Patent Grant
10655647
The present invention relates to a hydraulic drive system for a construction machine such as a hydraulic excavator, and in particular, to a hydraulic drive system for a construction machine that performs the load sensing control on the delivery flow rate of a hydraulic pump so that the delivery pressure of the hydraulic pump becomes higher than the maximum load pressure of a plurality of actuators by a target differential pressure.
Hydraulic drive systems for construction machines such as hydraulic excavators include those controlling the delivery flow rate of the hydraulic pump (main pump) so that the delivery pressure of the hydraulic pump becomes higher than the maximum load pressure of a plurality of actuators by a target differential pressure. This control is called “load sensing control”. In such a hydraulic drive system performing the load sensing control, the differential pressure across each of a plurality of flow control valves is kept at a prescribed differential pressure by use of a pressure compensating valve to make it possible during the combined operation (driving two or more actuators at the same time) to supply the hydraulic fluid to the actuators according to a ratio corresponding to the opening areas of the flow control valves irrespective of the magnitude of the load pressure of each actuator.
In such hydraulic drive systems performing the load sensing control, each pressure compensating valve is generally configured to fully close when the spool moving in the direction of decreasing the opening area reaches the stroke end, as described in Patent Literature 1, for example.
Meanwhile, Patent Literature 2 describes a hydraulic drive system that is configured so that each pressure compensating valve does not fully close even when the spool moving in the direction of decreasing the opening area reaches the stroke end.
Patent Literature 1: JP, A 2007-24103
Patent Literature 2: JP, A 7-76861
However, the conventional technologies described above involve the following problems:
As mentioned above, in the conventional hydraulic drive systems performing the load sensing control (such as the system described in the Patent Literature 1), the differential pressure across each of the flow control valves is kept at a prescribed differential pressure by use of a pressure compensating valve, making it possible during the combined operation (driving two or more actuators at the same time) to supply the hydraulic fluid to the actuators according to the ratio corresponding to the opening areas of the flow control valves irrespective of the load pressures.
However, since the delivery flow rate of the hydraulic pump has a certain upper limit (available maximum delivery flow rate), a state in which the delivery flow rate of the hydraulic pump is insufficient (hereinafter referred to as “saturation”) occurs when the hydraulic pump reaches the available maximum delivery flow rate during the combined operation driving two or more actuators at the same time.
In the hydraulic drive system described in the Patent Literature 1, differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators (hereinafter referred to as “load sensing differential pressure”) is lead to the pressure receiving part of each pressure compensating valve (for operating the valve in the direction of increasing the opening area) as a target compensation differential pressure. By setting the target compensation differential pressures of the pressure compensating valves at the same value equivalent to the load sensing differential pressure, the differential pressures across the flow control valves are kept at the load sensing differential pressure. With this configuration, when the saturation occurs during the combined operation (driving two or more actuators at the same time), the load sensing differential pressure also drops according to the degree of the saturation and the target compensation differential pressures of the pressure compensating valves (i.e., the differential pressures across the flow control valves) decrease uniformly. Consequently, the delivery flow rate of the hydraulic pump can be redistributed among the actuators according to the ratio among the demanded flow rates of the actuators.
However, in cases where the pressure compensating valves are configured to fully close at the stroke end in the direction of decreasing the opening area as in the hydraulic drive system of the Patent Literature 1, if the saturation occurs during combined operation with a great load pressure difference between two actuators, the pressure compensating valve on the low load pressure side can be restricted extremely or closed, by which the actuator on the low load side can be decelerated or stopped.
In the hydraulic drive system described in the Patent Literature 2, the pressure compensating valves are configured not to fully close at the stroke end in the direction of decreasing the opening area. Thus, the pressure compensating valve on the low load side is never restricted extremely or closed even when the saturation occurs during the aforementioned type of combined operation. Consequently, the deceleration/stoppage of the actuator on the low load side can be prevented.
Nevertheless, the hydraulic drive system of the Patent Literature 2 has the following problem: When the saturation occurs during combined operation in which the load pressure difference between two actuators becomes even greater, most of the delivery flow of the main pump is consumed by the actuator on the low load pressure side and this can cause stoppage of the actuator on the high load pressure side.
For example, when a non-travel actuator (e.g., the hydraulic cylinder for the boom, the arm or the bucket) is driven during the traveling of the construction machine, especially in a condition in which the travel load pressure tends to rise (e.g., ascending slope), the entire delivery flow from the hydraulic pump flows into actuators at lower load pressures (e.g., the boom cylinder, the arm cylinder and the bucket cylinder) than the travel motors, by which the traveling of the construction machine can be stopped.
Further, in combined operation of the traveling and the blade, quick operation on the blade during the traveling causes an instantaneous flow of the hydraulic fluid into the blade cylinder, which leads to deceleration/stoppage of the traveling and deterioration in the operational feel.
Besides the travel motors, the reserve actuator for an attachment (e.g., crusher) used in replacement with the bucket causes similar problems since the reserve actuator tends to rise to a high load pressure and the great load pressure difference occurs often in the combined operation with other actuators (e.g., the hydraulic cylinders for the boom, the arm and the bucket).
It is therefore the primary object of the present invention to provide a hydraulic drive system for a construction machine capable of achieving excellent operability in the combined operation by preventing the deceleration/stoppage of the actuator on the low load pressure side (by preventing the full closure of the pressure compensating valve on the low load pressure side) while also preventing the deceleration/stoppage of the high load pressure actuator (by securing a necessary amount of hydraulic fluid for the high load pressure actuator) when the saturation occurs in a hydraulic drive system performing the load sensing control due to the combined operation with a great load pressure difference between two actuators.
In this DESCRIPTION, a term “specific actuator” is used to mean an actuator whose load pressure rises to a high level (e.g., the travel motors and the reserve actuator for the crusher or the like) and which can stop due to the consumption of most of the delivery flow rate of the main pump by other actuators on the low load pressure side when the saturation occurs in a hydraulic drive system comprising pressure compensating valves of the type in which full closing of the valves is not attained at the stroke end in the direction of decreasing the opening area like the system described in the Patent Literature 2 due to combined operation with a great load pressure difference.
