Patent Application
20110173964
The present invention relates to hydraulic drive systems for hydraulic excavators or other construction machines and particularly to a hydraulic drive system for performing load sensing control so that the discharge pressure of a hydraulic pump can exceed the maximum load pressure of multiple actuators by a target differential pressure.
An example of such a hydraulic drive system is the one disclosed in Patent Document 1. The hydraulic drive system of Patent Document 1 has a main relief valve and unloading valve connected to a hydraulic fluid supply circuit through which pressurized hydraulic fluid flows from a hydraulic pump (i.e., main pump). The main relief valve is a type of safety valve and starts to operate when the loads on actuators are high and the pressure in the hydraulic fluid supply circuit (i.e., the discharge pressure of the hydraulic pump) reaches a relief pressure of, for example, 25 MPa during operation of flow control valves, whereby the circuit pressure can be prevented from exceeding the relief pressure. The unloading valve operates mainly when the flow control valves are not in operation (i.e., placed in neutral position) and control the pressure in the hydraulic fluid supply circuit (i.e., the discharge pressure of the hydraulic pump) such that it becomes higher than a target pressure for load sensing control (e.g., higher than 1.5 MPa) and lower than the relief pressure (e.g., set to 2.0 MPa), thereby reducing energy loss when the flow control valves are in neutral position.
Patent Document 2 also discloses a hydraulic drive system that is capable of switching the relief pressure of a main relief valve between a first value (standard value) and a second value for high-load operation which is greater than the first value.
The hydraulic drive system of Patent Document 1 that performs load sensing control is configured to return all the hydraulic fluid discharged from the hydraulic pump to a tank via the unloading valve when the flow control valves are in neutral position without control levers being operated. In this state, the discharge amount of the hydraulic pump is controlled to a minimum, but not zero, by load sensing control. The reason for reducing the discharge amount of the hydraulic pump to the minimum, but not to zero, when the control levers are not operated is to increase the initial responsiveness of the actuators when the control levers are operated to move the flow control valves from the neutral position. Even when the control levers are not operated (i.e., the flow control valves are in the neutral position), the hydraulic pump continues to discharge at the minimum flow rate; accordingly, the discharge pressure of the hydraulic pump is influenced by the control characteristics of the unloading valve during that time.
A pump-tilting control mechanism for controlling the tilting amount (i.e., displacement volume) of the hydraulic pump typically includes a torque tilting control unit for reducing the tilting amount of the hydraulic pump when the discharge pressure of the hydraulic pump is high, thereby reducing the discharge amount of the hydraulic pump. While the engine is stopped, the tilting amount of the hydraulic pump is being maximized by a spring included in the torque tilting control unit. Thus, at the time of engine start-up, the tilting of the hydraulic pump is changed from the largest to the smallest by load sensing control.
Hydraulic excavators or other construction machines are used in various environments; they are occasionally used when the ambient temperature is below the freezing point (e.g., as low as −10 degrees Celsius or below). When the engine is started by turning on a keyed starter switch in such a cold environment, the tilting amount of the hydraulic pump is reduced, as stated above, from the largest to the smallest by load sensing control, and the hydraulic pump discharges at the flow rate which corresponds to the tilting angle (displacement volume) of that time. However, under the ambient temperature being low, the hydraulic working fluid is subject to a considerable increase in viscosity, and the unloading valve becomes less responsive. Consequently, it will take more time for the unloading valve to open, causing high pressure to be stuck inside a hydraulic fluid supply line. The viscosity increase of the hydraulic working fluid also affects load sensing control, causing a response lag. During this response lag, the discharge amount of the hydraulic pump becomes excessively high. As a result, the pressure in the hydraulic fluid supply line (the discharge pressure of the hydraulic pump) becomes high and may occasionally reach as high as 10 MPa. Accordingly, the load on the hydraulic pump (hence the load on the engine) also becomes excessively high. This makes it impossible to increase the revolving speed of the engine even by rotating its starter, which degrades the engine start-up performance.
In the hydraulic drive system of Patent Document 2, the relief pressure of the main relief valve is switched, between the first value (standard value) and the second value for high-load operation, which is higher than the first value. Even if such a configuration is applied to a hydraulic drive system that performs load sensing control, high pressure is still present inside its hydraulic fluid supply line during cold engine start-up. Moreover, the load on the hydraulic pump (hence the load on the engine) also becomes excessively high, affecting the engine start-up performance.
A possible method for solving the above problems is to increase the responsiveness of the unloading valve so that it can be more responsive in a cold environment. When the control levers are moved gradually from their neutral position without the control levers being operated, however, the discharge pressure of the hydraulic pump gradually approaches the pressure set for the unloading valve. Thus, increasingly less working fluid returns to the tank via the unloading valve. If the unloading valve is highly responsive at this time, the control of the unloading valve becomes unstable, resulting in oscillation of the valve (i.e., valve hunting).
An object of the present invention is thus to provide a hydraulic drive system for a construction machine that is capable of reducing the load on its hydraulic pump during cold engine start-up without compromising the anti-hunting characteristics of an unloading valve and thereby improving engine start-up performance.
