Not Applicable
Not Applicable
1. Field of the Invention
The present invention relates to hydraulically powered equipment, such as off-road construction and agricultural vehicles, and more particularly to apparatus for reducing oscillation or wag when a hydraulically driven member on the equipment is decelerating, stopping, or reversing direction.
2. Description of the Related Art
With reference to
As the boom swings in one direction, pressurized fluid is introduced into one chamber of the boom swing cylinder 9, designated as the “driving chamber”, and fluid is exhausted from the other cylinder chamber, referred to as the “exhausting chamber”. When the boom swings in the opposite direction, the designation of the driving and exhausting chambers is reversed. When the operator suddenly stops the boom swing, inertia causes the motion of the backhoe boom assembly 3 to continue in the previously commanded direction. The amount of inertia is a function of the mass and extension position of the boom assembly 3 and the mass of any material carried in the bucket 4. This continued movement due to inertia compresses the hydraulic fluid in the previous exhausting chamber of the boom swing cylinder 9 and may produce cavitation in the previous driving cylinder chamber. Anti-cavitation valves typically are provided in the hydraulic system to overcome this latter problem.
Because the control valves for the cylinder are now closed, pressure in the previous exhausting chamber eventually increases to a magnitude that causes the boom assembly 3 motion to stop and recoil by moving in the opposite swing direction. This subsequent movement produces a reversal of the pressure conditions, wherein the previous driving chamber of the boom swing cylinder 9 becomes pressurized. When the boom motion in the opposite swing direction creates a sufficiently high pressure in the previous driving chamber, another reversal of the swing motion occurs. As a result, the boom assembly swing oscillates until inherent dampening provided by other forces ultimately brings the assembly to a stop. This phenomenon is known as “bounce” or “wag” and increases the time required to properly position the boom 6, thereby adversely affecting equipment productivity. The wag also is disconcerting to the machine operator. A similar motion phenomenon occurs when other types of hydraulically driven members stop.
In essence, the wag is a manifestation of the inability of the boom velocity to promptly respond to, or track, changes in the position of the valve that controls the flow of fluid to the swing cylinder. In other words, the valve closes when motion of the boom is to terminate, however the load force acting on the boom does not allow the velocity of the boom to decrease fast enough.
Various approaches have been utilized to minimize this wag. For example, U.S. Pat. No. 4,757,685 employs a separate relief valve for each hydraulic conduit connected to the swing cylinder chambers to vent fluid to a tank line when excessive pressure occurs in the associated chamber. Additional fluid is supplied from the supply line through makeup valves to counteract cavitation in the cylinder as the swing stops.
U.S. Pat. No. 5,025,626 describes a cushioned swing circuit which also has relief and make-up valves connected to the hydraulic lines for the boom swing cylinder. This circuit also incorporates a cushion valve which in an open position provides a fluid path between the cylinder hydraulic lines. That path includes a flow restriction orifice. The cushion valve is resiliently biased into the shut position by a spring and a mechanism opens the cushion valve for a predetermined time period when the pressure differential between the cylinder chambers exceeds a given threshold. Both of these previous solution attempts required additional valves and other components.
U.S. Pat. No. 6,705,079 describes another solution to the swing wag problem in which a sensor detects pressure in the hydraulic actuator. This pressure signal from the sensor is employed to determine the rate at which the pressure in the hydraulic actuator changes. When the rate of change of the pressure is less than a defined threshold after receiving a stop command, pressure in the hydraulic actuator is relieved, such as by opening a control valve connected to the hydraulic actuator.
However, there still is a desire to improve the responsiveness of the boom velocity to changes in the position of the valve and the resultant flow of fluid to the associated hydraulic cylinder, and in particular to provide a simplified mechanism for reducing wag.
The present method controls deceleration of a hydraulically driven machine member. Motion of a hydraulic actuator connected to the machine member is designated by a command, which may specify a desired velocity for example. A parameter value is produced that indicates the ability of the machine member motion to respond to change in fluid flow applied to the hydraulic actuator, which change results from alteration of the position of a valve controlling that flow. That ability is represented by the magnitude of a load force that is exerted on the hydraulic actuator by the machine member, and in particular is denoted by fluid pressure from the hydraulic actuator. The parameter value is used to configure a variable filter function applied to the command to produce a filtered command which is employed to control flow of fluid to the hydraulic actuator. The filter function controls the rate at which the motion command goes to zero to stop the machine member so that the command does not close the related valve faster than a rate the actuator and machine member are able to operate.
In one aspect of the control method, the amount of load force acting on the hydraulic actuator is employed to derive a filter frequency that defines the rate the motion command decreases to zero. The filter frequency varies inversely with changes in the load force. However, the filter frequency is preferably set to a predefined constant value when the magnitude of the load force is less than a first threshold. The filtering frequency may also be set to another predefined constant value when the magnitude of the load force is greater than a second threshold.
