Patent Application
20150240455
The present invention relates generally to hydraulic driving systems for construction machines such as hydraulic excavators. More particularly, the invention is directed to hydraulic driving systems for construction machines, each of the systems being configured to subject a delivery rate of hydraulic fluid from a hydraulic pump to load-sensing control so that a fluid delivery pressure of the hydraulic pump becomes higher by a target differential pressure than a load pressure of an actuator to which the highest load pressure is to be assigned among a plurality of actuators.
Some of the hydraulic driving systems for construction machines such as hydraulic excavators are designed to control a flow rate of a hydraulic fluid as delivered from a hydraulic pump (a main pump). Accordingly, a fluid delivery pressure of the hydraulic pump becomes higher by a target differential pressure than a load pressure of an actuator to which the highest load pressure is to be assigned among a plurality of actuators. Such flow rate control is called load-sensing control. The hydraulic driving systems in which the load-sensing control is performed are adapted to maintain a predetermined differential pressure across each of a plurality of flow control valves via a pressure compensating valve disposed for the flow control valve independently. During combined operations control for simultaneously driving the actuators, the hydraulic driving systems can thus supply the hydraulic fluid to the actuators at a ratio commensurate with an opening area of each flow control valve, irrespective of a magnitude of the actuator load pressures.
Patent Document 1, for example, describes such a hydraulic driving system adapted to perform the load-sensing control. The hydraulic driving system described in Patent Document 1 is configured so that a differential pressure (hereinafter referred to as the load-sensing differential pressure) between a fluid delivery pressure of a hydraulic pump and a load pressure of an actuator to which the highest load pressure is to be assigned among a plurality of actuators is guided as a target compensation differential pressure to pressure-receiving portions constructed so as to operate pressure compensating valves in a direction to increase in opening area. The hydraulic driving system is also configured so that the target compensation differential pressure across each of the pressure compensating valves is set to be the same value equivalent to the load-sensing differential pressure. Thus, a differential pressure across each of a plurality of flow control valves is held at the load-sensing differential pressure level. During combined operations control for simultaneously driving the actuators, therefore, even if the fluid delivery pressure of the hydraulic pump is insufficient (this state is hereinafter referred to as saturation), a decrease in load-sensing differential pressure according to a particular degree of the saturation uniformly reduces the target compensation differential pressures of the pressure compensating valves (i.e., the differential pressures across the flow control valves), thus enabling a delivery rate of the hydraulic fluid from the hydraulic pump to be redistributed to a ratio of the flow rates demanded from the actuators.
In addition, the pressure compensating valves of the hydraulic driving systems in which the load-sensing control is performed are usually configured so that as described in Patent Document 1, the valve will fully close when a spool operates in a direction to reduce an opening area of the valve and reaches a stroke end of the spool.
In contrast to the above, Patent Document 2 describes a hydraulic driving system configured so as not to fully close a pressure compensating valve even after a spool has operated in a direction to reduce an opening area of the valve and reached a stroke end of the spool.
Patent Document 1: JP-2007-24103-A
Patent Document 2: JPH7-76861-A
The following problems, however, exist in the above conventional art.
As discussed above, the conventional hydraulic driving systems in which the load-sensing control is performed, such as the one described in Patent Document 1, each include pressure compensating valves, whereby the system can supply a hydraulic fluid to a plurality of actuators at a ratio commensurate with an opening area of flow control valves, irrespective of the load pressures applied during the combined operations control for simultaneously driving the actuators.
In addition, for the hydraulic driving system described in Patent Document 1, the load-sensing control differential pressure is set as a target compensation differential pressure. Thus, even if saturation occurs during the combined operations control for simultaneously driving the plurality of actuators, a flow rate of the hydraulic fluid delivered from a hydraulic pump can be redistributed at a ratio of the flow rates demanded from the actuators.
For the hydraulic driving system described in Patent Document 1, however, since the pressure compensating valves are each constructed so as to fully close at the stroke end of the spool as operated in the direction to reduce the opening area of the valve, if saturation occurs during the combined operations control likely to generate a significant difference in load pressure between any two actuators, the pressure compensating valve lower in load pressure may be excessively reduced in opening area or excessively closed. The actuator undergoing the lower load pressure is therefore likely to slow down and/or even stop operating.
For the hydraulic driving system described in Patent Document 2, since the pressure compensating valve is constructed so as not to fully close at the stroke end of the spool as operated in the direction to reduce the opening area of the valve, even if saturation occurs during such combined operations control as discussed above, the pressure compensating valve lower in load pressure does not excessively reduce the opening area, nor does the valve fully close. A slowdown and/or stop of an actuator lower in load pressure can therefore be prevented.
The hydraulic driving system described in Patent Document 2, however, has a problem in that if saturation occurs during the combined operations control likely to generate a particularly significant difference in load pressure between any two actuators, since the pressure compensating valve of the actuator lower in load pressure does not close, a large portion of the fluid delivered from a main pump may be absorbed by the actuator lower in load pressure. The actuator undergoing the higher load pressure may therefore slow down and/or even stop operating.
For example, when either a boom, arm, or bucket hydraulic cylinder of a construction machine is driven for a change in a posture of a front working implement during slope climbing, a very high load pressure is usually applied to a track motor and a particularly significant difference in load pressure occurs between the track motor and the actuator (hydraulic cylinder) of the front working implement. Hence a hydraulic fluid delivered from a hydraulic pump may flow into the actuator of the front working implement that undergoes the lower load pressure, and the vehicle may stop traveling.
In addition, even when the vehicle is traveling along a level ground surface, if a blade is abruptly operated for a change in a posture of the blade during traveling, a particularly significant difference in load pressure between the track motor and the blade cylinder occurs as in the above case. In this case, a large portion of the hydraulic fluid delivered from the hydraulic pump may flow into the blade cylinder, which is the actuator having the lower load pressure. This situation may lead to a slowdown of traveling and undermine an operation feeling.