To achieve the above object, the invention is 1) a hydraulic drive system comprising: an engine; a variable displacement hydraulic pump driven by the engine; a plurality of actuators driven by pressurized hydraulic fluid discharged from the hydraulic pump; a plurality of flow control valves for controlling the flow rate of the pressurized hydraulic fluid supplied from the hydraulic pump to the actuators; maximum load pressure detecting means for detecting a maximum load pressure from among the load pressures of the actuators when the actuators are in operation and detecting tank pressure when the actuators are not in operation, thereby outputting the detected pressure as a signal pressure; load sensing control means for controlling the displacement volume of the hydraulic pump such that the discharge pressure of the hydraulic pump becomes higher than the signal pressure by a target differential pressure; an unloading valve connected to a hydraulic fluid supply line through which the pressurized hydraulic fluid discharged from the hydraulic pump is supplied to the flow control valves, and operative to open to return the hydraulic fluid discharged from the hydraulic pump to a tank when the discharge pressure of the hydraulic pump is higher than the signal pressure by a pressure set for the unloading valve; a main relief valve connected to the hydraulic fluid supply line and operative to open to return the hydraulic fluid discharged from the hydraulic pump to the tank when the discharge pressure of the hydraulic pump is higher than a first pressure set as relief pressure, thereby limiting the maximum pressure in the hydraulic fluid supply line to the first pressure or below; and relief pressure altering means for manually switching the relief pressure of the main relief valve between the first pressure and a second pressure for engine start-up that is lower than the first pressure and allows the relief valve to return the hydraulic fluid discharged from the hydraulic pump to the tank in conjunction with the unloading valve when the actuators are not in operation.
In the above system 1), the relief pressure altering means is manually operated to switch the relief pressure of the main relief valve from the first pressure (standard pressure) to the second pressure for engine start-up, which is lower than the first pressure. Thus, the main relief valve is allowed to return the hydraulic fluid discharged from the hydraulic pump to the tank in conjunction with the unloading valve when the actuators are not in operation. Consequently, during cold engine start-up, it is possible to prevent a decrease in the responsiveness of the unloading valve and a response lag of load sensing control which are attributable to a viscosity rise in the working fluid and thereby also prevent high pressure from staying inside the hydraulic fluid supply line. It is therefore possible to reduce the load on the hydraulic pump and improve the start-up performance of the engine.
Moreover, since both of the unloading valve and the main relief valve are used to return the hydraulic fluid discharged from the hydraulic pump to the tank, the responsiveness of the unloading valve need not be increased much, whereby the anti-hunting characteristics of the unloading valve are not compromised.
As above, the invention makes it possible to reduce the load on the hydraulic pump during cold engine start-up without compromising the anti-hunting characteristics of the unloading valve, thereby improving the engine start-up performance.
2) In the above hydraulic drive system 1), the main relief valve preferably includes a spring for biasing a valve body of the main relief valve in a valve-closing direction to set the relief pressure of the main relief valve. Further, the relief pressure altering means preferably includes: a biasing force altering unit installed behind the spring of the main relief valve and having a hydraulic fluid chamber for altering the biasing force of the spring by changing the hydraulic pressure in the hydraulic fluid chamber so that the relief pressure of the main relief valve can be switched between the first pressure and the second pressure; valve means for selectively connecting the hydraulic fluid chamber of the biasing force altering unit to a pilot hydraulic fluid source and to the tank; and manual control means for controlling the valve means.
In accordance with the hydraulic drive system 2), the manual control means is operated to control the valve means, thereby selectively connecting the hydraulic fluid chamber of the biasing force altering unit to the pilot hydraulic fluid source and to the tank so that the biasing force of the spring can be changed. Therefore, the relief pressure of the main relief valve can be switched easily and reliably between the first pressure and the second pressure.
3) Preferably, the hydraulic drive system 2) further comprises: a pilot pump; a primary pilot pressure generator connected to a discharge hydraulic line of the pilot pump for generating a primary pilot pressure based on a hydraulic fluid discharged from the pilot pump; a primary pilot pressure hydraulic line into which the primary pilot pressure generated by the primary pilot pressure generator is introduced; a plurality of remote control valves connected to the primary pilot pressure hydraulic line for generating, based on the primary pilot pressure introduced into the primary pilot pressure hydraulic line, control pilot pressures to actuate the respective flow control valves; a gate lock lever installed at the entrance of a cab and operated between lock position and unlock position; and a gate lock valve installed between the primary pilot pressure generator and the primary pilot pressure hydraulic line for disconnecting the primary pilot pressure generator from the primary pilot pressure hydraulic line and connecting the primary pilot pressure hydraulic line to the tank when the gate lock lever is operated in the lock position and for connecting the primary pilot pressure generator to the primary pilot pressure hydraulic line when the gate lock lever is operated in the unlock position, wherein the pilot hydraulic fluid source comprises the pilot pump and the primary pilot pressure generator, the valve means comprises the gate lock valve, and the manual control means comprises the gate lock lever.
In accordance with the hydraulic drive system 3), because the gate lock valve (the valve means) and the gate lock lever (the manual control means) that constitute means for controlling the biasing force altering unit are existing ones, it is possible to achieve a less costly machine configuration with fewer components. In addition, no special control is required to switch the relief pressure of the main relief valve between the first pressure and the second pressure because controlling the gate lock lever to change the state of the gate lock valve changes the state of the biasing force altering unit simultaneously.
4) In any of the above hydraulic drive systems 1) to 3), the second pressure for engine start-up is preferably higher than a pressure equivalent to the target differential pressure for the load sensing control means and smaller than double the pressure set for the unloading valve.
In accordance with the above hydraulic drive system 4), the second pressure for engine start-up is set higher than a pressure equivalent to the target differential pressure for the load sensing control means. Thus, the load sensing control means is prevented from maximizing the displacement volume of the hydraulic pump, which reduces fuel consumption.
Further, the second pressure for engine start-up is set lower than double the pressure set for the unloading valve. Thus, during cold engine start-up, the load on the hydraulic pump can be reduced reliably, thereby improving the engine start-up performance.
5) In any of the above hydraulic drive systems 1) to 3), the second pressure for engine start-up is preferably a pressure that allows the main relief valve to open to return the hydraulic fluid discharged from the hydraulic pump to the tank in conjunction with the unloading valve when the actuators are not in operation under the ambient temperature below freezing point.