Another aspect is to utilize a digital filter, in which case configuring the filter involves determining a set of filter coefficients. In the preferred embodiment, the filter coefficients are derived in response to the selected filtering frequency.
Although the present invention is being described in the context of use on a backhoe as shown in
With initial reference to
The supply conduit 14 and the tank return conduit 18 are connected to a plurality of hydraulic functions 19 and 20 on the backhoe. Separate hydraulic functions are provided for swinging the boom 6, raising the boom, moving the arm 5 and pivoting the bucket 4. The hydraulic function 20 for swinging the boom is illustrated in detail and other functions 19 have similar components and operation. The hydraulic system 10 is a distributed type in that the valves for each function and control circuitry for operating those valves are located adjacent to the associated hydraulic actuator. For example, those components for controlling boom swing are located at or near the swing cylinder 9 or the pivot joint 8.
In the boom swing function 20, the supply conduit 14 is connected to node “s” of a valve assembly 25, which also has a node “t” that is connected to the tank return conduit 18. The valve assembly 25 includes a workport node “a” connected by a first hydraulic conduit 30 to the head chamber 26 of the boom swing cylinder 9, and has another workport node “b” coupled by a second conduit 32 to the rod chamber 27 of boom swing cylinder 9. Four electrohydraulic proportional (EHP) valves 21, 22, 23, and 24 control the flow of hydraulic fluid between the nodes of the valve assembly 25 and thus control fluid flow to and from the boom swing cylinder 9. The first EHP valve 21 is connected between nodes “s” and “a”, and controls fluid flow between the supply conduit 14 and the head chamber 26 of the boom swing cylinder 9. The second EHP valve 22, is connected between nodes “s” and “b” and controls flow of fluid between the supply conduit 14 and the cylinder rod chamber 27. The third EHP valve 23 is connected between node “a” and node “t” and controls EHP flow between the head chamber 26 and the return conduit 18. The fourth EHP valve 24, between nodes “b” and “t”, controls fluid flow between the rod chamber 27 and the return conduit 18.
The hydraulic components for the boom swing function 20 also include two pressure sensors 36 and 38 which detect the pressures Pa and Pb within the head and rod chambers 26 and 27, respectively, of boom swing cylinder 9. Another pressure sensor 40 measures the pump supply pressure Ps at node “s”, while pressure sensor 42 detects the return conduit pressure Pr at node “t”. Pressure sensors 40 and 42 may not be present on all the hydraulic functions.
The pressure sensors 36, 38, 40 and 42 for the boom swing function 20 provide input signals to a function controller 44 which produces signals that operate the four electrohydraulic proportional valves 21-24. The function controller 44 is a microcomputer based circuit which receives other input signals from a computerized system controller 46, as will be described. A software program executed by the function controller 44 responds to those input signals by producing output signals that selectively open the four electrohydraulic proportional valves 21-24 by desired amounts to properly operate the boom swing cylinder 9.
The system controller 46 supervises the overall operation of the hydraulic system by receiving operator input signals from joysticks 47 and exchanging signals with the function controllers 44 and a pressure controller 48. The signals are exchanged among those controllers over a communication network 55 using a conventional message protocol. This enables the control functions for the hydraulic system 10 to be distributed among the different controllers 44, 46 and 48.
With reference to
The velocity commands for the swing cylinder 9 and the other hydraulic actuators 11 are sent to a setpoint routine 62 that determines the desired pressures for the supply and return conduits 14 and 18. Specifically, the setpoint routine 62 ascertains a supply pressure required by each hydraulic function 19 and 20 and selects the greatest of those pressures as the supply conduit pressure setpoint Ps. The setpoint routine 62 also determines a return conduit pressure setpoint Pr in a similar manner. These pressure setpoints Ps and Pr are applied as inputs to the pressure controller 48 that also receives signals from a supply conduit pressure sensor 49 at the outlet of the pump, a return conduit pressure sensor 51, and a tank pressure sensor 53. The pressure controller 48 responds to those inputs by operating the unloader valve 17 to regulate supply conduit pressure and the tank control valve 16 to control the return conduit pressure to achieve the desires setpoint pressures.
The velocity command for the swing cylinder 9 also is sent from the mapping routine 50 to the associated function controller 44 where it is applied to a valve opening program 56 comprises software that determines how to operate the EHP valves 21-24 in assembly 25 to achieve the commanded velocity of the piston rod 43. The swing direction designated by the velocity command denotes which two of the valves EHP valves 21-24 are activated and an amount that those valves are to open to convey fluid to and from the swing cylinder 9. Specifically valves 21 and 24 are opened to extend the piston rod 43 from the swing cylinder, and valves 22 and 22 are opened to retract the piston rod.
