponding to the tension of the spring 30 at the moment under consideration, and this until this variation becomes zero, that is to say, a new balance is reached between the tension of the spring and the pressure p' in the chambers 23 and 26.
In practice, the determination of the characteristics of the spring for obtaining the desired relation is effected experimentally, point by point. A spring of very weak rigidity will give a cubic capacity which varies little as a function of the pressure and as a result one will come close to a pump which cancels out its rate of flow at constant pressure; in any case, a spring of constant rigidity will give a torque on the shaft of which the value will decrease in proportion as the pressures increase, that is to say, as the cubic capacities become smaller; in order to obtain a constant torque, it will be necessary to use a spring of which the rigidity is increased at the same time as the amount of flexion.
It will be noted that, as a function of the chosen law concerning pressures and cubic capacities or as a function of the technological facilities, the spring can be a single spring or there may be several separate springs or even springs in two groups: one between the slide valve 29 and the shaft 4, the spring 30, and the other between the slide valve 29 and the rod 24, as is shown in FIG. 3, this being the spring 30a.
EXAMPLE
A double delivery pump has been developed which comprises six equal pistons 3 with a diameter of 31.00 mm, delivering at maximum cubic capacity (maximum inclination of the plate 6) two equal rates of flow of 50 cc/revolution. The piston 8 has a diameter of 60 mm, the rod 24 has a diameter of 18 mm and the slide valve 29 a diameter of 6 mm.
Assuming that these parts are balanced in the position shown in FIG. 2, with P 1 = 350 bars and P 2 = 250 bars, it is easily calculated that: ##SPC1##
The spring 30 is calibrated so as to balance this pressure of 242.7 bars which is acting on the slide valve 29.
If the pressure P 1 rises to 400 bars, as a result of an increasing force in one of the installations supplied by the pump, an identical calculation shows that the pressure p rises to p' = 262.9 bars approximately.
The spring 30 is then no longer sufficient to hold the slide valve 29, which is moved forward, and, as has been explained the piston 8 starts moving. As the speed of equilibrium is reached over a very short travel and as the displacements of the valve are negligible, it can be considered that the calibration of the spring 30 remains unchanged. The valve stops being opened, this when the pressure in the chamber 26 has fallen again to 242.7 bars. There then exists in the chamber 23 a transient pressure p't, which has the value:
π/4 2 2/(18 - 6) × 242.7 + π/4 2 2/(60 - 18) × p't = 1.5 π/.4 2/31 (400 + 250)
that is to say: ##SPC2##
p't = 961 × 975/3,276 - 288 × 242.7/3,276 264.7 bars
The pressure drop in the calibration 24a is thus 264.7 - 242.7 = 22 bars and the speed of the piston 8 is proportional to this quantity.
The movement of the piston 8 is continued and the spring 30 is compressed. When a compression is reached for which it is balancing, for example, 260 bars, a calculation identical with the preceding calculation shows that the pressure p't in the chamber 23 is then equivalent to 263.1 bars. The pressure drop in the calibration 24a is not equivalent to more than 263.1 - 260.0 = 3.1 bars, and it is seen that the speed of the piston 8 is decreased by about 86% with respect to the initial speed. In addition, the piston 8 tends progressively towards a zero speed, which it reaches at the point when the spring balances the pressure of 262.9 bars.
So as to ensure a better stability of the parts, it is possible to provide a spring 39 in the chamber 23, the purpose of which is to hold the piston 8 bearing on the stud 18. The previous calculation still remains valid, because the influence of such a spring is negligible, in the same way as that of the springs 7 of the pistons.
It is seen that the calibration 24a makes it possible to obtain a very good control of the speed of the piston 8 as a function of the error remaining to be made good, that is to say, a very good damping of the readjustments which are constantly necessary in order to maintain the selected relationship between the cubic capacities and the pressures. In addition, it is seen that the effective length of the calibration 24a is increased when the cubic capacities decrease; thus, a like variation in pressure will cause an initial speed which is smaller in proportion as the cubic capacity is smaller; moreover, experience shows that it is necessary for this to be so, so as to maintain everywhere the same quality as regards stability, especially for example, in the operation with a constant torque P 1 .V 1 + P 2 .V 2 = constant, V 1 and V 2 representing the cylinders of the two circuits.