In accordance with the hydraulic drive system 5), the load on the hydraulic pump can be reduced reliably during cold engine start-up, thereby improving the engine start-up performance.
In accordance with the invention, even during cold engine start-up, it is possible to prevent a response lag of load sensing control and a decrease in the responsiveness of an unloading valve which are attributable to a viscosity rise in the working fluid and thereby also prevent a pressure increase inside a hydraulic fluid supply line. It is therefore possible to reduce the load on a hydraulic pump and improve the engine start-up performance during cold engine start-up.
It is further possible to reduce the load on the hydraulic pump during cold engine start-up without compromising the anti-hunting characteristics of the unloading valve, thereby improving the engine start-up performance.
An embodiment of the present invention will now be described with reference to the accompanying drawings.
<Overall Configuration>
The hydraulic drive system of
The control valve block 4 comprises the following components: multiple valve sections 4a, 4b, . . . ; shuttle valves 6a, 6b, . . . ; an unloading valve 9; a main relief valve 13; and a differential-pressure detecting valve 11. The valve sections 4a, 4b, . . . include, respectively, pressure compensating valves 41a, 41b, . . . and flow control valves (main spools) 42a, 42b, . . . , all of which are connected to a hydraulic fluid supply line 8 through which the pressurized hydraulic fluid flows from the hydraulic pump 2 and control the flow (i.e., flow rate and direction) of the pressurized hydraulic fluid supplied from the pump 2 to the actuators 5a, 5b, . . . . The shuttle valves 6a, 6b, . . . are connected, respectively, to the load ports 44a, 44b, . . . (described later) of the flow control valves 42a, 42b, . . . and detect the highest pressure from among those of the load ports 44a, 44b, . . . . Specifically, when the actuators 5a, 5b, . . . are in operation, the shuttle valves 6a, 6b, . . . detect the highest load pressure (maximum load pressure Plmax) from among those of the actuators 5a, 5b, . . . ; when those are not in operation, the shuttle valves 6a, 6b, . . . detect the pressure of a tank T. The shuttle valves 6a, 6b, . . . output the detected pressure as a signal pressure to a signal pressure hydraulic line 7. The unloading valve 9, also connected to the hydraulic fluid supply line 8, controls the discharge pressure of the hydraulic pump 2. Specifically, when the discharge pressure of the pump 2 is higher by more than a given amount (i.e., target differential pressure) than the signal pressure of the signal pressure hydraulic line 7 (i.e., than the maximum load pressure Plmax when the actuators 5a, 5b, . . . are in operation or than the tank pressure when the actuators 5a, 5b, . . . are not in operation), then, the unloading valve 9 opens to return the hydraulic fluid discharged from the pump 2 to the tank T, so that the discharge pressure of the pump 2 cannot be higher than the signal pressure by more than the target differential pressure. The main relief valve 13, also connected to the hydraulic fluid supply line 8, is adapted to limit the maximum pressure in the hydraulic fluid supply line 8 to a first pressure (described later) or below. Specifically, when the discharge pressure of the hydraulic pump 2 is higher than the first pressure (set as a relief pressure), the main relief valve 13 opens to return the hydraulic fluid discharged from the pump 2 to the tank T. The differential-pressure detecting valve 11 outputs as an absolute pressure the differential pressure between the discharge pressure of the hydraulic pump 2 and the signal pressure of the signal pressure hydraulic line 7, that is, the differential pressure between the discharge pressure of the pump 2 and the maximum load pressure Plmax (LS differential pressure) when the actuators 5a, 5b, . . . are in operation or the differential pressure between the discharge pressure of the pump 2 and the tank pressure when the actuators 5a, 5b, . . . are not in operation.
The hydraulic pump 2 is provided with a pump-tilting control mechanism 30 for controlling the tilting amount (i.e., displacement volume) of the pump 2. The pump-tilting control mechanism 30 includes a torque tilting control unit 30a and an LS tilting control unit 30b. The torque tilting control unit 30a decreases the tilting amount (hereinafter referred to as “tilting”, as needed) of the hydraulic pump 2 when the discharge pressure of the pump 2 is high, thereby decreasing the discharge amount of the pump 2. The LS tilting control unit 30b controls the tilting of the hydraulic pump 2 by load sensing such that the discharge pressure of the pump 2 becomes higher by a given amount (i.e., target differential pressure) than the signal pressure of the signal pressure hydraulic line 7 (i.e., than the maximum load pressure Plmax when the actuators 5a, 5b, . . . are in operation or than the tank pressure when the actuators 5a, 5b, . . . are not in operation).
The torque tilting control unit 30a includes a torque control actuator 31a and a spring 31b. The torque control actuator 31a receives the discharge pressure of the hydraulic pump 2 and operates to decrease the tilting of the pump 2. The spring 31b, on the other hand, operates to increase the tilting of the hydraulic pump 2. When the discharge pressure of the hydraulic pump 2 is high enough for the torque of the pump 2 to exceed the maximum permissible torque the spring 31b can absorb, the torque control actuator 31a reduces the tilting of the pump 2 to decrease the discharge amount of the pump 2, so that the torque of the pump 2 cannot exceed the maximum permissible torque for the spring 31b.
The LS tilting control unit 30b includes an LS control valve 32 and an LS control actuator 33. The LS control valve 32 generates a control pressure to be supplied to the LS control actuator 33, based on a primary pilot pressure from a primary pilot pressure generator 20, described later. The LS control actuator 33 controls the tilting of the hydraulic pump 2 in response to the control pressure.
