Patent Application
20160265561
The present invention relates to a hydraulic drive system for a construction machine such as a hydraulic excavator. In particular, the present invention relates to a hydraulic drive system for a construction machine having at least two variable displacement hydraulic pumps in which one of the hydraulic pumps includes a pump control unit (regulator) for performing at least torque control and another one of the hydraulic pumps includes a pump control unit (regulator) for performing load sensing control and torque control.
In hydraulic drive systems for construction machines such as hydraulic excavators, widely used today are those equipped with a regulator for controlling the displacement (flow rate) of a hydraulic pump such that the delivery pressure of the hydraulic pump becomes higher by a target differential pressure than the maximum load pressure of a plurality of actuators. This type of control is called “load sensing control.” Such a hydraulic drive system for a construction machine equipped with a regulator for performing the load sensing control is described in Patent Document 1, in which a two-pump load sensing system including two hydraulic pumps each designed to perform the load sensing control is described.
The regulator of a hydraulic drive system for a construction machine performs torque control such that the absorption torque of a hydraulic pump does not exceed the rated output torque of the prime mover and prevents stoppage of the prime mover caused by excessive absorption torque (engine stall), generally by decreasing the displacement of the hydraulic pump as the delivery pressure of the hydraulic pump increases. In cases where the hydraulic drive system is equipped with two hydraulic pumps, the regulator of one hydraulic pump performs the torque control by taking in not only the delivery pressure of its own hydraulic pump but also a parameter regarding the absorption torque of the other hydraulic pump (total torque control) in order to prevent the stoppage of the prime mover and efficiently utilize the rated output torque of the prime mover.
For example, in Patent Document 2, the total torque control is performed by leading the delivery pressure of one hydraulic pump to the regulator of the other hydraulic pump via a pressure reducing valve. The set pressure of the pressure reducing valve is constant and has been set at a value simulating the maximum torque of the torque control performed by the regulator of the other hydraulic pump. With these features, in work in which only one or more actuators related to the one hydraulic pump are driven, the one hydraulic pump can efficiently use almost all of the rated output torque of the prime mover. Further, in work with a combined operation in which an actuator related to the other hydraulic pump is also driven at the same time, the total absorption torque of the pumps does not exceed the rated output torque of the prime mover and the stoppage of the prime mover can be prevented.
In Patent Document 3, in order to perform the total torque control on two hydraulic pumps of the variable displacement type, the tilting angle of the other hydraulic pump is detected as output pressure of a pressure reducing valve, and the output pressure is led to the regulator of the one hydraulic pump. In Patent Document 4, control precision of the total torque control is increased by detecting the arm length of a pivoting arm in place of the tilting angle of the other hydraulic pump.
Patent Document 1: JP-2011-196438-A
Patent Document 2: Japanese Patent No. 3865590
Patent Document 3: JP-1991-007030-B
Patent Document 4: JP-1995-189916-A
The total torque control becomes possible also in the two-pump load sensing system described in Patent Document 1 by incorporating the technology of the total torque control described in Patent Document 2 into the two-pump load sensing system of Patent Document 1. However, in the total torque control in Patent Document 2, the set pressure of the pressure reducing valve has been set at a constant value simulating the maximum torque of the torque control of the other hydraulic pump as mentioned above. Accordingly, the efficient use of the rated output torque of the prime mover can be achieved when the other hydraulic pump is in an operational state of undergoing the limitation by the torque control and operating at the maximum torque of the torque control in the combined operation in which actuators related to the two hydraulic pumps are driven at the same time. However, when the other hydraulic pump is in an operational state of not undergoing the limitation by the torque control and performing the displacement control by means of the load sensing control, even though the absorption torque of the other hydraulic pump is lower than the maximum torque of the torque control, the output pressure of the pressure reducing valve simulating the maximum torque is led to the regulator of the one hydraulic pump and the absorption torque of the one hydraulic pump is erroneously controlled to decrease more than necessary. Thus, it has been impossible to perform the total torque control with high precision.
The technology of Patent Document 3 attempts to increase the precision of the total torque control by detecting the tilting angle of the other hydraulic pump as the output pressure of the pressure reducing valve and leading the output pressure to the regulator of the one hydraulic pump. However, differently from the common method of calculating the torque of a pump as the product of the delivery pressure and the displacement, namely, (delivery pressure x pump displacement)/2π, the system of Patent Document 3 leads the delivery pressure of the one hydraulic pump to one of two pilot chambers of a stepped piston, leads the output pressure of the pressure reducing valve (delivery rate-proportional pressure of the other hydraulic pump) to the other pilot chamber of the stepped piston, and controls the displacement of the one hydraulic pump by using the sum of the delivery pressure and the delivery rate-proportional pressure as the parameter of the output torque. Thus, the technology of Patent Document 3 has a problem in that a considerably great error occurs between the calculated torque and the actually used torque.
In Patent Document 4, the control precision of the total torque control is increased by detecting the arm length of the pivoting arm in place of the tilting angle of the other hydraulic pump. However, the regulator in Patent Document 4 has extremely complex structure in which the pivoting arm and a piston arranged in a regulator piston relatively slide with each other while transmitting force. Thus, in order to make a structure having sufficient durability, components such as the pivoting arm and the regulator piston have to be strengthened and the downsizing of the regulator becomes difficult. Especially in small-sized hydraulic excavators whose rear end radius is small, that is, hydraulic excavators of the so-called small tail swing radius type, the space for storing the hydraulic pumps is small and the installation is difficult in some cases.
The object of the present invention is to provide a hydraulic drive system for a construction machine including at least two variable displacement hydraulic pumps, in which one of the hydraulic pumps includes a pump control unit for performing at least the torque control and the other hydraulic pumps performs the load sensing control and the torque control, capable of efficiently utilizing the rated output torque of the prime mover by performing the total torque control with high precision through precise detection of the absorption torque of the other hydraulic pump by use of a purely hydraulic structure and feedback of the absorption torque to the one hydraulic pump's side.