In addition, it has to be noted that this arrangement can be utilised in the case where the separate rates of flow are produced by pistons of different diameters, the previous calculations remaining applicable without any change.
FIGS. 9 to 12 relate to a modified constructional form concerned with the means for permitting the selection of the highest of the delivery pressures and of directing it on to the control valve.
In these FIGURES, the elements which correspond to those described in connection with FIGS. 1 to 8 bear the same references.
In the pump which is shown in the previous FIGURES, the valve receives the high pressure through the conduit 32 which discharges at the middle of the bearing stud 18, which encloses the chamber 16 in which obtains the highest of the delivery pressures.
In accordance with a modified construction which is shown in FIGS. 9 to 12, the strongest delivery pressure is selected by non-return valves 40 placed in the pump body 1 immediately downstream of the pistons 3.
The pump comprises 12 pistons 3 which are grouped in threes so that the pump has four independent delivery flows: the first delivery is supplied by the pistons 3a, the second by the pistons 3b, the third by the piston 3c and the fourth by the pistons 3d; the deliveries of the pistons of a single group are connected with one another downstream of the non-return valves 42a, 42b, 42c, 42d by conduits 41, such as 41a for the pistons 3a.
The delivery of the group a is thus taken by the non-return valve 40a, that of the group b by the non-return valve 40b, and so on, and downstream of the closure member, the non-return valves 40a, 40b, 40c and 40d are interconnected by a conduit 43. Thus, the pressure obtaining in the conduit 43 is always the highest of the four delivery pressures.
This pressure, thus selected, is carried through a conduit 44 drilled in the casing 1-2 of the pump as far as a passage 45 drilled in the jack 8, opening on to the slide member 29 of the valve.
The connection between the passage 44 and the passage 45 is obtained by means of a rotating joint formed by a ring 46 mounted with friction on the shaft 4, the said ring 46 being immobilised by a stop member 47.
The chamber 29a of the slide valve 29 communicates, by means of the passage 31, firstly with the chamber 23 situated behind the jack and, secondly, with the central space of the supporting stud 18'.
This supporting stud 18' is preferably biconical and receives, on the one hand, the spherical head of the jack 8 and, on the other hand, the spherical part of a hub 48 placed at the back of the swash plate 6. Formed in the body of this hub 48 is a passage 35a which communicates with the passage 35.
This modification has two advantages as compared with the embodiment in FIGS. 1 to 8: firstly, the means for selecting the highest delivery pressure are no longer actually placed inside the swash plate, which is more difficult to achieve from the point of view of machining operations, and secondly the hydraulic balancing of the supports of the jack 8 and of the hub or boss 48 on the stud 18' is much better. Actually, in the pump as shown in FIGS. 1 to 8, the pressure obtaining inside the stud 18 is always the highest of the delivery pressures, and this makes necessary the determination of the contact circle between jack 8 and stud 18 as a function of the highest possible delivery pressure, and the result of this is a hydraulic over-balancing of the bearing of the jack on its stud; this over-balancing necessitates the presence of a calibrated leakage flow device, such as that bearing the references 19 to 22.
On the contrary, with the arrangement shown in FIGS. 9 and 10, the pressure obtaining inside the stud is always equal to that obtaining in the chamber 23 and hence practically always equal to the pressure acting on the said jack 8.
FIGS. 11 and 12 show a modified form of the system for selecting delivery pressures.
It may actually be proved to be advantageous in certain cases to have an automatic stopping of the movements of the plate 6 when it reaches the zero delivery position.
To this end, and as shown in FIGS. 11 and 12, the non-return valves 40 are placed upstream in the chambers of the pistons 3, and no longer downstream of the non-return valves 42.
As is illustrated in FIG. 12, it is possible to position a non-return valve 40 in each chamber or even to position only one non-return valve 40 per delivery flow, by preferably selecting four pistons which follow one another, the arrangement being similar to that of FIG. 10, but the conduits such as 41a are then no longer shown, since they are at the downstream level of the valves 42.
