HYDRAULIC VANE PUMP

Abstract: This invention offers advantages and alternatives over the prior art by providing a dual port hydrualic fixed-displacement pump (20) which exhibits improved efficiency by diverting a portion of the flow of discharged fluid to the low-pressure side of a hydraulic system through electronic valve control. According to the present invention, a pair of discharge ports (12, 14) are provided, namely a first discharge port and a second discharge port. Under all operating conditions (e.g., low and high pump speed operating conditions) the fluid flowing within the first discharge port and primary discharge passageway is exposed to the working pressure of the primary line, which represents a high pressure line. The second discharge port fluidly communicates with a secondary discharge passageway which is in selective fluid communication with a low pressure line connected to a low pressure area of the pump (20) under first operating conditions and is also in selective communication with the first discharge port and the primary discharge passageway under second operating conditions.

Description: HYDRAULIC VANNE PUMP TECHNICAL FIELD The present invention relates generally to hydraulic pumps. BACKGROUND OF THE INVENTION Generally, a fluid powered system, e.g., steering system or transmission system, which is of a hydraulic design uses hydraulic pressure and flow to provide the required fluid power to the system. However, the hydraulic fluid must be pumped and regulated. The hydraulic pump creates the hydraulic force and typically a flow " control valve regulates the flow. A conventional vane-type pump comprises a cam (pump) ring having a substantially elliptical cam surface, a rotor which is adapted to rotate within the cam ring and a plurality of vanes adapted to move back and forth within radial slits formed in the rotor. The cam ring is stationary and the outer edges of the vanes touch the inside of the surface of the cam ring. Because of the substantially elliptical shape of the cam ring, the vanes slide in and out of their slots and maintain contact with the inside surface of the cam ring as the rotor turns therein. The volume of each pumping cavity constantly changes due to the elliptically shaped cam ring. Volume increases as the vanes move through the rising portion of the cam ring, drawing fluid through an intake port. When the vanes move into the "falling" portion of the ring contour, the volume decreases and forces the fluid out through the discharge ports. An intake portion of the hydraulic pump receives low-pressure hydraulic fluid from a pump reservoir. Discharged fluid, under high pressure, flows to a desired system location (e.g., a steering gear to provide power assist). In fixed-displacement pumps, at low engine speeds, the operating system can handle the volume of hydraulic fluid provided by the pump. Since the pump is usually directly driven by a fanbelt, line pressure dramatically increases at higher speeds because the pump draws and discharges a greater volume of fluid per unit time as it runs faster. These conditions raise operating temperatures and reduce pump durability and operating life. The system lines and seals are also strained. In addition, the torque necessary to drive the pump increases at higher system back pressures which corresponds to additional horsepower (energy) being required to effectively overcome the system back pressure and distribute the fluid throughout the system, thereby wasting fuel to generate unneeded line pressures. A common prior art solution for fixed-displacement pumps has been to provide flow control mechanisms in the high-pressure side of the system that "short-circuit" the flow from the high-pressure side to the low- pressure side of the system at excessive operating pressures. These mechanisms require addition of another component in the system and therefore increase the cost and complexity of the hydraulic system. Another pump conventionally used is a variable-displacement pump. A variable-displacement pump provides a reduction in flow as a function of operating conditions and therefore requires more costly shaft support solutions. Additionally, since variable-displacement pumps are typically single stroke, the pumps require a larger package size to provide the same pumping capacity. Variable-displacement pump valving also make these pumps less efficient in the full displacement operating condition. There is a perceived need for a fixed-displacement hydraulic pump, preferably a vane-type pump, for use in a vehicle operating system, wherein the pump has improved energy efficiency while at the same time providing adequate hydraulic power and having a mechanism for reducing high pressures that is substantially integrated into the pump itself with minimal need for installation of additional components. SUMMARY OF THE INVENTION This invention offers advantages and alternatives over the prior art by providing a dual port hydraulic fixed-displacement pump that exhibits improved efficiency by limiting the volume of discharged fluid which is subjected to the line pressure of a hydraulic system through electronic valve control and wherein the actual valving mechanism is completely contained within the pump itself. According to the present invention, a pair of discharge ports are provided, namely a first discharge port and a second discharge port. The first discharge port fluidly communicates