(1) To achieve the above object, the present invention provides a hydraulic drive system for a construction machine that includes: a prime mover; a first hydraulic pump of a variable displacement type driven by the prime mover; a second hydraulic pump of the variable displacement type driven by the prime mover; a plurality of actuators driven by a hydraulic fluid delivered by the first and second hydraulic pumps; a plurality of flow control valves that control flow rates of the hydraulic fluid supplied from the first and second hydraulic pumps to the actuators; a plurality of pressure compensating valves each of which controls a differential pressure across a corresponding one of the flow control valves; a first pump control unit that controls a delivery flow rate of the first hydraulic pump; and a second pump control unit that controls a delivery flow rate of the second hydraulic pump. The first pump control unit includes a first torque control section that controls a displacement of the first hydraulic pump in such a manner that an absorption torque of the first hydraulic pump does not exceed a first maximum torque when at least one of a delivery pressure and the displacement of the first hydraulic pump increases and the absorption torque of the first hydraulic pump increases. The second pump control unit includes: a second torque control section that controls a displacement of the second hydraulic pump in such a manner that an absorption torque of the second hydraulic pump does not exceed a second maximum torque when at least one of a delivery pressure and the displacement of the second hydraulic pump increases and the absorption torque of the second hydraulic pump increases; and a load sensing control section that controls the displacement of the second hydraulic pump in such a manner that the delivery pressure of the second hydraulic pump becomes higher by a target differential pressure than a maximum load pressure of the actuators driven by the hydraulic fluid delivered by the second hydraulic pump when the absorption torque of the second hydraulic pump is lower than the second maximum torque. The first torque control section includes: a first torque control actuator that is supplied with the delivery pressure of the first hydraulic pump and controls the displacement of the first hydraulic pump so as to decrease the displacement of the second hydraulic pump and thereby decrease the absorption torque of the second hydraulic pump when the delivery pressure rises; and first biasing means that sets the first maximum torque. The second torque control section includes: a second torque control actuator that is supplied with the delivery pressure of the second hydraulic pump and controls the displacement of the second hydraulic pump so as to decrease the displacement of the second hydraulic pump and thereby decrease the absorption torque of the second hydraulic pump when the delivery pressure rises; and second biasing means that sets the second maximum torque. The load sensing control section includes: a control valve that changes load sensing drive pressure in such a manner that the load sensing drive pressure decreases as a differential pressure between the delivery pressure of the second hydraulic pump and the maximum load pressure decreases below the target differential pressure; and a load sensing control actuator that controls the displacement of the second hydraulic pump so as to increase the displacement of the second hydraulic pump and thereby increase the delivery flow rate of the second hydraulic pump as the load sensing drive pressure decreases. The first pump control unit further includes: a torque feedback circuit that is supplied with the delivery pressure of the second hydraulic pump and the load sensing drive pressure, modifies the delivery pressure of the second hydraulic pump based on the delivery pressure of the second hydraulic pump and the load sensing drive pressure to achieve a characteristic simulating the absorption torque of the second hydraulic pump in both of when the second hydraulic pump undergoes a limitation by the control by the second torque control section and operates at the second maximum torque and when the second hydraulic pump does not undergo the limitation by the control by the second torque control section and the load sensing control section controls the displacement of the second hydraulic pump, and outputs the modified pressure; and a third torque control actuator that is supplied with an output pressure of the torque feedback circuit and controls the displacement of the first hydraulic pump so as to decrease the displacement of the first hydraulic pump and thereby decrease the first maximum torque as the output pressure of the torque feedback circuit increases.
In the present invention configured as above, not only when the second hydraulic pump (the other hydraulic pump) is in an operational state of undergoing the limitation by the torque control and operating at the second maximum torque of the torque control but also when the second hydraulic pump is in an operational state of not undergoing the limitation by the torque control and performing the displacement control by means of the load sensing control, the delivery pressure of the second hydraulic pump is modified by the torque feedback circuit to achieve a characteristic simulating the absorption torque of the second hydraulic pump, and the first maximum torque is modified by the third torque control actuator to decrease by an amount corresponding to the modified delivery pressure. With such features, the absorption torque of the second hydraulic pump is detected precisely by use of a purely hydraulic structure (torque feedback circuit). By feeding back the absorption torque to the first hydraulic pump's side (the one hydraulic pump's side), the total torque control can be performed precisely and the rated output torque of the prime mover can be utilized efficiently.
(2) Preferably, in the above hydraulic drive system (1), the torque feedback circuit includes a variable pressure reducing valve that is supplied with the delivery pressure of the second hydraulic pump, outputs the delivery pressure of the second hydraulic pump without change when the delivery pressure of the second hydraulic pump is lower than or equal to a set pressure, and reduces the delivery pressure of the second hydraulic pump to the set pressure and outputs the reduced pressure when the delivery pressure of the second hydraulic pump is higher than the set pressure. The variable pressure reducing valve includes a pressure receiving part that is also supplied with the load sensing drive pressure of the load sensing control section and decreases the set pressure as the load sensing drive pressure increases.
When a hydraulic pump performs the displacement control by means of the load sensing control, the position of a displacement changing member (swash plate) of the hydraulic pump, that is, the displacement (tilting angle) of the hydraulic pump, is determined by the equilibrium between resultant force of two pushing forces applied to the displacement changing member from a load sensing control actuator (LS control piston) on which the load sensing drive pressure acts and from a torque control actuator (torque control piston) on which the delivery pressure of the hydraulic pump acts and pushing force applied to the displacement changing member in the opposite direction from biasing means (spring) used for setting the maximum torque (
In the present invention, the torque feedback circuit is equipped with the variable pressure reducing valve and is configured such that the set pressure of the variable pressure reducing valve decreases as the load sensing drive pressure increases. Therefore, the maximum value of the output pressure of the torque feedback circuit (the delivery pressure of the second hydraulic pump via the variable pressure reducing valve) at times of increase in the delivery pressure of the second hydraulic pump changes so as to decrease as the load sensing drive pressure increases (
(3) Preferably, in the above hydraulic drive system (2), the torque feedback circuit further includes a first pressure dividing circuit including: a first fixed restrictor to which the delivery pressure of the second hydraulic pump is led; and a pressure control valve situated downstream of the first fixed restrictor and connected to a tank on a downstream side. The first pressure dividing circuit outputs pressure in a hydraulic line between the first fixed restrictor and the pressure control valve. The pressure control valve is configured such that the load sensing drive pressure of the load sensing control section is supplied to the pressure control valve and the pressure in the hydraulic line between the first fixed restrictor and the pressure control valve decreases as the load sensing drive pressure increases. The pressure in the hydraulic line between the first fixed restrictor and the pressure control valve is led to the variable pressure reducing valve as the delivery pressure of the second hydraulic pump.
As mentioned above, the ratio of increase of the absorption torque of a hydraulic pump at times of increase in the delivery pressure of the hydraulic pump decreases as the load sensing drive pressure increases.
In the present invention, the torque feedback circuit is equipped with the first pressure dividing circuit including the pressure control valve and is configured such that the output pressure of the first pressure dividing circuit decreases as the load sensing drive pressure increases. Therefore, the ratio of increase of the output pressure of the torque feedback circuit (output pressure of the first pressure dividing circuit) at times of increase in the delivery pressure of the second hydraulic pump changes so as to decrease as the load sensing drive pressure increases (
(4) Preferably, in the above hydraulic drive system (3), the pressure control valve is a variable restrictor valve configured such that an opening area thereof varies and increases as the load sensing drive pressure increases.
With such features, the ratio of increase of the output pressure of the torque feedback circuit at times of increase in the delivery pressure of the second hydraulic pump is modified so as to decrease as the load sensing drive pressure increases.
(5) Preferably, in the above hydraulic drive system (3), the pressure control valve is a variable relief valve configured such that a relief set pressure thereof decreases as the load sensing drive pressure increases.
Also with such features, the ratio of increase of the output pressure of the torque feedback circuit at times of increase in the delivery pressure of the second hydraulic pump is modified so as to decrease as the load sensing drive pressure increases.
(6) Preferably, in the above hydraulic drive system (3), the torque feedback circuit further includes: a second pressure dividing circuit including: a second fixed restrictor to which the delivery pressure of the second hydraulic pump is led; and a third fixed restrictor situated downstream of the second fixed restrictor and connected to the tank on the downstream side, the second pressure dividing circuit outputting a pressure in a hydraulic line between the second fixed restrictor and the third fixed restrictor; and a higher pressure selection valve that selects higher one of an output pressure of the variable pressure reducing valve and an output pressure of the second pressure dividing circuit and outputs the selected pressure. Output pressure of the higher pressure selection valve is led to the third torque control section.