In this FIGURE, the references 3c and 3d no longer designate the delivery pistons of the pump, but the bores in which the said pistons slide.
The present invention relates to improvements in hydraulic pumps of variable cubic capacity and particularly to pumps of the type such as described in U.S. Pat. No. 3,575,534 and having as its title "Constant torque hydraulic pump." The pump described in this patent is a pump having a rotatable swash plate, of which the inclination is modified by means of a hydraulic jack, of which the supply is controlled by a valve detecting the increases in pressures in the circuit supplied by the pump.
The present invention concerns a pump operating by the same general principle, but comprising a distribution system which is considerably improved as compared with that described in U.S. Pat. No. 3,575,534 of Feb. 5, 1969.
Actually, pumps providing separate rates of flow are being used to an ever-increasing extent: for example, with a pump having six pistons, it is possible to split up the flow of the pump into two separate flows, each supplied by three pistons, or even with a pump having nine pistons, into three flows each supplied by three pistons. It is quite obvious that, in these cases, it is necessary to have a valve capable of forming the sum of the pressures obtaining in the three circuits so as to act as a function of the torque necessary for driving the pump.
The distribution system described in U.S. Pat. No. 3,575,534 would make it necessary, when the pump is one having several flows, to have a valve body having as many stages as there are flows or deliveries; it would only be possible to obtain a satisfactory functioning of such a system at the cost of complex, delicate and troublesome means.
Another object of the invention is to permit the manufacture of the pumps to be standardised, a single pump head being capable of being used with different bodies having three, or six, or nine pistons.
A last object of the invention is to permit the supply of the pump to be achieved by an axial bore in the pump body, this considerably improving the volumetric yield thereof, but is not possible with a pump such as that which is described in U.S. Pat. No. 3,575,534.
BRIEF DESCRIPTION OF THE DRAWINGS
In the accompanying drawings, for facilitating the understanding of the invention and given by way of example:
FIG. 1 is a longitudinal section of a pump with 6 pistons, in accordance with the invention;
FIG. 2 is a partial view, as a longitudinal section along a plane perpendicular to the sectional plane of FIG. 1, in which a single piston is represented and this piston is offset by one-twelth of a revolution, so that the FIGURE may be more clear;
FIG. 3 is a partial view illustrating a modified form of the valve in FIG. 1;
FIG. 4 is a top plan view of the swash plate of FIG. 1;
FIG. 5 is a partial view of FIG. 1, illustrating the swash plate in the position which it occupies when the delivery is zero;
FIG. 6 is a sectional view along A--A of FIG. 5;
FIG. 7 is a partial view, as a section B--B of FIG. 6;
FIG. 8 is a partial view, as a section C--C of FIG. 6;
FIG. 9 is a longitudinal sectional view along the line A--A of FIG. 10 of a 12-piston pump, grouped into four deliveries, comprising the improvements according to the present addition;
FIG. 10 is a sectional view along the line B--B of FIG. 9;
FIG. 11 is a partial view as a longitudinal section of a pump similar to that of FIG. 9, comprising a modification;
FIG. 12 is a sectional view along the line DD of FIG. 11.
DETAILED DESCRIPTION
Referring to FIG. 1, it is seen that the pump is in two parts: the pump body 1, comprising the pumping pistons 3, and the pump head 2, comprising the driving shaft 4 which, by means of a pin 5, supports the swash plate 6. When the shaft 4 is driven in rotation by a motor (not shown) it drives the swash plate 6, which alternately pushes back the pistons 3, which are countersupported by springs 7. As the swash plate 6 is pivoted on a pin 5, its inclination can be modified, the travel of the pistons 3 decreasing as a function of the decrease in the angle of the plate 6.
The modifications in the inclination of the plate 6 are caused by a hydraulic jack formed by a piston 8 sliding in a bore 8a drilled axially in the shaft 4 and possibly provided with a sleeve or liner 8b, so that during operation, the plate receives the thrust of the pistons 3 on one side and thrust of the piston 8 on the other side.