The above drawbacks may also occur with elements other than the track motor. For example, a standby actuator provided on an attachment such as a crusher used in exchange for the bucket tends to increase in load pressure and a difference in load pressure increases particularly during the combined operations control where the standby actuator is driven simultaneously with any other actuator, for example the hydraulic cylinder of the boom, arm, or bucket. These increases in load pressure are also likely to cause problems similar to those described above.
An object of the present invention is to provide a hydraulic driving system for a construction machine in which the load-sensing control is performed. If saturation occurs during combined operations control that generates a significant difference in load pressure between any two actuators, the hydraulic driving system prevents full closing of a pressure compensating valve undergoing the lower load pressure, and hence a slowdown and stop of the actuator lower in load pressure. In addition, if saturation occurs during the combined operations control that generates a particularly significant difference in load pressure between any two actuators, the hydraulic driving system ensures a necessary supply of hydraulic fluid to the actuator higher in load pressure, thereby preventing a slowdown and stop of the actuator higher in load pressure, and thus providing appropriate combined-operations controllability.
To achieve the above object, in an aspect of the present invention, a hydraulic driving system for a construction machine includes: a variable-displacement type of hydraulic pump; a plurality of actuators each driven by a hydraulic fluid delivered from the hydraulic pump; a plurality of flow control valves that each control a flow rate of the hydraulic fluid supplied from the hydraulic pump to a corresponding one of the actuators; a plurality of operating devices disposed in association with the actuators, each of the operating devices including a remote control valve configured to generate an operating pilot pressure for driving a corresponding one of the flow control valves; a plurality of pressure compensating valves each for controlling a differential pressure across a corresponding one of the flow control valves independently; and a pump control unit for controlling a capacity of the hydraulic pump by means of load-sensing control so that a fluid delivery pressure of the hydraulic pump becomes higher by a target differential pressure than a load pressure of an actuator to which the highest load pressure is to be assigned among the plurality of actuators. In the hydraulic driving system, the pressure compensating valves are each a pressure compensating valve of a type not fully closing at a stroke end of the valve as operated in a direction to decrease in opening area. The plurality of actuators include a specific actuator that undergoes a higher load pressure during combined operations control when the specific actuator is driven simultaneously with actuators other than the specific actuator. A control valve is disposed in hydraulic fluid line portions upstream or downstream relative to a pressure compensating valve of the actuator other than the specific actuator, the control valve reducing a flow passage area of the hydraulic fluid line portion upon operation of a specific operating device, among the plurality of operating devices, that relates to the specific actuator.
When pressure compensating valves each of the type not fully closing at the stroke end of the valve as operated in the direction to decrease in opening area are arranged in this way as the plurality of pressure compensating valves, even if saturation occurs during the combined operations control that generates a significant difference in load pressure between any two actuators, full closing of a pressure compensating valve undergoing the lower load pressure is prevented and hence the actuator lower in load pressure can be prevented from slowing down and stopping.
In addition, a control valve is disposed in fluid line portions upstream or downstream relative to a pressure compensating valve of the actuators other than the specific actuator. The specific actuator is an actuator which undergoes a higher load pressure during simultaneous driving with another actuator by combined operations control. The control valve reduces a flow passage area of the hydraulic fluid line portion in response to the operation of a specific operating device, one of the plurality of operating devices that relates to the specific actuator. Thus, when the specific operating device is operated, the control valve reduces the flow passage area of the hydraulic fluid line portion. Accordingly, if saturation occurs during a combined operations control in which the specific actuator and the actuators other than the specific actuator have a significant difference in load pressure, then a flow rate of the hydraulic fluid supplied to the actuator other than the specific actuator, or the actuator undergoing the lower load pressure, is suppressed. This ensures a necessary supply of hydraulic fluid to the specific actuator, or the actuator undergoing the higher load pressure, thereby prevent a slowdown or stop of the specific actuator, or the actuator undergoing the higher load pressure, and thus provide appropriate combined-operations controllability.
The plurality of pressure compensating valves are each disposed in a corresponding one of a plurality of parallel hydraulic fluid lines branching from a supply fluid line connected to the hydraulic pump, and the hydraulic fluid line portion with the control valve disposed therein is one of the parallel hydraulic fluid lines and is where, for example, the pressure compensating valve relating to the actuator other than the specific actuator is disposed.
Accordingly, when the specific operating device is operated, a flow rate of the hydraulic fluid supplied only to the actuator corresponding to the parallel hydraulic fluid line will be suppressed and flow rates of the fluid supplied to the other actuators will not be suppressed. Controllability can therefore be prevented from decreasing, even if part of the other actuators decreases in speed during the combined operations control of the specific actuator and at least one of the other actuators.
The hydraulic fluid line portion with the control valve disposed therein may be a portion of the supply fluid line, and the hydraulic fluid line portion may lie upstream relative to a branching position of the parallel hydraulic fluid lines having the pressure compensating valves of the other actuators arranged therein.
Thus when the actuator other than the specific actuator is present in plurality, flow rates of the hydraulic fluid supplied to the actuators other than the specific actuator will also be suppressed with one control valve and the advantageous effects described above will be obtained. This will in turn reduce the number of constituent parts needed and yield the effects less expensively.
The hydraulic driving system further includes a shuttle valve serving as an operations detector to detect the operations on a specific operating device, and the shuttle valve detects an operating pilot pressure generated by the remote control valve of the specific operating device and outputs a hydraulic signal commensurate with the detected pilot pressure. The control valve in this case can be a hydraulic fluid pressure control valve that controls the fluid pressure according to the particular hydraulic signal. The hydraulic driving system additionally includes a pressure sensor that detects the operating pilot pressure generated by the remote control valve of the specific operating device and then outputs an electrical signal commensurate with the operating pilot pressure. The control valve in this case can be a solenoid-operated control valve that operates in accordance with the electrical signal.
The hydraulic driving system may further include a manual selector adapted to be switched between its first position and its second position. The system may also include a controller. When the manual selector is in the first position, the controller activates a function of the control valve that reduces the flow passage area of the hydraulic fluid line portion in response to the operation of the specific operating device. When the manual selector is switched to the second position, the controller deactivates the function of the control valve that reduces the flow passage area of the hydraulic fluid line portion in response to the operation of the specific operating device.