with a primary discharge passageway and discharge outlet which is connected to a primary line for distributing the fluid throughout the system. Under all operating conditions, e.g., low and high pump speed operating conditions, the fluid flowing within the first discharge port and primary discharge passageway is exposed to the working pressure of the primary line, which represents a high pressure line. The second discharge port fluidly communicates with a secondary discharge passageway which is in selective fluid communication with a low pressure line connected to a low pressure area of the pump (e.g., a reservoir) under first operating conditions and is also in selective communication with the first discharge port and the primary discharge passageway under second operating conditions. The first operating conditions comprise high speed operating conditions (e.g., pump speeds above 2500 rpm) where pump output exceeds system fluid demands and the second operating conditions comprise low speed operating conditions where system demands require full pump capacity. A flow control valve is disposed within the pump and acts to direct the fluid flowing within the secondary discharge passageway according to either a second discharge path, wherein the fluid is directed to the low pressure line and the low pressure reservoir or sump of the system, or a third discharge path, wherein the fluid is directed to the primary discharge passage and is subjected to the high pressure line of the system. In an exemplary embodiment, the flow control valve comprises an electrically controlled valve which is designed to actuate when the fluid flowing within the secondary discharge passageway reaches a predetermined flow rate. Upon actuation, all of the fluid flowing through the secondary discharge passageway is directed to the low pressure line instead of the high pressure line of the primary discharge passageway. The result is that the high pressure side of the system is partially or completely "short-circuited" to the low pressure line, thereby reducing the pressure differential of the system. As a result of reducing the pressure differential of the system, the torque to drive the pump is significantly reduced and thus a considerable reduction in horsepower is achieved because all of the fluid is not exposed to the high back pressure of the primary line. In practice, the flow control valve is actuated under high pump speed operating conditions (e.g., above 2500 rpm) where the pump output significantly exceeds system demands. Under low pump speed operating conditions when full pump capacity is needed, the flow control valve is not actuated and all of the fluid from the secondary discharge passageway is directed to the high pressure side. The above-described and other features and advantages of the present invention will be appreciated and understood by those skilled in the art from the following detailed description, drawings, and appended claims. BRIEF DESCRIPTION OF THE DRAWINGS Figure 1 is a schematic diagram of an exemplary hydraulic system of the present invention, in this case a hydraulic steering system. Figure 2 is a schematic diagram of a typical hydraulic steering system of the prior art, here shown with a dual-output pump. Figure 3 is a cross a cross sectional view of a conventional single-output pump illustrating the design of the discharge ports of the pump; Figure 4 is a cross sectional elevational view of an exemplary vane-type pump in accordance with the present invention; Figure 5 is a cross sectional view taken along the line 5-5 of Figure 4; Figure 6 is a front plan and a side perspective view of a exemplary electronic valve. Figure 7 is a sectional side view of an exemplary flow control valve of Figure 6 mounted within a pump . DETAILED DESCRIPTION OF THE INVENTION Referring to Figure 1 there is shown a hydraulic steering system wherein a vane-type pump 20 having a discharge outlet 86 in fluid communication with a high-pressure side, or line, 2 of the system. The high- pressure line 2 feeds into a power steering unit 3 or other load. The diagram depicts a power steering system, but it is to be understood that the teachings of this invention apply to any system that requires hydraulic power and, in that respect, the power steering unit 3 may be substituted with any system load. Fluid exits the load 3 into the low-pressure side 4 of the system, which is in fluid communication with a reservoir 5. The low-pressure lines 4 feed into one or more input inlets 6 of the pump 20. The vane pump 20 operates in a manner typical of vane-type pumps. An elliptical cam 44 is provided within which a rotor 46 is rotatably mounted. The rotor 46 has a plurality of slots in which are slideably mounted a plurality of radial vanes 9, the spacing between each set of vanes defining a fluid chamber 60. The vanes may be spring-biased in an outwardly radial direction or may simply be forced in that direction by centrifugal force. Regardless, the vanes 9 are compelled to follow the contour of the cam 44, thereby causing the chambers 60 between the vanes to expand and contract. Charging ports are positioned to allow fluid from the low-pressure line 4 to be drawn into the expanding chambers. Discharge ports are positioned to receive fluid forced from contracting chambers for transfer to the high-pressure lines 2 via the discharge outlet 86. To maintain continuity and symmetry, it is