Patent Application
20130167957
Not Applicable
Not Applicable
1. Field of the Invention
The present invention relates to hydraulic pumps, such as those that have pistons that move radially against an eccentric shaft, and more particularly to mechanisms for controlling the flow of fluid through the cylinders in which the pistons move.
2. Description of the Related Art
A common type of radial piston pump comprises a body with a plurality of cylinders radially disposed around a drive shaft that is rotated by an external motor or engine. A separate piston is slideably received within each cylinder, thereby defining a chamber at the interior of the cylinder. The drive shaft has an eccentric cam and the pistons are biased by springs to ride against that cam. As the cam rotates, the pistons slide reciprocally within the respective cylinders, thereby reducing and expanding the volume of the cylinder chambers in a cyclical manner. The smallest volume occurs at the top dead center point of the piston cycle and the largest volume occurs at the bottom dead center point.
An inlet port supplies fluid to an inlet passage that has a separate inlet into each cylinder. Every cylinder also has an outlet that is coupled by a separate outlet check valve to an outlet passage that leads to the outlet port of the pump. U.S. Pat. No. 3,434,428 discloses a pump of this configuration. The pump in that patent also has a throttle plate with apertures associated with the inlets for the cylinders. The throttle plate is rotated by an actuator to vary alignment of the apertures with the inlets and thereby alter the amount of fluid flowing between the common inlet passage and each cylinder inlet.
With this type of pump, as the piston moves from the top dead center point, fluid is not initially drawn into the expanding cylinder chamber because the location of the piston blocks the inlet. The piston has to move a considerable distance from the top dead center point before the inlet is unblocked and fluid from the inlet passage is drawn into the expanding cylinder chamber. After the bottom dead center point, the volume of the cylinder chamber begins reducing, however, the inlet still is open which prevents outlet check valve from opening. Here too, the piston must move some distance before the piston blocks the inlet and causes pressure in the cylinder chamber to increase. As the piston starts to pump, the sealing land of the piston is low in the cylinder and high pressure fluid leakage occurs thereby making this form of aspiration initially in-efficient. Eventually the pressure rises to a level that forces the outlet valve to open an outlet path through which the fluid is exhausted from the cylinder chamber. That exhausting continues until the piston again reaches the top dead center point.
A drawback of this type of pump is that during a dead portion of the piston cycle, between bottom dead center point and when the inlet becomes closed, no pumping action occurs. Specifically, fluid is neither being expelled from the cylinder nor being drawn into the cylinder during that dead portion, which can be a third of the piston cycle as shown in FIG. 6 of the U.S. Pat. No. 3,434,428. This inactive time and initial short sealing length results in a sizeable inefficiency. In addition this type of pump requires a relatively long piston stroke to accommodate the dead portion of the piston cycle, which increases the diameter of the pump.
These prior radial piston pumps also had a relatively large diameter due to the outlet valves and the outlet passage being located radially outward from each cylinder. For many machines, the amount of space for the pump is limited, thus it is desirable to reduce the size of the pump. More specifically, many times the pump is mounted alongside an engine or transmission and the radial space is limited preventing the installation of typical radial piston pumps.
A pump includes a cylinder block with an inlet port, an outlet port, a plurality of cylinders disposed radially in the cylinder block. A plurality of inlet passages are each connected between the inlet port and a different one of the plurality of cylinders, and a plurality of outlet passages each connected between the outlet port and a different one of the plurality of cylinders. A separate piston is slideably located in a each of the plurality of cylinders and drive shaft is rotatably received in the cylinder block for driving the piston reciprocally the cylinders.
A separate inlet check valve is located in each of the plurality of inlet passages and allows fluid flow only in a direction from the inlet port into one of the plurality of cylinders. A separate outlet check valve is located in each of the plurality of outlet passages and allow fluid flow only in a direction from one of the plurality of cylinders into the outlet port.
A throttle plate communicates with each of the plurality of inlet passages for varying a rate of fluid flow through the inlet passages. In one embodiment, the throttle plate extends across each of the plurality of inlet passages and has a plurality of control apertures there through. The throttle plate is moveable to alter alignment of the control apertures with the inlet passages and thereby vary a cross sectional area through which fluid flows in the inlet passages. This provides a variable orifice in each inlet passage.
One aspect of the present pump is that the flow area of the variable orifice is directly related to the magnitude of the fluid flow there through. Broadly speaking, as the throttle plate is moved from a position corresponding with the variable orifice being fully open to a position with the variable orifice is fully closed, the average rate of change of the flow area of the variable orifice relative to movement of the throttle plate is greater during a first half of the travel distance between the fully open and fully closed positions. For example, the flow area of the variable orifice decreases at least 80 percent in the first half of the throttle plate travel distance from the fully open position. This rapid closure rate of the variable orifice occurs in what is referred to as the first section of the throttle plate rotation. Thereafter, the rate of change of the flow area decreases significantly slower, requiring that the throttle plate move through the second half of travel distance to reduce the flow area the remaining 20 percent.
