JP5373310B2 - Work machine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a working machine which can easily and rapidly perform load control for avoiding an engine stall caused by an overload on the engine 7. <P>SOLUTION: This working machine comprises: a rack position sensor for detecting an engine load in association with a fuel injection pump; an engine revolution sensor for detecting the number of the revolutions of the engine; a hydraulic servo mechanism 40 for adjusting a swash plate angle of a hydraulic pump motor 31; a hydraulic adjustment mechanism 75 for adjusting a swash plate angle of a hydraulic pump 32; and a controller 101 for controlling the drive of both the mechanisms 40 and 75. The controller 101 adjusts both the mechanisms 40 and 75 in an interlocking manner so that the number of the revolutions of the engine can agree with the number of revolutions, set by a throttle lever, on the basis of information on detection by the rack position sensor and the engine revolution sensor. <P>COPYRIGHT: (C)2010,JPO&amp;INPIT

Description

  The present invention relates to a work machine such as a backhoe, and more particularly to a configuration for executing load control of the work machine.

  A backhoe, which is an example of a work machine, generally includes a variable displacement type first hydraulic pump that supplies hydraulic oil to an arm hydraulic cylinder and the like, and a variable displacement type second hydraulic pump that supplies hydraulic oil to a boom hydraulic cylinder and the like. A hydraulic pump, an engine for operating both hydraulic pumps, and an operating lever for manually operating each hydraulic cylinder are provided. This type of backhoe performs a predetermined work by driving each hydraulic cylinder with hydraulic oil supplied from each hydraulic pump in accordance with the operation amount of the operation lever.

  Patent Document 1 discloses an example of a backhoe. Patent Document 1 relates to a previous application by the applicant of the present application. In the backhoe of Patent Document 1, the pilot pressure is used to calculate the pressure on the discharge port side of the hydraulic pump and the pressure on the suction port side of the hydraulic cylinder. By adjusting the difference within a predetermined range, the amount of hydraulic fluid supplied to the hydraulic cylinder is kept substantially constant regardless of the load applied to the hydraulic cylinder (the working speed of the hydraulic cylinder is kept substantially constant). Load sensing control is adopted.

Moreover, in the backhoe of Patent Document 1, when the engine load becomes a predetermined value or more, the amount of hydraulic oil discharged from both hydraulic pumps is suppressed using the pilot pressure, so that the engine stall due to engine overload is suppressed. Load control is also used to avoid this.
JP 2006-177397 A

  However, each control disclosed in Patent Document 1 is a feedback control using a pilot pressure in spite of having a plurality of hydraulic pumps, so that the hydraulic system for the feedback control becomes long and complicated. Tend to. Then, since it is necessary to ensure a sufficient supply amount of hydraulic oil to each hydraulic cylinder, each hydraulic pump has to have a large capacity, resulting in an increase in cost. In general, feedback control using pilot pressure has room for improvement in terms of response speed.

  Therefore, the present invention has a technical problem to improve such a current situation.

The invention of claim 1 is a variable displacement type first hydraulic pressure supply device that supplies hydraulic oil to a first hydraulic actuator, a variable displacement type second hydraulic pressure supply device that supplies hydraulic oil to a second hydraulic actuator, A work machine having an engine for operating both hydraulic pressure supply devices, a load detection means for detecting an engine load in relation to the fuel supply device for the engine, and a rotation speed detection means for detecting the engine speed A first flow rate adjusting device for adjusting a discharge flow rate of the first hydraulic pressure supply device, a second flow rate adjusting device for adjusting a discharge flow rate of the second hydraulic pressure supply device, and driving of both the flow rate adjusting devices. control means and further comprising a, wherein, when the engine load exceeds the set set load advance, on the basis of the detected information and said load detecting means and said rotational speed detecting means d As Jin speed coincides with the target rotational speed, by adjusting in conjunction with the two flow controllers, wherein those of suppressing the amount of hydraulic oil from the hydraulic pressure supply device.

According to a second aspect of the present invention, in the work machine according to the first aspect, throttle operating means for operating the fuel supply device and changing the target rotational speed to a control unit of a body on which the engine is mounted. A working unit operating means for manually operating each of the hydraulic actuators, and the control means is configured to operate the throttle operating means when the target rotational speed becomes less than the original set rotational speed. In addition, each flow rate adjusting device is adjusted so that the amount of hydraulic oil from each hydraulic pressure supply device with respect to the operation amount of the working unit operating means is reduced.

According to the present invention, the control means for controlling the driving of the first flow rate adjusting device for adjusting the discharge flow rate of the first hydraulic pressure supply device and the second flow rate adjusting device for adjusting the discharge flow rate of the second hydraulic pressure supply device is the engine. When the load exceeds a preset set load , the two flow rate adjusting devices are linked so that the engine rotational speed matches the target rotational speed based on the detection information of the load detecting means and the rotational speed detecting means. by adjusting Te, wherein the suppressing amount of hydraulic oil from the hydraulic pressure supply device, even if the engine load is large, while maintaining the engine load value near the set value which is the upper limit, can the engine drive, While the engine stall can be surely suppressed, the energy loss is reduced and the engine output can be used efficiently.

