DIRECT ACTUATED VALVE CONTROL HYDRAULIC PUMP AND MOTOR

Abstract

A control device for controlling a displacement pump or motor, comprising a variable mechanical device, comprising at least two masks, wherein the at least two masks are phased relative to each other and a shaft, and are configured to create a desired cam profile. A cam follower is operatively connected to at least one of an intake valve an output valve of a displacement pump or motor.

Description

TECHNICAL FIELD

The present disclosure generally relates to pumps and motors, and in particular to use of a variable mechanical device to directly actuate valves of a hydraulic pump/motor unit to achieve flow diverting and flow limiting displacement control.

BACKGROUND

This section introduces aspects that may help facilitate a better understanding of the disclosure. Accordingly, these statements are to be read in this light and are not to be understood as admissions about what is or is not prior art.

A large problem facing hydraulic fluid pumps is low overall efficiency especially at low displacements. Digital displacement pumping techniques have been developed to mitigate such inefficiencies. Current digital pump/motors use solenoid poppet valves or a solenoid valve latching technique. The most common units (swashplate axial units) use valve plates. Current digital units also require significant and expensive controllers for the solenoids used in the other techniques. Therefore, improvements are needed in the field.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 shows a diagram of a flow diverting system according to one embodiment.

FIG. 2A shows a state diagram for valve cycling in a first state.

FIG. 2B shows a state diagram for valve cycling in a second state.

FIG. 3 shows a diagram of a pump being mechanically actuated using a cam follower.

FIG. 4A shows a diagram of a variable cam having two masks in a first rotation state.

FIG. 4B shows a diagram of a variable cam having two masks in a second rotation state.

FIG. 5A shows a first portion of a simulation model of a half masking cam pump according to one embodiment.

FIG. 5B shows a second portion of a simulation model of a half masking cam pump according to one embodiment.

FIG. 5C shows a third portion of a simulation model of a half masking cam pump according to one embodiment.

FIG. 6 is a graph showing transition length and compression angle of the cam profile.

FIG. 7A is a graph showing an efficiency contour plot at 20% displacement in a first example.

FIG. 7B is a graph showing an efficiency contour plot at 40% displacement in a first example.

FIG. 7C is a graph showing an efficiency contour plot at 60% displacement in a first example.

FIG. 7D is a graph showing an efficiency contour plot at 80% displacement in a first example.

FIG. 7E is a graph showing an efficiency contour plot at 100% displacement in a first example.

FIG. 8A is a graph showing an efficiency contour plot at 20% displacement in a second example.

FIG. 8B is a graph showing an efficiency contour plot at 40% displacement in a second example.

FIG. 8C is a graph showing an efficiency contour plot at 60% displacement in a second example.

FIG. 8D is a graph showing an efficiency contour plot at 80% displacement in a second example.

FIG. 8E is a graph showing an efficiency contour plot at 100% displacement in a second example.

FIG. 9 is a graph showing pressure angles for an example cam according to one embodiment.

FIG. 10 is a graph showing operating conditions on the distribution of losses at 20% displacement.

FIG. 11 shows an example on/off cam actuator and pump arrangement according to one embodiment.

FIG. 12 shows an example on/off check pumping cam actuator and pump arrangement according to one embodiment.

FIG. 13 shows an example cam FC bi-directional cam actuator and pump arrangement according to one embodiment.

FIG. 14 shows an example fixed cam FD bi-directional actuator and pump arrangement according to one embodiment.

FIG. 15 shows an example directional on/off cam actuator and pump arrangement according to one embodiment.

FIG. 16 shows an example bi-directional cam actuator and pump arrangement according to one embodiment.

DETAILED DESCRIPTION

For the purposes of promoting an understanding of the principles of the present disclosure, reference will now be made to the embodiments illustrated in the drawings, and specific language will be used to describe the same. It will nevertheless be understood that no limitation of the scope of this disclosure is thereby intended.

In response to the unmet need, the design disclosed herein addresses such issues by using state of the art operating strategies and directly actuating the valves using a variable, mechanical device which reduces lubricating gaps and consequently leakage and therefore increases efficiency compared with conventional devices. Additionally, in many cases the hydraulic fluid is not needed for lubrication at all which allows for many different fluids to be used within the unit. This technique is applicable to many different types of devices including but not limited to radial piston, inline piston, and wobble plate piston units. While any variable mechanical actuation could be used, the current design utilizes two half masks which are phased relative to one another and the shaft in order to create the desired cam profile. This is then followed by the valve poppet to open and close the valve allowing for significant and consistent actuation stroke with minimal control input and power losses.