preferred that there be two charging ports and two discharge ports, each positioned 180 degrees from each other around the central axis of the rotor 46. Note that in the embodiment shown, there is only one discharge outlet 86, thereby indicating that the discharge ports from the cam are fluidly connected within the pump, as will be described in more detail with regard to Figures 3 and 5 below. Also, in this drawing it is apparent that the charging ports are probably separate and distinct because the low-pressure line bifurcates and enters the pump an two separate inlets. This is a matter of design choice. The pump could just as easily be provided with a single input inlet in internal fluid communication with both charging ports. The system of the invention shown in Figure 1 may be compared to a typical system of the prior art as shown in Figure 2. Here, a prior art flow-control mechanism 15 is shown inserted in the high-pressure side 2 of the system. This external unit is configured to "short-circuit" the high-pressure side 2 to the low-pressure side 4 of the system via a bypass line 140 at high operating pressures such as are found when the pump 10 is run at high RPM, thereby preventing excessive pressure differential across the pump and the load (i.e., between the high and low-pressure sides). Because the bypass system of the present invention is integrated with the pump 20, the bypass line 140 is connected directly to the pump 20 in Figure 1 as contrasted with Figure 2. Figure 3 is a cross sectional view of a typical conventional vane-type pump such as that depicted in Figure 2. The vane-type pump is generally indicated at 10 and comprises a pump having dual internal discharge ports that join at a common discharge outlet 16. As is known in the art, the vanes within the rotor and the cam ring (not shown) define pumping chambers. More specifically, the space between the rotor, ring and any two adjacent vanes defines a single pumping chamber. The rotor is driven by a drive shaft 11. The rotation of the rotor and movement of the vanes causes the volume of each pumping cavity to constantly change due to the shape of the cam ring which is typically oval-shaped (elliptical). As the vanes move through the "rising" portion of the cam ring, the volume of each pumping cavity increases resulting in the fluid being drawn through a charging port of the pump. Conversely when the vanes move into the "falling" portion of the cam ring contour, the volume of each pumping cavity decreases. Decreased volume within the pumping cavity causes an increase in pressure within each pumping cavity resulting in the fluid being forced out of the pumping cavity and through the discharge ports 12, 14 of the pump. The illustrated vane-type pump 10 shown in Figure 3 includes a first discharge port 12 and a second discharge port 14. In this design, first and second discharge ports 12 and 14 are routed to a common discharge outlet 16. The fluid flow path from the first and second discharge ports 12 and 14 is generally indicated by directional arrows 18. In this example, pump 10 is required to force the fluid through the common discharge outlet 16 and the fluid works against system back pressure. Because of the back pressure which is observed in the system, in order for pump 10 to effectively distribute the fluid through the overall system, pump 10 must force the fluid at such a flow rate that the fluid overcomes the back pressure of the system and is therefore effectively distributed throughout the system. In this type of pump design, the fluid passing through both the first and second discharge ports 12 and 14 must work against the system back pressure. The operating system in which pump 10 is being used requires a certain fluid flow rate so that a sufficient amount of fluid is pumped throughout the system for proper operation thereof and in this design, pump 10 distributes all the fluid throughout the system. As is known, the energy consumption of the pump is linked to the amount of torque required to drive the unit and as the torque increases, an increase in horsepower is likewise observed and energy consumption rises. With this type of pump design, at higher operating conditions, the pump output in terms of forcing fluid through the system at a certain flow rate exceeds the system demands. Such a pump is termed a fixed-displacement pump because as the speed of the pump increases, the flow rate correspondingly increases. Consequently, at high pump speeds, the flow rate is unnecessarily high and the flow rate of the fluid exceeds the demands of the system. Pump 10 is therefore operating at less than efficient conditions because all of the discharged fluid is exposed to the working line pressure of the system. It should also be noted that the useful work done by the pump will be related to the pressure differential between the high and low-pressure sides of the system. In automotive uses, a high RPM often indicates that the vehicle is travelling at high speed. At high speed, however, less hydraulic work is required for certain systems, such as power steering systems. The extra pressure differential between the high and low-pressure lines at high RPM is therefore unneeded and the strain on the system and energy expended to overcome the increased pump torque is wasted. Referring to Figures 1, 4-7. According to the present invention, a dual port hydraulic fixed-displacement pump is made more