Each hydraulic pump has a minimum displacement that is determined by the structure of the hydraulic pump. When the hydraulic pump is at the minimum displacement, the absorption torque of the hydraulic pump at times of increase in the delivery pressure of the hydraulic pump increases at the smallest gradient (ratio of increase) (
In the present invention, by setting the output characteristic of the second pressure dividing circuit to be identical with the output characteristic of the first pressure dividing circuit supplied with the load sensing drive pressure that sets the second hydraulic pump at its minimum displacement (i.e., making the setting such that the opening area of the second fixed restrictor is equal to that of the first fixed restrictor and the throttling characteristic of the third fixed restrictor is identical with that of the pressure control valve supplied with the load sensing drive pressure that sets the second hydraulic pump at the minimum displacement), when the second hydraulic pump is at the minimum displacement, the output pressure of the second pressure dividing circuit is selected by the higher pressure selection and the pressure is outputted as the output pressure of the torque feedback circuit in the entire delivery pressure range of the second hydraulic pump.
Further, by setting the opening areas of the second and third fixed restrictor in conformity with the minimum ratio of increase of the absorption torque with the increase in the delivery pressure of the second hydraulic pump at times when the second hydraulic pump is at the minimum displacement, the output pressure of the second pressure dividing circuit takes on a characteristic of proportionally increasing at the minimum ratio of increase as the delivery pressure of the second hydraulic pump increases (
Furthermore, with such features, the total torque consumption of the first hydraulic pump and the second hydraulic pump does not become excessive and the stoppage of the prime mover can be prevented in combined operations of an actuator related to the first actuator and an actuator related to the second hydraulic pump in which the load pressure of the actuator related to the second hydraulic pump becomes high and the demanded flow rate is extremely low (e.g., combined operation of boom raising fine operation and swing operation or arm operation in load lifting work).
According to the present invention, not only when the second hydraulic pump (the other hydraulic pump) is in the operational state of undergoing the limitation by the torque control and operating at the second maximum torque of the torque control but also when the second hydraulic pump is in the operational state of not undergoing the limitation by the torque control and performing the displacement control by means of the load sensing control, the delivery pressure of the second hydraulic pump is modified by the torque feedback circuit to achieve a characteristic simulating the absorption torque of the second hydraulic pump, and the first maximum torque is modified by the third torque control actuator to decrease by an amount corresponding to the modified delivery pressure. With such features, the absorption torque of the second hydraulic pump is detected precisely by use of a purely hydraulic structure (torque feedback circuit). By feeding back the absorption torque to the first hydraulic pump's side (the one hydraulic pump's side), the total torque control can be performed precisely and the rated output torque of the prime mover can be utilized efficiently.
Referring now to the drawings, a description will be given in detail of preferred embodiments of the present invention.
Referring to
The control valve unit 4 includes flow control valves 6a, 6b, 6c, 6d, 6e, 6f, 6g, 6h, 6i and 6j, pressure compensating valves 7a, 7b, 7c, 7d, 7e, 7f, 7g, 7h, 7i and 7j, operation detection valves 8a, 8b, 8c, 8d, 8f, 8g, 8i and 8j, main relief valves 114, 214 and 314, and unloading valves 115, 215 and 315. The flow control valves 6a, 6b, 6c, 6d, 6e, 6f, 6g, 6h, 6i and 6j are connected to the first through third hydraulic fluid supply lines 105, 205 and 305 and control the flow rates of the hydraulic fluid supplied to the actuators 3a- 3h from the first and second delivery ports 102a and 102b of the main pump 102 and the third delivery port 202a of the main pump 202. Each pressure compensating valve 7a-7j controls the differential pressure across a corresponding flow control valve 6a- 6j such that the differential pressure becomes equal to a target differential pressure. Each operation detection valve 8a, 8b, 8c, 8d, 8f, 8g, 8i, 8j strokes together with the spool of a corresponding one of the flow control valves 6a-6j in order to detect the switching of the flow control valve. The main relief valve 114 is connected to the first hydraulic fluid supply line 105 and controls the pressure in the first hydraulic fluid supply line 105 such that the pressure does not reach or exceed a set pressure. The main relief valve 214 is connected to the second hydraulic fluid supply line 205 and controls the pressure in the second hydraulic fluid supply line 105 such that the pressure does not reach or exceed a set pressure. The main relief valve 314 is connected to the third hydraulic fluid supply line 305 and controls the pressure in the third hydraulic fluid supply line 305 such that the pressure does not reach or exceed a set pressure. The unloading valve 115 is connected to the first hydraulic fluid supply line 105. When the pressure in the first hydraulic fluid supply line 105 becomes higher than a pressure (unloading valve set pressure) defined as the sum of the maximum load pressure of the actuators driven by the hydraulic fluid delivered from the first delivery port 102a and a set pressure (prescribed pressure) of its own spring, the unloading valve 115 shifts to the open state and returns the hydraulic fluid in the first hydraulic fluid supply line 105 to a tank. The unloading valve 215 is connected to the second hydraulic fluid supply line 205. When the pressure in the second hydraulic fluid supply line 205 becomes higher than a pressure (unloading valve set pressure) defined as the sum of the maximum load pressure of the actuators driven by the hydraulic fluid delivered from the second delivery port 102b and a set pressure (prescribed pressure) of its own spring, the unloading valve 215 shifts to the open state and returns the hydraulic fluid in the second hydraulic fluid supply line 205 to the tank. The unloading valve 315 is connected to the third hydraulic fluid supply line 305. When the pressure in the third hydraulic fluid supply line 305 becomes higher than a pressure (unloading valve set pressure) defined as the sum of the maximum load pressure of the actuators driven by the hydraulic fluid delivered from the third delivery port 202a and a set pressure (prescribed pressure) of its own spring, the unloading valve 315 shifts to the open state and returns the hydraulic fluid in the third hydraulic fluid supply line 305 to the tank.
The control valve unit 4 further includes a first load pressure detection circuit 131, a second load pressure detection circuit 132, a third load pressure detection circuit 133, and differential pressure reducing valves 111, 211 and 311. The first load pressure detection circuit 131 includes shuttle valves 9d, 9f, 9i and 9j which are connected to load ports of the flow control valves 6d, 6f, 6i and 6j connected to the first hydraulic fluid supply line 105 in order to detect the maximum load pressure P1max1 of the actuators 3a, 3b, 3d and 3f. The second load pressure detection circuit 132 includes shuttle valves 9b, 9c and 9g which are connected to load ports of the flow control valves 6b, 6c and 6g connected to the second hydraulic fluid supply line 205 in order to detect the maximum load pressure P1max2 of the actuators 3b, 3c and 3g. The third load pressure detection circuit 133 includes shuttle valves 9e and 9h which are connected to load ports of the flow control valves 6a, 6e and 6h connected to the third hydraulic fluid supply line 305 in order to detect the load pressure (maximum load pressure) P1max3 of the actuators 3a, 3e and 3h. The differential pressure reducing valve 111 outputs the difference (LS differential pressure) between the pressure P1 in the first hydraulic fluid supply line 105 (i.e., the pressure in the first delivery port 102a) and the maximum load pressure P1max1 detected by the first load pressure detection circuit 131 (i.e., the maximum load pressure of the actuators 3a, 3b, 3d and 3f connected to the first hydraulic fluid supply line 105) as absolute pressure P1s1. The differential pressure reducing valve 211 outputs the difference (LS differential pressure) between the pressure P2 in the second hydraulic fluid supply line 205 (i.e., the pressure in the second delivery port 102b) and the maximum load pressure P1max2 detected by the second load pressure detection circuit 132 (i.e., the maximum load pressure of the actuators 3b, 3c and 3g connected to the second hydraulic fluid supply line 205) as absolute pressure P1s2. The differential pressure reducing valve 311 outputs the difference (LS differential pressure) between the pressure P3 in the third hydraulic fluid supply line 305 (i.e., the delivery pressure of the main pump 202 or the pressure in the third delivery port 202a) and the maximum load pressure P1max3 detected by the third load pressure detection circuit 133 (i.e., the load pressure of the actuators 3a, 3e and 3h connected to the third hydraulic fluid supply line 305) as absolute pressure P1s3. The absolute pressures P1s1, P1s2 and P1s3 outputted by the differential pressure reducing valves 111, 211 and 311 will hereinafter be referred to as LS differential pressures P1s1, P1s2 and P1s3 as needed.
