GB2441773A

GB2441773A - Pump Control System

Info

Publication number: GB2441773A
Application number: GB0618147A
Authority: GB
Inventors: Stephen Mark Hodge; Kevin Johanson
Original assignee: CONCENTRIC VFP Ltd
Current assignee: CONCENTRIC VFP Ltd
Priority date: 2006-09-15
Filing date: 2006-09-15
Publication date: 2008-03-19
Anticipated expiration: 2026-09-15
Also published as: GB0618147D0; GB2441773B

Abstract

An engine lubricant pump control system and a method of operating such a system comprising a variable output positive displacement pump, a sensor that measures either engine speed or lubricant pressure and control means that uses said measured value to increase lubricant delivery pressure at a first rate AB proportional to engine speed; then maintain lubricant delivery pressure approximately constant BC; then increase lubricant delivery pressure at a second rate CDE proportional to engine speed; the steps taking space sequentially at predetermined engine speeds as the speed increases from minimum to maximum. The pump may be of split rotor gerotor type and may be controlled electronically via a stepper motor or hydraulically via a piston moved rack and pinion. A spool valve (Fig. 4) is spring biased to prevent piston movement until a threshold pressure after which the pump is reduced in capacity in proportion to output pressure. At the end of the stroke pressure rises again proportional to pump speed.

Description

1

2441773

Pump Control System

The present invention relates to a control system for use with variable flow pumps and particularly pumps where the flow rate may be controlled independent of speed; for example gerotor type pumps.

When designing a reciprocating engine, it is important to ensure that there is a sufficient flow rate of oil through each component to prevent metal to metal contact, and also typically to dissipate waste heat released from engine. This inevitably leads to a situation where there is excessive oil being delivered at higher speeds which is normally dumped through the regulator valve so as to protect components which may be damaged by high pressures.

In it's simplest form, an engine lubrication system generally comprises a series of flow channels that provide fluid communication between individual engine components which require lubrication. These flow channels and associated components present resistance to flow. There may be restrictors in the flow passageways in order to direct the oil to where it is needed when the flow channels are divided.

Thus, a typical system of flow channels may link the following engine components: lubrication oil sump; oil pump; oil filter(s); cylinder block (crankshaft, big and little end bearings); and cylinder head (camshaft bearings, cam followers etc); there being return flow back to the oil sump.

The overall lubrication system has a total flow resistance nominally equal to the sum of all the individual resistances of the flow circuit. Thus, under most engine conditions the flow resistance is approximately constant; although it reduces as the engine ages as a result of increased clearances between moving parts. The resulting pressure drop depends both on the flow rate and the viscosity of the fluid being circulated. Therefore, the fluid flow rate required to attain a specified overall system pressure drop will be approximately constant once an engine has reached a steady operating temperature; however on start-up etc., as the oil temperature rises it becomes less viscous and so the flow rate required to generate a specified pressure drop increases. In a conventional system this is managed by a spill-back (pressure relief) valve which opens when the desired pressure has been reached. With hot oil this means that the regulator valve will open at a slightly higher speed/higher flow condition.

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The system is self compensating in that as the oil temperature rises more oil is allowed to flow into the system to maintain the desired pressure and less is dumped to sump.

Diesel engines typically have standard positive displacement type oil lubrication pumps. In such pumps the flow rate delivered by the pump is proportional to pump speed. Typically, such pumps are driven directly by the diesel engine and therefore deliver a lubricant flow rate that is proportional to engine speed. Such an arrangement works well for low to medium engine speeds; however at higher engine speeds spill back is necessary in order to prevent excessive oil delivery pressures being generated; a condition that may lead to damage of the engine lubrication system; for example oil gasket seals. Typically, spill back is provided by a pressure relief valve that opens at a specified pressure and thereby returns oil to the oil sump or to the oil pump inlet via a bypass conduit rather than via the engine lubrication channels. While such systems work in as much as they protect the engine against damage they are inefficient from an energy standpoint as a significant proportion of the energy used to drive the pump is lost as a result of such spill back. This situation leads to reduced overall engine efficiency.