The magnitude of the velocity command and the measured pressures (Pa, Pb, Pr, Ps) are utilized by the valve opening routine to determine the amount that each of the selected valves is to be opened to convey the amount of fluid flow necessary achieve the desired velocity of the piston 28. U.S. Pat. No. 6,775,974 describes one embodiment of the valve opening program 56. The resultant signals, indicating the amount that the EHP valves 21-24 are to open, are supplied to a set of valve drivers 58 which apply the appropriate magnitude of electric current to operate each of the two selected valves.
The valve opening program 56 includes a software routine that mitigates wag of the boom assembly 3 that otherwise could occur when swing cylinder is desired to stop. With reference to
When the backhoe operator desires to stop the boom swing, the joystick 47 is released and allowed to return to its center, neutral position. In this position, the mapping routine 50 produces a zero velocity command which is transmitted to the function controller 44 for the swing operation. If the function controller 44 simply responded to the zero velocity command by immediately shutting the valves, a swing wag could occur, especially if the boom assembly 3 had a relatively large inertia. That function controller 44, however, is programmed to reduce swing wag by low pass filtering the velocity command and thereby control the rate at which the EHP valves close in response to the velocity command. A dynamically varying filter function is utilized so that the swing decelerates in a controlled fashion under both relatively small and very large loads. Preferably a digital second order filter function is used.
In order that the filtering performs satisfactorily over a wide range of load force, the filter is disabled if the increasing pressure in the cylinder chamber, which tends to brake the swing motion, exceeds a predefined threshold level. When this happens the frequency of the low pass filter is decreased to almost a frozen state which has the effect of maintaining the EHP valves 21-24 in the existing open position. The filter and thus the EHP valves stay in the “frozen state” until the breaking cylinder chamber pressure falls below the predefined threshold level, at which time the filter is re-enabled and continues to decay to zero. By disabling the filter while the hydraulic function is going over a relief pressure setting for the hydraulic cylinder, the position of the EHP valves are closely coupled to the speed of the piston 28. In other words, the valves only close at a rate the machine system will support. A major advantage is that this solution to the swing wag problem does not require any additional components for the hydraulic system 10 and merely involves programming the function controller with the appropriate software routine.
With reference to
Then at step 76 a determination is made whether the newly calculated value for ΔP LOAD is less than the first, or lower, threshold ΔP LOAD1 (see
However, if neither expression at step 76 or 80 is true, meaning that the value of ΔP LOAD is between the two pressure differential thresholds inclusively, the program execution advances to step 84 to calculate a value for the filter frequency. That frequency is produced by solving a series of equations, the first of which produces a value designated TEMP1 which is equal to the value of ΔP LOAD minus the first threshold value ΔP LOAD1. Another value designated TEMP2 equals the difference between the two pressure differential thresholds and is derived by subtracting the first threshold ΔP LOAD1 from the second threshold ΔP LOAD2. Next a ratio is calculated by dividing TEMP2 into TEMP1 and squaring the result. A temporary frequency value, (FREQ TEMP) is produced by first subtracting the maximum frequency value (FREQ MAX) from the minimum frequency value (FREQ MIN) which produces a negative value that then is multiplied by the previously calculated ratio. The anti-wag frequency (AWFREQ) is produced at the final calculation step by summing the maximum frequency (FREQ MAX) with the negative value of the variable FREQ TEMP. The program execution then advances to step 86. As the hydraulic actuator (e.g. swing cylinder 9) slows, the pressure differential ΔP LOAD changes and step 84 dynamically changes the anti-wag frequency (AWFREQ) in a corresponding manner until the boom assembly 3 stops.
Upon entering step 86 of the filter function 68, the newly derived value for the anti-wag frequency (AWFREQ) is used to determine the coefficients for the filter function. Preferably, a biquadratic digital filter is employed to filter the velocity command. The filter function for a biquadratic filter is given by the expression:
where y(n) is the filter function output referred to as a filtered velocity command, terms A1, A2, B0, B1 and B2 are filter coefficients, x(n) is the present value of the velocity command, x(n−1) and x(n−2) are the previous two values of the velocity command, and y(n−1) and y(n−2) are the previous two values of the output of the filter.
The filter coefficients are defined according to the equations provided at that step 86 in
Thus, the filter routine varies the filter frequency depending upon the load force that the backhoe boom assembly 3 exerts on the hydraulic actuator, i.e. the swing cylinder 9 and piston 28. This frequency variation conforms to the filter function graphically depicted in
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.
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