Patent Application
20070224067
The invention relates to a variable displacement sliding vane pump having a slide whose position is controlled by a differential pressure between a constant pressure source and a variable pressure source, the differential pressure equilibrating a spring force applied to the slide to establish a desired flow rate and pressure.
A lubrication system for an engine pressurizes and distributes lubrication fluid to the engine lubrication circuits. It employs a rotor and a slide with multiple vanes and cavities which can vary the volume of fluid delivered to the oil circuits.
The slide is eccentrically offset from the rotor to create fluid chambers defined by the vanes, rotor, and inner surface of the slide. A compression spring positions the slide to create large fluid chambers as the default. When the engine requires less volume of fluid or less oil pressure by the pump, a pressure regulator directs fluid from the pump output line to a regulating chamber in the pump. Pressure in the regulating chamber pivots the slide against the force of the spring to more closely align the centers of the rotor and slide, thereby reducing the size of the fluid chambers. This reduces the amount of fluid drawn into the pump from the fluid reservoir and likewise, the amount of fluid output by the pump and thereby reduces the oil pressure as well.
There are two ways to control pump output. The first way is to direct line pressure to the regulating chamber via the pressure regulator to decrease pump output. The second way is to remove pressure from the regulating chamber via the pressure regulator by exhausting fluid to increase pump output.
Representative of the art is U.S. Pat. No. 4,342,545 (1982) to Schuster discloses a variable displacement vane type pump having a pivotally mounted ring member controllable to vary the eccentricity between the rotor and the ring thus controlling the pump displacement. The ring is positioned on the pivot such that the center thereof is always located in one quadrant relative to axes through the pivot point and the center of the pump rotor to continually maintain the net ring reaction force, due to internal pressure, directed to one side of the pivot connection in opposition to the displacement control pressure, which is impressed on a portion of the outer surface of the ring, whereby control stability throughout the displacement range is improved.
What is needed is a variable displacement sliding vane pump having a slide whose position is controlled by a differential pressure between a constant pressure source and a variable pressure source, the differential pressure equilibrating a spring force applied to the slide to establish a desired flow rate and pressure. The present invention meets this need.
The primary aspect of the invention is to provide a variable displacement sliding vane pump having a slide whose position is controlled by a differential pressure between a constant pressure source and a variable pressure source, the differential pressure equilibrating a spring force applied to the slide to establish a desired flow rate and pressure.
Other aspects of the invention will be pointed out or made obvious by the following description of the invention and the accompanying drawings.
The invention comprises a variable displacement sliding vane pump comprising a pump body, inlet and outlet ports formed in said pump body, a drive shaft rotatably mounted in said pump body, a rotor driven by said drive shaft and co-axially aligned therewith, a plurality of radially extending vanes slidably disposed in said rotor, a pivot disposed in said pump body, a slide pivotally disposed on said pivot in said pump body and having a central axis eccentric to the axis of said rotor, a plurality of fluid chambers defined by said rotor, said vanes, and said slide that are successively connected to said inlet and outlet ports, a spring acting on said slide to urge said slide in one direction, a first chamber and a second chamber, each suitable for receiving a fluid pressure and each disposed between said pump body and an outer surface of said slide, the first chamber in fluid communication with a pump outlet discharge pressure, and a valve operable to selectively pressurize and depressurize the second chamber.
The accompanying drawings, which are incorporated in and form a part of the specification, illustrate preferred embodiments of the present invention, and together with a description, serve to explain the principles of the invention.
Pump shaft 16 is rotatably mounted in body 10. A splined end 160 of pump shaft 16 engages rotor 13. As rotor 13 rotates vanes 14 are urged outwardly by a pair of vane control rings 17 and centripetal force into a sliding engagement with inner surface 120 of slide 12.
Slide 12 is pivotally engaged with the body at a pivot member 18. Slide 12 pivots at pivot member 18 within cavity 11 thereby describing an arc which defines the operating range of motion of slide 12.
The position of each vane 14 is a function of the position of slide 12 with respect to ring 17. Ring 17 occupies a space determined by the ends of vanes 14. Ring 17 is substantially concentric with inner surface 120.
The position of ring 17 with respect to rotor 13 determines the radial position of each vane 14 in each slot 15, which in turn, determines a given slide 12 position as compared to the position of the axis of rotation of rotor 13. This relationship determines the volume of each of the chambers 21 between the inlet port 19 and the outlet port 20 for a given engine speed and hence a given slide 12 position.