The LS control valve 32 has a pressure receiver 32a located on the side in which the control pressure is increased to reduce the tilting of the hydraulic pump 2 and also has a pressure receiver 32b located on the side in which the control pressure is decreased to increase the tilting of the pump 2. Supplied to the pressure receiver 32a is the output pressure of the differential-pressure detecting valve 11, that is, the differential pressure between the discharge pressure of the hydraulic pump 2 and the maximum load pressure Plmax (LS differential pressure) when the actuators 5a, 5b, . . . are in operation or the differential pressure between the discharge pressure of the pump 2 and the tank pressure when the actuators 5a, 5b, . . . are not in operation (the latter differential pressure is equal to the discharge pressure of the hydraulic pump 2 when the tank pressure is assumed to be zero). Supplied to the pressure receiver 32b is the output pressure of an engine revolution counter circuit 49, described later. Based on the output pressure of the engine revolution counter circuit 49, the pressure receiver 32b sets a target differential pressure for load sensing control to 1.5 MPa, for example.
When the output pressure of the differential-pressure detecting valve 11 received by the pressure receiver 32a is higher than the target differential pressure for load sensing control determined by the pressure receiver 32b based on the output pressure of the engine revolution counter circuit 49, the LS control valve 32 increases the control pressure to decrease the tilting of the hydraulic pump 2, thereby reducing the discharge amount (hence the discharge pressure) of the hydraulic pump 2. Conversely, when the foregoing output pressure of the differential-pressure detecting valve 11 received by the pressure receiver 32a is lower than the target differential pressure for load sensing control determined by the pressure receiver 32b based on the output pressure of the engine revolution counter circuit 49, the LS control valve 32 decreases the control pressure to increase the tilting of the hydraulic pump 2, thereby increasing the discharge amount (hence the discharge pressure) of the pump 2. Accordingly, the LS control valve 32 controls the tilting of the hydraulic pump 2 such that when the actuators 5a, 5b, . . . are in operation, the LS differential pressure becomes equal to the target differential pressure (that is, the discharge pressure of the pump 2 becomes higher than the maximum load pressure Plmax by the target differential pressure) and such that when the actuators 5a, 5b, . . . are not in operation, the discharge pressure of the pump 2 becomes equal to the target differential pressure (that is, the discharge pressure of the pump 2 becomes higher than the tank pressure, which is approximately zero, by the target differential pressure).
The flow control valves 42a, 42b, . . . are valves of the closed center type and can be actuated by operation of the respective control levers not illustrated, and the operation amount of each of the control levers determines the opening area of meter-in throttle 43a or 43b. As stated above, the flow control valves 42a, 42b, . . . include the load ports 44a, 44b, . . . , respectively. When the actuators 5a, 5b, . . . are in operation (that is, when the flow control valves 42a, 42b, . . . are in operation), the load ports 44a, 44b, . . . communicate with the downstream side of the meter-in throttles 43a or 43b, so that the load pressures of the actuators 5a, 5b, . . . are extracted to the load ports 44a, 44b, . . . , respectively. When the actuators 5a, 5b, . . . are not in operation (that is, when the flow control valves 42a, 42b, . . . are not in operation or are in neutral position), the load ports 44a, 44b, . . . communicate with the tank T, so that the tank pressure is extracted to the load ports 44a, 44b, . . . .
The pressure compensating valves 41a, 41b, . . . are the upstream type compensating valves that are installed upstream of the meter-in throttles 43a or 43b of the flow control valves 42a, 42b, . . . for controlling the differential pressures across the meter-in throttles 43a or 43b of the flow control valves 42a, 42b, . . . . The pressure compensating valve 41a has a pressure receiver 46a located on the valve-closing side and a pressure receiver 46b located on the valve-opening side, with the pressure receivers 46a and 46b facing each other, and also has a pressure receiver 46c located on the valve-opening side. Supplied to the pressure receivers 46a and 46b are the upstream pressure and the downstream pressure, respectively, of the meter-in throttle 43a or 43b of the flow control valve 42a. Supplied to the pressure receiver 46c is the output pressure of the differential-pressure detecting valve 11, that is, the differential pressure between the discharge pressure of the hydraulic pump 2 and the maximum load pressure Plmax (LS differential pressure) when the actuators 5a, 5b, . . . are in operation or the differential pressure between the discharge pressure of the pump 2 and the tank pressure when the actuators 5a, 5b, . . . are not in operation. Using the output pressure of the differential-pressure detecting valve 11 as a target compensatory differential pressure, the pressure compensating valve 41a controls the differential pressure across the flow control valve 42a. Likewise, the pressure compensating valve 41b includes pressure receivers 47a, 47b, and 47c and is structurally the same as the pressure compensating valve 41a. The rest of the pressure compensating valves also have the same configuration as the pressure compensating valves 41a and 41b. With the above configuration of the pressure compensating valves 41a, 41b, . . . , the differential pressures across the meter-in throttles 43a or 43b of the flow control valves 42a, 42b, . . . are controlled to the same level, and the pressurized hydraulic fluid can be supplied in proportion to the opening areas of the meter-in throttles of the flow control valves 42a, 42b, . . . , regardless of how large or small the load pressure is. Moreover, by using the output pressure of the differential-pressure detecting valve 11 (i.e., the LS differential pressure between the discharge pressure of the hydraulic pump 2 and the maximum load pressure Plmax when the actuators 5a, 5b, . . . are in operation or the differential pressure between the discharge pressure of the pump 2 and the tank pressure when the actuators 5a, 5b, . . . are not in operation) as the target compensatory differential pressure to control the differential pressure across the flow control valve 42a, the pressurized hydraulic fluid can be supplied in proportion to the opening areas of the meter-in throttles 43a or 43b of the flow control valves 42a, 42b, . . . even if the discharge amount of the pump 2 is below the demanded flow rate, i.e., in a saturated state.