The term “directly connected” as used herein means that the associated components are connected together by a conduit without any intervening element, such as a valve, an orifice or other device, which restricts or controls the flow of fluid beyond the inherent restriction of any conduit. References herein to directional relationships and movement, such as top and bottom or left and right, refer to the relationship and movement of the components in the orientation illustrated in the drawings, which may not be the orientation of the components as attached to machinery.
With reference to
With particular reference to
The tubular sleeve 39 that partially forms the cylinder 36 enables the inlet and outlet check valves 33 and 34 to be placed closer to the longitudinal axis 25 of the drive shaft 40. Note that the inlet and outlet check valves 33 and 34 are within the closed curved perimeter defined by the exterior side surface 38 of the cylinder block 30. In prior configurations the valves had to be outward from the top dead center position of the piston in order to receive the fluid forced out of the cylinder chamber 37. As shown in
Referring again to both to
A separate piston assembly 51 is slideably received within each of the cylinders 36. Every piston assembly 51 comprises a piston 52 and a piston rod 54. The piston rod 54 extends between the piston 52 and the cam bearing 46. The piston rod 54 has a curved shoe 56 which abuts the outer race 48 of the cam bearing 46. The shoe 56 is wider than the shaft of the piston rod creating a flange portion. A pair of annular retaining rings 58 extend around the cam 44 engaging the flange portion of each piston rod shoe 56, thereby holding the piston rods 54 against the cam bearing 46, which is particularly beneficial during the intake stroke portion of a pumping cycle. The retaining rings 58 eliminate the need for a spring to bias the piston assembly 51 against the cam bearing 46. The curved shoe 56 evenly distributes the piston load over a wide area of the cam bearing 46. As the drive shaft 40 and cam 44 rotate within the cylinder block 30, the outer race 48 of the cam bearing 46 remains relatively stationary. The outer race 48 rotates at a very slow rate in comparison to the speed of the drive shaft and the inner race 47. Therefore, there is little relative motion between each piston shoe 56 and the cam bearing's outer race 48.
The piston 52 is cup-shaped having an interior cavity 53 which opens toward the drive shaft 40. An end of the piston rod 54 is received within that interior cavity 53 and has a partially spherical head 60 that fits into a mating partially spherical depression 62 in the piston 52. The head of the piston 52 may have an aperture 50 there through to convey hydraulic fluid from the cylinder chamber 37 to lubricate the interface between a spherical head 60 and the piston 52. The piston rod 54 is held against the piston 52 by an open single bushing or a split bushing 55 and a snap ring 57 that rests in an interior groove in the piston's interior cavity 53. As the piston rod 54 follows the eccentric motion of the cam 44 and the piston 52 in turn follows by sliding within the cylinder 36. The bushing and snap ring arrangement allows the spherical head 60 of the piston rod to pivot with respect to the piston 52 when a rotational moment is imposed onto the piston rod 54 by rotation of the cam 44. Because of that pivoting, the rotational moment is not transferred into the piston 52, thereby minimizing the lateral force between the piston and the wall of the cylinder 36.
With continuing reference to
Rotation of the eccentric cam 44 causes each piston 52 to move cyclically within the respective cylinder 36, away from the sealing cup 24 during a fluid intake phase and then toward the sealing cup 24 during a fluid exhaust phase. Because of the radial arrangement of the cylinders 36, at any point in time some pistons 52 are in the intake phase while other pistons are in the exhaust phase.
The piston 52 illustrated in
Thereafter, further rotation of the eccentric cam 44 moves the piston 52 into the exhaust phase during which the piston moves outward, away from the center axis 25. That motion initially compresses the fluid in the cylinder chamber 37, thereby increasing the pressure of that fluid. Soon the pressure in the cylinder chamber 37 is approximately that same as the pressure in the inlet passage 26, at which point the associated spring closes the inlet first check valve 33. Eventually, the cylinder chamber pressure exceeds the pressure in the outlet gallery 32 and forces the outlet check valve 34 open, releasing the fluid from the cylinder chamber 37 into the outlet gallery and to the outlet port 29.
When continued rotation of the eccentric cam 44 moves the piston 52 to the top dead center position shown in
Because the inlet and outlet check valves 33 and 34 open and close almost immediately at the top dead center and bottom dead center positions, essentially the entire piston cycle is use to draw fluid into the cylinder chamber and then expel that fluid. This is in contrast to prior pumps that had throttle plates, but relied on the position of the piston to open and close an inlet opening into the cylinder. Those prior pumps had a dead region, which is some cases was one third the piston cycle, during which fluid was neither being drawn into or expelled from the cylinder chamber. Thus with the present pump configuration an equivalent fluid volume can be pumped by each piston cycle with less piston stroke distance. This feature contributes to the compact size of the present pump.