In particular, according to the second aspect of the present invention, when the target rotational speed is less than the original set rotational speed due to the operation of the throttle operating means, the control means controls the hydraulic pressure relative to the operation amount of the working unit operating means. Since each flow rate adjustment device is adjusted so that the hydraulic oil amount from the supply device is reduced, the hydraulic oil amount from each hydraulic pressure supply device with respect to the operation amount of the working unit operating means is a decrease in engine speed. It will be reduced proportionally.

  Accordingly, since the hydraulic actuators can be operated in a small amount by the operation of the working unit operating means, the working unit of the work machine can be finely adjusted, the operability is improved, and the work efficiency is improved. Play.

  DESCRIPTION OF EMBODIMENTS Hereinafter, an embodiment embodying the present invention will be described based on drawings (FIGS. 1 to 10) applied to a backhoe as a work machine. 1 is a side view of the backhoe, FIG. 2 is a plan view inside the cabin, FIG. 3 is a hydraulic system diagram of the backhoe, FIG. 4 is an enlarged explanatory view around the closed loop hydraulic circuit, FIG. 5 is an enlarged explanatory view around the hydraulic adjustment mechanism, FIG. 6 is a functional block diagram of the control means, FIG. 7 is a flow chart of load control, FIG. 8 is a flow chart of a fine adjustment control subroutine, FIG. 9 is an explanatory view showing another example of the hydraulic system, and FIG. It is a functional block diagram.

(1). Outline of Backhoe First, an outline of the backhoe 1 will be described with reference mainly to FIGS. 1 and 2.

  A backhoe 1, which is an example of a work machine, includes a crawler-type traveling device 2 having a pair of left and right traveling crawlers 3 (shown only on the left side in FIG. 1), and a swivel base 4 (airframe) provided on the traveling device 2. I have. The swivel base 4 is configured to be capable of horizontal swivel over 360 ° in all directions by a swivel motor 9 (see FIG. 2). A soil discharge plate 5 is mounted on the front portion of the traveling device 2 so as to be rotatable up and down.

  The swivel 4 is equipped with a cabin 6 and an engine 7 as a control unit. A working unit 10 having a boom 11, an arm 12, and a bucket 13 for excavation work is provided at the front of the swivel 4. Inside the cabin 6, there is a control seat 8 on which an operator is seated, a throttle lever 116 as a throttle operating means for setting and maintaining the output rotational speed of the engine 7, and lever switch groups 117 to 120 (working operation) A lever 117, an arm operation lever 118, a bucket operation switch 119 and a boom operation lever 120), a load factor setting dial 113 as a load factor setting means, and the like are arranged (see FIG. 2).

  The boom 11, which is a component of the working unit 10, is formed in a shape that protrudes forward at the tip side and is bent in a square shape when viewed from the side. The base end portion of the boom 11 is pivotally attached to a boom bracket 14 attached to the front portion of the swivel base 4 so as to be swingable about a horizontal boom shaft 15. On the inner surface (front surface) side of the boom 11, a one-rod double-acting boom cylinder 16 is disposed for swinging it up and down. The cylinder side end of the boom cylinder 16 is pivotally supported by the front end of the boom bracket 14 so as to be rotatable. The rod side end portion of the boom cylinder 16 is pivotally supported by a front bracket 17 fixed to the front surface side (dent side) of the bent portion of the boom 11. The boom cylinder 16 corresponds to a first hydraulic actuator described in the claims.

  A base end portion of a long rectangular tube-like arm 12 is pivotally attached to a tip end portion of the boom 11 so as to be swingable about a lateral arm shaft 19. A one-rod double-acting arm cylinder 20 for swinging and swinging the arm 12 is disposed on the upper front side of the boom 11. The cylinder side end portion of the arm cylinder 20 is pivotally supported by a rear bracket 18 fixed to the back side (projecting side) of the bent portion of the boom 11. The rod side end of the arm cylinder 20 is pivotally supported by an arm bracket 21 fixed to the base end side outer surface (front surface) of the arm 12.

  A bucket 13 as an attachment for excavation is pivotally attached to the distal end portion of the arm 12 so that the bucket 13 can be swung around a lateral bucket shaft 22. On the outer surface (front surface) side of the arm 12, a one-rod double-acting bucket cylinder 23 for scrambling and rotating the bucket 13 is disposed. A cylinder side end portion of the bucket cylinder 23 is pivotally supported by the arm bracket 21. The rod side end of the bucket cylinder 23 is pivotally supported by the bucket 13 via a connecting link 24 and a relay rod 25. The swing motor 9, the arm cylinder 20, and the bucket cylinder 23 correspond to a second hydraulic actuator described in the claims.

(2). Next, the structure of the hydraulic system 30 of the backhoe 1 will be described with reference to FIGS.