The unloading valve 9 has a pressure receiver 9a and spring 9c located on the valve-closing side and a pressure receiver 9b located on the valve-opening side, with the pressure receivers 9a and 9b facing each other. The pressure receiver 9a is connected to the signal pressure hydraulic line 7 via a signal pressure hydraulic line 10. The pressure receiver 9a receives the signal pressure detected by the shuttle valves 6a, 6b, . . . (i.e., the maximum load pressure Plmax when the actuators 5a, 5b, . . . are in operation or the tank pressure when the actuators 5a, 5b, . . . are not in operation) while the pressure receiver 9b receives the discharge pressure of the hydraulic pump 2, i.e., the pressure of the hydraulic fluid supply line 8. The pressure receiver 9a has an area of Aa, and the pressure receiver 9b an area of Ab, where Aa is equal to Ab. The spring 9c sets a target differential pressure for the unloading valve 9 to 2.0 MPa, for example. With the above configuration, the unloading valve 9 opens to return the hydraulic fluid discharged from the hydraulic pump 2 to the tank T when the discharge pressure of the pump 2 is higher than the signal pressure of the signal pressure hydraulic line 7 (i.e., than the maximum load pressure Plmax when the actuators 5a, 5b, . . . are in operation or than the tank pressure when the actuators 5a, 5b, . . . are not in operation) by more than the target differential pressure set by the spring 9c, so that the discharge pressure of the pump 2 cannot be higher than the signal pressure by more than the target differential pressure.
The main relief valve 13 includes a spring 13a located on the valve-closing side and a pressure receiver 13b located on the valve-opening side. The pressure receiver 13b receives the discharge pressure of the hydraulic pump 2 (i.e., the pressure of the hydraulic fluid supply line 8). When the discharge pressure of the pump 2 exceeds the relief pressure set by the spring 13a, the main relief valve 13 opens to return the pressurized hydraulic fluid inside the hydraulic fluid supply line 8 to the tank T, so that the discharge pressure of the pump 2 cannot exceed the relief pressure. The main relief valve 13 is also provided with a biasing force altering unit 60, described later, for changing the biasing force of the spring 13a to switch the relief pressure of the main relief valve 13 between a first pressure (a standard pressure of, for example, 25 MPa) and a second pressure for engine start-up (e.g., 3 MPa).
The differential-pressure detecting valve 11 has a pressure receiver 11a located on the pressure-increasing side and pressure receivers 11b and 11c located on the pressure-reducing side. The pressure receiver 11a receives the discharge pressure of the hydraulic pump 2 while the pressure receivers 11b and 11c receive, respectively, the signal pressure of the signal pressure hydraulic line 7 and the output pressure of the differential-pressure detecting valve 11. Exploiting the balance among those pressures and using the primary pilot pressure from the primary pilot pressure generator 20 (described later), the differential-pressure detecting valve 11 generates and outputs as an absolute pressure the differential pressure between the discharge pressure of the hydraulic pump 2 and the signal pressure of the signal pressure hydraulic line 7.
The output port of the differential-pressure detecting valve 11 is connected to the pressure receiver 32a of the LS control valve 32 of the pump-tilting control mechanism 30 via signal pressure hydraulic lines 15 and 16, so that the output pressure of the differential-pressure detecting valve 11 can be supplied to the pressure receiver 32a. The output port of the differential-pressure detecting valve 11 is connected also to the pressure receivers 46c, 47c, . . . of the pressure compensating valves 41a, 41b, . . . via the signal pressure hydraulic line 15 and signal pressure hydraulic lines 17 and 18, so that the output pressure of the differential-pressure detecting valve 11 can be supplied to the pressure receivers 46c, 47c, . . . as the target compensatory differential pressure.
The actuators 5a, 5b, . . . could be boom cylinders, arm cylinders, or the like for a hydraulic excavator. A hydraulic excavator according to the invention also has other actuators including a swing motor, right and left travelling cylinders, a bucket cylinder, and the like.
The hydraulic drive system of the present embodiment includes the engine revolution counter circuit 49 and the primary pilot pressure generator 20 as stated above and further includes a gate lock valve 23.
The engine revolution counter circuit 49 includes a flow-rate detecting valve 50 and a differential-pressure detecting valve 51. The flow-rate detecting valve 50 includes a variable throttle 50a. The upstream side of the throttle 50a is connected to a discharge hydraulic line 3a that extends from the pilot pump 3 while the downstream side of the throttle 50a is connected to a hydraulic line 3c that extends from the primary pilot pressure generator 20.
The flow-rate detecting valve 50 detects the discharge amount of the pilot pump 3 as a change of differential pressure across the throttle 50a. Because the discharge amount of the pilot pump 3 varies with changes of the revolving speed of the engine 1, detecting the discharge amount of the pilot pump 3 allows detection of the revolving speed of the engine 1. For instance, a decrease in the revolving speed of the engine 1 leads to a decrease in the discharge amount of the pilot pump 3 and also to a decrease in differential pressure across the throttle 50a.
The variable throttle 50a is configured such that its orifice area changes in a continuous manner. The flow-rate detecting valve 50 further includes a pressure receiver 50b, which operates to open the valve 50, and a pressure receiver 50c and spring 50d, which operate to reduce the orifice area of the valve 50. The pressure receiver 50b receives the upstream pressure of the throttle 50a (i.e., the pressure of the discharge hydraulic line 3a) while the pressure receiver 50c receives the downstream pressure of the throttle 50a (i.e., the pressure of the hydraulic line 3c). The throttle 50a changes its own orifice area when the differential pressure across the throttle 50a changes.
