VARIABLE OUTPUT OIL PUMP

Abstract: The invention provides a variable output pump of the gerotor type in which one rotor (16) meshes in two side by side annulii (14, 12) which can be aligned for maximum output and turned in opposite directions to be non-aligned for minimum output. The movement is effected by a gear train (20, 24, 29) displaced by a rack (40) which is hydraulically operated by pump pressure. Hence high pressure displaces the rack to result in lower output against lower pressure. A variation is described in which a control spool diverts fluid to either side of the piston according to need and this is displaced in one direction by the derived fluid pressure and in the other direction additionally by spring pressure transferred from the piston position by a lever system. This can give desired pump output/speed results at reduced power consumption.

Description: This invention relates to variable output oil pumps of the kind described for example in EPA 0076033 in which a single male lobed rotor with n lobes extends through and is meshed with a pair of side-by-side female lobed annulii of n + x lobes (usually but not essentially x = 1) and each annulus is journalled in a separate eccentric ring and having an external gear, each eccentric ring being journalled in the body of the pump. The two gears can be turned in opposite directions by a pinion meshed with the two gears and driven by a rack displaced in one direction by pressure derived from the pump itself and urged by a torsion spring in the opposite direction so that the eccentrics go from identical positions, when the annulii are aligned, to opposite generally mirror image non-aligned positions. This movement is relative to the fixed inlet and outlet ports of the pump and the effect is to go from maximum volume to minimum volume of the pumping chambers formed between the lobes: hence such movement changes the pump from maximum to minimum output or vice versa. But the actual output is also (at least approximately) proportional to the rotational speed. The intent is to provide variable output between the extremes, necessary because the pumped fluid requirement, for example as i.c. engine lubricating oil, is not proportional to pump speed when the pump is driven, for example gear driven, from the engine. EP-A-0051192 describes a variable output vane pump having a pair of rings with oval shaped inner contours mounted side-by-side for relative rotation from a position where the ovals are in register to a moved position where they are out of register. This has a similar effect in terms of pump output to that in the lobed rotor gerotor pump mentioned above. Movement is caused by a valve spool balanced between the discharge fluid pressure of the pump and the force exerted on the spool by a valve spring. This is so that with no discharge pressure the rings are moved to a full delivery position. As discharge pressure builds up it acts against the spool and when high enough to overcome the valve spring communication is opened to port discharge pressure to a piston which in turn is acted upon by a torsion spring and when the discharge pressure is sufficient the piston moves to turn the rings towards the minimum displacement position. The force of the valve spring is adjusted so that when the pump discharge pressure reaches a maximum the valve spool uncovers a passage which ensures that the piston moves the rings to a zero displacement position. The problem with these arrangements is that the output/drive speed relationship tends to be linear since this is not necessarily required there is power wastage. The problem is solved according to the present invention by balancing the hydraulic pressure causing displacement of the pump parts in one direction against spring pressure tending to return the parts, but arranging for the spring pressure to be increased by the hydraulic pressure. This results in a modified output/speed graph which results in power saving. For the hydraulic pressure, the actual pump outlet pressure may be used, i.e. the outlet may be connected to a piston to drive said gears, or the oil pressure in a main lubricant gallery of the engine may be used, for example. The invention is more particularly described with reference to the accompanying drawings in which: Figure 1 is a diagrammatic elevation showing drive arrangements; Figure 2 is a sectional plan on line 2-2 of Figure 1; Figure 3 shows the drive arrangements; and Figure 4 is an illustrative performance graph. Turning now to the drawings and particularly Figures 1 and 2 thereof, the part 10 forms a casing for the pump to house the eccentrics 12 which in turn receive the internally toothed annuli 14 surrounding the rotor 16. The illustrated rotor 16 has five teeth and the annulus 14 has six teeth. However other formations are possible. The rotor is driven by shaft 18. The rotor 16 is a single component but the annulus 14 is a pair of components located axially end-to-end, and each annulus is located in a corresponding eccentric. As best seen in Figure 2, the eccentrics are each provided with straight cut spur pinion teeth 20, 22, and between the two sets of pinion teeth, two roller bearing cages 24 extend having two axially extending end-to-end sets of rollers 26, 28 with individual cage sets to allow contra-rotation. The needle roller bearings are effective between the two eccentric components 12 and the casing 10. The location of the needle rollers and the teeth is diagrammatically indicated in Figure 1 by the reference 20/24. The drive arrangements are best seen in Figure 2. Drive shaft 27 is pinned to straight cut pinion 29 meshed with the gear ring 22. It is also keyed at 30 to a further such pinion 32 which is in turn meshed with pinion 34 journalled on shaft 36 and meshed with gear ring 20. It will be appreciated that when the shaft 27 turns, pinions 29 and 34 turn in opposite directions and likewise for the gear rings 20, 22 and hence the two eccentrics 12. A clock spring 38 or another torsion spring is or may be provided and connected to the shaft 27 for example to return the same to a position in which the eccentricity is at a maximum. In order to provide a drive for the shaft 27, pinion 29 is further meshed with rack 40. Said rack is fast with piston 104 (Figure 3) which also carries piston 102 located in the main drive cylinder 46. Parallel to the main cylinder is a control bore housing spool 114 and the two are interconnected by a system of passageways. The spool 114 has a pair of axially spaced waists 62, 64, a through bore 66, and a transverse bore 68. The control bore is exposed to pump outlet pressure by means of the axially extending passage 130 communicating (in this instance) with the outlet port of the pump. As mentioned earlier, the control bore could be communicated into alternative pressure sources e.g. an engine gallery. Figure 3 shows the rack in the extreme position in which the rotors are aligned for maximum volume output, e.g. as a startup. Spring 106 is a light compression spring extending between the piston and an end abutment 108 fixed to a drive pin 110 slidable in bush 112. Spool 114 comprises nose 116 which projects out of the pump body parallel to drive pin 110. A third parallel pin 120 provides a fulcrum at 112 for lever 124 which abuts both pins. A light spring 126 is provided to return the spool to the illustrated position when the pump is not running. The control spool is waisted and ported, and the control cylinder is ported and connected to the main cylinder so that fluid in passage 130 connected to the pump outlet port communicates via passages 66, 68, 132 to reach one end of the main cylinder 100, or via passage 138 to reach the other end of the main cylinder. Exhaust is via passage 136 and waist 62. In operation, at startup the arrangement is as illustrated. The pump is at maximum volume setting. As the drive speed increases the volumetric output increases, and so does pressure in the system connecting via bore 130. The porting ensures that the communicated pressure holds the piston 102 in the rotor aligned full volume position. When the pressure displaces the spool to the left the first result is to close the port 132 communicating fluid to the left hand end of the main cylinder. At the same time the nose 116 via lever 124 displaces end stop 108 so as to begin loading the spring 106. Further small movement of the spool connects port 132 with the exhaust port 136 via the major waist of the spool whilst connecting the pressure source in 130 via port 138 to the right hand end of the main cylinder so that the piston 102 is displaced leftwards. At the same time the spool nose 116 further displaces the lever to displace the spring cap 108 to the right. From this point on the rack piston floats in position, being balanced by derived pressure acting on it and displacing it leftwards, but, via the spring 106 tending to displace the control spool rightwards, i.e. reduce or cut off the source of the derived pressure causing the movement. Also, the land 140 may be of the same width as the port 138 which means that small movements one way or the other connect port 138 alternately to exhaust or supply (136 or 140). By these means it is now possible to stabilize pressure (and volumetric output) over a range of speeds. Moreover by suitable selection of parameters particularly for spring 106, it is possible to allow the rack to travel progressively as pressure rises before the floating and stabilized pressure regime becomes effective which ran result in considerable power savings. The power saving point is illustrated in Figure 4. Turning now to Figure 4, this is a generalised graph showing the pump output pressure/driven speed relationship for three different pumps. The line A-B is the possible straight line relationship for a constant output pump, for example the pump of the present invention if the rack were to be immobilised. The line A, C, D, E shows the results where the part A, C, D results from small progressive rack movement until the spring becomes effective and the part D, E from the effect of the spring. The line A, C, E shows the effect of the Figure 3 version with the portion C, E resulting from the floating or balance action of the rack piston. (It is assumed that point E is a desired pressure/speed relationship to be attained by the pump). Effectively the single and cross-hatched area is the saving in power achieved as compared to a fixed volume pump.