The above problem may be largely overcome using variable output positive displacement type pumps. Thus, when it becomes necessary to limit lubricant delivery pressure as engine speed increases it is possible to maintain approximately constant delivery pressure even though pump speed is increasing. For example, the variable output gerotor type pump can achieve this result by appropriately varying an index angle between a first rotor portion and a second rotor portion. Such systems offer the advantage that there is a much reduced wastage of energy used to drive the pump, as spill back is eliminated. Therefore, using such systems it is possible to control a pump so that initially pump output is proportional to engine speed and then pump output is substantially constant as engine speed is further increased.

There are also situations where eventually, after passing through a substantially constant output phase, it is desirable to increase pump output as engine speed further increases; for example, when the engine employs piston cooling jets to protect the engine against overheating at high speeds. When the piston cooling jets are opened, it is necessary to increase the pump output in order to maintain the desired pressure in the lubrication system. Such control is difficult to achieve using known control systems and in particular the simple pump plus spill back type system.

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An aim of the present invention is to provide an improved lubricant pump control system that: avoids the use of spill back; provides two regions where pump flow rate increases with engine speed and an intermediate region where pump output and/or pump delivery pressure is substantially constant.

In one aspect the invention provides an engine lubricant pump control system comprising a variable output positive displacement pump, a sensor that provides an indication of engine speed, lubricant delivery pressure or both and control means that respond to the sensor and vary the output of the pump to:- (a) increase lubricant delivery pressure at a first rate as engine speed increases; then (b) maintain lubricant delivery pressure approximately constant over a range of engine speeds; then (c) increase lubricant delivery pressure at a second rate as engine speed increases; steps (a) to (c) taking space sequentially at predetermined values of lubricant delivery pressure as the engine speed is increased from minimum to maximum.

The lubricant delivery pressure may be the delivery pressure of the pump or a pressure in a selected portion within the lubricant flow-path through the engine.

Preferably, the positive displacement pump comprises a gerotor type pump with a split rotor.

The control means may comprise a spool valve in fluid communication with either the pump outlet and/or a part of the engine lubrication system. Thus, the control means may be energised by and may respond to the lubricant delivery pressure so communicated.

The control means may also comprise a spool valve that is energised by oil pump outlet pressure and responds to the lubricant pressure at a selected position within the lubricant flow-path through the engine.

A control system may also comprise a cylinder, piston, rack and pinion that in combination respond to the lubricant delivery pressure and wherein resultant movement of the pinion varies the offset angle (13 ) between the two portions of the split rotor.

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The control means may respond to measured values or indications of one or more of engine speed, pump outlet pressure, engine oil gallery pressure and accordingly adjust the delivery pressure to a value less than the maximum value available at a particular engine speed.

In another aspect the invention comprises a method of controlling a positive displacement variable output pump, used in an engine lubricant system, wherein engine speed, lubricant pressure or both are sensed/measured and control means that on the basis of this sensed/measured value at low engine speeds ensures that lubricant delivery pressure increases with engine speed at a first rate; at medium engine speeds ensures that the lubricant delivery pressure is approximately constant; and at highest speeds ensures that the lubricant delivery pressure increases with engine speed at a second rate.

The lubricant delivery pressure may be the delivery pressure of the pump or a pressure in a selected portion of the engine lubrication system.

Preferably, the variable output positive displacement pump comprises a gerotor type pump with a split rotor.

The control means may also comprise a spool valve that is energised by oil pump outlet pressure and responds to the engine oil gallery pressure.

A control system may also comprise a cylinder, piston, rack and pinion that in combination respond to the oil pump outlet pressure or engine oil gallery pressure and wherein resultant movement of the pinion varies the offset angle (13 ) between two portions of the split rotor.

The control means may respond to measured values or indications of one or more of engine speed, pump outlet pressure, engine oil gallery pressure and accordingly adjust the lubricant delivery pressure to a value less than the maximum value available at a particular engine speed.