Thus an operator can freely select whether to use a function of the present invention according to his or her needs or preference.
The specific actuator is for example a track motor that drives a track structure of the construction machine, and each of the actuators other than the specific actuator is for example one of the hydraulic cylinders which actuate the front working implement of the construction machine, or otherwise the blade cylinder that actuates the blade.
Thus during climbing of an upslope, when any one of the hydraulic cylinders is driven for a change in a posture of the front working implement, the control valve suppresses a flow rate of the hydraulic fluid supplied to the particular hydraulic cylinder. A necessary amount of fluid is then reliably supplied to the track motor, a slowdown and stop of traveling are prevented, and appropriate combined-operations controllability is consequently obtained. In addition, if or when the blade is abruptly operated for a change in a posture of the blade during traveling along a level ground surface, the control valve suppresses a flow rate of the hydraulic fluid supplied to the blade cylinder. A necessary amount of fluid is then reliably supplied to the track motor, a slowdown of traveling is prevented, and an operation feeling is improved.
In accordance with the present invention, in the hydraulic driving system in which the load-sensing control is performed, if saturation occurs during the combined operations control that generates a significant difference in load pressure between any two actuators, the system prevents a slowdown and stop of the actuator with the lower load pressure by preventing full closing of the pressure compensating valve with the lower load pressure. Additionally, if saturation occurs during the combined operations control likely to generate a particularly significant difference in load pressure between any two actuators, the hydraulic driving system ensures the necessary supply of hydraulic fluid to the actuator higher in load pressure, thereby preventing a slowdown and stop of the actuator higher in load pressure, and thus providing appropriate combined-operations controllability.
Hereunder, embodiments of the present invention will be described in accordance with the accompanying drawings.
An appearance of a hydraulic excavator is shown in
Referring to
A hydraulic driving system according to a first embodiment of the present invention is shown in
First, a basic configuration of the hydraulic driving system according to the present embodiment is described.
The hydraulic driving system according to the present embodiment includes: an engine 1; a main hydraulic pump (hereinafter, referred to simply as main pump) 2 that is driven by the engine 1; a pilot pump 3 that operates in association with the main pump 2 and is driven by the engine 1; a plurality of actuators 5, 6, 7, 8, 9, 10, 11, and 12 that are each driven by a hydraulic fluid delivered from the main pump 2, more specifically the actuators being a left track motor 5, a right track motor 6, a swing motor 7, a blade cylinder 8, a swing cylinder 9, a boom cylinder 10, an arm cylinder 11, and a bucket cylinder 12; and a control valve 4. The hydraulic excavator employing the hydraulic driving system according to the present embodiment is a hydraulic mini-excavator, for example.
The control valve 4 includes: a plurality of valve sections 13, 14, 15, 16, 17, 18, 19, and 20 that are each connected to a supply fluid line 2a of the main pump 2 and independently control a direction and flow rate of the hydraulic fluid supplied from the main pump 2 to a corresponding one of the actuators; a plurality of shuttle valves 22a, 22b, 22c, 22d, 22e, 22f, and 22g that each select a maximum load pressure PLmax, the highest of load pressures upon the actuators 5, 6, 7, 8, 9, 10, 11, 12, and outputs the maximum load pressure to a signal fluid line 21; a main relief valve 23 connected to an intra-valve supply fluid line 4a connected to the supply fluid line 2a of the main pump 2, the valve 23 being disposed to limit a maximum pump pressure that is a maximum fluid delivery pressure of the main pump 2; a differential-pressure reducing valve 24 connected to a pilot hydraulic fluid source 33 described later herein, and adapted to receive pressures of the supply fluid line 4a and the signal fluid line 21 as pressure signal inputs, and then output an absolute pressure that is a differential pressure PLS between a fluid delivery pressure (pump pressure) Pd of the main pump 2 and the maximum load pressure PLmax; and an unloading valve 25 connected to the intra-valve supply fluid line 4a and functioning to receive the pressures of the supply fluid line 4a and the signal fluid line 21 as pressure signal inputs, then after the differential pressure PLS between the pump pressure Pd and the maximum load pressure PLmax has exceeded a constant value preset via a spring 25a, return a portion of the delivered fluid flow rate within the main pump 2 to a tank T, and maintain the differential pressure PLS at a level equal to or less than the constant value preset via the spring 25a. The unloading valve 25 and the main relief valve 23 are connected at respective exit ends to an intra-valve tank fluid line 29 and further connected to the tank T via the fluid line 29.
The valve section 13 includes a flow control valve 26a and a pressure compensating valve 27a, the valve section 14 includes a flow control valve 26b and a pressure compensating valve 27b, the valve section 15 includes a flow control valve 26c and a pressure compensating valve 27c, the valve section 16 includes a flow control valve 26d and a pressure compensating valve 27d, the valve section 17 includes a flow control valve 26e and a pressure compensating valve 27e, the valve section 18 includes a flow control valve 26f and a pressure compensating valve 27f, the valve section 19 includes a flow control valve 26g and a pressure compensating valve 27g, the valve section 20 includes a flow control valve 26h and a pressure compensating valve 27h. Each of the pressure compensating valves 27a to 27h is disposed in a corresponding independent one of a plurality of parallel hydraulic fluid lines 41a to 41f branching, at an upstream side of the flow control valves 26a to 26h, from the intra-valve supply fluid line 4a connected to the supply fluid line 2a of the main pump 2.
The flow control valves 26a to 26h independently control the direction and flow rate of the hydraulic fluid supplied from the main pump 2 to the actuators 5 to 12, respectively. The pressure compensating valves 27a to 27h independently control differential pressures existing across the flow control valves 26a to 26h, respectively.