Body 12 defines a pair of kidney shaped ports 19, 20 which comprise an inlet port and an outlet port, respectively, for the pump 100. A plurality of chambers 21 are formed by the vanes 14, rotor 13 and inner surface 120. Chambers 21 rotate with rotor 13 and expand and contract during rotation, as is well-known in vane type pumps.
Inlet port 19 accepts fluid from a source or reservoir such as an engine oil system, not shown, and passes the fluid to the chambers 21 in turn as rotor 13 rotates. Vanes 14 move the fluid in chambers 21 from the inlet port 19 to the outlet port 20. As can be seen in
The position of slide 12 is established by the combined effect of the control pressure in each for two chambers, namely, chamber 22 and chamber 23 acting in balance with the spring force from spring 31. Chamber 22 extends about a portion of the outer circumference of slide 12 from seal member 24 disposed in a groove 26 to seal member 25 disposed in a groove 27, each formed in slide 12. Each seal member 24 and 25 is urged outwardly against surface 28 by a resilient backing member 29, 30 respectively. Chamber 23 extends about a portion of the outer circumference of slide 12 from seal member 24 to pivot member 18.
Spring 31 acts in opposition to the sum of the fluid pressures in chambers 22 and 23 such that as the total pressure in chambers 22 and 23 increases, and therefore the torque of the slide around the pivot member increases, the pump slide 12 will move clockwise about pivot member 18. The combined torque caused by the pressure in chamber 22, 23 is balanced by the spring force of spring 31.
The fluid pressure in chamber 22 is supplied by fluid in ultimate communication with the outlet port 20 of pump 100 and is therefore subject to the outlet pressure of pump 100 or from a feedback channel to the engine gallery, see
Under most operating conditions, the axis of slide 12, and therefore of inner surface 120, moves between position 32 during low engine speed conditions to position 33 during high engine speed conditions. As vanes 14 are rotated from the inlet port 19 to outlet port 20 a pressure transition takes place with the chambers 21.
Since the inner surface 120 is subjected to the internal pressure generation in chambers 21, slide 12 is inherently unbalanced during operation. The net resultant reaction force due to the internal pressure generation passes through the central axis of inner surface 120. It will be appreciated that the reaction forces always provides a counterclockwise moment about axis 18 which is in opposition to the clockwise moment generated by the control pressure in chambers 22 and 23.
The pressures in chamber 22, 23 are balanced against the force of spring 31 so that the displacement of the pump, and as a result the flow, may be adjusted by varying the chamber pressures. The inventive pump controls both displacement and oil flow for two or more outlet pressure levels based upon the pump outlet pressure or the engine gallery pressure.
Typically the desirable pressure level in the pump for each chamber is the pressure level required to produce the proper oil flow and pressure for all engine speeds and load conditions. In some cases, at lower rpm's the engine does not require a high oil pressure level, therefore a somewhat lower pressure is acceptable, and therefore the flow is reduced as well. The lower operating pressure and flow is achieved by pressurizing chamber 23.
The required magnitude of the lower oil pressure depends upon different engine parameters, including whether it is a gas or diesel engine, the engine complexity, engine speed and load.
The inventive pump provides two levels of control. The first is pressure control over a given speed range due to the variable vane pump function. The second is based upon the ability of the pump to change between two (or more) pressure levels by use of two (or more) pressure chambers 22, 23, controlling the position of slide 12.
A cover 70 is secured to the housing 10 by a plurality of fasteners 37. Leakage from the chambers 21 radially outwardly past the cover 70 is prevented by surface to surface contact.
Connected to the main oil gallery 204 is conduit 205 which connects to chamber 22 through conduit 34, see
The total force exerted against spring 31 by slide 12 is the sum of the torques created by the fluid pressure in chamber 22 plus the fluid pressure in chamber 23, both acting about pivot member 18.