  The hydraulic system 30 of the backhoe 1 shown in FIG. 3 includes a variable displacement hydraulic pump motor 31 as a first hydraulic supply device, a variable displacement hydraulic pump 32 as a second hydraulic supply device, and a fixed displacement pilot. And a pump 33. An output shaft 27 protruding from the engine 7 passes through the pumps 31 to 33, and the pumps 31 to 33 are configured to be driven by the rotation of the output shaft 27. That is, the rotating shaft (output shaft 27) for driving the pumps 31 to 33 is a single common shaft.

  The hydraulic pump motor 31 is for supplying hydraulic oil to the boom cylinder 16 as a first hydraulic actuator that requires a relatively large driving force. The hydraulic pump 32 is for supplying hydraulic oil to the swing motor 9, the arm cylinder 20, and the bucket cylinder 23 as the second hydraulic actuator. The pilot pump 33 is for applying a pilot pressure to a hydraulic pressure adjusting mechanism 75 described later.

(2-1). Structure of Closed Loop Hydraulic Circuit First, the closed loop hydraulic circuit 34 that drives the boom cylinder 16 as the first hydraulic actuator will be described. The closed-loop hydraulic circuit 34 includes the above-described single-rod double-acting boom cylinder 16 and a hydraulic pump motor 31 that is a swash plate type axial piston pump motor. The boom cylinder 16 and the hydraulic pump motor 31 are connected in a closed loop by a bottom side oil passage 35 and a rod side oil passage 36. In this case, the bottom oil chamber 37 of the boom cylinder 16 is connected to the hydraulic pump motor 31 via the bottom side oil passage 35, and the rod oil chamber 38 of the boom cylinder 16 is connected to the hydraulic pump motor 31 via the rod side oil passage 36. It is connected.

  For this reason, when the boom 11 is lowered, the hydraulic oil from the boom cylinder 16 is supplied to the hydraulic pump motor 31 via the bottom side oil passage 35. As a result, the hydraulic pump motor 31 exhibits an action as a motor. The driving of the hydraulic pump 32 and the pilot pump 33 on the common output shaft 27 is partially or entirely borne. That is, since the driving load of the hydraulic pump 32 and the pilot pump 33 is reduced by the motor-like operation of the hydraulic pump motor 31, the work amount of the entire engine can be reduced, and the effect of improving the fuel efficiency is high.

  The closed loop hydraulic circuit 34 includes a hydraulic servo mechanism 40 as a first flow rate adjusting device that controls the inclination angle (swash plate angle) of the movable swash plate 31 a in the hydraulic pump motor 31. The hydraulic servo mechanism 40 includes a single-rod double-acting adjustment cylinder 41 that changes the inclination angle of the movable swash plate 31a, an adjustment pump 42 that supplies hydraulic oil to the adjustment cylinder 41, and supply of hydraulic oil from the adjustment pump 42. And a 4-port 3-position switching electromagnetic servo valve 43 for adjusting the direction and supply amount.

  The inlet port 43 a of the electromagnetic servo valve 43 is connected to the hydraulic oil tank 44 via the adjustment pump 42, and the outlet port 43 d is directly connected to the hydraulic oil tank 44. The bottom port 43 b of the electromagnetic servo valve 43 is connected to the bottom oil chamber 45 of the double action adjustment cylinder 41, and the rod side port 43 c is connected to the rod oil chamber 46 of the double action adjustment cylinder 41. The tip of the piston rod 47 in the double-action adjusting cylinder 41 is linked to the movable swash plate 31 a of the hydraulic pump motor 31.

  The electromagnetic servo valve 43 has a neutral state and a hydraulic oil supply state to the bottom oil chamber 45 of the double acting adjustment cylinder 41 by excitation of an electromagnetic solenoid corresponding to an operation amount of the boom operation lever 120 disposed in the cabin 6. The double-acting adjustment cylinder 41 is configured to be switched to a hydraulic oil supply state to the rod oil chamber 46. When the electromagnetic servo valve 43 is switched and driven by operating the boom operation lever 120, the double-action adjusting cylinder 41 expands and contracts, and the inclination angle (swash plate angle) of the movable swash plate 31a in the hydraulic pump motor 31 is changed and adjusted. Then, the supply direction and supply amount of hydraulic oil from the hydraulic pump motor 31 to the boom cylinder 16 are adjusted. As a result, the expansion / contraction direction and the expansion / contraction amount of the boom cylinder 16 are changed steplessly, and the boom 11 swings up and down.

  Between the bottom side oil passage 35 and the rod side oil passage 36, a three-port three-position switching type direction switching valve 50 is disposed. In the direction switching valve 50, the amount of hydraulic oil flowing out from one oil chamber 37 due to the pressure receiving area difference between the two oil chambers 37, 38 in the boom cylinder 16 is greater than the amount of hydraulic oil flowing into the other oil chamber 38. This is to discharge the surplus when there are many.

  The first inlet port 50 a of the direction switching valve 50 is connected to the bottom side oil passage 35 via the first inlet oil passage 51, and the second inlet port 50 b is connected to the rod side oil passage 36 via the second inlet oil passage 52. It is connected. The outlet port 50 c of the direction switching valve 50 is connected to the hydraulic oil tank 44 via the drain oil passage 53.