To achieve the above object, the present invention provides a hydraulic drive system for a construction machine, comprising: a variable displacement type hydraulic pump; a plurality of actuators which are driven by hydraulic fluid delivered from the hydraulic pump; a plurality of flow control valves which control flow rates of the hydraulic fluid supplied from the hydraulic pump to the actuators; a plurality of operating devices provided corresponding to the actuators and including remote control valves for generating operation pilot pressures for driving the flow control valves; a plurality of pressure compensating valves which respectively control differential pressures across the flow control valves; and a pump control system which performs load sensing control on displacement of the hydraulic pump so that delivery pressure of the hydraulic pump becomes higher than maximum load pressure of the actuators by a target differential pressure. The pressure compensating valves are of the type in which full closing of the valves is not attained at the stroke end in the direction of decreasing the opening area. The hydraulic drive system comprises a pilot primary pressure circuit which supplies pilot primary pressure, as pressure of a pilot hydraulic pressure source, to the remote control valves of the operating devices. The pilot primary pressure circuit includes a first circuit which supplies the pilot primary pressure to the remote control valves of one or more specific operating devices among the plurality of operating devices corresponding to one or more specific actuators and a second circuit which supplies the pilot primary pressure to the remote control valves of operating devices other than the specific operating devices. When the specific operating devices are not operated, the second circuit supplies the pilot primary pressure directly to the remote control valves of the operating devices other than the specific operating devices. When the specific operating devices are operated, the second circuit reduces the pilot primary pressure and supplies the reduced pilot primary pressure to the remote control valves of the operating devices other than the specific operating devices.
In the hydraulic drive system configured as above, the pressure compensating valves are of the type in which full closing of the valves is not attained at the stroke end in the direction of decreasing the opening area. Therefore, even when the saturation occurs due to the combined operation with a great load pressure difference between two actuators, the full closure of the pressure compensating valve on the low load pressure side is prevented, by which the deceleration/stoppage of the actuator on the low load pressure side can be prevented.
Further, the second circuit supplies the pilot primary pressure directly to the remote control valves of the operating devices other than the specific operating devices when the specific operating devices are not operated, while reducing the pilot primary pressure and supplying the reduced pilot primary pressure to the remote control valves of the operating devices other than the specific operating devices when the specific operating devices are operated. Therefore, the inflow of the hydraulic fluid into the actuators corresponding to the operating devices other than the specific operating devices is suppressed. Consequently, even when the saturation occurs during combined operation in which the specific actuator is on the high load pressure side and the load pressure difference is great, the necessary amount of hydraulic fluid for the specific actuator (high load pressure actuator) is secured, the deceleration/stoppage of the specific actuator is prevented, and excellent operability in the combined operation is achieved.
In the present invention, the second circuit can be implemented in various configurations.
For example, the second circuit may include: a third circuit which directly supplies the pilot primary pressure; a fourth circuit which reduces the pilot primary pressure and supplies the reduced pilot primary pressure; and a selector valve which makes a selection from pressure of the third circuit and pressure of the fourth circuit and supplies the selected pressure to the remote control valves of the operating devices other than the specific operating devices.
In this case, the fourth circuit may include a pressure reducing valve which reduces the pilot primary pressure. The fourth circuit may also be configured to include a restrictor circuit which reduces the pilot primary pressure.
The second circuit may also be configured to include: a fifth circuit having a pilot-operated pressure reducing valve and leading the pilot primary pressure directly to the remote control valves of the operating devices other than the specific operating devices when pilot pressure lead to the pilot-operated pressure reducing valve is at a first pressure, while reducing the pilot primary pressure and leading the reduced pilot primary pressure to the remote control valves of the operating devices other than the specific operating devices when the pilot pressure lead to the pilot-operated pressure reducing valve is switched to a second pressure; and a sixth circuit having a selector valve which switches the pilot pressure lead to the pilot-operated pressure reducing valve between the first pressure and the second pressure.
Preferably, the hydraulic drive system further comprises an operation detection device which detects operation of the specific operating devices corresponding to the specific actuators. When the operation detection device detects no operation of the specific operating devices, the second circuit supplies the pilot primary pressure directly to the remote control valves of the operating devices other than the specific operating devices. When the operation detection device detects the operation of the specific operating devices, the second circuit reduces the pilot primary pressure and supplies the reduced pilot primary pressure to the remote control valves of the operating devices other than the specific operating devices.
The hydraulic drive system may further comprise shuttle valves which detect the operation pilot pressures generated by the remote control valves of the specific operating devices corresponding to the specific actuators and output the detected operation pilot pressures as hydraulic signals as the operation detection device. In this case, the selector valve is a hydraulic selector valve which is switched by the hydraulic signals.
Alternatively, the hydraulic drive system may further comprise a pressure sensor which outputs an electric signal by detecting the operation pilot pressures generated by the remote control valves of the specific operating devices corresponding to the specific actuators as the operation detection device. In this case, the selector valve is a solenoid selector valve which operates according to the electric signal.
The hydraulic drive system may further comprise a manual selection device which can be switched between a first position and a second position. When the manual selection device is at the first position, the second circuit enables the function of reducing the pilot primary pressure when the specific operating devices are operated. When the manual selection device is switched to the second position, the second circuit disables the function of reducing the pilot primary pressure when the specific operating devices are operated.
According to the present invention, when the saturation occurs in a hydraulic drive system performing the load sensing control due to the combined operation with a great load pressure difference between two actuators, the deceleration/stoppage of the actuator on the low load pressure side is prevented by preventing the full closure of the pressure compensating valve on the low load pressure side, while also preventing the deceleration/stoppage of the high load pressure actuator by securing a necessary amount of hydraulic fluid for the high load pressure actuator. Consequently, excellent operability in the combined operation is achieved.
Referring now to the drawings, a description will be given in detail of preferred embodiments of the present invention.
<Hydraulic Excavator>
Referring to
(Basic Configuration)
First, the basic configuration of the hydraulic drive system according to this embodiment will be described below.