efficient by limiting the volume of the discharged fluid, e.g. oil, that is subjected to the line pressure of the hydraulic system. That is to say that a portion of the discharge flow from the high-pressure side is "shorted" to the low-pressure side so as to reduce the pressure differential across the pump. More specifically, the present invention may be incorporated into a number of types of pumping assemblies, including piston pumps, vane-type pumps and gear pumps; however, for the purpose of illustration only, the present invention is described with reference to an exemplary dual port hydraulic fixed-displacement vane-type pump. It being understood that one of skill would appreciate that the improved efficiency dual discharge port design of the present invention may be incorporated into these other pump assemblies besides the illustrated vane-type pump. The exemplary vane-type pump is generally indicated at 20 in Figures 1, 4, and 5. As previously discussed, the term "fixed-displacement pump" refe s to a pump in which an increase in the speed of the pump leads to a corresponding increase in the flow rate of the discharged fluid. Referring to Figure 4, vane-type pump 20 includes a pump housing 22 having an internal housing cavity 24 with a large opening 26 at one end thereof and a smaller opening 28 at the other end thereof. A drive shaft 30 extends through the smaller opening 28 and is rotatably supported in a shaft bearing 51 which is secured in the opening 28 and is contacted by a shaft seal 32 also secured within the opening 28. Adequate shaft support is placed in the assembly to deal with bending loads which result from the unbalanced condition when pump 20 is operating in a fuel efficient mode. The shaft seal 32 functions to prevent atmospheric air from entering the pump 20 and fluid leakage from pump 20. Within the housing cavity 24 is a vane pump assembly, generally designated at 40, and includes a pressure plate 42, a cam ring 44, a rotor 46, a plurality of vanes (not shown), and an end cover 49 and thrust plate 50. The end cover 49 cooperates with annular seal ring 52 and a locking ring 54 to close the large opening 26. The rotor 46 includes a plurality of slots in which the plurality of vanes are slidably disposed as is known in the art. The plurality of vanes contact the inner surface of cam ring 44 so as to provide a plurality of peripheral pumping chambers 60 which expand and contract upon the rotation of rotor 46 when it is driven by a drive shaft 30. The thrust plate 50 includes discharge porting arrangements as will be described in greater detail hereinafter to effectively direct the forced fluid from vane assembly 40 to discharge passageways and outlets of the pump 20 which act to distribute the fluid to the other components of the system. The discharged fluid from the pumping chambers 60 of the vane assembly 40 passes through the thrust plate 50 to first and second discharge ports 80 and 82, respectively, which in turn are in fluid communication with a pump discharge passage (not shown in Figure 4) formed in pump 20. Referring now to Figure 5 in which a cross-sectional view of the exemplary pump 20 from Figure 4 is shown. Figure 5 illustrates the dual fluid discharge port design of the pump 20. First discharge port 80 fluidly communicates with a discharge outlet 86 which serves to route the discharged fluid to the high-pressure side 2 of the system. As in the conventional pump 10 shown in Figure 3, the first discharge port 80 is part of a primary discharge passageway 90 for the fluid to flow in response to the pumping action. In Figure 5, a primary discharge path in which the fluid flows from first discharge port 80 is illustrated by directional arrows 92. Because the first discharge port 80 is directly connected to the pump discharge outlet 86, this primary discharge passageway 90 is exposed to working line pressure of the system under all operating conditions of the pump. In other words, at either low speed or high speed operating conditions, pump 20 must work against the high pressure side of the system in order to effectively distribute fluid according to the primary discharge path 92 as the fluid is distributed throughout the system. According to the present invention, second discharge port 82 partially defines a second discharge path for the fluid to flow in response to the action of pump 20. In the exemplary and illustrated embodiment, second discharge port 82 fluidly communicates with a secondary discharge passageway 110 so that fluid flowing through second discharge port 82 is directed to secondary discharge passageway 110 and on to an intake port 125 of an electric valve. The electric valve has two output ports that it may switch the flow of fluid to. The first switched port is in fluid communication with the low-pressure side of the system and the second switched port is in fluid communication with first discharge port 80 and permits the discharged fluid from secondary discharge passageway 110 to join the fluid flowing from first discharge port 80 and out through discharge outlet 86 under selective operating conditions, as will be described in greater detail hereinafter. Secondary discharge passageway 110 includes a flow control valve 120. Flow control valve 120 is mounted to direct the fluid flowing from discharge port 82 into the valve common port 125 of the valve 120. The valve will then either direct the flow out