  Further, the direction switching valve 50 is connected to the bottom side oil passage 35 via the bottom side pilot oil passage 54, and is connected to the rod side oil passage 36 via the rod side pilot oil passage 55. For this reason, the pressure in the bottom side pilot oil passage 54 is substantially the same as the pressure in the bottom side oil passage 35, and hence the pressure in the bottom oil chamber 37 in the boom cylinder 16, and the pressure in the rod side pilot oil passage 55 is The pressure in the rod-side oil passage 36 and thus the pressure in the rod oil chamber 38 in the boom cylinder 16 are substantially the same.

  In this case, the direction switching valve 50 has a neutral state, a hydraulic oil discharge state from the bottom side oil passage 35 to the drain oil passage 53, and the rod side oil passage 36 according to the pressure difference between the pilot oil passages 54 and 55. Is switched to a hydraulic oil discharge state to the drain oil passage 53. Due to the action of the direction switching valve 50, the imbalance in the amount of hydraulic fluid flowing between the bottom side and rod side oil passages 35 and 36 is corrected, and the boom cylinder 16 smoothly expands and contracts.

  A charge relief circuit 60 having two relief valves 63 and 64 and two check valves 65 and 66 is disposed between the bottom side oil passage 35 and the rod side oil passage 36. Basically, if the pressure in one oil passage 35 (36) becomes too high, the charge relief circuit 60 does not supply hydraulic oil to one oil chamber 37 (38) in the boom cylinder 16, but the other oil. The overload of the closed loop hydraulic circuit 34 is prevented by letting it escape to the path 36 (35) and the hydraulic oil tank 44. When the amount of hydraulic oil in the closed-loop hydraulic circuit 34 is small, the hydraulic oil is supplied from the hydraulic oil tank 44 via the charge relief circuit 61 by the self-priming force of the hydraulic pump motor 31.

  In the embodiment, a pair of bypass oil passages 61 and 62 are connected in parallel to the bottom side oil passage 35 and the rod side oil passage 36. A first relief valve 63 for releasing the pressure in the bottom side oil passage 35 and a second relief valve 64 for releasing the pressure in the rod side oil passage 36 are provided in the cylinder side bypass oil passage 61. ing. In the pump-side bypass oil passage 62, a first check valve 65 that opens only in the direction of the bottom-side oil passage 35 and a second check valve 66 that opens only in the direction of the rod-side oil passage 36 are provided. .

  A discharge oil passage 67 connects between the relief valves 64 and 65 in the cylinder-side bypass oil passage 62 and between the check valves 66 and 67 in the pump-side bypass oil passage 63. The leading end of the discharge oil passage 67 is connected to the middle portion of the drain oil passage 53. Accordingly, the drain oil passage 67 communicates with the hydraulic oil tank 44 via the drain oil passage 53.

(2-2). Next, the charge hydraulic circuit 74 for driving the second hydraulic actuator (the swing motor 9, the arm cylinder 20, and the bucket cylinder 23) will be described. The charge hydraulic circuit 74 includes the swing motor 9, the single rod double acting arm cylinder 20, the single rod double acting bucket cylinder 23, the hydraulic pump 32 which is a swash plate type axial piston pump, and a fixed capacity. And a hydraulic pressure adjusting mechanism 75 as a second flow rate adjusting device.

  The suction side of the hydraulic pump 32 communicates with the hydraulic oil tank 44 via the suction oil passage 76. The swing motor 9, the arm cylinder 20, and the bucket cylinder 23 are branched and connected to the charge oil passage 77 extending from the discharge side of the hydraulic pump 32 via corresponding flow control valve units 78, 79, and 80.

  The turning flow control valve unit 78 corresponds to the operation amount of the turning operation lever 117 disposed in the cabin 6, and the neutral state and the operating oil supply state to one motor oil passage 81 for the turning motor 9. The hydraulic oil supply state to the other motor oil passage 82 is switched and driven. When the flow control valve unit 78 for turning is switched by operating the turning operation lever 117, the supply direction and the supply amount of hydraulic oil from the hydraulic pump 32 to the turning motor 9 are adjusted. As a result, the rotation direction and the rotation amount of the turning motor 9 are changed steplessly, and the turntable 4 turns horizontally.

  The arm flow control valve unit 79 corresponds to the operation amount of the arm operation lever 118 disposed in the cabin 6, and is in a neutral state and the hydraulic oil supply state to the bottom side oil passage 83 with respect to the arm cylinder 20. It is configured to be switched to a hydraulic oil supply state to the rod side oil passage 84. When the arm flow control valve unit 79 is switched and operated by operating the arm operation lever 118, the supply direction and supply amount of hydraulic oil from the hydraulic pump 32 to the arm cylinder 20 are adjusted. As a result, the direction and amount of expansion and contraction of the arm cylinder 20 are changed steplessly, and the arm 12 swings up and down.