The hydraulic drive system in this embodiment comprises an engine 1, a main hydraulic pump (hereinafter referred to as a “main pump”) 2 which is driven by the engine 1, a pilot pump 3 which is driven by the engine 1 in conjunction with the main pump 2, a plurality of actuators 5, 6, 7, 8, 9, 10, 11 and 12 which are driven by the hydraulic fluid delivered from the main pump 2 (i.e., the left and right travel motors 5 and 6, the swing motor 7, the blade cylinder 8, the swing cylinder 9, the boom cylinder 10, the arm cylinder 11 and the bucket cylinder 12), and a control valve 4. The hydraulic excavator in this embodiment is a mini-excavator, for example.
The control valve 4 includes a plurality of valve sections 13, 14, 15, 16, 17, 18, 19 and 20, a plurality of shuttle valves 22a, 22b, 22c, 22d, 22e, 22f and 22g, a main relief valve 23, a differential pressure reducing valve 24, and an unload valve 25. The valve sections 13, 14, 15, 16, 17, 18, 19 and 20 are connected to a supply line 2a of the main pump 2. Each valve section 13, 14, 15, 16, 17, 18, 19, 20 controls the direction and the flow rate of the hydraulic fluid supplied from the main pump 2 to each actuator. The shuttle valves 22a, 22b, 22c, 22d, 22e, 22f and 22g select the highest load pressure PLmax from the load pressures of the actuators 5, 6, 7, 8, 9, 10, 11 and 12 (hereinafter referred to as “the maximum load pressure PLmax”) and output the maximum load pressure PLmax to a signal hydraulic line 21. The main relief valve 23 is connected to an in-valve supply line 4a which is connected to the supply line 2a of the main pump 2 and limits the maximum delivery pressure of the main pump 2 (maximum pump pressure). The differential pressure reducing valve 24 is connected to a pilot hydraulic pressure source 33 (explained later), receives the pressures in the supply line 4a and the signal hydraulic line 21 as signal pressures, and outputs the differential pressure PLS between the delivery pressure (pump pressure) Pd of the main pump 2 and the maximum load pressure PLmax as an absolute pressure. The unload valve 25 is connected to the in-valve supply line 4a, receives the pressures in the supply line 4a and the signal hydraulic line 21 as signal pressures, and keeps the differential pressure PLS within a constant value that is set by a spring 25a by returning part of the delivery flow of the main pump 2 to a tank T when the differential pressure PLS between the pump pressure Pd and the maximum load pressure PLmax exceeds the constant value set by the spring 25a. The outlet side of the unload valve 25 and the outlet side of the main relief valve 23 are connected to an in-valve tank line 29 and connected to the tank T via the line 29.
The valve section 13 is formed of a flow control valve 26a and a pressure compensating valve 27a. The valve section 14 is formed of a flow control valve 26b and a pressure compensating valve 27b. The valve section 15 is formed of a flow control valve 26c and a pressure compensating valve 27c. The valve section 16 is formed of a flow control valve 26d and a pressure compensating valve 27d. The valve section 17 is formed of a flow control valve 26e and a pressure compensating valve 27e. The valve section 18 is formed of a flow control valve 26f and a pressure compensating valve 27f. The valve section 19 is formed of a flow control valve 26g and a pressure compensating valve 27g. The valve section 20 is formed of a flow control valve 26h and a pressure compensating valve 27h.
Each flow control valve 26a-26h controls the direction and the flow rate of the hydraulic fluid supplied from the main pump 2 to each actuator 5-12. Each pressure compensating valve 27a-27h controls the differential pressure across each flow control valve 26a-26h.
Each pressure compensating valve 27a-27h has a valve-opening pressure receiving part 28a, 28b, 28c, 28d, 28e, 28f, 28g, 28h for setting a target differential pressure. The output pressure of the differential pressure reducing valve 24 is lead to the pressure receiving parts 28a-28h. A target compensation differential pressure is set to the pressure receiving parts 28a-28h according to the absolute pressure of the differential pressure PLS between the hydraulic pump pressure Pd and the maximum load pressure PLmax (hereinafter referred to as “absolute pressure PLS”). By controlling the differential pressures across the flow control valves 26a-26h at the same value (PLS) as above, the pressure compensating valves 27a-27h carry out control so that the differential pressures across the flow control valves 26a-26h equal the differential pressure PLS between the hydraulic pump pressure Pd and the maximum load pressure PLmax. As a result, in the combined operation in which two or more actuators are driven at the same time, the delivery flow rate (delivery flow) of the main pump 2 can be properly distributed according to the opening area ratio among the flow control valves 26a-26h irrespective of the magnitude of the load pressure of each actuator 5-12, by which excellent operability in the combined operation can be secured. Further, in the saturation state in which the delivery flow rate of the main pump 2 is less than the demanded flow rate, the differential pressure PLS drops according to the degree of the supply deficiency. Accordingly, the differential pressures across the flow control valves 26a-26h (controlled by the pressure compensating valves 27a-27h) drop at the same ratio and the flow rates through the flow control valves 26a-26h decrease at the same ratio. Therefore, also in this case, the delivery flow rate (delivery flow) of the main pump 2 can be properly distributed according to the opening area ratio among the flow control valves 26a-26h and excellent operability in the combined operation can be secured.