The differential-pressure detecting valve 51 outputs as an absolute pressure the differential pressure across the throttle 50a, which changes in response to the engine revolving speed, whereby the engine speed can be detected. The differential-pressure detecting valve 51 has a pressure receiver 51a located on the pressure-increasing side and pressure receivers 51b and 51c located on the pressure-reducing side. The pressure receiver 51a receives the upstream pressure of the throttle 50a while the pressure receivers 51b and 51c receive, respectively, the downstream pressure of the throttle 50a and the output pressure from the differential-pressure detecting valve 51. Exploiting the balance among those pressures and using the primary pilot pressure from the primary pilot pressure generator 20, the differential-pressure detecting valve 51 generates and outputs as an absolute pressure the differential pressure across the throttle 50a.
The output port of the differential-pressure detecting valve 51 is connected to the pressure receiver 32b of the LS control valve 32 via a signal pressure hydraulic line 53, so that the output pressure of the differential-pressure detecting valve 51 can be supplied to the pressure receiver 32b as the target differential pressure for load sensing control. By thus directing the differential pressure across the throttle 50a to the pressure receiver 32b of the LS control valve 32 and setting that differential pressure as the target differential pressure for load sensing control, saturation phenomena can be overcome on an engine-speed basis. This leads also to finer and more precise machine maneuverability when the engine revolving speed is low. JP-10-196604-A has a detailed description of the above.
The primary pilot pressure generator 20 includes a pilot relief valve 21 connected to the hydraulic line 3c. The pilot relief valve 21 maintains the pressure in the hydraulic line 3c at a fixed value (e.g., 4.0 MPa), thereby generating the primary pilot pressure. The downstream side of the hydraulic line 3c is connected to a primary pilot pressure hydraulic line 3b via the gate lock valve 23. Also connected to the primary pilot pressure hydraulic line 3b are remote control valves (not illustrated) that are actuated by the above-mentioned control levers and generates, based on the pressure from the primary pilot pressure generator 20 (i.e., the primary pilot pressure), control pilot pressures for controlling the respective flow control valves 42a, 42b, . . . .
The gate lock valve 23 is positioned between the hydraulic line 3c and the primary pilot pressure hydraulic line 3b and controlled by a gate lock lever 24 located at the cab entrance of the hydraulic excavator. The gate lock lever 24 is operated between the lock position (OFF position) that allows the operator to get in/out of the cab and the unlock position (ON position) that does not allow the operator to do so. When the gate lock lever 24 is placed in the lock position (OFF), the gate lock valve 23 is also placed in that position (i.e., moved to the right of
In the present embodiment, the biasing force altering unit 60 for the main relief valve 13 is connected to the primary pilot pressure hydraulic line 3b via a hydraulic line 22. When the gate lock valve 23 is in the unlock position, the biasing force altering unit 60 sets the relief pressure to the first pressure (a standard pressure of, for example, 25 MPa). When the gate lock valve 23 is in the lock position, the biasing force altering unit 60 sets the relief pressure to the second pressure for engine start-up (e.g., 3 MPa).
The biasing force altering unit 60, the gate lock valve 23, and the gate lock lever 24 constitute relief pressure altering means for manually (with the use of the gate lock lever 24) switching the relief pressure of the main relief valve 13 between the first pressure (a standard pressure of, for example, 25 MPa) and the second pressure for engine start-up (e.g., 3.0 MPa) that is lower than the first pressure and allows the main relief valve 13 to return the hydraulic fluid discharged from the hydraulic pump 2 to the tank in conjunction with the unloading valve 9 when the actuators 5a, 5b, . . . are not in operation.
The pilot pump 3 and the primary pilot pressure generator 20 constitute a pilot hydraulic fluid source. The gate lock valve 23 constitutes valve means for selectively connecting the hydraulic fluid chamber 69 of the biasing force altering unit 60 (described later) to the pilot hydraulic fluid source and to the tank T. The gate lock lever 24 constitutes manual control means for controlling the valve means (i.e., the gate lock valve 23).
The second pressure for engine start-up is set low to such an extent that when the actuators 5a, 5b, . . . are not in operation (when the flow control valves 42a, 42b, . . . are all in neutral position) under the ambient temperature below the freezing point, the main relief valve 13 can open to return the hydraulic fluid discharged from the hydraulic pump 2 to the tank T in conjunction with the unloading valve 9. Preferably, the second pressure is higher than a pressure equivalent to the target differential pressure for load sensing control (e.g., higher than 1.5 MPa) and smaller than double the pressure set for the unloading valve 9 (e.g., smaller than 4.0 MPa (2.0 MPa times 2)).
<Detailed Structure of the Main Relief Valve 13>
The main relief valve 13 comprises the following components: a housing 64; a valve body 65; and a support 70. The housing 64 has a valve chamber 61 therein and an input port 62 and an output port 63 therethrough. The valve body 65 is located inside the housing 64 and used to open or close the input port 62. The support 70 has an inlet 66 that communicates with the input port 62 and an outlet 67 that communicates with the output port 63 and is used to secure the housing 64. The above-mentioned spring 13a of the main relief valve 13 is installed inside the housing 64 in such a way as to bias the valve body 65 in the valve-closing direction. The pressure receiver 13b, mentioned above, of the relief valve 13 is installed on the downstream side of the input port 62 where the valve body 65 is seated. The inlet 66 is connected to the hydraulic fluid supply line 8 while the outlet 67 is connected to the tank T.
The biasing force altering unit 60 is installed behind the spring 13a located inside the housing 64. The biasing force altering unit 60 includes a piston 68 and the hydraulic fluid chamber 69, mentioned above. The piston 68 is installed inside the housing 64 such that the piston 68 can move in axial directions of the housing 64 (to the right and left of
As illustrated in
As illustrated in
<Structure of the Hydraulic Excavator>
The boom 111 and the arm 112 are moved by expanding or contracting a boom cylinder 5a and an arm cylinder 5b (the boom cylinder 5a and the arm cylinder 5b correspond to the actuators 5a and 5b, respectively, of
The lower travel structure 101 travels by the rotation of left and right travel motors 118a and 118b while the swing post 103 rotates by the expansion or contraction of a swing cylinder 119. The hydraulic circuit diagram of
Inside the cab 108 is a cab seat 121 on which the operator is seated. Installed on the right and left sides of the cab seat 121 are, respectively, a control lever device 122 having bucket/boom control levers and a control lever device 123 having swing/arm control levers. Also, the gate lock lever 24 is installed at the entrance of the cab seat 121. The solid line of
—Operation—
Described next is the operation of the above hydraulic excavator on which the hydraulic drive system of the present embodiment is mounted.