Claims: 1. A pump of the kind comprising a single male lobed rotor (16) with n lobes which extends through and is meshed with a pair of side by side female lobed annulii of n + x lobes, and each annulus is journalled in a separate eccentric ring (12) having an external gear (20, 22) and journalled in the body of the pump, a hydraulically driven piston (102) for turning the gears in opposite directions so as to take the annulii between aligned identical positions for maximum pump output and non-aligned mirror image positions for minimum pump output, said movement being relative to fixed inlet and outlet ports of the pump, said piston being hydraulically driven in one direction by pressure derived from the pump output, and with spring means (38) for returning the parts, characterised in that said piston (102) is exposed to the hydraulic pressure on both sides via passages (132, 138) connected to a spool valve (114), said piston (102) is also urged by a spring (106) and a mechanical connection (116, 124, 110) is provided between the spool and an abutment (108) for said piston spring (106) so that increased hydraulic pressure tending to displace the piston in the direction which turns the annulii to reduce pump output causes the spring abutment (108) to displace in the direction which increases the spring resistance to such movement. 2. Apparatus as claimed in Claim 1 wherein the said means for turning the gears is connected to a main piston (102) and cylinder, and said spool valve is parallel to the main piston and cylinder and said mechanical arrangement comprises a lever pivoted between the main cylinder and the spool, connected to the spool to be displaced thereby at one end of the lever, and connected to said spring abutment at the opposite end of the lever.

Inventors: CHILD, ROBIN EDWARD
Applicants: CONCENTRIC PUMPS LIMITED
Application year: 1988
Main classification: F04C15
IPC codes: F04C2, F04C14