As is clear from the symbol representation in
The hydraulic drive system further comprises an engine revolution speed detection valve 30, a pilot hydraulic pressure source 33, and operating devices 34a, 34b, 34c, 34d, 34e, 34f, 34g and 34h. The engine revolution speed detection valve 30 is connected to a supply line 3a of the pilot pump 3 and outputs absolute pressure corresponding to the delivery flow rate of the pilot pump 3. The pilot hydraulic pressure source 33 is connected to the downstream side of the engine revolution speed detection valve 30. The pilot hydraulic pressure source 33 has a pilot relief valve 32 which maintains the pressure in a pilot line 31 at a constant level. The operating devices 34a, 34b, 34c, 34d, 34e, 34f, 34g and 34h are connected to the pilot line 31. The operating devices 34a, 34b, 34c, 34d, 34e, 34f, 34g and 34h are respectively equipped with remote control valves 34a-2, 34b-2, 34c-2, 34d-2, 34e-2, 34f-2, 34g-2 and 34h-2 (see
The engine revolution speed detection valve 30 includes a restrictor element (fixed restrictor part) 30f which is arranged in a hydraulic line connecting the supply line 3a of the pilot pump 3 to the pilot line 31, a flow rate detection valve 30a which is connected in parallel with the restrictor element 30f, and a differential pressure reducing valve 30b. The input side of the flow rate detection valve 30a is connected to the supply line 3a of the pilot pump 3, while the output side of the flow rate detection valve 30a is connected to the pilot line 31. The flow rate detection valve 30a has a variable restrictor part 30c which increases its opening area with the increase in the flow rate. The hydraulic fluid delivered from the pilot pump 3 can flow into the pilot line 31 through either the restrictor element 30f or the variable restrictor part 30c of the flow rate detection valve 30a. In this case, differential pressure that increases with the increase in the flow rate occurs across the restrictor element 30f and the variable restrictor part 30c of the flow rate detection valve 30a. The differential pressure reducing valve 30b outputs the differential pressure as absolute pressure Pa. Since the delivery flow rate of the pilot pump 3 changes according to the revolution speed of the engine 1, the delivery flow rate of the pilot pump 3 and the revolution speed of the engine 1 can be measured by detecting the differential pressure across the restrictor element 30f and the variable restrictor part 30c. The variable restrictor part 30c is configured so as to reduce the degree of increase of the differential pressure with the increase in the flow rate, by increasing the opening area with the increase in the flow rate (i.e., with the increase in the differential pressure).
The main pump 2 is a variable displacement type hydraulic pump. The main pump 2 is equipped with a pump control system 35 for controlling the tilting angle (displacement) of the main pump 2. The pump control system 35 includes a pump torque control unit 35A and an LS control unit 35B.
The pump torque control unit 35A includes a torque control tilting actuator 35a. The torque control tilting actuator 35a limits the input torque of the main pump 2 so as not to exceed preset maximum torque, by driving the swash plate 2s (variable displacement member) of the main pump 2 to reduce its tilting angle (displacement) when the delivery pressure of the main pump 2 becomes high. By this operation, the power consumption of the main pump 2 is limited and the stoppage of the engine 1 due to the overload (engine stall) is prevented.
The LS control unit 35B includes an LS control valve 35b and an LS control tilting actuator 35c.
The LS control valve 35b has pressure receiving parts 35d and 35e opposing each other. To the pressure receiving part 35d, the absolute pressure Pa generated by the differential pressure reducing valve 30b of the engine revolution speed detection valve 30 is lead via a hydraulic line 40 as the target differential pressure of the load sensing control (target LS differential pressure). To the pressure receiving part 35e, the absolute pressure PLS (i.e., the differential pressure PLS between the delivery pressure Pd of the main pump 2 and the maximum load pressure PLmax) generated by the differential pressure reducing valve 24 is lead as feedback differential pressure. When the absolute pressure PLS exceeds the absolute pressure Pa (PLS>Pa), the LS control valve 35b leads the pressure of the pilot hydraulic pressure source 33 to the LS control tilting actuator 35c. When the absolute pressure PLS falls below the absolute pressure Pa (PLS<Pa), the LS control valve 35b connects the LS control tilting actuator 35c to the tank T. When the pressure of the pilot hydraulic pressure source 33 is lead thereto, the LS control tilting actuator 35c drives the swash plate 2s of the main pump 2 to decrease the tilting angle of the main pump 2. When connected to the tank T, the LS control tilting actuator 35c drives the swash plate 2s of the main pump 2 to increase the tilting angle of the main pump 2. By this operation, the tilting angle (displacement) of the main pump 2 is controlled so that the delivery pressure Pd of the main pump 2 becomes higher than the maximum load pressure PLmax by the absolute pressure Pa (target differential pressure).
Incidentally, since the absolute pressure Pa is a value changing according to the engine revolution speed, actuator speed control according to the engine revolution speed becomes possible by using the absolute pressure Pa as the target differential pressure of the load sensing control and setting the target compensation differential pressure of the pressure compensating valves 27a-27h by using the absolute pressure PLS of the differential pressure between the delivery pressure Pd of the main pump 2 and the maximum load pressure PLmax.
The preset pressure of the spring 25a of the unload valve 25 has been set to be slightly higher than the absolute pressure Pa (the target differential pressure of the load sensing control) that is generated by the differential pressure reducing valve 30b of the engine revolution speed detection valve 30 when the engine 1 is at its rated maximum revolution speed.
The operating device 34a includes a control lever 34a-1 and a remote control valve 34a-2. The remote control valve 34a-2 has a pair of pressure reducing valves PVa and PVb. When the control lever 34a-1 is operated rightward in
The operating devices 34b-34h are also configured in the same way. Specifically, each operating device 34b-34h includes a control lever 34b-1, 34c-1, 34d-1, 34e-1, 34f-1, 34g-1, 34h-1 and a remote control valve 34b-2, 34c-2, 34d-2, 34e-2, 34f-2, 34g-2, 34h-2. When the control lever 34b-1, 34c-1, 34d-1, 34e-1, 34f-1, 34g-1, 34h-1 is operated rightward in
(Characteristic Configuration)
Next, a configuration that is characteristic of the hydraulic drive system according to this embodiment will be described below.
The hydraulic drive system according to this embodiment comprises, as its characteristic configuration, a pilot primary pressure circuit 40 which supplies the pilot primary pressure (i.e., the pressure of the pilot hydraulic pressure source 33) to the remote control valves 34a-2, 34b-2, 34c-2, 34d-2, 34e-2, 34f-2, 34g-2 and 34h-2 of the operating devices 34a, 34b, 34c, 34d, 34e, 34f, 34g and 34h. The pilot primary pressure circuit 40 includes a first circuit 41 which supplies the pilot primary pressure to the remote control valves 34a-2 and 34b-2 of the travel operating devices 34a and 34b and a second circuit 42 which supplies the pilot primary pressure to the remote control valves 34c-2-34h-2 of the operating devices 34c-34h other than the travel operating devices (hereinafter referred to simply as “non-travel operating devices).