When there is an odd number of pistons 3, for example five pistons, the thrust exerted by the front face of the plate is alternately caused by two or three pistons, and the result thereof is a high frequency pulsation of the pressure obtaining at the rear of the piston 8, which pulsation cancels out the frictional effects of the various movable elements of the system. By way of example, with a pump having five pistons, the thrust experienced by the front face of the swash plate oscillates by 2.5 ± 20%, and this, when the shaft 4 is turning at 1,500 r.p.m., corresponds to a pulsation at a frequency of 250 cycles per second.
When there is an even number of pistons 3, for example, six pistons, and when the pump is a double-acting pump, the swash plate 6 also experiences high frequency pressure pulsations when the two deliveries are not at the same pressure. When the two deliveries are at the same pressure, or when the pump has a single delivery, there is also a high frequency pulsation which arises from the fact that, because of compressibility phenomena which are shown at the pressures under consideration, the travel of the pistons with a rise in pressure is not symmetrical with that with a fall in pressure, relatively to the line of maximum slope of the plate.
Referring to FIGS. 2 and 4, it is seen that the swash plate 6 comprises a central blind hole 9 which communicates by way of a passage 10 with a crescent-shaped passage 11, over which the studs 12 of the pistons 3 pass at the time of rotation of the plate 6. When the studs 12 pass over the crescent-shaped passage 11, the pistons 3 are in their suction stroke and the liquid surrounding the plate 6 in the chamber 13 is drawn into the bores in which the said pistons 3 are sliding.
A plurality of bores 14 have been placed along the path traversed by the studs 12 in the part corresponding to the delivery, each bore being provided with a non-return valve 15 and communicating with a central chamber 16 which is drilled in the rear face of the swash plate 6 (FIG. 2).
As shown in FIG. 4, the bores 14 are spaced from one another at a distance such that any one stud of a group of studs corresponding to a rate of flow always has its central orifice communicating with a bore.
The result of this arrangement is that when the pump is a pump having several rates of flow, the pressure obtaining in the chamber 16 is always the strongest of the pressures obtaining in the circuits connected to the said pump.
By referring to FIGS. 5 and 6, it will be seen that the pin 5 on which the swash plate 6 is pivoted comprises along its longitudinal axis a duct 34 which communicates with the said chamber 16 through a duct 35 and a passage 35a, drilled in the body of the swash plate 6. As is shown in FIGS. 7 and 8, there are passages 36 and 37 radiating from the two ends of the duct 34 towards the bearings 38 carried by the supporting shaft 4, in which the pin 5 is journalled. The passages 36 are directed towards the face of the plate 6, against which the pistons 3 are adapted to bear, and the passages 37 are directed towards the face of the plate 6, against which the piston 8 is bearing. The purpose of this lack of symmetry is to balance at least partially the torsional force to which the pin 5 is subjected, because the force of the pistons 3 on the plate 6 is unsymmetrical.
Referring to FIG. 1, it is seen that the chamber 13 is supplied with liquid originating from a reservoir through a duct 17 in the pump body 1 along the axis of this latter, this being favourable to the supply of the blind hole 9 and consequently of the crescent-shaped passage 11.
This arrangement in which the liquid arrives through a duct opening perpendicularly of the centre of the chamber 13 assists the supply to the crescent-shaped passage 11 by a centrifugal action of the hydraulic liquid, and cancels out the distrubing effect which is due to the movement of the pistons.
The piston 8 bears against the rear face of the swash plate 6 by means of a stud 18 which covers the chamber 16.
This stud 18 comprises an annular recess 19 facing the plate 6 and an annular recess 20 facing the spherical head of the piston 8. These two recesses are supplied with oil under pressure coming from the chamber 16 through a conduit 21 comprising a calibrated passage 22.
The piston 8 is extended at its rear end by a rod 24 which slides in a bore 25 which discharges into a chamber 26 closed by a plug 27.
The rod 24 comprises a cylindrical shoulder 24a which, with the bore 25, defines an annular calibrated passage of variable length.