  The bucket flow control valve unit 80 includes a neutral state, a hydraulic oil supply state to the bottom side oil passage 85 with respect to the bucket cylinder 23, and a rod side oil passage corresponding to the operation amount of the sliding bucket operation switch 119. The hydraulic oil supply state to 86 is switched and driven. When the bucket flow control valve unit 80 is switched and driven by operating the bucket operation switch 119, the supply direction and supply amount of hydraulic oil from the hydraulic pump 32 to the bucket cylinder 23 are adjusted. As a result, the expansion / contraction direction and the expansion / contraction amount of the bucket cylinder 23 are changed steplessly, and the bucket 13 at the distal end portion of the arm 12 scoops and rotates around the lateral bucket shaft 22.

  The hydraulic adjustment mechanism 75 as the second flow rate adjusting device is for controlling the inclination angle (swash plate angle) of the movable swash plate 32a in the hydraulic pump 32, and is a single rod that changes the inclination angle of the movable swash plate 32a. Single-acting adjustment cylinder 90, three-port two-position pressure servo valve 91 for adjusting the supply direction and supply amount of hydraulic oil from hydraulic pump 32 to single-acting adjustment cylinder 90, and pilot pump 33 And a pressure adjusting electromagnetic valve 92 for adding the pilot pressure to the pressure servo valve 91.

  The oil path side first port 91a of the pressure servo valve 91 is connected between the hydraulic pump 32 in the charge oil path 77 and the branch portion 87 to each flow control valve unit 78-80, and the oil path side second port 91b is Directly connected to the hydraulic oil tank 44. The cylinder side port 91 c of the pressure servo valve 91 is connected to the bottom oil chamber 93 of the single acting adjustment cylinder 90. In the rod chamber of the single-acting adjustment cylinder 90, a return spring 95 that urges the piston rod 94 in the direction of shortening is housed. The tip of the piston rod 94 in the single-acting adjustment cylinder 90 is linked to the movable swash plate 32 a of the hydraulic pump 32.

  The pressure servo valve 91 is connected between the hydraulic pump 32 and the branching portion 87 in the charge oil passage 77 via the first pilot oil passage 96, while each flow rate via the second pilot oil passage 97. It is connected to control valve units 78-80. For this reason, the pressure in the first pilot oil passage 96 is substantially the same as the pressure on the discharge side of the hydraulic pump 32, and the pressure in the second pilot oil passage 97 is in the swing motor 9, arm cylinder 20, and bucket cylinder 23. It is almost the same as the highest pressure on the suction side.

  The pressure servo valve 91 is also connected to the discharge side of the pressure adjusting electromagnetic valve 92 via a forced pilot oil passage 98. The suction side of the pressure adjusting electromagnetic valve 92 is connected to the discharge side of the pilot pump 33. The suction side of the pilot pump 33 is connected to a midway portion of the suction oil passage 76.

  In this case, the pressure servo valve 91 basically has a hydraulic oil supply state to the bottom oil chamber 93 of the single acting adjustment cylinder 91 according to the pressure difference between the first and second pilot oil passages 96 and 97, and the bottom oil. The hydraulic oil is discharged from the chamber 93 and switched to a state where it is discharged.

  When the pressure difference between the first and second pilot oil passages 96 and 97 deviates from a predetermined range set in advance and the pressure servo valve 91 is switched and driven, the single-acting adjustment cylinder 90 expands and contracts to move in the hydraulic pump 32. The inclination angle (swash plate angle) of the swash plate 32a is changed and adjusted, and the amount of hydraulic oil supplied from the hydraulic pump 32 to each second hydraulic actuator (the swing motor 9, the arm cylinder 20 and the bucket cylinder 23) is adjusted. .

  As a result, the difference between the pressure on the discharge side of the hydraulic pump 32 and the pressure on the suction side of each second hydraulic actuator 8, 20, 23 is adjusted within a predetermined range and applied to each second hydraulic actuator 8, 20, 23. Regardless of the magnitude of the load, the amount of hydraulic oil supplied to each second hydraulic actuator 8, 20, 23 is held substantially constant. That is, the load sensing function that keeps the operating speed of each of the second hydraulic actuators 8, 20, and 23 substantially constant works.

  The pressure adjusting electromagnetic valve 92 is configured to adjust the pilot pressure applied to the pressure servo valve 91 via the first pilot oil passage 96 by excitation of an electromagnetic solenoid based on control information from the controller 101 described later. ing. As a result of the action of the pressure adjusting electromagnetic valve 92, the pressure servo valve 91 is switched to the hydraulic oil supply state and the single action adjusting cylinder 90 changes / adjusts the inclination angle of the movable swash plate 32a in the hydraulic pump 32, The amount of hydraulic oil from the hydraulic pump 32 is forcibly reduced. Note that the pressure adjusting solenoid valve 92 is normally set so that pilot pressure is not applied to the pressure servo valve 91 (when there is no control information from the controller 101).

(3). Configuration for Executing Load Control Next, a configuration for executing backhoe load control will be described with reference to FIG.

  The controller 101 and the electronic governor controller 102 as control means mounted on the backhoe 1 temporarily store a ROM for storing a control program and data, a control program and data, in addition to a CPU that executes various arithmetic processes and controls. A RAM for storage, an input / output interface, and the like are provided. The controller 101 is connected to the battery 104 via a key switch 103 for applying power.