through a first valve switched port into bypass line 140 or through a second valve switched port into a high- pressure connecting conduit 114 according to a flow path indicated by directional arrow 130 and on to the discharge outlet 86 to join with flow path 92. In other words, flow control valve 120 dictates whether the fluid flowing from the secondary discharge port goes to the high-pressure side 2 of the system or is diverted to the low-pressure side 4 via the bypass line 140. Note that because valve 120 controls only one discharge port, the maximum fluid flow that can be diverted is 50% of the total, assuming equality of the two discharge ports 80, 82. Referring to Figures 6 and 7, flow control valve 120 may comprise any number of suitable valves which are designed to actuate upon application of an electronic signal, which will preferably be designed to coincide with the occurrence of a predetermined event, such as when the fluid flowing through the system exceeds a predetermined rate, or when the pressure differential across the pump exceeds a predetermined level, or when the pump or car engine exceeds a predetermined RPM, or other suitable indicia as well as any suitable combinations thereof. Detection of such conditions may be effected by sensors and circuitry as are known in the art. Referring to Figure 6, a preferred control valve 120 will have an actuator 121 and an input terminal 122 for receiving an electric signal for controlling the actuator 121. There is also provided at least one common port 125 and two switched ports, a first switched port 123 and a second switched port 124. When the actuator is in a first condition, the first switched port 123 will be in fluid communication with the common port 125. When the actuator is in a second condition, the second switched port 124 is in fluid communication with the common port 125. In a preferred embodiment, the actuator 121 is a solenoid-type actuator having only a first and second condition such that fluid flow is either entirely directed through one switched port or the other. This is the simplest embodiment and perhaps the most reliable and inexpensive. In another embodiment, the actuator is an electric motor type actuator such that the flow of fluid is split amongst the two switched ports as the actuator moves between the first and second conditions. Hence the total fluid flow through the common port may be divvied up between the two switched ports in any proportion, so long as the proportions flowing through both switched ports add to 100% of the total. This embodiment gives the greatest control over the system pressures and performance, but is also more expensive and prone to breakdown. Referring to Figure 7, flow control valve 120 comprises a electronically controlled valve which is designed to actuate when the fluid flowing within the secondary discharge passageway 110 reaches a predetermined flow rate. Prior to that event, the actuator 121 is in the first condition so that fluid coming from the second discharge port 82 (not shown in this Figure) enters the valve common port 125 and passes through the second valve switched port 124 as depicted by flow arrow 130. In this state, the pump behaves like any prior art dual port pump like that shown in Figure 3. Upon actuation of flow control valve 120 toward the second condition, the fluid flowing through the valve is diverted to the bypass line 140 which fluidly communicates with the low-pressure side of the system. A cable conduit 126 is provided to carry control wires to the valve's input terminal 122. As a result of diverting, or "shorting" fluid flow to the low- pressure side of the system, the pressure differential across the pump (i.e., between the low and high-pressure sides of the system) is reduced. The degree of reduction is dependent upon the proportion of fluid diverted, which is entirely under the control of the designer. As a result of the pressure drop, the torque required to drive pump 20 is significantly reduced, a considerable reduction in required horsepower is achieved, and improved efficiency and lower operating and repair costs are achieved. Also, fuel economy savings to a vehicle are realized along with a reduction in operating temperatures and noise. Because control of the valve is electronic, the designer of the system has great flexibility to determine what first set of operating conditions will cause the valve to be in the first condition and what second set of operating conditions will trigger the valve to switch over toward the second condition. The switch may occur in response to such operating conditions as excessive pressure differential between the high and low-pressure sides, excessive pump rate of rotation (RPM), or excessive rate of fluid flow, among others. The designer may opt for a continuous valve so that the change from the first to the second condition occurs smoothly and continuously. The designer then has very fine control over the system and may tailor very specific responses to changes in system operating conditions. It will be understood that a person skilled in the art may make modifications to the preferred embodiment shown herein within the scope and intent of the claims. While the present invention has been described as carried out in a specific embodiment thereof, it is not intended to be limited thereby but is intended to cover the invention broadly within the scope and spirit of the claims.