The pressure compensating valves 27a to 27h each include one of valve-opening end pressure receiving portions 28a, 28b, 28c, 28d, 28e, 28f, 28g, and 28h for setting target differential pressures. The output pressure from the differential pressure reducing valve 24 is guided to the pressure receiving portions 28a to 28h, and then a target compensation differential pressure is set according to the particular absolute pressure of the differential pressure PLS between the hydraulic pump pressure Pd and the maximum load pressure PLmax. The absolute differential pressure is hereinafter referred to as the absolute pressure PLS. In this way, each of the individual differential pressures across a corresponding one of the flow control valves 26a to 26h is controlled to equal the same value of differential pressure PLS, so that the pressure compensating valves 27a to 27h provide pressure control to ensure that each differential pressure across the corresponding one of the flow control valves 26a to 26h equals the differential pressure PLS between the hydraulic pump pressure Pd and the maximum load pressure PLmax. Thus during the combined operations control where a plurality of actuators are driven at the same time, the fluid delivery rate of the main pump 2 can be distributed according to a particular opening-area ratio of the flow control valves 26a to 26h, irrespective of a magnitude of the load pressures of the actuators 5 to 12, thereby to provide appropriate combined-operations controllability. In addition, under a saturation state causing the fluid delivery rate of the main pump 2 to fall short of a demanded flow rate, the differential pressure PLS decreases according to a particular degree of the undersupply. Accordingly, each differential pressure across the corresponding one of the flow control valves 26a to 26h controlled by the pressure compensating valves 27a to 27h, respectively, decreases at the same rate and thus the flows of the fluid through the flow control valves 26a to 26h also decrease at the same time. Even under these situations, appropriate combined-operations controllability can be obtained since the fluid delivery rate of the main pump 2 is distributed according to the particular opening-area ratio of the flow control valves 26a to 26.
As can be seen from their symbol representation in
The hydraulic driving system also includes: an engine speed detection valve 30 connected to a supply fluid line 3a of a pilot pump 3 and configured to output an absolute pressure according to a flow rate of the fluid delivered from the pilot pump 3; a pilot hydraulic fluid source 33 with a pilot relief valve 32 connected to a downstream end of the engine speed detection valve 30 and functioning to maintain a constant pressure inside a pilot hydraulic fluid line 31; and operating devices 34a, 34b, 34c, 34d, 34e, 34f, 34g, and 34h, which include, as shown in
The engine speed detection valve 30 includes a restriction element (fixed restrictor) 30f disposed in a fluid line connecting the supply fluid line 3a of the pilot pump 3 to the pilot hydraulic fluid line 31, a flow detection valve 30a connected in parallel to the restriction element 30f, and a differential-pressure reducing valve 30b. The flow detection valve 30a is connected at its inlet side to the supply fluid line 3a of the pilot pump 3, and at its outlet side to the pilot hydraulic fluid line 31. The flow detection valve 30a includes a variable restrictor 30c that increases an opening area of its own as the flow rate of the fluid passing through the restrictor 30c increases. The fluid that has been delivered from the pilot pump 3 flows through both of the restriction element 30f and the variable restrictor 30c of the flow detection valve 30a, and then flows into the pilot hydraulic fluid line 31. At this time, a differential pressure that increases with increases in the flow rate of the passing fluid occurs in the restriction element 30f and in the variable restrictor 30c of the flow detection valve 30a, and the differential-pressure reducing valve 30b outputs the particular differential pressure as an absolute pressure Pa. Since the flow rate of the delivered fluid from the pilot pump 3 changes with the engine speed, detection of both the differential pressure across the restriction element 30f and the differential difference across the variable restrictor 30c allows detection of the fluid delivery rate of the pilot pump 3, and hence, detection of the engine speed. Additionally the fixed restrictor 30c is constructed so that as the flow rate of the passing fluid increases (i.e., as the differential pressure increases), the restrictor increases an opening area of its own, thus rendering an increase rate of the differential pressure more gentle as the flow rate of the passing fluid increases.
The main pump 2 is a variable-displacement type of hydraulic pump, including a pump control unit 35 to control a tilting angle (capacity) of the pump. The pump control unit 35 includes a pump torque controller 35A and a load-sensing (LS) controller 35B.
The pump torque controller 35A includes a torque control tilting actuator 35a, and the torque control tilting actuator 35a drives a swash plate (capacity varying member) 2s of the main pump 2 to reduce the tilting angle (capacity) of the main pump 2 with increases in the fluid delivery pressure of the main pump 2 and limit an input torque of the main pump 2 under a previously set maximum torque value. This control limits horsepower consumption within the main pump 2 and prevents the engine 1 from coming to a stop, or engine stall, due to overload.
The LS controller 35B includes an LS control valve 35b and an LS control tilting actuator 35c.
The LS control valve 35b includes opposed pressure-receiving portions 35d and 35e. The absolute pressure Pa that the differential-pressure reducing valve 30b of the engine speed detection valve 30 has generated is guided as a load-sensing control target differential pressure, or a target LS differential pressure, into the pressure-receiving portion 35d via a fluid line 40. The absolute pressure PLS that the differential-pressure reducing valve 24 has generated (i.e., the differential pressure PLS between the fluid delivery pressure Pd of the main pump 2 and the maximum load pressure PLmax) is guided as a feedback differential pressure into the pressure-receiving portion 35e. As the absolute pressure PLS increases above the absolute pressure Pa (i.e., PLS>Pa), the LS control valve 35b guides the pressure of the pilot hydraulic fluid source 33 to the LS control tilting actuator 35c, and as the absolute pressure PLS decreases below the absolute pressure Pa (i.e., PLS<Pa), the LS control valve 35b makes the LS control tilting actuator 35c communicate with the tank T. Upon receiving the pressure guided from the pilot hydraulic fluid source 33, the LS control tilting actuator 35c drives the swash plate 2s of the main pump 2 to reduce the tilting angle of the main pump 2, and upon being made to communicate with 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. The tilting angle (capacity) of the main pump 2 is thus controlled so that the fluid delivery pressure Pd of the main pump 2 is higher than the maximum load pressure PLmax by the absolute pressure Pa, the target differential pressure.