The second circuit 42 is configured as below. When the travel operating devices 34a and 34b are not operated, the second circuit 42 supplies the pilot primary pressure directly to the remote control valves 34c-2-34h-2 of the non-travel operating devices 34c-34h. When the travel operating devices 34a and 34b are operated, the second circuit 42 reduces the pilot primary pressure and supplies the reduced pilot primary pressure to the remote control valves 34c-2-34h-2 of the non-travel operating devices 34c-34h.
The travel motors 5 and 6 are specific actuators, and the travel operating devices 34a and 34b are specific operating devices corresponding to the specific actuators (the travel motors 5 and 6) among the operating devices 34a-34h. In this DESCRIPTION, the term “specific actuator” means an actuator of the following type: In combined operation in which the specific actuator and another actuator (latter actuator) are driven at the same time, the latter actuator stays on the low load pressure side and the load pressure of the specific actuator rises to such an extent that the pressure compensating valve of the latter actuator (actuator on the low load side) operates to a position close to the stroke end.
The hydraulic drive system according to this embodiment further comprises an operation detection device 43 which detects the operation of the travel operating devices 34a and 34b. The operation detection device 43 includes shuttle valves 48a, 48b and 48c for detecting the operation pilot pressures generated by the remote control valves 34a-2 and 34b-2 of the travel operating devices 34a and 34b (travel operation pilot pressures) and outputting the detected travel operation pilot pressures as a hydraulic signal. The second circuit 42 includes a third circuit 44 for directly supplying the pilot primary pressure, a fourth circuit 45 for reducing the pilot primary pressure and supplying the reduced pilot primary pressure, and a selector valve 46 for making a selection from (switching between) the pressure of the third circuit 44 and the pressure of the fourth circuit 45 and supplying the selected pressure to the remote control valves 34c-2-34h-2 of the non-travel operating devices 34c-34h. The fourth circuit 45 includes a pressure reducing valve 47 for reducing the pilot primary pressure. The selector valve 46 includes a pilot pressure receiving part 46a to which the hydraulic signal from the shuttle valves 48a, 48b and 48c is lead via a hydraulic line 48d.
When the control levers 34a-1 and 34b-1 of the travel operating devices 34a and 34b are not operated and no travel operation pilot pressure is generated, the selector valve 46 is situated at a first position (rightward in
When the control levers 34a-1 and 34b-1 of the travel operating devices 34a and 34b are not operated, no travel operation pilot pressure is generated, and thus the selector valve 46 is situated at the first position (rightward in
In contrast, when the control levers 34a-1 and 34b-1 of the travel operating devices 34a and 34b are operated, the travel operation pilot pressure is generated and the selector valve 46 is switched to the second position (leftward in
As a result, the spool stroke of the non-travel flow control valve 26c-26h increases from 0 only to an intermediate stroke Str corresponding to the operation pilot pressure Ppa, that is, the maximum stroke of the non-travel flow control valve 26c-26h is limited to the intermediate stroke Str (characteristic B2 shown in
(Operation of Basic Configuration)
First, the operation of the basic configuration of the hydraulic drive system according to this embodiment will be explained.
<When all Control Levers are at Neutral Positions>
When the control levers 34a-1-34h-1 of all the operating devices 34a-34h are at their neutral positions, all the flow control valves 26a-26h are at their neutral positions and no hydraulic fluid is supplied to the actuators 5-12. When the flow control valves 26a-26h are at the neutral positions, the maximum load pressure PLmax detected by the shuttle valves 22a-22g equals the tank pressure.
The hydraulic fluid delivered from the main pump 2 is supplied to the supply lines 2a and 4a and increases the pressure in the supply lines 2a and 4a. The supply line 4a is equipped with the unload valve 25. When the pressure in the supply line 2a becomes the preset pressure of the spring 25a or more higher than the maximum load pressure PLmax (in this case, the tank pressure), the unload valve 25 opens, returns the hydraulic fluid in the supply line 2a to the tank, and thereby limits the increase in the pressure in the supply line 2a. By the above operation, the delivery pressure of the main pump 2 is controlled to be at a minimum pressure Pmin.
The differential pressure reducing valve 24 is outputting the differential pressure PLS between the delivery pressure Pd of the main pump 2 and the maximum load pressure PLmax (the tank pressure in this case) as the absolute pressure. The LS control valve 35b of the LS control unit 35B of the main pump 2 is supplied with the output pressure of the engine revolution speed detection valve 30 and the output pressure of the differential pressure reducing valve 24. When the delivery pressure of the main pump 2 rises and the output pressure of the differential pressure reducing valve 24 exceeds the output pressure of the engine revolution speed detection valve 30, the LS control valve 35b is switched to the rightward position in
<When Control Lever is Operated>
When the control lever for any driven member (assumed here to be the control lever 34f-1 of the operating device 34f for the boom) is operated, the flow control valve 26f for the boom is switched, the hydraulic fluid is supplied to the boom cylinder 10, and the boom cylinder 10 is driven.
The flow rate through the flow control valve 26f is determined by the opening area of the meter-in restrictor of the flow control valve 26f and the differential pressure across the meter-in restrictor. The differential pressure across the meter-in restrictor is controlled by the pressure compensating valve 27f to be equal to the output pressure of the differential pressure reducing valve 24. Therefore, the flow rate through the flow control valve 26f (i.e., driving speed of the boom cylinder 10) is controlled according to the operation amount of the control lever.
Meanwhile, the load pressure of the boom cylinder 10 is detected by the shuttle valves 22a-22g as the maximum load pressure and is transmitted to the differential pressure reducing valve 24 and the unload valve 25.
When the load pressure of the boom cylinder 10 is lead to the unload valve 25 as the maximum load pressure, the cracking pressure of the unload valve 25 (at which the unload valve 25 starts opening) rises accordingly. When the pressure in the supply line 2a transiently becomes the preset pressure of the spring 25a or more higher than the maximum load pressure, the unload valve 25 opens and thereby returns the hydraulic fluid in the supply line 4a to the tank. By this operation, the pressure in the supply lines 2a and 4a is prevented from exceeding the maximum load pressure PLmax by the preset pressure of the spring 25a or more (i.e., prevented from exceeding the sum of the maximum load pressure PLmax and the preset pressure of the spring 25a).
