Abstract:
A machine member driven by a hydraulic actuator may oscillate, or wag, when the hydraulic actuator decelerates or stops. The degree of oscillation is a function of the machine member&#39;s ability to track a deceleration command, which ability varies with changes in the position of the machine member and the load force acting thereon. To reduce the oscillation, a command that controls operation of the hydraulic actuator is filtered using a filter function that changes with the machine member&#39;s load. The load force exerted on the hydraulic actuator which in turn can be designated by fluid pressure that results from the hydraulic actuator. Preferably, the frequency of the filter function is varied inversely with the magnitude of the actuator load force.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS  
       [0001]     Not Applicable  
       STATEMENT REGARDING FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT  
       [0002]     Not Applicable  
       BACKGROUND OF THE INVENTION  
       [0003]     1. Field of the Invention  
         [0004]     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.  
         [0005]     2. Description of the Related Art  
         [0006]     With reference to  FIG. 1 , a backhoe  2  is a common type of earth moving equipment that has a boom assembly  3  comprising a bucket  4  attached to the end of an arm  5  which in turn is coupled by a boom  6  to the frame of a tractor  7 . Three hydraulic cylinders  11  form actuators that are operated independently to move the bucket, arm, and boom. A pivot joint  8  allows the boom assembly  3  to swing left and right with respect to the rear end of the tractor  7 . A hydraulic boom swing cylinder  9  is attached to the boom  6  on one side of the tractor  8  and provides the driving force that swings the boom assembly. On larger backhoes, a pair of hydraulic cylinders are attached on opposite sides of the tractor  7  to swing the boom. Hydraulic fluid is supplied to the boom swing cylinder  9  through valves that are controlled by the backhoe operator.  
         [0007]     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.  
         [0008]     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.  
         [0009]     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.  
         [0010]     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.  
         [0011]     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.  
         [0012]     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.  
         [0013]     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.  
       SUMMARY OF THE INVENTION  
       [0014]     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.  
         [0015]     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.  
         [0016]     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. 
     
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0017]      FIG. 1  is a side view of a backhoe incorporating the present invention;  
         [0018]      FIG. 2  is a schematic diagram of the hydraulic system for the backhoe;  
         [0019]      FIG. 3  is a control diagram for the hydraulic system;  
         [0020]      FIG. 4  graphically depicts a filter function that is applied to a backhoe boom swing command to prevent the boom wag upon stopping; and  
         [0021]      FIG. 5  is a flowchart of a software routine that implements the filter function.  
     