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The invention will now be described by reference to the following diagramatic figures in which:

Figure 1 shows illustrative flow characteristics of a conventional lubricant pump fitted with spill back;

Figure 2 show one example of flow and pressure characteristics for a system according to the present invention;

Figure 3 shows a split rotor gerotor pump that may be used in the invention; and

Figures 4 and 5 show a mechanical/hydraulic system for controlling a pump according to the present invention.

Figure 1 shows flow illustrative data for a conventional positive displacement lubricant pump fitted with spill back. Over the engine speed range 0 to 750 RMP the flow rate increases linearly from 0 to 40 litres/min. If spill back were absent the flow rate would continue to increase linearly. However, in order to prevent excessive delivery pressures spill back has been added; comprising a pressure relief valve that opens at a pressure of 4.0 bar (in this illustrative case). Thus when the pump outlet pressure reaches 4.0 bar (at 750 RPM in this case) the relief valve opens. This results in the pump outlet pressure remaining approximately constant over the remainder of the engine speed range; that is 750 to 2500 RPM.

Figure 2 shows illustrative flow and pressure characteristics for a control system according to the invention employing a variable flow positive displacement pump. All of the specific values of lubricant flowrate, engine speed, and pump outlet pressure shown in Figure 2 are merely illustrative and as such the invention is not limited to such values or ranges of values. Thus, such values may vary considerably; for example, as a result of the configuration of oil lubrication channels within an engine; lubricant properties; and engine operating conditions such as lubricant temperature. Thus, when an engine is started from cold, the relatively high initial viscosity of the lubricant leads to higher pressure losses in the oil lubrication system than for the same engine speed etc., when the engine has reached normal operating temperature.

Over the engine speed range 0 to 750 RMP the pump throughput increases linearly (see portion A-B) with engine speed at a rate Pi (LPM/RPM). The gradient (Pi) of the portion A-B is a characteristic of the particular pump and depends upon pump dimensions. At engine

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speeds between 750 and 1500 RPM the pump delivery pressure remains substantially constant (see portion B-C). At engine speeds above 1500 RPM the pump delivery pressure again starts to increase; but at a slower rate (P2) than the initial linear increase (see portions C-D-E). The right hand ordinate axis also shows illustrative pump outlet pressures. In this example pump outlet pressure is proportional to the square of the pump throughput. Pump outlet pressure is also dependent upon the hydraulic resistance of the engine lubrication system flow channels and the temperature of oil circulating at any given time. Hence the pump outlet pressure corresponding to a particular engine speed will vary according with oil temperature.

Flow characteristics according to Figure 2 may be obtained by use of a split rotor gerotor pump with a suitable control system. Figure 3 shows the main features of such a pump 10. Thus pump 10 comprises a drive shaft 70, a first rotor portion 64, and a second rotor portion 66. In use, the angular orientation 13 of the first rotor portion 64 may be adjusted relative to the second rotor portion 66. This is achieved by rotationally mounting second rotor portion 66 on an eccentric shaft 76 such that the axis of rotation of the second rotor portion 66 may be changed, relative to that for the first rotor portion 64, by axial movement of eccentric shaft 76; achieved by axial movement of connected drive shaft 132; this driveshaft being moved by gears 106 and 108. Typically 13 may be varied over the range 0 to 180°. Thus, when (3=0° the first and second rotor portions are aligned and the pump output (at a given speed) is a maximum. Conversely, when 13=180° the first and second rotor portions are misaligned to a maximum extent and the pump output (at a given speed) is minimised. The workings of a split gerotor pump of the type shown in Figure 3 are explained in more detail in patent application GB 2313411 A. The variation in flow rate over the range of 13 depends upon the relative thickness of the first and second rotor portions. If the ratio of thickness of the first rotor portion to the second rotor portion is 75:25 then varying 13 from 0 to 180° will result in a 50% reduction in flow rate. Thus, the illustrative characteristics shown in Figure 2 could be obtained by use of such a split so that as engine speed increases from 750 to 1500 RPM 13 changes from 0 to 180°and so results in a flow of 50% of what would have occurred with, for example, a single (non-split) rotor positive displacement pump.