Claims: CLAIMS What is claimed is: 1. A hydraulic fluid pump (20) for use with a fluid powered system comprising: a first pump discharge outlet (86) for delivering fluid to a high-pressure side of the system from the pump (20); a pump inlet (6) for accepting fluid from a low-pressure side of the system; a pumping mechanism to drive fluid received from said pump inlet (6) through one or more discharge ports (12,14); an internal conduit in fluid communication with said one or more discharge ports (12,14) and said discharge outlet (86); a flow control valve (120) disposed within said pump (20), wherein actuation of said flow control valve (120) diverts a portion of fluid flow from said conduit to said low-pressure side of the system. 2. The apparatus of claim 1, wherein said flow control valve (120) comprises: an electronically controlled actuator (121); a common port (125); a first switched port (123); a second switched port (124); wherein said actuator (121) is controlled to switch from a first condition to a second condition wherein: said first condition is characterized by said common port (125) being in fluid communication with said first switched port (123); said second condition is characterized by said common port (125) being in fluid communication with said second port; and wherein: said common port (125) is in fluid communication with one or more of said discharge ports (12,14); said first switched port (123) is in fluid communication with said discharge outlet (86); and said second switched port (124) is in fluid communication with said low-pressure side of the system. 3. The apparatus of claim 1, wherein said flow control valve (120) is actuated to divert fluid to said low-pressure side when said pump (20) exceeds a predetermined rate of rotation. 4. The apparatus of claim 3 wherein said predetermined level is one where said rate of rotation exceeds system demand. 5. The apparatus of claim 1, wherein said flow control valve (120) is actuated to divert fluid to said low-pressure side when a measured pressure differential between said high-pressure side and low- pressure side exceeds a predetermined level. 6. The apparatus of claim 5 wherein said predetermined level is one where said pressure differential exceeds system demand. 7. The apparatus of claim 1 wherein said flow control valve (120) is actuated to divert fluid to said low-pressure side when a measured rate of fluid flow exceeds a predetermined level. 8. The apparatus of claim 7 wherein said predetermined level is one where said rate of fluid flow exceeds system demand. 9. The apparatus of claim 2 wherein said actuator (121) is a solenoid. 10. The apparatus of claim 2 wherein said flow control valve (120) may exist in only said first condition or said second condition. 11. The apparatus of claim 2 wherein said flow control valve (120) may smoothly and continuously change states between said first condition and said second condition. 12. The apparatus of claim 1, wherein the fluid powered system comprises a vehicle operating system selected from the group consisting of a power steering system, a transmission assembly, and a hydraulic engine cooling system.

Inventors: ADEN DAVID R, DAVISON JAMES LEROY, KELBEY RYAN G, RYTLEWSKI THOMAS C, ADEN, DAVID, R., DAVISON, JAMES, LEROY, KELBEY, RYAN, G., RYTLEWSKI, THOMAS, C.
Applicants: DELPHI TECH INC, DELPHI TECHNOLOGIES, INC.
Application year: 2001
Main classification: F04C15
IPC codes: B62D5, B62D6, F04C2, F04C11, F04C14, F04C15