The absolute pressure Pa here is a value that changes according to the particular engine speed. Use of the absolute pressure Pa as the target differential pressure for load-sensing control, therefore, allows control of an actuator speed appropriate for the engine speed, by setting the target compensation differential pressure of the pressure compensating valves 27a to 27h as per the absolute pressure PLS of the differential pressure between the fluid delivery pressure Pd of the main pump 2 and the maximum load pressure PLmax.
The spring 25a of the unloading valve 25 is set to have a pressure slightly higher than the absolute pressure Pa (target differential pressure for load-sensing control) that the differential-pressure reducing valve 30b of the engine speed detection valve 30 has generated at a rated maximum engine speed.
The operating device 34a includes a control lever 34a-1 and a remote control valve 34a-2, and the remote control valve 34a-2 includes a pair of pressure reducing valves, PVa and PVb. Manipulating the control lever 34a-1 in a rightward direction of
The operating devices 34b to 34h are also constructed similarly to and operate as with the operating device 34a. That is to say, the operating devices 34b to 34h include control levers 34b-1, 34c-1, 34d-1, 34e-1, 34f-1, 34g-1, and 34h-1, respectively, and remote control valves 34b-2, 34c-2, 34d-2, 34e-2, 34f-2, 34g-2, and 34h-2, respectively. Manipulating the control levers 34b-1, 34c-1, 34d-1, 34e-1, 34f-1, 34g-1, and 34h-1 in a rightward direction of
Next, constituent elements characterizing the hydraulic driving system according to the present embodiment are described below.
The hydraulic driving system according to the present embodiment includes control valves 100f, 100g, and 100h, as part of the elements characterizing the system. The control valve 100f is disposed in a parallel fluid line 41f that is a fluid line portion lying at an upstream side of the pressure compensating valve 27f for the boom. The control valve 100g is disposed in a parallel fluid line 41g that is a fluid line portion lying at an upstream side of the pressure compensating valve 27g for the arm. The control valve 100h is disposed in a parallel fluid line 41h that is a fluid line portion lying at an upstream side of the pressure compensating valve 27h for the bucket. The control valves 100f, 100g, and 100h reduce flow passage areas of the parallel fluid lines 41f, 41g, and 41h when the operating devices 34a and 34b for traveling are operated.
The control valves 100f, 100g, and 100h each have a fully open communicating position in which the valve fully opens to communicate, and a restricting position in which the valve reduces an opening area. When no operations are being carried out upon the operating devices 34a and 34b for traveling, the control valves 100f, 100g, and 100h are in their fully open communicating positions shown at left positions of the valves in
The hydraulic driving system according to the present embodiment further includes an operations detector 43 that detects any operations on the operating devices 34a and 34b for traveling. As shown in
Reference symbol Xa in
During slope climbing, when at least one of the boom cylinder 10, the arm cylinder 11, and the bucket cylinder 12 is driven by combined operations control, the difference in load pressure between the track motor 5 or 6 and one of the boom cylinder 10, the arm cylinder 11, and the bucket cylinder 12, becomes particularly significant and the pressure compensating valve of the actuator with the lower load pressure, namely one of the boom cylinder 10, the arm cylinder 11, and the bucket cylinder 12, operates nearly to the stroke end in the direction that the opening area decreases. If saturation occurs during the combined operations control where the difference in load pressure tends to become particularly significant, a large portion of the fluid delivered from the main pump is likely to be absorbed by the actuator lower in load pressure, with the result that the track motor 5 or 6 is likely to stop operating. The actuator that undergoes the higher load pressure during the combined operations control likely to generate the particularly significant difference in load pressure may be hereinafter referred to as the specific actuator. In addition to the track motors, examples of the specific actuator include, as described later herein, a standby actuator provided on an attachment such as a crusher.
First, operation of the basic elements constituting the hydraulic driving system according to the present embodiment is described below.
When the control levers 34a-1 to 34h-1 of all operating devices 34a to 34h are in their neutral positions, all flow control valves 26a to 26h are also in the respective neutral positions and the hydraulic fluid is not supplied to the actuators 5 to 12. Additionally, when all flow control valves 26a to 26h are in the neutral positions, the maximum load pressure PLmax detected by the shuttle valves 22a to 22g will be equal to the tank pressure.
The fluid that has been delivered from the main pump 2 is supplied to the supply fluid lines 2a and 4a, which increases the pressures in the supply fluid lines 2a and 4a. In the supply fluid line 4a is disposed the unloading valve 25, which, when the pressure in the supply fluid line 2a increases by at least the preset pressure of the spring 25a above the maximum load pressure PLmax (in the above case, the tank pressure), opens to return the hydraulic fluid within the supply fluid line 2a to the tank and limit an increase in the internal pressure of the supply fluid line 2a. This controls the fluid delivery pressure of the main pump 2 to the minimum pressure Pmin.
The differential pressure PLS between the fluid delivery pressure of the main pump 2 and the maximum load pressure PLmax is output as the absolute pressure from the differential-pressure reducing valve 24. The output pressure of the engine speed detection valve 30 and that of the differential-pressure reducing valve 24 are guided into the LS control valve 35b of the LS controller 35B within the main pump 2. When the fluid delivery pressure of the main pump 2 increases and the output pressure of the differential-pressure reducing valve 24 increases above that of the engine speed detection valve 30, the LS control valve 35b switches to a position shown as the right position in
When a driven member such as the control lever 34f-1 of the operating device 34f for the boom is operated, the flow control valve 26f for the boom switches, then the hydraulic fluid is supplied to the boom cylinder 10, and the boom cylinder 10 is driven.
The flow rate of the fluid through the flow control valve 26f is dictated by the opening area of the meter-in restrictor of the flow control valve 26f and a differential pressure detected across the meter-in restrictor. The differential pressure across the meter-in restrictor is controlled, by the pressure compensating valve 27, to equal the output pressure of the differential-pressure reducing valve 24. Accordingly the flow rate of the fluid through the flow control valve 26f (hence a driving speed of the boom cylinder 10) is controlled according to the particular amount of operation of the control lever.