When the boom cylinder 10 starts moving, the pressure in the supply lines 2a and 4a drops temporarily. At this point, the output pressure of the differential pressure reducing valve 24 drops because the difference between the pressure in the supply line 2a and the load pressure of the boom cylinder 10 is outputted as the output pressure of the differential pressure reducing valve 24.
The LS control valve 35b of the LS control unit 35B of the main pump 2 is supplied with the output pressure of the engine revolution speed detection valve 30 and the output pressure of the differential pressure reducing valve 24. When the output pressure of the differential pressure reducing valve 24 falls below the output pressure of the engine revolution speed detection valve 30, the LS control valve 35b is switched to the leftward position in
When the control levers of operating devices for two or more driven members (assumed here to be the control lever 34f-1 of the operating device 34f for the boom and the control lever 34g-1 of the operating device 34g for the arm) are operated, the flow control valves 26f and 26g are switched and the hydraulic fluid is supplied to the boom cylinder 10 and the arm cylinder 11 to drive the boom cylinder 10 and the arm cylinder 11.
The higher one of the load pressures of the boom cylinder 10 and the arm cylinder 11 is detected by the shuttle valves 22a-22g as the maximum load pressure PLmax and is transmitted to the differential pressure reducing valve 24 and the unload valve 25.
The operation when the maximum load pressure PLmax detected by the shuttle valves 22a-22g is lead to the unload valve 25 is equivalent to that in the case where the boom cylinder 10 is driven alone. The cracking pressure of the unload valve 25 rises according to the rise in the maximum load pressure PLmax, and the pressure in the supply lines 2a and 4a is prevented from exceeding the maximum load pressure PLmax by the preset pressure of the spring 25a or more (i.e., prevented from exceeding the sum of the maximum load pressure PLmax and the preset pressure of the spring 25a).
The LS control valve 35b of the LS control unit 35B of the main pump 2 is supplied with the output pressure of the engine revolution speed detection valve 30 and the output pressure of the differential pressure reducing valve 24. Similarly to the case where the boom cylinder 10 is driven alone, the delivery pressure of the main pump 2 (the pressure in the supply lines 2a and 4a) is controlled to be the output pressure of the engine revolution speed detection valve 30 (target differential pressure) higher than the maximum load pressure PLmax (i.e., to be higher than the maximum load pressure PLmax by the output pressure of the engine revolution speed detection valve 30 (target differential pressure)) and the so-called load sensing control for supplying the flow rate (flow) demanded by the flow control valves 26f and 26g to the boom cylinder 10 and the arm cylinder 11 is carried out.
The output pressure of the differential pressure reducing valve 24 is lead to the pressure compensating valves 27a-27h as the target compensation differential pressure. The pressure compensating valves 27f and 27g perform control so that the differential pressure across the flow control valve 26f and the differential pressure across the flow control valve 26g equal the differential pressure between the delivery pressure of the main pump 2 and the maximum load pressure PLmax. This makes it possible to supply the hydraulic fluid to the boom cylinder 10 and the arm cylinder 11 according to the ratio between the opening areas of the meter-in restrictor parts of the flow control valves 26f and 26g irrespective of the magnitude of the load pressures of the boom cylinder 10 and the arm cylinder 11.
In this case, when the delivery flow rate of the main pump 2 falls below the flow rate demanded by the flow control valves 26f and 26g (saturation state), the output pressure of the differential pressure reducing valve 24 (the differential pressure between the delivery pressure of the main pump 2 and the maximum load pressure PLmax) drops according to the degree of the saturation. Since the target compensation differential pressure of the pressure compensating valves 27a-27h also drops accordingly, the delivery flow rate (delivery flow) of the main pump 2 can be redistributed properly at the ratio between the flow rates demanded by the flow control valves 26f and 26g.
Further, since the pressure compensating valves 27a-27h are configured not to fully close at the stroke end in the direction of decreasing the opening area (leftward in
<When Engine Revolution Speed is Reduced>
The operation described above is the operation at times when the engine 1 is rotating at its maximum rated revolution speed. When the revolution speed of the engine 1 is reduced to a lower speed, the output pressure of the engine revolution speed detection valve 30 drops correspondingly and thus the target differential pressure of the LS control valve 35b of the LS control unit 35B also drops similarly. Further, the target compensation differential pressure of the pressure compensating valves 27a-27h also drops similarly as a result of the load sensing control. Thus, with the reduction in the engine revolution speed, the delivery flow rate of the main pump 2 and the demanded flow rate of the flow control valves 26a-26h decrease. Consequently, the driving speeds of the actuators 5-12 are prevented from increasing too much and the fine-tuning operability when the engine revolution speed is reduced can be improved.
(Operation of Characteristic Configuration)
Next, the operation of the characteristic configuration of the hydraulic drive system according to this embodiment will be explained below.
Also when the control levers 34a-1 and 34b-1 of the travel operating devices 34a and 34b are operated, the flow control valves 26a and 26b are switched and the hydraulic fluid is supplied to the travel motors 5 and 6 similarly to the above-described case of combined operation. Meanwhile, the delivery flow rate of the main pump 2 is controlled by the load sensing control, the flow rate (flow) demanded by the flow control valves 26a and 26b is supplied to the travel motors 5 and 6, and the hydraulic excavator travels.
When the control lever for any one of the boom, the arm and the bucket (assumed here to be the control lever 34g-1 of the operating device 34g for the arm) is operated during the traveling of the hydraulic excavator in order to change the posture of the front work implement, the flow control valve 26g is switched, the hydraulic fluid is supplied also to the arm cylinder 11, and the arm cylinder 11 is driven.
At this point, the travel operation pilot pressure has been generated due to the operator's operation on the control levers 34a-1 and 34b-1 of the travel operating devices 34a and 34b, the selector valve 46 has been switched to the second position (leftward in
Incidentally, in the conventional configuration in which the pressure compensating valves are of the type in which full closing of the valves is not attained at the stroke end in the direction of decreasing the opening area, when another driven member (e.g., the boom, arm or bucket) is operated during the traveling (especially in a condition in which the travel load pressure tends to rise (e.g., ascending slope)), the pressure compensating valve of the low-load actuator (e.g., the boom cylinder, arm cylinder or bucket cylinder at lower load pressure than the travel motors) is still open even after reaching the stroke end. Thus, there are cases where all the delivery flow rate (delivery flow) of the hydraulic pump flows to the low-load actuator and the traveling of the hydraulic excavator is decelerated or stopped.