Efficiency is the main attraction of the herein disclosed methods and devices. Traditional unit efficiency has good efficiency at high displacement which drops off quickly as displacement is reduced. Using the flow diverting and flow limiting variable displacement strategies combined with this technology, efficiency will more closely scale with displacement and will lead to greater efficiency especially in displacement controlled systems. Also, as one embodiment of a function of the flow limiting/diverting strategies, the piston chamber is only pressurized when doing work further reducing losses. Additionally, actuating the valves mechanically compared to electronically can supply larger flow areas (less metering losses) and greater consistency in actuation. The spring return on the follower also reclaims some energy for actuation. Compared to standard pump/motors (namely swashplate axial units) leakage losses are greatly reduced as hydraulic fluid is not necessarily used for lubrication of moving parts. Port plates removed in favor of positively sealing valves. This also allows the herein described methods and devices to be used with non-standard hydraulic fluids such as but not limited to water. Reduces the amount of controls needed. Additionally, the herein disclosed methods and devices can be implemented on most fixed displacement, piston type units with minimal expense as long as the piston chambers are stationary.

According to one embodiment, a Partial Flow-Diverting strategy is utilized to provide variable displacement. As the name implies, with this technique excess flow is diverted back to the low pressure port rather than pumping it to the high pressure side; varying the amount of flow which is diverted back allows us to achieve variable displacement. The flow diverting strategy is pictorially illustrated in FIG. 1. As the piston 100 is moving down starting at top dead center (TDC), valve 01 opens allowing the fluid to flow (top right of FIG. 1) and completely fill the displacement chamber (bottom right of FIG. 1). Valve 101 would still be held open as the piston moves upwards (bottom left of FIG. 1) towards the top dead center (TDC); holding valve 101 open would divert the flow back to the low pressure port. The displacement of the unit is defined by the piston position at which valve 101 closes during the upward motion of the piston. When the right amount of fluid is available in the displacement chamber, valve 101 closes and the fluid is pressurized; then valve 102 opens (upper left of FIG. 1) and the high pressure fluid is pumped into the high pressure line.

According to one embodiment, the repetitive nature of the valve actuation is exploited to allow for mechanical actuation. A state analysis was performed on the valves and several configurations from a simple pump to full four-quadrant pump/motors were deemed viable. FIGS. 2A and 2B show examples of the state diagrams for the valve cycling. Each diagram represents the state the valve must be in, in this example open or closed, at the corresponding rotation angle.

FIG. 3 shows one example embodiment of a pump 300 with a low pressure intake valve 302 being mechanically actuated using a cam follower 304 and a high pressure output valve implemented as a check valve 306. In the illustrated configuration, the speed of the valve actuation will be proportional to the speed of the shaft of the pump 300 and the high actuation forces of the cam allow for valves with larger orifice areas to be used. In order to achieve variable displacement of the pump 300 the cam profile needs to be varied. To accomplish this, two 50% high cam profiles, or masks, are provided. This allows the effective cam profile to be varied from 100% to 50% high by phasing one mask relative to the other. FIG. 4 illustrates one example of a variable cam having two halves or masks 402 and 404, with each mask having a 50% high profile. The two halves 402 and 404 can be rotated or phased with respect to one another. In FIG. 4A, the two halves are shown in phase with each other, thereby providing a 50% overall high profile. In FIG. 4B, they are shown phased to provide an overall high profile of approximately 85%. The cam follower 304 is implemented as a roller follower to reduce the friction and size of the cams. The roller follower is connected to the valve 302 to achieve opening or closing of the valve as the cam rotates.

In order to determine the most efficient configuration for the half masking cam, a simulation of the illustrated pump was created in Matlab/Simscape. The kinematics and pumping chambers used in this model are based on those created for simulation of an electronically controlled digital pump motor. A single piston version of the mechanically actuated Simulink model is shown in FIGS. 5A, 5B, and 5C. A single piston is shown for clarity; the simulation was run with three pistons.

The model of FIGS. 5A, 5B and 5C allows for complete manipulation of the cam profiles and allows these to be tested under any operating parameter within the scope of pump operation. The main variables to be manipulated are the transition length, compression angle, transition type, pressure difference, speed, and displacement.

Transition type is the path that is followed to increase or decrease the cam profile from the low state to the high state. While many different transition types may be used, harmonic is used in the simulation as it was determined to have the lowest peak acceleration and jerk. Other transition types, such as cycloidal and parabolic may also be used.

Transition length is the amount of degrees the transition type takes to transition from the low to high state. An ideal valve would have a transition length of zero resulting in no addition metering as the valve is opening; thus efficiency is inversely proportional to the amount of transition degrees. Transition length affects the acceleration of the cam follower as well as the pressure angle of the cam. A rule of thumb for industry is the pressure angle should not exceed 30° in high speed applications to prevent seizing. The pressure angle can be calculated using equation (1) below where e is eccentricity, x is the cam displacement profile, and r_b is the base radius of the cam.

$\begin{array}{cc}{\alpha }_p={\tan }^{-1}\left(\frac{x^{\prime }\left(\theta \right)-e}{x\left(\theta \right)+\sqrt{r_b^2-e^2}}\right)& \left(1\right)\end{array}$

Compression angle is the amount of degrees of rotation the piston cylinder needed to decompress the fluid as the piston chamber transitions from expulsion to the high pressure line to intake from the low pressure line at TDC. FIG. 6 shows transition length (θ_t) and compression angle (θ_c) on the cam profile.