To the aforementioned unloading valve 115, the maximum load pressure P1max1 detected by the first load pressure detection circuit 131 is led as the maximum load pressure of the actuators driven by the hydraulic fluid delivered from the first delivery port 102a. To the aforementioned unloading valve 215, the maximum load pressure P1max2 detected by the second load pressure detection circuit 132 is led as the maximum load pressure of the actuators driven by the hydraulic fluid delivered from the second delivery port 102b. To the aforementioned unloading valve 315, the maximum load pressure P1max3 detected by the third load pressure detection circuit 133 is led as the maximum load pressure of the actuators driven by the hydraulic fluid delivered from the third delivery port 202a.
The LS differential pressure P1s1 outputted by the differential pressure reducing valve 111 is led to the pressure compensating valves 7d, 7f, 7i and 7j connected to the first hydraulic fluid supply line 105 and to the regulator 112 of the main pump 102. The LS differential pressure P1s2 outputted by the differential pressure reducing valve 211 is led to the pressure compensating valves 7b, 7c and 7g connected to the second hydraulic fluid supply line 205 and to the regulator 112 of the main pump 102. The LS differential pressure P1s3 outputted by the differential pressure reducing valve 311 is led to the pressure compensating valves 7a, 7e and 7h connected to the third hydraulic fluid supply line 305 and to the regulator 212 of the main pump 202.
The actuator 3a is connected to the first delivery port 102a via the flow control valve 6i, the pressure compensating valve 7i and the first hydraulic fluid supply line 105, and to the third delivery port 202a via the flow control valve 6a, the pressure compensating valve 7a and the third hydraulic fluid supply line 305. The actuator 3a is a boom cylinder for driving a boom of the hydraulic excavator, for example. The flow control valve 6a is used for the main driving of the boom cylinder 3a, while the flow control valve 6i is used for the assist driving of the boom cylinder 3a. The actuator 3b is connected to the first delivery port 102a via the flow control valve 6j, the pressure compensating valve 7j and the first hydraulic fluid supply line 105, and to the second delivery port 102b via the flow control valve 6b, the pressure compensating valve 7b and the second hydraulic fluid supply line 205. The actuator 3b is an arm cylinder for driving an arm of the hydraulic excavator, for example. The flow control valve 6b is used for the main driving of the arm cylinder 3b, while the flow control valve 6j is used for the assist driving of the arm cylinder 3b.
The actuators 3d and 3f are connected to the first delivery port 102a via the flow control valves 6d and 6f, the pressure compensating valves 7d and 7f and the first hydraulic fluid supply line 105, respectively. The actuators 3c and 3g are connected to the second delivery port 102b via the flow control valves 6c and 6g, the pressure compensating valves 7c and 7g and the second hydraulic fluid supply line 205, respectively. The actuators 3d and 3f are, for example, a bucket cylinder for driving a bucket of the hydraulic excavator and a left travel motor for driving a left crawler of a lower track structure of the hydraulic excavator, respectively. The actuators 3c and 3g are, for example, a swing motor for driving an upper swing structure of the hydraulic excavator and a right travel motor for driving a right crawler of the lower track structure of the hydraulic excavator, respectively. The actuators 3e and 3h are connected to the third delivery port 102a via the flow control valves 6e and 6h, the pressure compensating valves 7e and 7h and the third hydraulic fluid supply line 305, respectively. The actuators 3e and 3h are, for example, a swing cylinder for driving a swing post of the hydraulic excavator and a blade cylinder for driving a blade of the hydraulic excavator, respectively.
The upper part of
The opening area characteristic of the flow control valve 6a for the main driving of the boom cylinder 3a has been set such that the opening area increases as the spool stroke increases beyond the dead zone 0-S1, the opening area reaches the maximum opening area A1 at an intermediate stroke S2, and thereafter the maximum opening area A1 is maintained until the spool stroke reaches the maximum spool stroke S3. The opening area characteristic of the flow control valve 6b for the main driving of the arm cylinder 3b has also been set similarly.
The opening area characteristic of the flow control valve 6i for the assist driving of the boom cylinder 3a has been set such that the opening area remains at zero until the spool stroke reaches an intermediate stroke S2, increases as the spool stroke increases beyond the intermediate stroke S2, and reaches the maximum opening area A2 just before the spool stroke reaches the maximum spool stroke S3. The opening area characteristic of the flow control valve 6j for the assist driving of the arm cylinder 3b has also been set similarly.
The lower part of
The meter-in channel of each flow control valve 6a, 6i of the boom cylinder 3a has the opening area characteristic explained above. Consequently, the meter-in channels of the flow control valves 6a and 6i of the boom cylinder 3a have a combined opening area characteristic in which the opening area increases as the spool stroke increases beyond the dead zone 0-S1 and the opening area reaches the maximum opening area A1+A2 just before the spool stroke reaches the maximum spool stroke S3. The combined opening area characteristic of the flow control valves 6b and 6j of the arm cylinder 3b has also been set similarly.
Here, the maximum opening area A3 regarding the flow control valves 6c, 6d, 6e, 6f, 6g and 6h of the actuators 3c-3h shown in
Returning to
At times other than a travel combined operation for driving the actuator 3f as the left travel motor (hereinafter referred to as a “left travel motor 3f” as needed) and/or the actuator 3g as the right travel motor (hereinafter referred to as a “right travel motor 3g” as needed) and at least one of the actuators 3a, 3b, 3c and 3d other than the left and right travel motors connected to the first or second hydraulic fluid supply line 105 or 205 at the same time, the travel combined operation detection hydraulic line 53 is connected to the tank via at least one of the operation detection valves 8a, 8b, 8c, 8d, 8f, 8g, 8i and 8j, by which the pressure in the hydraulic line 53 becomes equal to the tank pressure. When the travel combined operation is performed, the operation detection valves 8f and 8g and at least one of the operation detection valves 8a, 8b, 8c, 8d, 8i and 8j stroke together with corresponding flow control valves and the communication between the travel combined operation detection hydraulic line 53 and the tank is interrupted, by which the operation detection pressure (operation detection signal) is generated in the hydraulic line 53.
When the travel combined operation is not performed, the first selector valve 40 is positioned at a first position (interruption position) as the lower position in
When the travel combined operation is not performed, the second selector valve 146 is positioned at a first position as the lower position in
When the travel combined operation is not performed, the third selector valve 246 is positioned at a first position as the lower position in
Incidentally, the left travel motor 3f and the right travel motor 3g are actuators driven at the same time and achieving a prescribed function by having supply flow rates equivalent to each other when driven at the same time. In this embodiment, the left travel motor 3f is driven by the hydraulic fluid delivered from the first delivery port 102a of the split flow type main pump 102, while the right travel motor 3g is driven by the hydraulic fluid delivered from the second delivery port 102b of the split flow type main pump 102.
In
The prime mover revolution speed detection valve 13 includes a flow rate detection valve 50 which is connected between the hydraulic fluid supply line 31a of the pilot pump 30 and the pilot hydraulic fluid supply line 31b and a differential pressure reducing valve 51 which outputs the differential pressure across the flow rate detection valve 50 as absolute pressure Pgr.