<When the Gate Lock Lever is in the Lock Position>
After a day's work, the operator turns off a keyed starter switch not illustrated to stop the engine 1. At this time, the operator places the gate lock lever 24 in the lock position for safety purposes, thereby also placing the gate lock valve 23 in the lock position, which position allows the primary pilot pressure hydraulic line 3b to communicate with the tank T so that the flow control valves 42a, 42b, . . . cannot be controlled. When the engine 1 stops, the hydraulic pump 2 does not discharge any pressurized hydraulic fluid; thus, the spring 31b of the torque tilting control unit 30a works to maximize the tilting of the pump 2.
Before the hydraulic excavator is operated for a day's work, the gate lock lever 24 is in the lock position, and the tilting (i.e., displacement volume) of the hydraulic pump 2 is the largest. Since the gate lock lever 24 is in the lock position and the gate lock valve 23 allows the primary pilot pressure hydraulic line 3b to communicate with the tank T, the piston 68 of the biasing force altering unit 60 extends the spring 13a and keeps its force weak, as illustrated in
To start operation of the hydraulic excavator for a day's work, the operator first turns on the keyed starter switch, not illustrated, thereby starting the engine 1. Right after the engine start-up, the LS control valve 32 starts to control the tilting (displacement volume) of the hydraulic pump 2 (i.e., perform load sensing control) such that the signal pressure received by the pressure receiver 32a from the signal pressure hydraulic line 16 is equal to a target differential pressure set by the pressure receiver 32b (e.g., 1.5 MPa). Because, right after the start-up, the control levers are not operated and the flow control valves 42a, 42b, . . . are thus in neutral position, the signal pressure from the signal pressure hydraulic line 7 (i.e., the output pressure of the shuttle valves 6a, 6b, . . . ) is the tank pressure, and the signal pressure from the signal pressure hydraulic line 16 (i.e., the output pressure of the differential-pressure detecting valve 11) is approximately equal to the discharge pressure of the hydraulic pump 2. Since the tilting of the hydraulic pump 2 is the largest right after the start-up of the engine 1, the discharge pressure of the hydraulic pump 2 will increase transiently, exceeding the target differential pressure for load sensing control. Therefore, the LS control valve 32 reduces the tilting of the hydraulic pump 2 from the largest to the smallest so that the discharge pressure of the pump can be equal to the target differential pressure, thereby reducing the discharge amount of the pump 2 to a minimum possible value but not to zero. The reason for reducing the discharge amount of the pump 2 to the minimum possible value, but not to zero, even when the flow control valves 42a, 42b, . . . are in neutral position without the control levers being operated is to increase the responsiveness of the actuators when the control levers are operated to move the flow control valves 42a, 42b, . . . from the neutral position.
Thus controlling the tilting (discharge amount) of the hydraulic pump 2 allows the unloading valve 9 to open to return the hydraulic fluid discharged from the hydraulic pump 2 (i.e., the pressurized hydraulic fluid inside the hydraulic fluid supply line 8) to the tank when the discharge pressure of the pump 2 exceeds the pressure set for the unloading valve 9 (i.e., target differential pressure).
When the ambient temperature is below the freezing point (e.g., as low as −10 degrees Celsius), the working fluid is considerably high in viscosity during engine start-up. In such a case, the responsiveness of the unloading valve 9 decreases, and it will take more time for the unloading valve 9 to open, causing high pressure to be stuck inside the hydraulic fluid supply line 8. The viscosity increase of the working fluid also affects load sensing control, causing a response lag. During this response lag, the discharge amount of the hydraulic pump 2 becomes excessively high. As a result, the pressure in the hydraulic fluid supply line 8 (the discharge pressure of the hydraulic pump) becomes high and may occasionally reach as high as 10 MPa. For this reason, the load on the hydraulic pump 2 (hence the load on the engine 1) is conventionally too high, affecting engine start-up.
In the present embodiment, however, when the gate lock lever 24 is in the lock position (OFF), the relief pressure of the main relief valve 13 is set, as stated above, to the second pressure for engine start-up (e.g., 3.0 MPa), which is lower than the first pressure (the standard pressure, e.g., 25 MPa). Consequently, when the discharge pressure of the hydraulic pump 2 reaches the lower second pressure, the main relief valve 13 starts to open, thereby returning the hydraulic fluid discharged from the pump 2 to the tank.
By thus allowing the main relief valve 13 to open besides the unloading valve 9, the discharge pressure of the hydraulic pump 2 can be prevented from becoming excessively high especially when the ambient temperature is low, whereby engine start-up performance can be improved.
If the engine start-up second pressure for the main relief valve 13 is set lower than a pressure equivalent to the target differential pressure for load sensing control (e.g., lower than 1.5 MPa), the LS tilting control unit 30b (load sensing control means) controls the displacement volume of the hydraulic pump 2 in such a way as to maximize it, which increases fuel consumption. In the present embodiment, by contrast, the second pressure for the main relief valve 13 is set higher than a pressure equivalent to the target differential pressure for load sensing control. Thus, the load sensing control means is prevented from maximizing the displacement volume of the hydraulic pump 2, which leads to less fuel consumption.