  The key switch 103 is a rotary switch that can be rotated by a predetermined key inserted into the keyhole, and is arranged in the cabin 6 although not shown. The key switch 103 is also connected to a starter 105 for starting the engine 7.

  An electronic governor controller 102 that controls the engine speed R is connected to the controller 101. The electronic governor controller 102 includes a fuel injection pump 106 with an electronic governor 107 as a fuel supply device, an engine rotation sensor 108 as a rotation speed detecting means for detecting the engine rotation speed, and fuel injection from the rack position of the fuel injection pump 106. A rack position sensor 109 as load detecting means for detecting the amount, a throttle potentiometer 110 for detecting the operation position of the throttle lever 116, and a throttle solenoid 111 for adjusting the rack position of the fuel injection pump 106 are connected.

  When the throttle lever 116 is manually operated, the electronic governor controller 102 controls the throttle solenoid 111 based on the detection information of the throttle potentiometer 110 so that the engine speed R becomes the set speed Rs (target speed) of the throttle lever 116. The rack position of the fuel injection pump 106 is adjusted by driving. For this reason, the engine speed R is held at a value corresponding to the position of the throttle lever 116.

  As an output-related device, the controller 101 adds an electromagnetic servo valve 43 for adjusting the supply direction and supply amount of hydraulic oil from the adjustment pump 42 and a pilot pressure from the pilot pump 33 to the pressure servo valve 91. A pressure adjusting electromagnetic valve 92 is connected. Also, the controller 101 includes a boom potentiometer 112 for detecting the operation position of the boom operation lever 120 as an input-related device, and a load factor setting dial 113 for manually setting the set load factor Zs of the engine during load control. Is connected.

  In the embodiment, the amount of hydraulic oil discharged from the hydraulic pump motor 31 and the hydraulic pump 32 is suppressed so that the engine speed R matches the set speed Rs based on detection results of the engine load factor Z and the set load factor Zs. Load control is adopted. Here, the engine load factor Z is obtained by calculating the ratio of the engine load during work with the engine load detected by the rack position sensor 109 being 100%, and is the engine load factor in the idling state. Z becomes 0 (zero). The load factor setting dial 113 is configured to be able to arbitrarily adjust the set load factor Zs in the range of 70 to 100%.

  In the ROM of the controller 101, a relational expression for executing proportional-integral control (PI control) based on the difference between the engine load factor Z and the set load factor Zs and the difference between the set rotational speed Rs and the engine rotational speed R. Is stored in advance. As a relational expression in this case, the following expression 1 is adopted.

Here, k1 to k4 are proportional constants, the first and third terms on the right side are proportional terms, and the second and fourth terms on the right side are integral terms. If the control calculation value PI is less than 0 (PI <0), it is replaced with 0 and the state variables (Z-Zs, Rs-R) of the integral term are reset. If the control calculation value PI is equal to or greater than the predetermined value A (PI> A), it is replaced with A as it is. That is, the control range of the control calculation value PI is set to 0 ≦ PI ≦ A.

(4). Description of Load Control Next, an example of load control will be described with reference to the flowcharts of FIGS.

  The controller 101 serving as the control means suppresses the amount of hydraulic oil discharged from the hydraulic pump motor 31 and the hydraulic pump 32 so that the engine rotational speed R matches the set rotational speed Rs. Execute load control to avoid the stall. Here, it is assumed that the set rotational speed Rs is preset by the throttle lever 116 and the set load factor Zs is preset by the load factor setting dial 113.

  First, following the start of load control, a set load factor Zs that is a set value of the load factor setting dial 113, a detected value (engine load) of the rack position sensor 109, and a set rotation speed that is a set value of the throttle lever 116 Rs and the detected value R of the engine rotation sensor 108 are read (step S1), and the current engine load factor Z is calculated based on the detected value of the rack position sensor 109 (step S2).

  Next, after calculating the control calculation value PI from the set load factor Zs and the current engine load factor Z, the set rotation speed Rs and the current engine rotation speed R according to the above-described equation 1 (step S3), the control calculation value Based on PI, a command current value Ip to the pressure adjusting electromagnetic valve 92 and a command current value Ir to the electromagnetic servo valve 43 are calculated (step S4). Examples of the arithmetic expression employed in step S4 include the following expressions 2 and 3. Here, L1 and L2 are proportional constants.

The command current value Ip to the pressure adjusting solenoid valve 92 is the product of the proportional constant and the power root of the control calculation value PI. The pressure servo valve 91 is indirectly controlled via the pressure adjusting solenoid valve 92. This is because the structure is adopted.

  Then, based on the command current value Ip (control information) from the controller 101, the electromagnetic solenoid of the pressure adjusting electromagnetic valve 92 is excited to add pilot pressure to the pressure servo valve 91 via the first pilot oil passage 96. The electromagnetic solenoid of the electromagnetic servo valve 43 is excited on the basis of the command current value Ir (control information) from the controller 101, and the electromagnetic servo valve 43 is directly brought into the operating oil supply state to the bottom oil chamber 45 of the double-action adjusting cylinder 41. Switching driving is performed (step S5).