The flow rate detection valve 50 includes a variable restrictor part 50a whose opening area increases as the flow rate therethrough (delivery flow rate of the pilot pump 30) increases. The hydraulic fluid delivered from the pilot pump 30 passes through the variable restrictor part 50a of the flow rate detection valve 50 and then flows to the pilot hydraulic line 31b's side. In this case, a differential pressure increasing as the flow rate increases occurs across the variable restrictor part 50a of the flow rate detection valve 50. The differential pressure reducing valve 51 outputs the differential pressure across the variable restrictor part 50a as the absolute pressure Pgr. Since the delivery flow rate of the pilot pump 30 changes according to the revolution speed of the prime mover 1, the delivery flow rate of the pilot pump 30 and the revolution speed of the prime mover 1 can be detected by the detection of the differential pressure across the variable restrictor part 50a. The absolute pressure Pgr outputted by the prime mover revolution speed detection valve 13 (differential pressure reducing valve 51) is led to the regulators 112 and 212 as target LS differential pressure. The absolute pressure Pgr outputted by the differential pressure reducing valve 51 will hereinafter be referred to as “output pressure Pgr” or “target LS differential pressure Pgr” as needed.
The regulator 112 (first pump control unit) includes a low-pressure selection valve 112a, an LS control valve 112b, an LS control piston 112c, torque control (power control) pistons 112d and 112e (first torque control actuators), and a spring 112u. The low-pressure selection valve 112a selects a pressure on the low pressure side from the LS differential pressure P1s1 outputted by the differential pressure reducing valve 111 and the LS differential pressure P1s2 outputted by the differential pressure reducing valve 211. The LS control valve 112b is supplied with the selected lower LS differential pressure P1s12 and the output pressure Pgr of the prime mover revolution speed detection valve 13 as the target LS differential pressure Pgr and changes load sensing drive pressure (hereinafter referred to as “LS drive pressure Px12”) such that the LS drive pressure Px12 decreases as the LS differential pressure P1s12 decreases below the target LS differential pressure Pgr. The LS control piston 112c is supplied with the LS drive pressure Px12 and controls the tilting angle (displacement) of the main pump 102 so as to increase the tilting angle and thereby increase the delivery flow rate of the main pump 102 as the LS drive pressure Px12 decreases. The torque control (power control) piston 112d (first torque control actuator) is supplied with the pressure in the first delivery port 102a of the main pump 102 and controls the tilting angle of the swash plate of the main pump 102 so as to decrease the tilting angle and thereby decrease the absorption torque of the main pump 102 when the pressure in the first delivery port 102a increases. The torque control (power control) piston 112e (first torque control actuator) is supplied with the pressure in the second delivery port 102b of the main pump 102 and controls the tilting angle of the swash plate of the main pump 102 so as to decrease the tilting angle and thereby decrease the absorption torque of the main pump 102 when the pressure in the second delivery port 102b increases. The spring 112u is used as biasing means for setting maximum torque T12max (see
The low-pressure selection valve 112a, the LS control valve 112b and the LS control piston 112c constitute a first load sensing control section which controls the displacement of the main pump 102 such that the delivery pressure of the main pump 102 (delivery pressure on the high pressure side of the first and second delivery ports 102a and 102b) becomes higher by a target differential pressure (target LS differential pressure Pgr) than the maximum load pressure of the actuators driven by the hydraulic fluid delivered from the main pump 102 (pressure on the high pressure side of the maximum load pressures P1max1 and P1max2).
The torque control pistons 112d and 112e and the spring 112u constitute a first torque control section which controls the displacement of the main pump 102 such that the absorption torque of the main pump 102 does not exceed the maximum torque T12max set by the spring 112u when the absorption torque of the main pump 102 increases due to an increase in at least one of the displacement of the main pump 102 and the delivery pressure of each delivery port 102a, 102b of the main pump 102 (the delivery pressure of main pump 102).
In
The first load sensing control section (the low-pressure selection valve 112a, the LS control valve 112b and the LS control piston 112c) functions when the absorption torque of the main pump 102 is lower than the maximum torque T12max and is not undergoing the limitation by the torque control by the first torque control section, and controls the displacement of the main pump 102 by means of the load sensing control.
The regulator 212 (second pump control unit) includes an LS control valve 212b, an LS control piston 212c (load sensing control actuator), a torque control (power control) piston 212d (second torque control actuator), and a spring 212e. The LS control valve 212b is supplied with the LS differential pressure P1s3 outputted by the differential pressure reducing valve 311 and the output pressure Pgr of the prime mover revolution speed detection valve 13 as the target LS differential pressure Pgr and changes load sensing drive pressure (hereinafter referred to as “LS drive pressure Px3”) such that the LS drive pressure Px3 decreases as the LS differential pressure P1s3 decreases below the target LS differential pressure Pgr. The LS control piston 212c (load sensing control actuator) is supplied with the LS drive pressure Px3 and controls the tilting angle (displacement) of the main pump 202 so as to increase the tilting angle and thereby increase the delivery flow rate of the main pump 202 as the LS drive pressure Px3 decreases. The torque control (power control) piston 212d (second torque control actuator) is supplied with the delivery pressure of the main pump 202 and controls the tilting angle of the swash plate of the main pump 202 so as to decrease the tilting angle and thereby decrease the absorption torque of the main pump 202 when the delivery pressure of the main pump 202 increases. The spring 212e is used as biasing means for setting maximum torque T3max (see
The LS control valve 212b and the LS control piston 212c constitute a second load sensing control section which controls the displacement of the main pump 202 such that the delivery pressure of the main pump 202 becomes higher by the target differential pressure (target LS differential pressure Pgr) than the maximum load pressure P1max3 of the actuators driven by the hydraulic fluid delivered from the main pump 202.
The torque control piston 212d and the spring 212e constitute a second torque control section which controls the displacement of the main pump 202 such that the absorption torque of the main pump 202 does not exceed the maximum torque T3max when the absorption torque of the main pump 202 increases due to an increase in at least one of the delivery pressure and the displacement of the main pump 202.
In
The second load sensing control section (the LS control valve 212b and the LS control piston 212c) functions when the absorption torque of the main pump 202 is lower than the maximum torque T3max and is not undergoing the limitation by the torque control by the second torque control section, and controls the displacement of the main pump 202 by means of the load sensing control.
Returning to
The arrows in
The details of the torque feedback circuit 112v will be explained below.
The torque feedback circuit 112v includes a first pressure dividing circuit 112r, a variable pressure reducing valve 112g, a second pressure dividing circuit 112s, and a shuttle valve (higher pressure selection valve) 112j. The first pressure dividing circuit 112r includes a first fixed restrictor 112i to which the delivery pressure of the main pump 202 is led and a variable restrictor valve 112h situated downstream of the first fixed restrictor 112i and connected to the tank on the downstream side. The first pressure dividing circuit 112r outputs the pressure in a hydraulic line 112m between the first fixed restrictor 112i and the variable restrictor valve 112h. The variable pressure reducing valve 112g is supplied with the output pressure of the first pressure dividing circuit 112r (the pressure in the hydraulic line 112m), outputs the output pressure of the first pressure dividing circuit 112r without change when the pressure in the hydraulic line 112m is lower than or equal to a set pressure, and reduces the output pressure of the first pressure dividing circuit 112r to the set pressure and outputs the reduced pressure when the output pressure is higher than the set pressure. The second pressure dividing circuit 112s includes a second fixed restrictor 112k to which the delivery pressure of the main pump 202 is led and a third fixed restrictor 112l situated downstream of the second fixed restrictor 112k and connected to the tank on the downstream side. The second pressure dividing circuit 112s outputs the pressure in a hydraulic line 112n between the second fixed restrictor 112k and the third fixed restrictor 112l. The shuttle valve (higher pressure selection valve) 112j selects a pressure on the high pressure side from the output pressure of the variable pressure reducing valve 112g and the output pressure of the second pressure dividing circuit 112s and outputs the selected higher pressure. The output pressure of the shuttle valve 112j is led to the torque feedback piston 112f as the output pressure of the torque feedback circuit 112v.