With reference to
The hydraulic pump 10 further includes an actuator 100 for rotating the throttle plate 90 within the cylinder block 30. For that purpose, a tab 98 projects outward from the outer edge of the throttle plate 90 and into an actuator bore 102 in the cylinder block 30. The actuator bore 102 has a control port 104 to which a hydraulic conduit from a control circuit connects. An actuator piston 108 is slideably received in the actuator bore 102 and engages the tab 98 of the throttle plate 90. Pressurized fluid applied to the control port 104 drives the piston to the right in the actuator bore 102 (see
The angular position of the throttle plate 90 within the cylinder block 30 determines the alignment of the control apertures 95 in the throttle plate with the transmission apertures 94 in the transition plate 91. Varying that alignment alters the degree to which those apertures overlap and thus alters the cross sectional area through which fluid is able to flow between the inlet gallery 31 and the cylinders 36 during the piston cycle intake phase. In other words, the adjustable alignment of the transmission and control apertures 94 and 95 forms a variable orifice in that flow path provided by the inlet passages 26. Both the control apertures 95 and the transmission apertures 94 have unique shapes so that fluid flow varies in a specific manner to regulate the displacement of the pump 10 and maintain the output pressure at a desired level.
The variation in the rate of orifice area change is determined by the unique shape of the transverse cross section of the control apertures 95 in the throttle plate 90. Transverse cross section as used herein means a cross section across a control aperture in a plane that is transverse to the direction that fluid flows through the aperture. As shown in
From the fully aligned position in
Upon reaching the intermediate position in
The amount of this flow can be proportionally controlled by controlling the rotational position of the throttle plate 90 and thus the amount of that aperture overlap. As the rotation of the throttle plate 90 continues the tapered aperture regions 97 cause the flow area to change at a smaller rate than occurred during previous motion to reach that intermediate position from the fully aligned position of the transmission and control apertures 94 and 95. Now for each given incremental distance that the actuator piston 108 moves and for each given incremental angle change of the throttle plate, an relatively smaller change in flow area occurs than happened previously. Therefore, the rate that the open area of the control apertures 95 changes decreases as that open area becomes smaller.
Continued activation of the control actuator 100, results in the throttle plate 90 eventually reaching the position illustrated in
Thereafter the rate of change of the flow area decreases significantly slower, requiring that the actuator piston 108 move the remaining 50 percent of the travel distance (the second section 126 of the throttle plate rotation) to reduce the flow area the final 20 percent to the fully closed position. Thus during the second section of throttle plate rotation, the piston and the throttle plate decreases the flow from 20 percent of the maximum flow to zero flow over the same amount (i.e., 50 percent) of throttle plate rotation as it takes to decrease the flow from 100 percent to 20 percent of the maximum flow. In other words, at a constant rate of rotation of the throttle plate 90, the flow area of the variable orifice changes from the maximum flow area to about 20 percent of that maximum flow area at a rate that is at least twice as fast and the rate at which the flow area changes between about 20 percent of that maximum flow area and zero flow area. Therefore from the fully aligned aperture position, rotation of the throttle plate initially produces a relatively rapid decrease in flow area and then the flow area decrease occurs at a slower rate the as aperture motion approached the closed position. The inverse rates of change occur as the throttle plate 90 moves clockwise in the drawings and the variable orifice, formed by the degree of alignment of the control apertures 95 with the transmission apertures 94, opens greater amounts.
The use of a throttle plate 90 to control the amount of flow between the inlet gallery 31 and the inlet passages 26 enables the displacement of the pump 10 to be dynamically varied. When the throttle plate apertures 95 are only partially aligned with the transition plate transmission apertures 94, the amount of fluid flowing into the cylinder chamber 37 during the intake phase of each piston cycle is reduced. As a result, the piston 52 reaches bottom dead center without the cylinder chamber 37 being completely filled with hydraulic fluid. Thus, a portion of the total effective piston displacement is lost. The amount of lost displacement does not vary significantly as a function of the speed of the pump, since the average pressure drop across the throttle plate is constant for typical pump speeds of 800 to 2500 RPM.
The present pump configuration with the rotatable throttle plate 90 provides variable throttle choking at the input of each inlet check valve. This has a significant advantage over a pump that has throttle choking at a single place for all the cylinders, such as between the inlet port 28 and the inlet gallery 31. With the per inlet check valve choking arrangement of the present pump 10, the fluid volume between the throttle plate and the inlet check valve is relatively small and results in improved consistency and dynamic response in both starting and stopping fluid flow.
The foregoing description was primarily directed to a preferred embodiment of the invention. Although some attention was given to various alternatives within the scope of the invention, it is anticipated that one skilled in the art will likely realize additional alternatives that are now apparent from disclosure of embodiments of the invention. Accordingly, the scope of the invention should be determined from the following claims and not limited by the above disclosure.