In contrast, in this embodiment, even when the control lever 34g-1 of the arm operating device 34g is operated to the limit, the meter-in opening area of the flow control valve 26g is limited to Astr and the demanded flow rate of the flow control valve 26g is restricted as explained above. Accordingly, the flow rate of the hydraulic fluid flowing into the low load pressure actuator decreases. Consequently, a necessary amount of hydraulic fluid for the travel motors 5 and 6 is secured, the deceleration/stoppage of the traveling is prevented, and excellent operability in the combined operation is achieved.
Also when the control lever 34d-1 of the operating device 34d for the blade is operated quickly during the traveling, in the conventional configuration in which the pressure compensating valves are of the type in which full closing of the valves is not attained at the stroke end in the direction of decreasing the opening area, the hydraulic fluid instantaneously flows into the blade cylinder 8 and the traveling of the hydraulic excavator is decelerated or stopped. The deceleration/stoppage of the traveling causes a cenesthesic shock and deteriorates the operational feel. In contrast, in this embodiment, the demanded flow rate of the flow control valve 26d for the blade is restricted similarly to the above case where the control lever of the operating device for the boom, arm or bucket is operated during the traveling in order to change the posture of the front work implement. Consequently, the necessary amount of hydraulic fluid for the travel motors 5 and 6 is secured, the deceleration/stoppage of the traveling is prevented, and the operational feel is improved.
(Effect)
As described above, according to this embodiment, when the saturation occurs during combined operation with a great load pressure difference between two actuators, the full closure of the pressure compensating valve on the low load pressure side is prevented, by which the deceleration/stoppage of the actuator on the low load pressure side is prevented. Further, in the travel combined operation including the driving of the travel motors 5 and 6 (specific actuators), the operation pilot pressures of the non-travel actuators are restricted. Consequently, the inflow of the hydraulic fluid into the non-travel actuators is suppressed, the necessary amount of hydraulic fluid for the travel motors is secured, the deceleration/stoppage of the traveling is prevented, and the operability in the travel combined operation is improved.
Specifically, the hydraulic drive system in this embodiment comprises a pilot primary pressure circuit 40A. A second circuit 42A of the pilot primary pressure circuit 40A includes a fifth circuit 52 and a sixth circuit 54. The fifth circuit 52 has a pilot-operated pressure reducing valve 51. The sixth circuit 54 has a selector valve 53 which switches the pilot pressure lead to a pilot pressure receiving part 51a of the pilot-operated pressure reducing valve 51 between the pressure of the pilot hydraulic pressure source 33 (first pressure) and the tank pressure (second pressure). When the pilot pressure lead to the pilot pressure receiving part 51a of the pilot-operated pressure reducing valve 51 is the pressure of the pilot hydraulic pressure source 33, the fifth circuit 52 leads the pilot primary pressure directly to the remote control valves 34c-2-34h-2 of the non-travel operating devices 34c-34h. When the pilot pressure lead to the pilot pressure receiving part 51a of the pilot-operated pressure reducing valve 51 is switched to the tank pressure, the fifth circuit 52 reduces the pilot primary pressure and leads the reduced pilot primary pressure to the remote control valves of the non-travel operating devices.
In this embodiment configured as above, when the control levers 34a-1 and 34b-1 of the travel operating devices 34a and 34b are not operated, the pressure of the pilot hydraulic pressure source 33 is lead to the pilot-operated pressure reducing valve 51 via the selector valve 53 and thus the pressure on the outlet side of the pilot-operated pressure reducing valve 51 is not reduced and the pressure of the pilot hydraulic pressure source 33 (pilot primary pressure) is supplied to the remote control valves 34c-2-34h-2 of the non-travel operating devices 34c-34h. Consequently, the spool strokes (meter-in opening areas) of the flow control valves 26c-26h are not restricted and normal operations such as the excavating operation can be carried out.
When the control levers 34a-1 and 34b-1 of the travel operating devices 34a and 34b are operated, the travel operation pilot pressure is lead to a pilot pressure receiving part 53a of the selector valve 53, the selector valve 53 is switched, and the hydraulic fluid which has been lead to the pilot pressure receiving part 51a of the pilot-operated pressure reducing valve 51 is interrupted. Accordingly, the primary pilot pressure which is lead to the remote control valves 34c-2-34h-2 of the non-travel operating devices is reduced by the pilot-operated pressure reducing valve 51, the spool strokes (meter-in opening areas) of the flow control valves 26c-26h are restricted, and their demanded flow rate is restricted. Consequently, the necessary amount of hydraulic fluid for the travel motors 5 and 6 is secured, the stoppage of the traveling is prevented, and excellent operability in the combined operation is achieved.
As above, also in this embodiment, effects similar to those of the first embodiment can be achieved.
Specifically, the hydraulic drive system in this embodiment comprises a pilot primary pressure circuit 40B. A second circuit 42B of the pilot primary pressure circuit 40B includes a third circuit 61 for directly supplying the pilot primary pressure, a fourth circuit 62 for reducing the pilot primary pressure and supplying the reduced pilot primary pressure, and a selector valve 63 for making a selection from (switching between) the pressure of the third circuit 61 and the pressure of the fourth circuit 62 and supplying the selected pressure to the remote control valves of the non-travel operating devices. The fourth circuit 62 includes a restrictor circuit 64 for reducing the pilot primary pressure. The restrictor circuit 64 includes a hydraulic line 64b whose upstream end is connected to the pilot line 31 and downstream end is connected to the tank T via a low-pressure relief valve 64a, two fixed restrictors 64c and 64d which are arranged in the hydraulic line 64b, and a hydraulic line 64e which is connected to a point between the two fixed restrictors 64c and 64d. An intermediate pressure obtained by pressure reduction by the two fixed restrictors 64c and 64d is lead to the hydraulic line 64e.