The correct compression angle is necessary to reclaim the energy input into the fluid and maximize efficiency. For an ideal valve the compression angle would be calculated using the following equations. While this is still useful for estimation, the transition type and length will affect the ideal compression angle on a real machine. Equation (22) shows the relationship between the change in volume and the change in pressure.

$\begin{array}{cc}\mathrm{dP}=K\frac{\mathrm{dV}}{V}& \left(2\right)\end{array}$

where V is the effective volume of the piston chamber which is equivalent to the dead volume and the piston area, A_p, multiplied by the piston stroke ((3). It should be noted that at TDC, x=1 removing the second part of the equation. The change in volume, dV, is equivalent to the change in piston stroke multiplied by the piston area (4).

V=V _dead +A _p(l−x)  (3)

dV=Δx A _p  (4)

Combining these three equations to determine the compression angle for TDC, (5 is derived.

$\begin{array}{cc}\Delta x=\frac{\mathrm{dP}V_{\mathrm{dead}}}{KA_p}& \left(5\right)\end{array}$

Spring constant was calculated using (6 where m_f is the mass of the follower and x_pc is the spring pre-compression distance.

$\begin{array}{cc}k=\max \left(-\frac{x^{″}\left(\theta \right)*m_f}{x\left(\theta \right)+x_{\mathrm{pc}}}\right)& \left(6\right)\end{array}$

In one example, the following system parameters simulated.

• • Valve max opening area=100 mm²
  • Poppet stroke=4 mm
  • Piston cylinder radial gap=16 μm
  • Piston Area=314 mm²
  • Piston Stroke=30 mm
  • Transition Length=6-12°
  • Compression Angle=12-20°
  • Speed=500, 1000, 1500, and 2000 rpm
  • Pressure differential=69 and 138 bar
  • Displacement=20%, 40%, 60%, 80%, 100%
  • Number of Pistons=3.

The results of the simulation demonstrate an increase in overall efficiency. When the compression angle was optimized for the operating range of 69-138 bar, an average efficiency of 89.72% was reached at an 18° compression angle. The same method used for optimization for this operating range could be used with any desired parameters. FIGS. 7A-7F and 8A-8F show efficiency contour plots for varying displacements at a speed of 2000 rpm and differing pressures. By comparing these graphs, the following trends with respect to the mechanically actuated pump may be observed:

• • Efficiency decreases as the transition length increases.
  • Increase in pressure increases efficiency
  • Increase in pressure increases the ideal compression angle.
  • There is a distinct efficiency maximum at the ideal compression angle.

Ideally, the transition length would minimized to maximize efficiency, however, this is practically not feasible. In order to stay below the 30° pressure angle requirement while keeping the cam size relatively small, the following values were selected in one example:

• • Cam base diameter=140 mm
  • Roller diameter=26 mm
  • Eccentricity=20 mm
  • Spring pre-compression=2 mm
  • Follower mass=0.1 kg
  • Transition length=10 degrees

FIG. 9 shows the pressure angles for the cam selected for this simulation. Notice that the maximum value is 28.47 degrees for the rising section of the cam, while falling the pressure angle does reach 43.82 degrees. Eccentricity was used to shift the pressure angle in this way because during the falling section, there is less force acting in a direction that will prevent the cam from seizing.

Using the above parameters, FIG. 10 shows the effect of different operating conditions on the distribution of losses when operating at 20% displacement. Notice that viscous friction losses are the losses in the piston chamber due to friction and are relatively constant for this pump regardless of operating conditions. For these operating conditions, valve losses remain lower than the frictional losses of the pump though speed is a significant factor.

Referring to FIGS. 11, 12, 13, 14, 15, and 16, further example embodiments of mechanical cam actuators for use with the pumps described above are shown. It shall be appreciated that these configurations are only example configurations and are not intended to be limiting on the types of configurations claimed using the herein disclosed methods and devices.

Those skilled in the art will recognize that numerous modifications can be made to the specific implementations described above. The implementations should not be limited to the particular limitations described. Other implementations may be possible.

Claims

1. A method for achieving flow diverting and flow limiting variable displacement control of a displacement pump or motor, comprising utilizing a variable mechanical device to provide a control signal for an intake and output valve operatively connected to said displacement pump or motor.

2. The method of claim 1, wherein the variable mechanical device comprises a rotating cam connected to a rotating shaft and a cam follower operatively connected to at least one of said intake and output valve.

3. The method of claim 1, wherein the cam comprises at least two masks, wherein the two masks are phased relative to each other and a shaft, and are configured to create a desired cam profile for controlling at least one of said intake and output valve.

4. A control device for controlling a displacement pump or motor, comprising: a variable mechanical device, comprising at least two rotating cam masks,wherein the at least two masks are phased relative to each other and a shaft, and are configured to create a desired cam profile; anda cam follower operatively connected to at least one of an intake valve and an output valve of a displacement pump or motor.

5. The variable mechanical device of claim 3, wherein the at least two masks are half masks.