The LS drive pressure Px3 of the regulator 212 is led to a side of the variable restrictor valve 112h of the first pressure dividing circuit 112r in the direction for increasing the opening area of the valve. The variable restrictor valve 112h is configured such that the valve is fully closed when the LS drive pressure Px3 is at the tank pressure, the opening area increases (the pressure in the hydraulic line 112m between the first fixed restrictor 112i and the variable restrictor valve 112h decreases) as the LS drive pressure Px3 increases, and switches to the right-hand position in
The variable pressure reducing valve 112g is supplied with the LS drive pressure Px3 of the regulator 212. The variable pressure reducing valve 112g is configured such that its set pressure equals a preset maximum value (initial value) when the LS drive pressure Px3 is at the tank pressure, decreases as the LS drive pressure Px3 increases, and reaches a preset minimum value when the LS drive pressure Px3 has risen to the constant pilot primary pressure Ppilot of the pilot hydraulic fluid supply line 31b.
The torque feedback circuit 112v is configured such that the opening areas of the first fixed restrictor 112i and the second fixed restrictor 112k are equal to each other and the opening area of the third fixed restrictor 112l equals the maximum opening area of the variable restrictor valve 112h switched to the right-hand position in
In
When any one of the control levers of the actuators 3a, 3e and 3h related to the main pump 202 is operated by the full operation and a demanded flow rate determined by the opening area of the flow control valve (hereinafter referred to simply as “the demanded flow rate of the flow control valve”) is higher than or equal to the flow rate limited by the maximum torque T3 (
When any one of the control levers of the actuators 3a, 3e and 3h related to the main pump 202 is operated by a fine operation, the LS control valve 212b strokes from the left-hand position in
When all the control levers of the actuators 3a, 3e and 3h related to the main pump 202 are at the neutral positions and when any one of these control levers is operated but its operation amount is extremely small and the demanded flow rate of the flow control valve is lower than a minimum flow rate obtained at the minimum tilting angle q3min of the main pump 202, the LS control valve 212b is positioned at the left-hand position (rightward stroke end position) in
In
The output pressure Pn of the second pressure dividing circuit 112s is a pressure obtained by dividing the delivery pressure P3 of the main pump 202 according to the ratio between the opening areas of the second fixed restrictor 112k and the third fixed restrictor 112l. This pressure increases linearly and proportionally like the straight line An as the delivery pressure P3 of the main pump 202 increases. The opening area of the second fixed restrictor 112k of the second pressure dividing circuit 112s equals that of the first fixed restrictor 112i of the first pressure dividing circuit 112r. The opening area of the third fixed restrictor 112l of the second pressure dividing circuit 112s equals the maximum opening area of the variable restrictor valve 112h switched to the right-hand position in
In
The high pressure side of the output pressures of the variable pressure reducing valve 112g and the second pressure dividing circuit 112s is selected and outputted by the shuttle valve 112j as the output pressure of the torque feedback circuit 112v. Thus, the output pressure P3t of the torque feedback circuit 112v changes as shown in
Next, an explanation will be given of the function of the torque feedback circuit 112v correcting the delivery pressure of the main pump 202 to achieve a characteristic simulating the absorption torque of the main pump 202 and outputting the modified pressure.
When the main pump 202 performs the displacement control by means of the load sensing control, the position of the displacement changing member (swash plate) of the main pump 202, that is, the displacement (tilting angle) of the main pump 202, is determined by the equilibrium between resultant force of two pushing forces applied to the swash plate from the LS control piston 212c on which the LS drive pressure acts and from the torque control piston 212d on which the delivery pressure of the main pump 202 acts and pushing force applied to the swash plate in the opposite direction from the spring 212e serving as the biasing means for setting the maximum torque. Therefore, the tilting angle of the main pump 202 during the load sensing control changes not only depending on the LS drive pressure but also due to the influence of the delivery pressure of the main pump 202.
When any one of the control levers of the actuators 3a, 3e and 3h related to the main pump 202 is operated by the full operation and the delivery flow rate of the main pump 202 saturates and the LS drive pressure Px3 becomes equal to the tank pressure (e.g., boom raising full operation (c) which will be explained later), as the delivery pressure P3 of the main pump 202 increases, the tilting angle q3 of the main pump 202 changes like the characteristic Hq (Hqa, Hqb) shown in
When any one of the control levers of the actuators 3a, 3e and 3h related to the main pump 202 is operated by a fine operation and the LS drive pressure Px3 increases to an intermediate pressure between the tank pressure and the pilot primary pressure Ppilot (e.g., boom raising fine operation (b) and horizontally leveling work (f) which will be explained later), as the LS drive pressure Px3 increases like Px3b, Px3c and Px3d, the tilting angle q3 of the main pump 202 changes like the curves Iq, Jq and Kq in
In other words, when the delivery pressure P3 of the main pump 202 rises, the tilting angle q3 of the main pump 202 decreases like the curve Iq due to the influence of the increase in the delivery pressure P3 as mentioned above even if the LS drive pressure Px3 is constant at Px3b, for example. Thus, in a high pressure range of the delivery pressure P3, the tilting angle q3 becomes smaller than the tilting angle situated on the curve Hqb of T3max (
The same goes for the cases where the LS drive pressure Px3 is Px3c or Px3d. The tilting angle q3 decreases like the curves Jq and Kq due to the influence of the increase in the delivery pressure P3, and becomes even smaller than the tilting angle on the curve Iq in a high pressure range of the delivery pressure P3 (
When all the control levers of the actuators 3a, 3e and 3h related to the main pump 202 are at the neutral positions and when any one of these control levers is operated but its operation amount is extremely small and the demanded flow rate of the flow control valve is lower than the minimum flow rate obtained at the minimum tilting angle q3min of the main pump 202 (e.g., (a) operation when all control levers are at the neutral positions and (g) boom raising fine operation in load lifting work which will be explained later), the tilting angle q3 of the main pump 202 is maintained at the minimum tilting angle q3min determined by the structure of the main pump 202 as indicated by the straight line Lq in
Returning to
As is clear from the comparison between
To sum up, the torque feedback circuit 112v modifies the delivery pressure of the main pump 202 to achieve a characteristic simulating the absorption torque of the main pump 202 in both of when the main pump 202 (second hydraulic pump) undergoes the limitation by the torque control and operates at the maximum torque T3max of the torque control and when the main pump 202 does not undergo the limitation by the torque control and performs the displacement control by means of the load sensing control, and outputs the modified pressure.
Referring to
The upper swing structure 109 is provided with a cab 108 of the canopy type. Arranged in the cab 108 are a cab seat 121, left and right front/swing operating devices 122 and 123 (only the left side is shown in
Next, the operation of this embodiment will be explained below.