Several methods may be used to achieve the flow and pressure characteristics illustrated in Figure 2.

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Firstly, one or more of engine speed, pump outlet pressure or the pressure in part of the engine lubrication system such as the oil gallery may be measured and an electronic control system used to send a signal to the gerotor pump that results in an appropriate change (increase or decrease) in the value of B once engine speed reaches 750 RPM (in this illustrative case) or the lubricant delivery pressure reaches 4 bar (in this illustrative case). Typically the angle 6 would be varied by a step-motor driving adjustment shaft 132 in pump 10; movement of shaft 132 causing pivoting of eccentric shaft 76 of second rotor portion 66; and thus a change in the value of B. Thus, over portion B-C of the curve the angle B can be varied as required to keep the lubricant delivery pressure substantially constant. When the engine speed reaches 1500 RPM (see illustrative point C) the value of B is held constant. This results in the pump throughput and lubricant delivery pressure increasing again (along portion C-D of curve) as engine speed increases.

Secondly, a combined mechanical and hydraulic system may be used. Thus, the pump outlet or engine oil gallery pressure (used as the indication of lubricant delivery pressure) may when it reaches a pre-set value activate control means that results in the angle B being changed. In such systems it may be preferable to use the pump outlet pressure to provide the necessary drive force to energise the mechanical control means but to activate such movement in response to the pressure sensed at a particular point in the engine lubrication system such as the engine gallery pressure. For example; in general the pressure in the engine gallery is a more reliable indication of lubrication condition of the engine than the pump outlet pressure.

Figure 4 and 5 illustrate such a mechanical/hydraulic system.

Figure 4 shows a spool valve 100; cylinder 102; piston 104; rack 106 and pinion 108 in the position they take when the pump is operating over region A-B of the curve (see Figure 2). Spool 112 is held against rim 114 of the spool valve 100 by spring 116 which is under compression. Inlet duct 110 of the spool valve is always in fluid communication with the outlet of the pump and inlet duct 122 of the spool valve is always in fluid communication with the engine oil lubrication gallery, or other suitable sensing position. Spring 116 biases spool 112 towards rim 114; however when the oil gallery pressure communicated to inlet 122 is sufficient to overcome this bias spool 112 will move from the position shown in Figure 4 to that shown in Figure 5. This first occurs as pump speed increases beyond point B shown in Figure 2.

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As shown in Figure 4, inlet duct 110 is in fluid communication with face 120 of piston 104 via duct 118 and outlet duct 124 of the spool (which exhausts to an oil sump) is in fluid communication with face 126 of piston 104. This results in piston 104 and attached rack 106 being in an un-extended position as shown in Figure 4 and corresponds to a pinion 108 position at which 6 is zero or has a small value. Thus, when 13 is zero both gerotor rotors are aligned; the pump delivers maximum flow (at a given speed) and so hydraulically is equivalent to a standard (single rotor) gerotor pump.

When the oil gallery pressure is sufficient to move spool 112 to the position shown in Figure 5 then face 120 of piston 104 is no longer in fluid communication with the outlet of the pump but is vented to an oil sump via duct 118 and spool outlet 128. Further face 126 of piston 104 is no longer vented to an oil sump but is now in fluid communication with the oil pump outlet via inlet duct 110 and duct 130. The resulting pressure on face 126 of piston 104 causes piston 104 to start to move outwardly; eventually reaching the fully extended position shown in Figure 5. In this position pinion 108 has pivoted fixed axle 132 and so adjusted offset 13 angle to its maximum value. For example, the position shown in Figure 5 may correspond to 13=180°.

During control in region B-C, once piston 104 has moved the distance required to set 13 (via the rack and pinion mechanism) at an appropriate value then spool valve 100 will oscillate (or hunt) with small perturbations about a position intermediate of that shown in Figures 4 and 5.