Meanwhile, the load pressure upon the boom cylinder 10 is detected as a maximum load pressure by a corresponding one of the shuttle valves 22a to 22g, and then transmitted to the differential-pressure reducing valve 24 and the unloading valve 25.
When the load pressure of the boom cylinder 10 is guided into the unloading valve 25 as the maximum load pressure, the unloading valve 25 correspondingly raises a cracking pressure, or a pressure at which the unloading valve 25 begins to open, and then when the pressure in the supply fluid line 2a temporarily be higher by at least the preset pressure of the spring 25a than the maximum load pressure, the unloading valve 25 opens to return the hydraulic fluid within the supply fluid line 4a to the tank. Thus the pressure in the supply fluid lines 2a and 4a is controlled to be not higher, by the preset pressure set for the spring 25a, than the maximum load pressure PLmax.
Once the boom cylinder 10 has begun to operate, the pressure in the supply fluid lines 2a and 4a temporarily decreases. At this time, the output pressure of the differential-pressure reducing valve 24 also decreases since the difference in load pressure between the pressure of the supply fluid line 2a and the load pressure of the boom cylinder 10 is output as the output pressure of the differential-pressure reducing valve 24.
The output pressure of the engine speed detection valve 30 and that of the differential-pressure reducing valve 24 are introduced into the LS control valve 35b of the LS controller 35B of the main pump 2, and when the output pressure of the differential-pressure reducing valve 24 decreases below that of the engine speed detection valve 30, the LS control valve 35b switches to a position shown as the left position in
When at least two driven members, for example the control levers 34f-1 and 34g-1 of the operating device 34f for the boom and the operating device 34g for the arm are operated, the flow control valves 26f and 26g both switch, then the hydraulic fluid is supplied to the boom cylinder 10 and the arm cylinder 11, and the boom cylinder 10 and the arm cylinder 11 are driven.
Of the load pressures in the boom cylinder 10 and the arm cylinder 11, the higher pressure is detected as the maximum load pressure PLmax by the shuttle valves 22a to 22g and transmitted to the differential-pressure reducing valve 24 and the unloading valve 25.
The way the unloading valve 25 operates in this case when the maximum load pressure PLmax that the shuttle valves 22a to 22g have detected is guided to the unloading valve 25 is the same as developed when the boom cylinder 10 is driven independently. In other words, as the maximum load pressure PLmax increases, the cracking pressure of the unloading valve 25 also increases and the pressure in the supply fluid lines 2a and 4a is controlled to be not higher than the maximum load pressure PLmax by the preset pressure for the spring 25a.
The output pressure of the engine speed detection valve 30 and that of the differential-pressure reducing valve 24 are also introduced into the LS control valve 35b of the LS controller 35B of the main pump 2. In this case, as in the case with the independent driving of the boom cylinder 10, so-called load-sensing control is performed. That is to say, the fluid delivery pressure of the main pump 2 (i.e., the pressure in the supply fluid lines 2a and 4a) is controlled to be higher, by the output pressure of the engine speed detection valve 30 (i.e., the target differential pressure), than the maximum load pressure PLmax, and the fluid is supplied to the boom cylinder 10 and the arm cylinder 11 at the flow rates demanded from the flow control valves 26f and 26g.
The output pressure of the differential-pressure reducing valve 24 is introduced into the pressure compensating valves 27a to 27h as the target compensation differential pressure, and the pressure compensating valves 27f and 27g each control the differential pressure across the corresponding one of the flow control valves 26f and 26g respectively to equal the differential pressure between the fluid delivery pressure of the main pump 2 and the maximum load pressure PLmax. With this control, irrespective of the magnitude of the load pressures of the boom cylinder 10 and the arm cylinder 11, the hydraulic fluid can be supplied to the boom cylinder 10 and the arm cylinder 11 at a ratio commensurate with a meter-in restrictor opening area ratio between the flow control valves 26f and 26g.
At this time, if saturation occurs, in other words, if the flow rate of the fluid delivered from the main pump 2 does not satisfy the flow rate demands of the flow control valves 26f and 26g, the output pressure of the differential-pressure reducing valve 24 (i.e., the differential pressure between the fluid delivery pressure of the main pump 2 and the maximum load pressure PLmax) decreases according to a particular degree of the saturation. The decrease in the output pressure of the differential-pressure reducing valve 24 correspondingly reduces the target compensation differential pressures of the pressure compensating valves 27a to 27h, thus enabling the delivery flow rate of the hydraulic fluid from the main pump 2 to be redistributed to the ratio of the flow rates demanded from the flow control valves 26f and 26g.
The pressure compensating valves 27a to 27h are each constructed so that they do not fully close at the stroke end of the valve as operated in the direction that the opening area decreases. In addition to the above favorable effects, therefore, during the combined operations control where one of the boom cylinder 10 and the arm cylinder 11 is operated with the other being used, even if saturation occurs and the pressure compensating valve lower in load pressure operates through a long stroke in the direction that the opening area decreases, full closing of the pressure compensating valve lower in load pressure is prevented, which in turn prevents complete shutoff of the hydraulic fluid. Hence a slowdown and stop of the actuator with the lower load pressure can be prevented.
The operation described above applies when the engine 1 rotates at its maximum rated speed. On the other hand, when the engine speed is reduced, since the output pressure of the engine speed detection valve 30 correspondingly decreases, the LS control valve 35b of the LS controller 35B likewise decreases in target differential pressure. The pressure compensating valves 27a to 27h also experience a similar decrease in target compensation differential pressure after load-sensing control. Thus as the engine speed decreases, both the flow rate of the delivered fluid from the main pump 2 and the flow rates demanded from the flow control valves 26a to 26h decrease, which then enables the driving speeds of the actuators 5 to 12 to be appropriately maintained and fine (microscopic) operability/controllability at reduced engine speeds to be improved.
The following describes operation of the characteristic elements constituting the hydraulic driving system of the present embodiment.