First, the hydraulic fluid delivered from the fixed displacement pilot pump 30 driven by the prime mover 1 is supplied to the hydraulic fluid supply line 31a. The hydraulic fluid supply line 31a is equipped with the prime mover revolution speed detection valve 13. By using the flow rate detection valve 50 and the differential pressure reducing valve 51, the prime mover revolution speed detection valve 13 outputs the differential pressure across the flow rate detection valve 50 corresponding to the delivery flow rate of the pilot pump 30 as the absolute pressure Pgr (target LS differential pressure). The pilot relief valve 32 connected downstream of the prime mover revolution speed detection valve 13 generates the constant pressure (the pilot primary pressure Ppilot) in the pilot hydraulic fluid supply line 31b.
All the flow control valves 6a- 6j are positioned at their neutral positions since the control levers of all the operating devices are at their neutral positions. Since all the flow control valves 6a-6j are at the neutral positions, the first load pressure detection circuit 131, the second load pressure detection circuit 132 and the third load pressure detection circuit 133 detect the tank pressure as the maximum load pressures P1maxl, P1max2 and P1max3, respectively. These maximum load pressures P1maxl, P1max2 and P1max3 are led to the unloading valves 115, 215 and 315 and the differential pressure reducing valves 111, 211 and 311, respectively.
Due to the maximum load pressure P1max1, P1max2, P1max3 led to each unloading valve 115, 215, 315, the pressure P1, P2, P3 in each of the first, second and third delivery ports 102a, 102b and 202a is maintained at a pressure (unloading valve set pressure) as the sum of the maximum load pressure P1max1, P1max2, P1max3 and the set pressure Pun0 of the spring of each unloading valve 115, 215, 315. Here, the maximum load pressures P1maxl, P1max2 and P1max3 equal the tank pressure as mentioned above, and the tank pressure is approximately 0 MPa. Therefore, the unloading valve set pressure becomes equal to the set pressure Pun0 of the spring and the pressures P1, P2 and P3 in the first, second and third delivery ports 102a, 102b and 202a are maintained at PunO (minimum delivery pressure P3min). The pressure PunO is generally set slightly higher than the output pressure Pgr of the prime mover revolution speed detection valve 13 defined as the target LS differential pressure (Pun0>Pgr).
Each differential pressure reducing valve 111, 211, 311 outputs the differential pressure (LS differential pressure) between the pressure P1, P2, P3 in each of the first, second and third hydraulic fluid supply lines 105, 205 and 305 and the maximum load pressure P1max1, P1max2, P1max3 (tank pressure) as the absolute pressure P1s1, P1s2, P1s3. Since the maximum load pressures P1max1, P1max2 and P1max3 equal the tank pressure as mentioned above, relationships P1s1=P1−P1max1=P1=Pun0>Pgr, P1s2=P2−P1max2=P2=Pun0>Pgr, and P1s3=P3−P1max3=P3=Pun0>Pgr hold. The LS differential pressures P1s1 and P1s2 are led to the low-pressure selection valve 112a of the regulator 112, while the LS differential pressure P1s3 is led to the LS control valve 212b of the regulator 212.
In the regulator 112, the low pressure side is selected from the LS differential pressures P1s1 and P1s2 led to the low-pressure selection valve 112a and the selected lower pressure is led to the LS control valve 112b as the LS differential pressure P1s12. In this case, P1s12>Pgr holds irrespective of which of P1s1 or P1s2 is selected, and thus the LS control valve 112b is pushed leftward in
Meanwhile, the LS differential pressure P1s3 is led to the LS control valve 212b of the regulator 212. Since P1s3>Pgr holds, the LS control valve 212b is pushed rightward in
Further, since the LS drive pressure Px3 becomes equal to the pilot primary pressure Ppilot when all the control levers are at the neutral positions, the torque feedback circuit 112v takes on the setting of the straight line An in
When the control lever of the boom operating device (boom control lever) is operated in the direction of expanding the boom cylinder 3a (i.e., boom raising direction), for example, the flow control valves 6a and 6i for driving the boom cylinder 3a are switched upward in
When the operation on the boom control lever is a fine operation and the strokes of the flow control valves 6a and 6i are within S2 shown in
As above, in the boom raising fine operation, even if the flow control valve 6i for the assist driving is switched upward in
In contrast, when the flow selector valve 6a is switched upward in
Just after the control lever is operated at the start of the boom raising operation, the load pressure of the boom cylinder 3a is transmitted to the third hydraulic fluid supply line 305 and the pressure difference between two lines becomes almost zero, and thus the LS differential pressure P1s3 becomes almost equal to zero. Since the relationship P1s3<Pgr holds, the LS control valve 212b switches leftward in
Further, since the LS drive pressure Px3 takes on an intermediate pressure between the tank pressure and the pilot primary pressure Ppilot, the torque feedback circuit 112v takes on the setting indicated by the straight lines Bm and Bp in
For example, when the main pump 202 in the boom raising fine operation operates at the point X2 (P3a, q3b) in
With such features, even when the operation has shifted from the single operation of the boom raising fine operation to a combined operation of the boom raising fine operation and an operation driving any one of the actuators related to the main pump 102 (e.g., horizontally leveling work which will be explained later) and the control lever of the actuator is operated by the full operation, the first torque control section controls the tilting angle of the main pump 102 such that the absorption torque of the main pump 102 does not exceed T12max−T3gs, by which the sum of the absorption torque of the main pump 102 and the absorption torque of the main pump 202 is inhibited from exceeding the maximum torque T12max. Consequently, the stoppage of the prime mover 1 (engine stall) can be prevented.
When the boom control lever is operated by the full operation in the direction of expanding the boom cylinder 3a (i.e., boom raising direction), for example, the flow control valves 6a and 6i for driving the boom cylinder 3a are switched upward in
As mentioned above, the load pressure of the boom cylinder 3a is detected by the third load pressure detection circuit 133 as the maximum load pressure P1max3 via the load port of the flow control valve 6a. According to the maximum load pressure P1max3, the delivery flow rate of the main pump 202 is controlled such that P1s3 becomes equal to Pgr, and the hydraulic fluid is supplied from the main pump 202 to the bottom side of the boom cylinder 3a.
Meanwhile, the load pressure on the bottom side of the boom cylinder 3a is detected by the first load pressure detection circuit 131 as the maximum load pressure P1max1 via the load port of the flow control valve 6i and is led to the unloading valve 115 and the differential pressure reducing valve 111. Due to the maximum load pressure P1max1 led to the unloading valve 115, the set pressure of the unloading valve 115 rises to a pressure as the sum of the maximum load pressure P1maxl (the load pressure on the bottom side of the boom cylinder 3a) and the set pressure Pun0 of the spring, by which the hydraulic line for discharging the hydraulic fluid in the first hydraulic fluid supply line 105 to the tank is interrupted. Further, due to the maximum load pressure P1max1 led to the differential pressure reducing valve 111, the differential pressure (LS differential pressure) between the pressure P1 in the first hydraulic fluid supply line 105 and the maximum load pressure P1max1 is outputted by the differential pressure reducing valve 111 as the absolute pressure P1s1. The pressure P1s1 is led to the low-pressure selection valve 112a of the regulator 112 and the low pressure side is selected from P1s1 and P1s2 by the low-pressure selection valve 112a.