The piston 8 and the rod 24, together with the bore 8a, define an annular chamber 23. The said piston 8 carries a valve which receives the high pressure and the low pressure and which causes one or the other to communicate with the said chamber 23 for actuating the piston 8 in one direction or the other, so as to modify the inclination of the swash plate 6 and hence to adapt the rate of flow delivered by the pump.
In the example illustrated, this valve is formed by a slide valve 29 which slides in a bore 28 formed in the rod 24 and the piston 8, the said valve being counter-supported by a spring 30. In its middle portion, the slide valve 29 comprises an annular chamber 29a, into which discharges a conduit 31 communicating with the chamber 23; with the displacement of the slide valve 29, this chamber is capable of being brought into communication either with the chamber 16 at high pressure, by means of the conduit 32, or with the chamber 13 at low pressure, by means of the conduit 33.
When the shaft 4 is stationary, the parts in the position which is shown in FIG. 1; on the other hand, FIG. 2 illustrates the parts in an equilibrium position.
The swash plate 6 receives on one face a thrust which is caused by the pistons 3, which thrust is balanced by that of the piston 8. If the outlet pressures (in the case of a pump with 2 x 3 pistons delivering two independent rates of flow) are called P 1 and P 2 , and if s is the total section of the piston 8, the balancing pressure p in the chambers 23 and 26 is a function of P 1 and P 2 , in the form:
p s = P 1 S 1 + P 2 S 2 ,
s 1 and S 2 being the active means surfaces of the piston groups 3, that is to say, half the total surface of the assembly of pistons connected to the same flow.
In the particular case where the pump has two identical rates of flow, that is to say: S 1 = S 2 = S, then there is obtained
p = S/S (P 1 + P 2 ).
This pressure p acts on the slide valve 29 in opposition with the spring 30; in the balanced state, these two forces are equal.
The purpose of the non-return devices 15 will be noted, this being to select the higher of the two pressures P 1 and P 2 for supplying the valve and thus to permit that a displacement of the piston 8 is always able to correspond to a displacement of the said valve.
When one of the pressures P 1 and P 2 increases, the thrust on the plate 6 likewise increases, and this has the effect of increasing the balancing pressure p in the chambers 23 and 26. The effect of the increase in the pressure in the chamber 26 is to drive in the slide valve 29, which brings the conduits 32 and 31 into communication, in such a way that the high pressure obtaining in the chamber 16 is directed into the chamber 23. The piston 8 then starts to be displaced towards the right at a speed determined by the speed at which the slide valve is driven in, this movement of said valve being continued as long as there is lack of equilibrium between the pressure in the chamber 26 and the action of the spring 30.
As soon as the movement of the piston starts, the volume of the chamber 26 increases. The delivery which is necessary for making good this increase in volume is proportional to the speed of the jack. This delivery or flow comes from the chamber 23, while undergoing a pressure drop through the calibrated orifice 24a, this drop in pressure being proportional to the said flow, that is to say, to the speed of the jack. The signalling pressure in the chamber 26, which was p in the state of equilibrium and which was changed to, shall we say, p', because of the increase in P 1 or P 2 , is thus lowered in proportion as the slide valve moves inwardly and the jack is accelerated. However, all this has taken place over approximately a few hundredths of millimetres. This fall in the pressure in the chamber 26 reduces the lack of balance in the slide valve 29, and this finally immobilises it. This immobilisation is produced substantially when the pressure in the chamber 26 has once again become p, since the movement of the slide valve 29 and the bringing of the piston 8 to speed are produced over a few hundredths of a millimetre and the calibration of the spring 30 is thus not substantially affected. The initial pressure drop in the calibration 24a is then almost exactly p'-p, that is to say, the initial speed of the jack is substantially proportional to the initial spacing between p' and the pressure corresponding to the tension of the spring 30.
However, with the valve open, the movement of the piston 8 is continued. The spring 30 becomes increasingly compressed and continuously moves the slide valve 29 towards the neutral position. It should be shown that, at any moment, the speed of the jack is substantially proportional to the residual variation between p' and the pressure corresponding to the tension of the spring 30 at the moment under consideration, and this until this variation becomes zero, that is to say, a new balance is reached between the tension of the spring and the pressure p' in the chambers 23 and 26.