Further, if the second pressure for the main relief valve 13 is set higher than double the pressure set for the unloading valve 9, the load on the hydraulic pump 2 during engine start-up may not be reduced greatly when the ambient temperature is lower than −10 degrees Celsius. In the present embodiment, by contrast, the second pressure for the main relief valve 13 is set lower than double the pressure set for the unloading valve 9 and set, for example, to 3.0 MPa or thereabout, which is approximately 1.5 times the pressure set for the unloading valve 9. Thus, even when the ambient temperature is lower than −10 degrees Celsius, the load on the hydraulic pump 2 can be reduced reliably, thereby improving engine start-up performance.
<When the Gate Lock Lever is in the Unlock Position>
When the operator places the gate lock lever 24 in the unlock position (ON) after the engine start-up, the gate lock valve 23 is also switched to the unlock position, thereby connecting the discharge hydraulic line 3a of the pilot pump 3 to the primary pilot pressure hydraulic line 3b. Further, the piston 68 of the biasing force altering unit 60 contracts the spring 13a and keeps its force strong as illustrated in
Unless the control levers are operated after the placement of the gate lock lever 24 in the unlock position, the LS control valve 32 continues to minimize the tilting of the hydraulic pump 2 so that the discharge amount of the pump 2 can be a minimum possible value. The discharge pressure of the hydraulic pump 2 is maintained at the pressure set for the unloading valve because the unloading valve 9 opens to return the hydraulic fluid discharged from the pump 2 (the pressurized hydraulic fluid inside the hydraulic fluid supply line 8) to the tank when the discharge pressure of the pump 2 exceeds the pressure set for the unloading valve 9 (e.g., 2.0 MPa). Further, when the gate lock lever 24 is in the unlock position, the relief pressure of the main relief valve 13 is set to the first pressure (the standard pressure, e.g., 25 MPa). Accordingly, the main relief valve 13 does not open unless the discharge pressure of the hydraulic pump 2 reaches the first pressure.
In accordance with the above-described embodiment of the invention, the relief pressure altering means (the biasing force altering unit 60, the gate lock valve 23, and the gate lock lever 24) is manually operated to switch the relief pressure of the main relief valve 13 from the first pressure (the standard pressure, e.g., 25 MPa) to the second pressure for engine start-up (e.g., 3.0 MPa) which is lower than the first pressure. Thus, the main relief valve 13 is allowed to return the hydraulic fluid discharged from the hydraulic pump 2 to the tank T in conjunction with the unloading valve 9 if the discharge pressure of the pump 2 exceeds the pressure set for the unloading valve 9 (e.g., 2.0 MPa) without the actuators 5a, 5b, . . . being operated. Consequently, during cold engine start-up, it is possible to prevent a response lag of load sensing control and a decrease in the responsiveness of the unloading valve 9 which are attributable to a viscosity rise in the working fluid and thereby also prevent high pressure from staying inside the hydraulic fluid supply line 8. It is therefore possible to prevent a considerable boost in the discharge pressure of the hydraulic pump 2, reduce the load on the hydraulic pump 2, and improve the start-up performance of the engine 1.
Moreover, since both of the unloading valve 9 and the main relief valve 13 are used to return the hydraulic fluid discharged from the hydraulic pump 2 to the tank T, the responsiveness of the unloading valve 9 need not be increased much, whereby the anti-hunting characteristics of the unloading valve 9 are not compromised.
Further, the gate lock lever 24 (the manual control means) is operated to control the gate lock valve 23 (the valve means), thereby selectively connecting the hydraulic fluid chamber 69 of the biasing force altering unit 60 to the primary pilot pressure generator 20 or to the tank T so that the biasing force of the spring 13a can be changed. Therefore, the relief pressure of the main relief valve 13 can be switched easily and reliably between the first pressure and the second pressure.
Furthermore, because the gate lock valve 23 (the valve means) and the gate lock lever 24 (the manual control means) that constitute means for controlling the biasing force altering unit 60 are existing ones, it is possible to achieve a less costly machine configuration with fewer components. In addition, no special control is required to switch the relief pressure of the main relief valve 13 between the first pressure and the second pressure because controlling the gate lock lever 24 to change the state of the gate lock valve 23 changes the state of the biasing force altering unit 60 simultaneously.
The above-described embodiment of the invention can be modified or changed in various forms within the scope of the invention. For instance, while the biasing force altering unit 60 of the above embodiment is hydraulically driven, it can instead be solenoid-driven. In that case, the position of the gate lock lever 24 is detected electrically, and solenoid excitation and non-excitation are controlled. This provides the same advantages as those of the above embodiment (improved engine start-up performance during cold start-up and the like).
Further, while both of the gate lock valve 23 (the valve means) and the gate lock lever 24 (the manual control means) constitute the means for controlling the biasing force altering unit 60 in the above embodiment, it is instead possible to use dedicated valve means and manual control means, in which case, too, the same advantages as those of the above embodiment can be obtained.
In the above embodiment, the target differential pressure for load sensing control is set as a variable that changes in response to the engine revolving speed, based on the output pressure of the engine revolution counter circuit 49, and the target differential pressure for the unloading valve 9 is set as a constant by the spring 9c. Alternatively, the target differential pressure for the unloading valve 9 can also be set as a variable that changes in response to the engine revolving speed, based on the output pressure of the engine revolution counter circuit 49.
Furthermore, while the above embodiment has taken the hydraulic excavator as an example of a construction machine, the invention can be applied in the same manner to other construction machines such as cranes, wheel loaders, and the like.
| Number | Date | Country | Kind |
|---|---|---|---|
| 2008-282031 | Oct 2008 | JP | national |
| Filing Document | Filing Date | Country | Kind | 371c Date |
|---|---|---|---|---|
| PCT/JP2009/065754 | 9/9/2009 | WO | 00 | 3/25/2011 |