At or less than a first operating speed, valve 207 is OPEN thereby allowing the engine gallery pressure to enter chamber 23. The pressure in chamber 23 and combined with the pressure in chamber 22 causes slide 12 to pivot about pivot member 18 an arcuate distance to a position where the torque caused by the combined pressures in chambers 22, 23 is balanced by the spring force of spring 31. The pump characteristics with slide 12 in this position are shown by portion “A” of
At an operating condition greater than the first operating speed, valve 207 is closed thereby venting chamber 23 to ambient pressure (approximately 1 bar). The pressure in chamber 22 causes slide 12 to pivot about pivot member 18 an arcuate distance to an equilibrium position where the torque caused by the pressure in chamber 22 is balanced by the spring force of spring 31. Slide 12 pivots because as the pump speed increases, the pressure in chamber 22 also increases, thereby increasing the force exerted against spring 31. The pump characteristics with slide 12 in this position are shown by portion B of
In an alternate embodiment valve 207 may be opened to a partial position thereby causing slide 12 to move to a position that is intermediate position A and position B, causing an intermediate outlet pressure and flow. Placing valve 207 in any position between fully open and fully closed allows the pressure in chamber 23 to be variable, thereby providing a range of slide positions for a given pump outlet pressure.
In the case of a failure of valve 207, the pump will continue to operate in a passive mode (chamber 23 depressurized) while meeting all oil requirements of the engine. The passive operating mode is still more efficient that a fixed displacement pump. With valve 207 in operation the instant invention provides incremental horsepower reduction over the passive design.
The engine speed range is from 0 RPM to 8000 RPM. The outlet pressure range is from 0 bar to 6.00 bar. The pump flow rate range is from 0 liters/minute to 90.00 liters/minute.
For the purposes of illustration an engine speed of ˜3,500 RPM is selected to demonstrate the characteristics of the inventive pump. The transition between operating conditions “A” and “B” is depicted as the “switching point” in the center of the curves in the graph.
For engine speeds less than ˜3,500 RPM the maximum pump outlet pressure is approximately 2.6 bar. The maximum flow rate is approximately 20.0 liters/minute.
For engine speeds greater than ˜3,500 RPM the pump outlet pressure quickly transitions up to a minimum outlet pressure of approximately 4.9 bar at 7,500 RPM. The flow rate transitions to a maximum of approximately 28.0 liters/minute at 7,500 RPM.
At the transition point the step change in pressure is approximately 1.6 bar. The step change in flow is approximately 5 l/min.
The performance transition is caused by slide 12 pivoting about pivot 18 caused by deactivation of valve 207 venting chamber 23 to ambient atmospheric conditions. Valve 207 is controlled by an electric signal transmitted by an engine ECU, for example. Upon reaching the predetermined engine speed, in this case ˜3,5.00 RPM, ECU 208 (see
As described previously, the pressures in chambers 22, 23 create a torque and hence force which is greater than the combination of the force of the spring 31 and the fluid force in chambers 21, thereby causing spring 31 to compress. This causes slide 12 to pivot. By pivoting in the clockwise direction the flow rate and outlet pressure are each substantially decreased at the predetermined engine speed because pump displacement is reduced.
For the purposes of comparison, the dashed lines in portion A of
The engine speed range is from 0 RPM to 8000 RPM. The outlet pressure range is from 0 bar to 6.00 bar. The pump flow rate range is from 0 liters/minute to 90 liters/minute.
For the purposes of illustration an engine speed of ˜2,000 RPM is selected to demonstrate the characteristics of the inventive pump. The transition between operating conditions “A” and “B” is depicted as the “switching point” at approximately 2,000 RPM.
In this example, valve 207 is OFF at start up and for engine speeds less than 2,000 RPM, namely, chamber 23 is unpressurized and vented to ambient. For engine speeds less than approximately 2,000 RPM the maximum pump outlet pressure (Line Pressure) is approximately 3.6 bar. The maximum flow rate (Flow Rate) is approximately 25.0 liters/minute.
For engine speeds greater than approximately 2,000 RPM the pump outlet pressure (Line Pressure) quickly transitions down to a minimum outlet pressure of approximately 2.4 bar at 2,000 RPM up to 3.2 bar at approximately 7,500 RPM. The flow rate (Flow Rate) transitions to a maximum of approximately 23.0 liters/minute at 7,500 RPM.
At the transition point the step change in pressure is approximately 1.4 bar. The step change in flow is approximately 5 l/min.
The performance transition in this example is caused by slide 12 pivoting about pivot 18 caused by activation of valve 207 thereby pressuring chamber 23. Valve 207 is controlled by an electric signal transmitted by an engine ECU, for example. Upon reaching the predetermined engine speed, in this case approximately 2,000 RPM, ECU 208 (see
Although a form of the invention has been described herein, it will be obvious to those skilled in the art that variations may be made in the construction and relation of parts without departing from the spirit and scope of the invention described herein.