In practice, the determination of the characteristics of the spring for obtaining the desired relation is effected experimentally, point by point. A spring of very weak rigidity will give a cubic capacity which varies little as a function of the pressure and as a result one will come close to a pump which cancels out its rate of flow at constant pressure; in any case, a spring of constant rigidity will give a torque on the shaft of which the value will decrease in proportion as the pressures increase, that is to say, as the cubic capacities become smaller; in order to obtain a constant torque, it will be necessary to use a spring of which the rigidity is increased at the same time as the amount of flexion.
It will be noted that, as a function of the chosen law concerning pressures and cubic capacities or as a function of the technological facilities, the spring can be a single spring or there may be several separate springs or even springs in two groups: one between the slide valve 29 and the shaft 4, the spring 30, and the other between the slide valve 29 and the rod 24, as is shown in FIG. 3, this being the spring 30a.
EXAMPLE
A double delivery pump has been developed which comprises six equal pistons 3 with a diameter of 31.00 mm, delivering at maximum cubic capacity (maximum inclination of the plate 6) two equal rates of flow of 50 cc/revolution. The piston 8 has a diameter of 60 mm, the rod 24 has a diameter of 18 mm and the slide valve 29 a diameter of 6 mm.
Assuming that these parts are balanced in the position shown in FIG. 2, with P 1 = 350 bars and P 2 = 250 bars, it is easily calculated that: ##SPC1##
The spring 30 is calibrated so as to balance this pressure of 242.7 bars which is acting on the slide valve 29.
If the pressure P 1 rises to 400 bars, as a result of an increasing force in one of the installations supplied by the pump, an identical calculation shows that the pressure p rises to p' = 262.9 bars approximately.
The spring 30 is then no longer sufficient to hold the slide valve 29, which is moved forward, and, as has been explained the piston 8 starts moving. As the speed of equilibrium is reached over a very short travel and as the displacements of the valve are negligible, it can be considered that the calibration of the spring 30 remains unchanged. The valve stops being opened, this when the pressure in the chamber 26 has fallen again to 242.7 bars. There then exists in the chamber 23 a transient pressure p't, which has the value:
π/4 2 2/(18 - 6) × 242.7 + π/4 2 2/(60 - 18) × p't = 1.5 π/.4 2/31 (400 + 250)
that is to say: ##SPC2##
p't = 961 × 975/3,276 - 288 × 242.7/3,276 264.7 bars
The pressure drop in the calibration 24a is thus 264.7 - 242.7 = 22 bars and the speed of the piston 8 is proportional to this quantity.
The movement of the piston 8 is continued and the spring 30 is compressed. When a compression is reached for which it is balancing, for example, 260 bars, a calculation identical with the preceding calculation shows that the pressure p't in the chamber 23 is then equivalent to 263.1 bars. The pressure drop in the calibration 24a is not equivalent to more than 263.1 - 260.0 = 3.1 bars, and it is seen that the speed of the piston 8 is decreased by about 86% with respect to the initial speed. In addition, the piston 8 tends progressively towards a zero speed, which it reaches at the point when the spring balances the pressure of 262.9 bars.
So as to ensure a better stability of the parts, it is possible to provide a spring 39 in the chamber 23, the purpose of which is to hold the piston 8 bearing on the stud 18. The previous calculation still remains valid, because the influence of such a spring is negligible, in the same way as that of the springs 7 of the pistons.
It is seen that the calibration 24a makes it possible to obtain a very good control of the speed of the piston 8 as a function of the error remaining to be made good, that is to say, a very good damping of the readjustments which are constantly necessary in order to maintain the selected relationship between the cubic capacities and the pressures. In addition, it is seen that the effective length of the calibration 24a is increased when the cubic capacities decrease; thus, a like variation in pressure will cause an initial speed which is smaller in proportion as the cubic capacity is smaller; moreover, experience shows that it is necessary for this to be so, so as to maintain everywhere the same quality as regards stability, especially for example, in the operation with a constant torque P 1 .V 1 + P 2 .V 2 = constant, V 1 and V 2 representing the cylinders of the two circuits.