  Then, the pressure servo valve 91 is switched to the hydraulic oil supply state, and the single acting adjustment cylinder 90 changes / adjusts the inclination angle of the movable swash plate 32a in the hydraulic pump 32, so that the hydraulic oil amount from the hydraulic pump 32 is appropriately adjusted. To reduce. At the same time, the double-acting adjustment cylinder 41 changes and adjusts the inclination angle of the movable swash plate 31a in the hydraulic pump motor 31 by the switching drive of the electromagnetic servo valve 43, and the amount of hydraulic oil from the hydraulic pump motor 31 is also appropriately reduced. To do. As a result, the engine load is reduced and engine stall can be avoided.

  In the embodiment, the load control as described above is executed every moment, and when the actual engine load exceeds the set load, the load is reduced, and the engine speed R matches the set speed Rs. Since the amount of hydraulic oil discharged from the hydraulic pump motor 31 and the hydraulic pump 32 is controlled, the engine 7 can be driven while maintaining the engine load factor Z at a value near the upper limit set load factor Zs. For this reason, while being able to suppress engine stall reliably, energy loss can be reduced and the output of the engine 7 can be used efficiently.

  In the embodiment, since feedback control by the electromagnetic valves 43 and 92 is employed, the response speed is faster than the configuration of Patent Document 1. Further, since the hydraulic pressure is not used for controlling the electromagnetic valves 43 and 92, the hydraulic pump 32 is not burdened (the capacity of the hydraulic pump 32 need not be increased for feedback control).

  Moreover, since the load factor setting dial 113 is also arranged in the cabin, the reference value for determining whether the engine 7 is in an overload state can be arbitrarily set by the setting operation of the load factor setting dial 113 (70 to 100% in the embodiment). Range). For this reason, when executing the load control, it is possible to easily adopt an appropriate setting according to the work situation, the operator's preference, etc., and there is an advantage that the load control can be optimized.

  When the load control is being executed, the operator manually operates the throttle lever 116 in a direction to decrease the engine speed R, and the engine speed R becomes less than the original set speed Rs (step S6: If YES, the fine adjustment control subroutine is executed (step S7).

  As shown in FIG. 8, in the fine adjustment control subroutine, following the start, the set load factor Zs, the detected value (engine load) of the rack position sensor 109, the new set rotation speed Rs', and the engine rotation sensor 108 are detected. Is detected (step T1), and the current engine load factor Z ′ is calculated based on the detected value of the rack position sensor 109 (step T2).

  Next, after calculating the control calculation value PI ′ from the set load factor Zs and the current engine load factor Z ′, the new set rotation speed Rs ′ and the current engine rotation speed R ′ according to the above-described equation 1 (step) T3) A correction gain G is obtained from the new set rotational speed Rs ′ and the current engine rotational speed R ′ and the relational expression or control map stored in advance in the ROM of the controller 101 (step T4). The correction gain G in this case becomes smaller as the difference (Rs−R) between the new set speed Rs ′ and the current engine speed R ′ increases.

  After calculating the correction gain G in step T4, based on the correction gain G and the control calculation value PI ′, the command current value Ip ′ to the pressure adjusting electromagnetic valve 92 and the command current value Ir ′ to the electromagnetic servo valve 43 are calculated. Are calculated (step T5). Examples of the arithmetic expression employed in step T5 include the following expressions 4 and 5. Here, L1 ′ and L2 ′ are proportional constants.

Then, based on the command current value Ip ′ (control information) from the controller 101, the electromagnetic solenoid of the pressure adjusting electromagnetic valve 92 is excited to add pilot pressure to the pressure servo valve 91 via the first pilot oil passage 96, Similarly, the electromagnetic solenoid of the electromagnetic servo valve 43 is excited based on the command current value Ir ′ (control information) from the controller 101, and the electromagnetic servo valve 43 is directly supplied to the bottom oil chamber 45 of the double acting adjustment cylinder 41. The drive is switched to the state (step T6).

  Then, the pressure servo valve 91 is switched to the hydraulic oil supply state, and the single acting adjustment cylinder 90 changes / adjusts the inclination angle of the movable swash plate 32a in the hydraulic pump 32, so that the hydraulic oil amount from the hydraulic pump 32 is appropriately adjusted. To reduce. At the same time, the double-acting adjustment cylinder 41 changes and adjusts the inclination angle of the movable swash plate 31a in the hydraulic pump motor 31 by the switching drive of the electromagnetic servo valve 43, and the amount of hydraulic oil from the hydraulic pump motor 31 is also appropriately reduced. To do.

  As a result, the amount of hydraulic oil from the hydraulic pump motor 31 and the hydraulic pump 32 with respect to the operation amounts of the respective working unit operation means 117 to 120 (the turning operation lever 117, the arm operation lever 118, the bucket operation switch 119, and the boom operation lever 120) is Therefore, it decreases in proportion to the decrease in the engine speed R ′.