The control system of Figures 4 and 5 used with a split rotor pump provides a control system according to the present invention. Portion A-B of Figure 2 is determined by the size of the pump rotors. Control of lubricant delivery pressure in region B-C is provided by providing a spring 116 under compression that holds spool 112 against recess 114 until the gallery pressure that corresponds to an engine speed of (say) 750 RPM is reached. Point C is determined jointly by the relative width of the rotors and the variation in 13 allowed by full movement of rack 106. Thus; for example the length of portion B-C of Figure 2 could be provided by using a 75:25 rotor split and allowing 13 to vary between 0 and 180°. Other combinations such as a 50:50 split and a range of 13 between 0 and 90° would achieve a like result.

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In use, following start-up of the engine being controlled, the control system will operate at an appropriate point on line A-B-C-D of a curve of the type shown in Figure 2, according to the particular engine speed, lubricant temperature etc., prevailing at that instant. The relationships between flowrate, pump outlet pressure and engine speed shown in Figure 2 are by way of example only and will therefore vary according to the pump selected, the engine in use and on factors that affect oil delivery pressure; such as the engine lubrication channel hydraulic characteristics and the lubricant temperature; the latter affecting in particular lubricant viscosity and thus pressure drop across specific flow elements of the lubrication system.

Portion D-E of Figure 2 cannot be obtained using the above described system alone. However, partial spill-back using a relief valve with appropriate settings may be employed to provide this effect.

The system and method of the invention provides a simple means of controlling lubricant delivery pressure according to the characteristics of the type shown in Figure 2 while avoiding the energy loss disadvantage of the known prior art.

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Claims

Claims

1. An engine lubricant pump control system comprising a variable output positive displacement pump, a sensor that provides an indication of engine speed, lubricant delivery pressure or both and control means that respond to the sensor and vary the output of the pump to:-

(a) increase lubricant delivery pressure at a first rate as engine speed increases; then

(b) maintain lubricant delivery pressure approximately constant over a range of engine speeds; then

(c) increase lubricant delivery pressure at a second rate as engine speed increases;

steps (a) to (c) taking space sequentially at predetermined values of lubricant delivery pressure as the engine speed is increased from minimum to maximum.

2. A control system according to Claim 1 wherein the variable output positive displacement pump comprises a gerotor type pump with a split rotor.

3. A control system according to any preceding claim wherein the control means comprise a spool valve in fluid communication with either the pump outlet and/or a part of the engine lubrication system.

4. A control system according to any preceding claim wherein the control means comprise a spool valve that is energised by oil pump outlet pressure and responds to the engine oil gallery pressure.

5. A control system according to any preceding claim wherein the control means comprises a cylinder, piston, rack and pinion that in combination respond to the oil pump outlet pressure or engine oil gallery pressure and wherein resultant movement of the pinion varies the offset angle (Ji) between the two portions of the split rotor.

6. A method of controlling a positive displacement variable output pump, used in an engine lubricant system, wherein engine speed, lubricant pressure or both are sensed/measured and control means that on the basis of this sensed/measured value: at low engine speeds ensures that lubricant delivery pressure increases with engine

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speed at a first rate; at medium engine speeds ensures that the lubricant delivery pressure is approximately constant; and at highest speeds ensures that the lubricant delivery pressure increases with engine speed at a second rate.

7. A method according to Claim 6 wherein the variable output positive displacement pump comprises a gerotor type pump with a split rotor.

8. A method according to Claim 6 or 7 wherein the control means comprise a spool valve in fluid communication with either the pump outlet and/or a part of the engine lubrication system.

9. A method according to Claim 6 or 7 wherein the control means comprise a spool valve that is energised by oil pump outlet pressure and responds to the engine oil gallery pressure.

10. A method according to any of Claims 6 to 9 wherein the control means comprises a cylinder, piston, rack and pinion that in combination respond to the oil pump outlet pressure or engine oil gallery pressure and wherein resultant movement of the pinion varies the offset angle (B) between the two portions of the split rotor.

11. A method according to Claim 7 wherein the control mean generates and sends an electrical output to the gerotor pump; said electrical output being used to adjust to the offset angle (B) between the two portions of the split rotor.

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