When the control levers 34a-1 and 34b-1 of the operating devices 34a and 34b for traveling are operated, the flow control valves 26a and 26b both switch as in the combined operations control described above, and thereby the hydraulic fluid is supplied to the track motors 5 and 6. Additionally, the fluid delivery flow rate of the main pump 2 is controlled by load-sensing control, the fluid is supplied to the track motors 5 and 6 at the flow rates demanded from the flow control valves 26a and 26b, and the hydraulic excavator travels.
During traveling, when either the boom, the arm, or the bucket, for example the control lever 34g-1 of the operating device 34g for the arm is operated for a change in the posture of the front working implement, the flow control valve 26g switches, thereby the hydraulic fluid is also supplied to the arm cylinder 11 and the arm cylinder 11 is driven.
In a conventional system configuration with pressure compensating valves each of a type not fully closing at a stroke end of the valve as operated in a direction that an operating area decreases, during traveling control with a driven member, when another driven member (e.g., a boom, an arm, or a bucket) is operated, under the conditions that involve a high traveling load pressure particularly for climbing a slope, a pressure compensating valve of an actuator lower in load pressure than track motors, such as a boom cylinder, arm cylinder, or bucket cylinder, is open even after reaching the stroke end. A flow rate of a fluid delivered from a hydraulic pump may therefore be drawn into the actuator lower in load pressure, with the result that traveling may slow down and/or stop.
In contrast to the above conventional system configuration, in the present embodiment, when a full-stroke operation is being carried out upon the control lever 34a-1 or 34b-1 of the operating device 34a or 34b for traveling and the operating pilot pressure for traveling is being generated, the control valve 100f, 100g, or 100h switches to the restricting position shown as the right position in
On the other hand, the combined operations control for traveling along a level ground surface is usually conducted at low speeds and the load pressure upon the track motors 5 and 6 is usually not too high. Even during the low-speed combined operations control for traveling, when the control lever 34a-1 or 34b-1 of the operating device 34a or 34b for traveling is operated and the control valve 100f, 100g, or 100h switches to the restricting position, the flow rate of the fluid supplied to the boom cylinder 10, the arm cylinder 11, or the bucket cylinder 12 might be suppressed despite a low possibility that a large portion of the fluid delivered from the main pump 2 would be absorbed by the actuator having the lower load pressure. Operation of the front working implement 302 might consequently slow down to reduce working efficiency.
In the present embodiment, as described above, the amount of control lever operation, Xa, of the control valve 100f, 100g, or 100h is set to be a value close to ‘Full’, the maximum achievable operating stroke of the control lever. During the low-speed combined operations control for traveling on a level ground surface, therefore, when the control lever 34a-1 or 34b-1 of the operating device 34a or 34b for traveling is operated, the control valve 100f, 100g, or 100h does not switch to the restricting position. For this reason, the flow rate of the hydraulic fluid supplied to the boom cylinder 10, the arm cylinder 11, or the bucket cylinder 12 will not be suppressed. This will slow down the operation of the front working implement 302 and hence prevent working efficiency from decreasing.
As set forth above, in the present embodiment, even if saturation occurs during the combined operations control likely to generate a significant difference in load pressure between any two actuators, full closing of the pressure compensating valve lower in load pressure is prevented, which in turn prevents a slowdown and stop of the actuator lower in load pressure. In addition, during the combined operations control for traveling that includes the driving of the track motor 5 or 6 as the specific actuator, a flow of the hydraulic fluid into the boom cylinder 10, the arm cylinder 11, or the bucket cylinder 12 is suppressed, the necessary amount of hydraulic fluid is supplied to the track motor 5 or 6, and a slowdown and stop of traveling is prevented. The combined operations controllability for traveling can therefore be enhanced.
Furthermore, since the amount of control lever operation, Xa, of the control valve 100f, 100g, or 100h is set to be a value close to ‘Full’, the maximum achievable operating stroke of the control lever, the operation of the front working implement 302 is prevented from slowing down during the low-speed combined operations control for traveling on a level ground surface. As a result, working efficiency can be prevented from decreasing.
Moreover, the control valves 100f, 100g, and 100h are arranged in the parallel fluid lines 41f, 41g, and 41h. Thus, when the control lever 34a-1 or 34b-1 of the operating device 34a or 34b for traveling is operated, the flow rate of the hydraulic fluid supplied only to the actuator corresponding to the parallel fluid line 41f, 41g, or 41h (i.e., the boom cylinder 10, the arm cylinder 11, or the bucket cylinder 12) will be suppressed and the flow rates of the hydraulic fluid supplied to the other actuators will not be suppressed. During the combined operations control for driving the track motor 5 or 6 concurrently with any other actuator, reduction in operability/controllability due to a decrease in a speed of the other actuator can be prevented.
A hydraulic driving system according to a second embodiment of the present invention is shown in
More specifically, whereas the first embodiment shown in
The control valve 100, as with the control valves 100f, 100g, and 100h, has two positions, namely a fully open communicating position in which the valve fully opens to communicate, and a restricting position in which the valve reduces an opening area. When no operations are being carried out upon the operating devices 34a and 34b for traveling, the control valve 100 is in the fully open communicating position shown as an left position of the valve in
In the present embodiment of the above configuration, the operating pilot pressure for traveling is also generated when the operating device 34a or 34b for traveling is operated through a full stroke. The control valve 100 then switches to the restricting position shown as the lower position in
As set forth above, substantially the same advantageous effects as those of the first embodiment can also be obtained in the second embodiment.
In the present embodiment, since the flow rates of the hydraulic fluid supplied to a plurality of actuators are also suppressed with one control valve 100, the advantageous effects described above can be obtained, thus the number of constituent parts needed can be reduced, and the effects can be obtained less expensively.
A hydraulic driving system according to a third embodiment of the present invention is shown in
More specifically, the hydraulic driving system in the third embodiment includes solenoid-operated control valves 46f, 46g, and 46h instead of the hydraulic control valves 100f, 100g, and 100h in the first embodiment. The hydraulic driving system also includes a controller 71. The hydraulic driving system further includes an operations detector 43A having, in addition to the shuttle valves 48a, 48b, and 48c shown in
When the operating devices 34a and 34b for traveling are not operated and the driving signal is not output from the controller 71, the solenoid-operated control valves 46f, 46g, and 46h are in their communicating positions shown as left positions of the valves in
Hence, substantially the same advantageous effects as those of the first embodiment can also be obtained in the third embodiment.