  Accordingly, since the hydraulic actuators 9, 16, 20, and 23 can be operated in a small amount by the operation of each working unit operating means 117 to 120, the swivel base 4 and each working unit 10 can be finely adjusted, and the operability is improved. It also contributes to improving work efficiency.

(5). Another Example of Hydraulic System FIGS. 9 and 10 show another example of the hydraulic system in the backhoe. In the other example, a hydraulic pressure adjusting mechanism 175 as a second swash plate adjusting device is different from that of the above-described embodiment. In other words, the pressure adjusting electromagnetic valve 92 is omitted, and the pressure servo valve 91 is changed to a 3-port 2-position switching electromagnetic servo valve 191, which is different from the above-described embodiment. Therefore, a charging electromagnetic servo valve 191 is connected to the controller 101 instead of the pressure adjusting electromagnetic valve 92. Other configurations are the same as those of the above-described embodiment.

  The mode of load control in another example is basically the same as that of the above-described embodiment. In this case, the command current values Ip and Ip ′ to the charging electromagnetic servo valve 191 are products of proportionality constants and control operation values PI and PI ′. That is, the following formulas 2 ′ and 4 ′ are expressed by linear functions as in the above formulas 3 and 5. This is because the structure is such that the charging electromagnetic servo valve 92 is directly controlled without the pressure adjusting electromagnetic valve 92.

Even when such control is employed, the same operational effects as those of the above-described embodiment can be obtained. In particular, the pressure adjusting solenoid valve 92 is omitted, and the pressure servo valve 91 is changed to a 3-port 2-position switching electromagnetic servo valve 191. Therefore, the response speed is faster than that in the above-described embodiment. In addition, since the number of parts is small, it contributes to cost reduction.

(6). Others The present invention is not limited to the above-described embodiment, and can be embodied in various forms. For example, the present invention is not limited to a backhoe, but can also be applied to agricultural machines such as a combiner and special work vehicles such as a wheel loader. In addition, the configuration of each unit is not limited to the illustrated embodiment, and various modifications can be made without departing from the spirit of the present invention.

It is a side view of a backhoe. It is a top view in a cabin. It is a hydraulic system diagram of a backhoe. It is an expansion explanatory view around a closed loop hydraulic circuit. It is an expansion explanatory view around a hydraulic adjustment mechanism. It is a functional block diagram of a control means. It is a flowchart of load control. It is a flowchart of a fine adjustment control subroutine. It is explanatory drawing which shows another example of a hydraulic system. It is a functional block diagram of the control means in another example.

DESCRIPTION OF SYMBOLS 1 Backhoe as a working machine 4 Swivel base 6 Cabin 7 Engine 9 Turning motor 10 Working part 11 Boom 12 Arm 13 Bucket 16 Boom cylinder 20 Arm cylinder 23 Bucket cylinder 30 Hydraulic system 31 Hydraulic pump motor 31a as the first hydraulic supply device Movable Swash plate 32 Hydraulic pump 33 as second hydraulic supply device Pilot pump 34 Closed loop hydraulic circuit 40 Hydraulic servo mechanism 43 as first flow control device Electromagnetic servo valve 44 Hydraulic oil tank 74 Charge hydraulic circuit 75 Second hydraulic flow control device Hydraulic adjustment mechanism 91 Pressure servo valve 92 Pressure adjustment electromagnetic valve 101 Controller 102 as control means Electronic governor controller 106 Fuel injection pump 107 Electronic governor 108 Engine rotation sensor 109 Rack position sensor 113 Load factor setting unit Dial 116 throttle lever 120 boom operation lever

Claims (2)

A variable displacement type first hydraulic pressure supply device that supplies hydraulic oil to the first hydraulic actuator, a variable displacement type second hydraulic pressure supply device that supplies hydraulic oil to the second hydraulic actuator, and both the hydraulic pressure supply devices are operated. A working machine equipped with an engine,
A load detecting means for detecting an engine load in relation to the fuel supply device to the engine; a rotation speed detecting means for detecting an engine speed; and a first flow rate adjusting device for adjusting a discharge flow rate of the first hydraulic pressure supply device. And a second flow rate adjusting device for adjusting the discharge flow rate of the second hydraulic pressure supply device, and a control means for controlling the drive of the both flow rate adjusting devices,
The control means, when the engine load exceeds a preset set load , based on detection information of the load detection means and the rotation speed detection means so that the engine rotation speed matches the target rotation speed. , By controlling the flow rate adjusting devices in conjunction with each other , the amount of hydraulic oil from each hydraulic pressure supply device is suppressed ,
Work machine. A control unit of an airframe on which the engine is mounted has a throttle operation unit for operating the fuel supply device to change the target rotational speed, and a working unit operation unit for manually operating each hydraulic actuator. Provided,
When the target rotational speed is less than the original set rotational speed due to the operation of the throttle operating means, the control means determines the amount of hydraulic oil from each hydraulic pressure supply device relative to the operational amount of the working unit operating means. Adjusting each of the flow control devices to reduce,
The work machine according to claim 1.