The pressure of the pilot hydraulic pressure source 33 (pilot primary pressure) is maintained by the fixed restrictor 64c at a normal pressure which is set by the pilot relief valve 32 (see
When the control levers 34a-1 and 34b-1 of the travel operating devices 34a and 34b are operated, the travel operation pilot pressure is lead to a pilot pressure receiving part 63a of the selector valve 63, the selector valve 63 is switched, and the pressure reduced by the fixed restrictors 64c and 64d of the restrictor circuit 64 is lead to the remote control valves 34c-2-34h-2 of the non-travel operating devices. Accordingly, the spool strokes (meter-in opening areas) of the flow control valves 26c-26h are limited and their demanded flow rate is restricted. Consequently, the necessary amount of hydraulic fluid for the travel motors 5 and 6 is secured, the stoppage of the traveling is prevented, and excellent operability in the combined operation is achieved.
As above, also in this embodiment, effects similar to those of the first embodiment can be achieved.
Specifically, the hydraulic drive system in this embodiment comprises a pilot primary pressure circuit 40C. A second circuit 42C of the pilot primary pressure circuit 40C includes a solenoid selector valve 46C and a controller 71 instead of the hydraulic selector valve 46 in the first embodiment. An operation detection device 43C includes a pressure sensor 72 which outputs an electric signal by detecting the operation pilot pressures generated by the remote control valves of the travel operating devices (included in the plurality of operating devices). The electric signal from the pressure sensor 72 is inputted to the controller 71. The controller 71 converts the electric signal into a drive signal for the solenoid selector valve 46C and outputs the drive signal to a solenoid 46b of the solenoid selector valve 46C.
When the control levers 34a-1 and 34b-1 of the travel operating devices (specific operating devices) 34a and 34b are not operated and no drive signal is outputted from the controller 71, the solenoid selector valve 46C is situated at a first position (rightward in
As above, also in this embodiment, effects similar to those of the first embodiment can be achieved.
Incidentally, while this embodiment employs a solenoid selector valve instead of the selector valve 46 shown in
Specifically, the hydraulic drive system in this embodiment further comprises a manual selection device 81 which can be switched between a first position and a second position. The manual selection device 81 is implemented by, for example, a switch that outputs an electric signal corresponding to the switch position. Further, a second circuit 42D of a pilot primary pressure circuit 40D in this embodiment further includes a solenoid selector valve 83 which is arranged in the hydraulic line 48d (leading the hydraulic signal detected by the operation detection device 43 to the pilot pressure receiving part 46a of the selector valve 46) and operates according to the electric signal from the manual selection device (manual switch) 81.
When the manual selection device 81 is at the first position and no electric signal is outputted therefrom, the solenoid selector valve 83 is situated at a first position (rightward in
In this embodiment configured as above, the operator is allowed to freely select whether to use the control for restricting the demanded flow rate for the non-travel actuators according to the present invention or not based on the operator's preference or the type of the work/operation.
The embodiments described above can be modified in various ways within the spirit and scope of the present invention. For example, while a case where the specific actuators are the travel motors has been described in the above embodiments, equivalent effects can be achieved by the present invention even in cases where the specific actuators are actuators other than the travel motors as long as the hydraulic drive system comprises pressure compensating valves of the type in which full closing of the valves is not attained at the stroke end in the direction of decreasing the opening area and the specific actuators are actuators that can stop (due to the consumption of most of the delivery flow rate of the main pump by other actuators on the low load pressure side) when the saturation is caused by combined operation with a great load pressure difference. For example, the load pressure of the reserve actuator for an attachment like the crusher tends to rise to a high level. By employing the present invention while designating the reserve actuator as the specific actuator, it is possible to restrict the demanded flow rate for the other actuators and preferentially supply the hydraulic fluid to the reserve actuator at the times of combined operation with other actuators (boom, arm, bucket, etc.).
While the above embodiments have been described by taking a hydraulic excavator as an example of the construction machine, it is also possible to apply the present invention to other types of construction machines (hydraulic cranes, wheel excavators, etc.) and achieve equivalent effects.
| Number | Date | Country | Kind |
|---|---|---|---|
| 2012-089670 | Apr 2012 | JP | national |
| Number | Name | Date | Kind |
|---|---|---|---|
| 4508013 | Barbagli | Apr 1985 | A |
| 6397591 | Tsuruga | Jun 2002 | B1 |
| 6408622 | Tsuruga | Jun 2002 | B1 |
| 8857169 | Takahashi | Oct 2014 | B2 |
| 8919109 | Takahashi | Dec 2014 | B2 |
| 9080481 | Takebayashi | Jul 2015 | B2 |
| 9200431 | Mori | Dec 2015 | B2 |
| 9518593 | Mori | Dec 2016 | B2 |
| 9702379 | Takahashi | Jul 2017 | B2 |
| 9835180 | Takahashi | Dec 2017 | B2 |
| 9890801 | Takahashi | Feb 2018 | B2 |
| 9963856 | Takahashi | May 2018 | B2 |
| 10100495 | Takahashi | Oct 2018 | B2 |
| 10107311 | Takahashi | Oct 2018 | B2 |
| 10215198 | Takahashi | Feb 2019 | B2 |
| 10280592 | Takahashi | May 2019 | B2 |
| 20030200747 | Matsumoto | Oct 2003 | A1 |
| 20130287601 | Mori | Oct 2013 | A1 |
| 20140069091 | Franzoni et al. | Mar 2014 | A1 |
| 20150240455 | Takebayashi | Aug 2015 | A1 |
| 20160333900 | Takahashi | Nov 2016 | A1 |
| Number | Date | Country |
|---|---|---|
| 04-370402 | Dec 1992 | JP |
| 07-076861 | Mar 1995 | JP |
| 2003-156006 | May 2003 | JP |
| 2007-024103 | Feb 2007 | JP |
| 2009-167618 | Jul 2009 | JP |
| 2010-047983 | Mar 2010 | JP |
| Entry |
|---|
| International Preliminary Report on Patentability received in International Application No. PCT/JP2013/059946 dated Oct. 3, 2014. |
| Number | Date | Country | |
|---|---|---|---|
| 20180156242 A1 | Jun 2018 | US |
| Number | Date | Country | |
|---|---|---|---|
| Parent | 14383150 | US | |
| Child | 15883357 | US |