In addition, it has to be noted that this arrangement can be utilised in the case where the separate rates of flow are produced by pistons of different diameters, the previous calculations remaining applicable without any change.
FIGS. 9 to 12 relate to a modified constructional form concerned with the means for permitting the selection of the highest of the delivery pressures and of directing it on to the control valve.
In these FIGURES, the elements which correspond to those described in connection with FIGS. 1 to 8 bear the same references.
In the pump which is shown in the previous FIGURES, the valve receives the high pressure through the conduit 32 which discharges at the middle of the bearing stud 18, which encloses the chamber 16 in which obtains the highest of the delivery pressures.
In accordance with a modified construction which is shown in FIGS. 9 to 12, the strongest delivery pressure is selected by non-return valves 40 placed in the pump body 1 immediately downstream of the pistons 3.
The pump comprises 12 pistons 3 which are grouped in threes so that the pump has four independent delivery flows: the first delivery is supplied by the pistons 3a, the second by the pistons 3b, the third by the piston 3c and the fourth by the pistons 3d; the deliveries of the pistons of a single group are connected with one another downstream of the non-return valves 42a, 42b, 42c, 42d by conduits 41, such as 41a for the pistons 3a.
The delivery of the group a is thus taken by the non-return valve 40a, that of the group b by the non-return valve 40b, and so on, and downstream of the closure member, the non-return valves 40a, 40b, 40c and 40d are interconnected by a conduit 43. Thus, the pressure obtaining in the conduit 43 is always the highest of the four delivery pressures.
This pressure, thus selected, is carried through a conduit 44 drilled in the casing 1-2 of the pump as far as a passage 45 drilled in the jack 8, opening on to the slide member 29 of the valve.
The connection between the passage 44 and the passage 45 is obtained by means of a rotating joint formed by a ring 46 mounted with friction on the shaft 4, the said ring 46 being immobilised by a stop member 47.
The chamber 29a of the slide valve 29 communicates, by means of the passage 31, firstly with the chamber 23 situated behind the jack and, secondly, with the central space of the supporting stud 18'.
This supporting stud 18' is preferably biconical and receives, on the one hand, the spherical head of the jack 8 and, on the other hand, the spherical part of a hub 48 placed at the back of the swash plate 6. Formed in the body of this hub 48 is a passage 35a which communicates with the passage 35.
This modification has two advantages as compared with the embodiment in FIGS. 1 to 8: firstly, the means for selecting the highest delivery pressure are no longer actually placed inside the swash plate, which is more difficult to achieve from the point of view of machining operations, and secondly the hydraulic balancing of the supports of the jack 8 and of the hub or boss 48 on the stud 18' is much better. Actually, in the pump as shown in FIGS. 1 to 8, the pressure obtaining inside the stud 18 is always the highest of the delivery pressures, and this makes necessary the determination of the contact circle between jack 8 and stud 18 as a function of the highest possible delivery pressure, and the result of this is a hydraulic over-balancing of the bearing of the jack on its stud; this over-balancing necessitates the presence of a calibrated leakage flow device, such as that bearing the references 19 to 22.
On the contrary, with the arrangement shown in FIGS. 9 and 10, the pressure obtaining inside the stud is always equal to that obtaining in the chamber 23 and hence practically always equal to the pressure acting on the said jack 8.
FIGS. 11 and 12 show a modified form of the system for selecting delivery pressures.
It may actually be proved to be advantageous in certain cases to have an automatic stopping of the movements of the plate 6 when it reaches the zero delivery position.
To this end, and as shown in FIGS. 11 and 12, the non-return valves 40 are placed upstream in the chambers of the pistons 3, and no longer downstream of the non-return valves 42.
As is illustrated in FIG. 12, it is possible to position a non-return valve 40 in each chamber or even to position only one non-return valve 40 per delivery flow, by preferably selecting four pistons which follow one another, the arrangement being similar to that of FIG. 10, but the conduits such as 41a are then no longer shown, since they are at the downstream level of the valves 42.
In this FIGURE, the references 3c and 3d no longer designate the delivery pistons of the pump, but the bores in which the said pistons slide.