The present embodiment employs solenoid-operated control valves as a substitute for the control valves 100f, 100g, and 100h in the first embodiment. However, if a further solenoid-operated control valve is employed instead by the control valve 100 in
A hydraulic driving system according to a fourth embodiment of the present invention is shown in
More specifically, the hydraulic driving system in the fourth embodiment additionally includes a manual selector 81 adapted to be switched between its first position and its second position. The manual selector 81 is, for example, a switch that will output an appropriate electrical signal according to the switching position selected. The present embodiment further includes a solenoid-operated control valve 83 disposed in a fluid line 48 to guide the hydraulic signal detected by the operations detector 43 beforehand to the pressure receiving portions 101f, 101g, and 101h of the control valves 100f, 100g, and 100h. The solenoid-operated control valve 83 operates in accordance with the electrical signal output from the manual selector (manual switch) 81.
When the manual selector 81 is in the first position and the electrical signal is not output, the solenoid-operated control valve 83 is in a first position shown as a lower position of the valve in
Thus when the manual selector 81 is in the first position, the control valves 100f, 100g, and 100h activate the respective functions to reduce the flow passage areas of the parallel fluid lines 41f, 41g, and 41h in response to the operation of specific operating devices 34a and 34b for traveling. As in the above mentioned embodiments, therefore, supply of the hydraulic fluid to the boom cylinder 10, the arm cylinder 11, and the bucket cylinder 12 can be suppressed via the control valves 100f, 100g, and 100h during the combined operations control for traveling. When the manual selector 81 is switched to the second position, the control valves 100f, 100g, and 100h deactivate the respective functions of reducing the flow passage areas of the parallel fluid lines 41f, 41g, and 41h in response to the operation of the specific operating devices 34a and 34b for traveling. Even during the combined operations control for traveling, therefore, the suppression of the supply of the hydraulic fluid to the boom cylinder 10, the arm cylinder 11, and the bucket cylinder 12 is deactivated, which then enables substantially the same operation as achievable in conventional system configurations.
In the present embodiment having the above configuration, an operator can freely select whether to use a specific function of the present invention according to his or her needs or preference.
A hydraulic driving system according to a fifth embodiment of the present invention is shown in
More specifically, whereas the first embodiment shown in
The control valve 100d, as with the control valves 100f, 100g, and 100h, has two positions, namely a fully open communicating position and a restricting position in which the valve reduces an opening area. When no operations are being carried out upon the operating device 34a or 34b for traveling, the control valve 100d is in the fully open communicating position shown as a left position of the valve in
In the conventional system configurations with a pressure compensating valve of the type where the valve does not fully close at the stroke end in the opening-area reduction direction even during abrupt operation of a blade-operating device for traveling 34d, a possible instantaneous or momentary flow of the hydraulic fluid into the blade cylinder 8 may lead to a slowdown of traveling, hence causing a bodily sensory shock to the operator, and undermining his or her operation feeling.
In the present embodiment, however, as is the case in which the control lever of either the boom, the arm, or the bucket is operated for the front operation during traveling, since the control valve 100d suppresses the flow rate of the hydraulic fluid supplied to the blade cylinder 8, a necessary amount of hydraulic fluid is reliably supplied to the track motor 5 or 6 and a slowdown of traveling is prevented. An operation feeling can therefore be improved.
A hydraulic driving system according to a sixth embodiment of the present invention is shown in
In the second embodiment of
In the present embodiment of the above configuration, when the operating device 34a or 34b for traveling is operated through a full stroke, the operating pilot pressure for traveling is generated, whereby the control valve 100A switches to a restricting position shown as a lower position in
The embodiments that have been described above may each be changed and modified in various forms without departing from the spirit and scope of the present invention.
For example, in each embodiment of the present invention, the control valves that reduce the flow passage areas of the fluid line portions during operations on specific operating devices are employed and these control valves (e.g., 100f, 100g, and 100h) each have a fully open communicating position and a restricting position for reducing the opening area of the valve. Each of the control valves is constructed so that when no operations are being carried out upon the operating device 34a or 34b for traveling, the control valve is in the fully open communicating position, and so that when the operating device 34a or 34b for traveling is operated, the control valve is switched to the restricting position to reduce the flow passage area of the corresponding fluid line portion. This construction of the control valves, however, is not always limited.
In addition, while each embodiment of the present invention has been described taking a track motor as an example of a specific actuator, substantially the same advantageous effects can likewise be obtained for elements other than the track motor. To be more specific, substantially the same advantageous effects can likewise be obtained by applying the present invention to a hydraulic driving system having pressure compensating valves of a type not closing at a stroke end of the valve as operated in a direction to reduce its opening area, the system further having actuators including an actuator likely to stop operating if, during the combined operations control likely to generate a particularly significant difference in load pressure between any two actuators, saturation occurs and a large portion of the delivered fluid from the main pump is absorbed by the actuator with the lower load pressure. For example, since a load pressure upon a standby actuator provided on an attachment such as a crusher tends to increase, if the present invention is applied with a standby actuator as a specific actuator, then during the combined operations control where the standby actuator is driven simultaneously with actuators other than the specific actuator (e.g., the boom, the arm, or the bucket), the flow rate demanded from each of the actuators other than the specific actuator can be limited and the hydraulic fluid can be supplied to the standby actuator preferentially.
Furthermore, while each embodiment of the present invention has been described taking a hydraulic excavator as an example of a construction machine, substantially the same advantageous effects can likewise be obtained by applying the invention to other construction machines such as a hydraulic crane or wheeled excavator.
| Number | Date | Country | Kind |
|---|---|---|---|
| 2012-230071 | Oct 2012 | JP | national |
| Filing Document | Filing Date | Country | Kind |
|---|---|---|---|
| PCT/JP2013/077364 | 10/8/2013 | WO | 00 |