    
     DETAILED DESCRIPTION OF THE INVENTION  
       [0022]     Although the present invention is being described in the context of use on a backhoe as shown in  FIG. 1 , it has application on other types of machines in which large inertia machine functions are attached to the hydraulic actuator and exhibit controllability difficulties.  
         [0023]     With initial reference to  FIG. 2 , the elements of the boom assembly  3  of the backhoe  2  are moved by a novel hydraulic system  10  that includes the hydraulic actuators, such as the boom swing cylinder  9 . The hydraulic system  10  has a positive displacement pump  12  that is driven by a motor or engine (not shown) to draw fluid from a tank  15  and furnish the fluid under pressure to a supply conduit  14 . An unloader valve  17  (such as a proportional pressure relief valve) is connected between the supply conduit  14  and a tank return conduit  18  that leads to the system tank  15 . Operation of the unloader valve  17  regulates pressure in the supply conduit  14 . The novel technique for reducing wag described herein also can be implemented on a hydraulic system that employs a variable displacement pump or other types of hydraulic actuators.  
         [0024]     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 .  
         [0025]     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 .  
         [0026]     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.  
         [0027]     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 .  
         [0028]     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 .  
         [0029]     With reference to  FIG. 3  and the boom swing function  20 , the machine operator manipulates a joystick  47  to indicate desired swing motion for the boom assembly  3 . The output signal from the joystick  47  is applied to an input of a mapping routine  50  in the system controller  46 , which converts the signal indicating the joystick position into a signal denoting a desired velocity for the hydraulic actuator being controlled. The mapping function can be linear or have other shapes as desired. The mapping routine may be implemented by an arithmetic expression that is solved by the computer within system controller  46 , or the mapping may be accomplished by a look up table stored in the system controller&#39;s memory. The output of the mapping routine  50  is a velocity command indicating the direction and speed at which the swing cylinder  9  is desired to move the boom assembly.  
         [0030]     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.  
         [0031]     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.  
         [0032]     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.  
         [0033]     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  FIGS. 2 and 3 , assume that the backhoe operator has been swinging the boom assembly  3  in one direction. In this case, the signal from the joystick  47  for this machine operation indicates a desired velocity for the swing action. The velocity command is transmitted from the system controller  46  to the function controller  44  which controls the operation of the swing hydraulic cylinder  9 . Thus, the function controller  44  is producing signals that open either the first and fourth EHP valves  21  and  24  or the second and third EHP valves  22  and  23 , depending upon the direction of the swing.  
         [0034]     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.  
         [0035]     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.  
         [0036]      FIG. 4  graphically depicts the filter function  68  in terms of a relationship between the filter frequency and a load pressure differential (ΔP LOAD) between the two chambers  26  and  27  of the swing cylinder  9 . Those cylinder chamber pressures are measured by sensors  36  and  38  in  FIG. 2 . This pressure differential corresponds to the load force that the boom assembly  3  exerts on the swing cylinder  9 , which in turn corresponds to the responsiveness of the boom motion to variations in the fluid flow applied to a hydraulic actuator due to changes in the position of the respective control valve. As will be described, the filter frequency is varied as a function of changes in the load pressure differential as defined by the function for the filter depicted in  FIG. 4 . However, it should be understood that the illustrated filter function merely is exemplary and other functions and breakpoints can be utilized without departing from the concept of the present invention. In that regard, the upper and lower limits of the filter frequency have been selected as 1.1 Hz and 0.05 Hz, respectively. Applying these frequency boundaries to the filter function, defines two pressure differential thresholds. The first is a lower pressure threshold, ΔP LOAD  1 , below which the filter frequency remains constant at the maximum filter frequency (FREQ MAX). The second pressure threshold, designated ΔP LOAD 2 , is an upper threshold limit above which the filter frequency remains constant at a minimum frequency (FREQ MIN). For values of the load pressure differential between the first and second thresholds, the filter frequency varies as designated by the curved line in the graph.  
         [0037]     With reference to  FIG. 5 , upon the receipt of a velocity command, the function controller  44  applies an anti-wag filter routine  70  to that command. It should be understood that swing of the boom assembly  3  in one direction is arbitrarily defined as having a positive velocity, whereas swing movement in the opposite direction is designated as a negative velocity. Knowing the direction of the swing is necessary in order to determine a cylinder pressure differential value which has the proper arithmetic sign for use in subsequent calculations by the filter function. The filter routine  70  commences at step  71  where the direction of the present swing is ascertained by determining whether the velocity command is less than the prior filtered velocity command produced by the filter routine  70 . If that relationship is true, the program execution branches to step  72  at which the value of the measured pressure Pb within the rod chamber  27  of the swing cylinder is subtracted from the measured pressure Pa within the head chamber  26  to produce a differential pressure value, ΔP LOAD. Otherwise, if the expression within step  71  is false, the program execution branches to step  74  at which the measured head chamber pressure Pa is subtracted from the rod chamber pressure Pb to produce the differential pressure value, ΔP LOAD.  
         [0038]     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 LOAD 1  (see  FIG. 4 ). If that is the case, the program execution branches to step  78  at which a value for the anti-wag filter frequency (AWFREQ) is set to the maximum frequency value (FREQ MAX), which for example is 1.1 Hz. The program execution then jumps to step  86 . However, if the value of ΔP LOAD is not less than the first threshold, the program execution advances to step  80  at which a determination is made whether that value is greater than the second, or upper, threshold ΔP LOAD 2 . In this latter case, the program execution branches to step  82  at which the anti-wag filter frequency (AWFREQ) is set to the minimum frequency value (FREQ MIN), which for example is 0.05 Hz. Thereafter the program execution jumps to step  86 .  
         [0039]     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 TEMP 1  which is equal to the value of ΔP LOAD minus the first threshold value ΔP LOAD 1 . Another value designated TEMP 2  equals the difference between the two pressure differential thresholds and is derived by subtracting the first threshold ΔP LOAD 1  from the second threshold ΔP LOAD 2 . Next a ratio is calculated by dividing TEMP 2  into TEMP 1  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.  
         [0040]     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:  
               y   ⁡     (   n   )       =         B   ⁢           ⁢   0   *     x   ⁡     (   n   )         +     B   ⁢           ⁢   1   *     x   ⁡     (     n   -   1     )         +     B   ⁢           ⁢   2   *     x   ⁡     (     n   -   2     )               A   ⁢           ⁢   1   *     y   ⁡     (     n   -   1     )         +     A   ⁢           ⁢   2   *     y   ⁡     (     n   -   2     )                     (   1   )             
 
 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. 
 
         [0041]     The filter coefficients are defined according to the equations provided at that step  86  in  FIG. 5 . Specifically the value for coefficient A 0  is produced by multiplying the anti-wag frequency (AWFREQ) by a gain factor and adding an offset. However, it will be understood by those skilled in the art, that not only can the filter coefficients for a biquadratic filter be defined in a other manners, other types of filters and filter functions may be utilized to reduce the effects of swing wag. Next the defined filter coefficients are passed to the conventional digital biquadratic filter at step  88  to configure that filter. Then at step  90 , the filter function is applied to the present velocity command to produce the filtered velocity command, which is utilized by the valve opening program  56  in  FIG. 2  to produce the signals for operating the four electrohydraulic valves  21 - 24 .  
         [0042]     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  FIG. 4 , such that the greater the load force, the lower the filter frequency and hence, the slower the response of the valve assembly  25  to changes of the velocity command. By adapting, the filter function to the magnitude of the load force acting on the swing cylinder  9 , the command filtering is optimized. Under relatively small load force conditions, which produce a commensurate low amount of inertia, a relatively high filter frequency is employed. As the load force and resultant inertia increases, the filter frequency decreases to adequately control the valve assembly  25  to decelerate the boom sufficiently fast to avoid the wag.  
         [0043]     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.