Just after the control lever is operated at the start of the boom raising operation, the load pressure of the boom cylinder 3a is transmitted to the first hydraulic fluid supply line 105 and the pressure difference between two lines becomes almost zero, and thus the LS differential pressure P1s1 becomes almost equal to zero. On the other hand, the LS differential pressure P1s2 has been maintained at a level higher than Pgr in this case (P1s2=P2−P1max2=P2=Pun0>Pgr) similarly to the case where the control lever is at the neutral position. Thus, the LS differential pressure P1s1 is selected by the low-pressure selection valve 112a as the LS differential pressure P1s12 on the low pressure side and is led to the LS control valve 112b. The LS control valve 112b compares the LS differential pressure P1s1 with the target LS differential pressure Pgr. In this case, the LS differential pressure P1s1 is almost equal to zero as mentioned above and the relationship P1s1<Pgr holds. Therefore, the LS control valve 112b switches rightward in
In this case, the second hydraulic fluid supply line 205 is supplied with the hydraulic fluid at the same flow rate as the hydraulic fluid supplied to the first hydraulic fluid supply line 105. However, the hydraulic fluid supplied to the first hydraulic fluid supply line 105 is returned to the tank as a surplus flow via the unloading valve 215. In this case, the second load pressure detection circuit 132 is detecting the tank pressure as the maximum load pressure P1max2, and thus the set pressure of the unloading valve 215 becomes equal to the set pressure Pun0 of the spring and the pressure P2 in the second hydraulic fluid supply line 205 is maintained at the low pressure Pun0. Accordingly, the pressure loss occurring in the unloading valve 215 when the surplus flow returns to the tank is reduced and operation with less energy loss is made possible.
Here, while the main pump 202 delivers the hydraulic fluid at a flow rate according to the demanded flow rate of the flow control valve 6a, when the demanded flow rate is higher than or equal to the flow rate limited by the maximum torque T3 (
For example, when the main pump 202 in the boom raising full operation operates at the point X1 (P3a, q3a) on the curve 602 of the maximum torque T3max in
With such features, the first torque control section controls the tilting angle of the main pump 102 such that the absorption torque of the main pump 102 does not exceed T12max - T3max, by which the sum of the absorption torque of the main pump 102 and the absorption torque of the main pump 202 is inhibited from exceeding the maximum torque T12max. Consequently, the stoppage of the prime mover 1 (engine stall) can be prevented.
When the control lever of the arm operating device (arm control lever) is operated in the direction of expanding the arm cylinder 3b (i.e., arm crowding direction), for example, the flow control valves 6b and 6j for driving the arm cylinder 3b are switched downward in
When the operation on the arm control lever is a fine operation and the strokes of the flow control valves 6b and 6j are within S2 shown in
When the flow control valve 6b is switched downward in
Just after the control lever is operated at the start of the arm crowding operation, the load pressure of the arm cylinder 3b is transmitted to the second hydraulic fluid supply line 205 and the pressure difference between two lines becomes almost zero, and thus the LS differential pressure P1s2 becomes almost equal to zero. On the other hand, the LS differential pressure P1s1 has been maintained at a level higher than Pgr in this case (P1s1=P1−P1max1=P1=Pun0>Pgr) similarly to the case where the control lever is at the neutral position. Thus, the LS differential pressure P1s2 is selected by the low-pressure selection valve 112a as the LS differential pressure P1s12 on the low pressure side and is led to the LS control valve 112b. The LS control valve 112b compares the LS differential pressure P1s2 with the output pressure Pgr of the prime mover revolution speed detection valve 13 as the target LS differential pressure. In this case, the LS differential pressure P1s2 is almost equal to zero as mentioned above and the relationship P1s2<Pgr holds. Therefore, the LS control valve 112b switches rightward in
In this case, the first hydraulic fluid supply line 105 is supplied with the hydraulic fluid at the same flow rate as the hydraulic fluid supplied to the second hydraulic fluid supply line 205, and the hydraulic fluid supplied to the first hydraulic fluid supply line 105 is returned to the tank as a surplus flow via the unloading valve 115. At that time, the first load pressure detection circuit 131 detects the tank pressure as the maximum load pressure P1max1, and thus the set pressure of the unloading valve 115 becomes equal to the set pressure Pun0 of the spring and the pressure P1 in the first hydraulic fluid supply line 105 is maintained at the low pressure Pun0. Accordingly, the pressure loss occurring in the unloading valve 115 when the surplus flow returns to the tank is reduced and operation with less energy loss is made possible.
Further, since no actuator related to the main pump 202 is driven in this case, similarly to the case where all the control levers are at the neutral positions, the torque feedback circuit 112v takes on the setting of the straight line An in
When the arm control lever is operated by the full operation in the direction of expanding the arm cylinder 3b (i.e., arm crowding direction), for example, the flow control valves 6b and 6j for driving the arm cylinder 3b are switched downward in
As explained in the above chapter (d), the load pressure on the bottom side of the arm cylinder 3b is detected by the second load pressure detection circuit 132 as the maximum load pressure P1max2 via the load port of the flow control valve 6b, and the unloading valve 215 interrupts the hydraulic line for discharging the hydraulic fluid in the second hydraulic fluid supply line 205 to the tank. Since the maximum load pressure P1max2 is led to the differential pressure reducing valve 211, the LS differential pressure P1s2 is outputted and is led to the low-pressure selection valve 112a of the regulator 112.
Meanwhile, the load pressure on the bottom side of the arm cylinder 3b is detected by the first load pressure detection circuit 131 as the maximum load pressure P1max1 (=P1max2) via the load port of the flow control valve 6i and is led to the unloading valve 115 and the differential pressure reducing valve 111. Due to the maximum load pressure P1max1 led to the unloading valve 115, the hydraulic line for discharging the hydraulic fluid in the first hydraulic fluid supply line 105 to the tank is interrupted by the unloading valve 115. Further, since the maximum load pressure P1max1 is led to the differential pressure reducing valve 111, the LS differential pressure P1s1 (=P1s2) is led to the low-pressure selection valve 112a of the regulator 112.
Just after the control lever is operated at the start of the arm crowding operation, the load pressure of the arm cylinder 3b is transmitted to the first and second hydraulic fluid supply lines 105 and 205 and the pressure difference between two lines becomes almost zero in regard to each hydraulic fluid supply line, and thus both of the LS differential pressures P1s1 and P1s2 become almost equal to zero. Thus, P1s1 or P1s2 is selected by the low-pressure selection valve 112a as the LS differential pressure P1s12 on the low pressure side and the LS differential pressure P1s12 is led to the LS control valve 112b. In this case, both of P1s1 and P1s2 are almost equal to zero as mentioned above and the relationship P1s12<Pgr holds. Therefore, the LS control valve 112b switches rightward in
Further, since no actuator related to the main pump 202 is driven also in this case, similarly to the case where all the control levers are at the neutral positions, the torque feedback circuit 112v takes on the setting of the straight line An in
The horizontally leveling work is a combination of the boom raising fine operation and the arm crowding full operation. As for the movement of the actuators, the horizontally leveling operation is implemented by expansion of the arm cylinder 3b and expansion of the boom cylinder 3a.
In the horizontally leveling work, the boom raising is a fine operation. Thus, as explained in the chapter (b), the opening area of the meter-in channel of the flow control valve 6a for the main driving of the boom cylinder 3a becomes smaller than or equal to A1 and the opening area of the meter-in channel of the flow control valve 6i for the assist driving of the boom cylinder 3a is maintained at zero. The load pressure of the boom cylinder 3a is detected by the third load pressure detection circuit 133 as the maximum load pressure P1max3 via the load port of the flow control valve 6a, and the hydraulic line for discharging the hydraulic fluid in the third hydraulic fluid supply line 305 to the tank is interrupted by the unloading valve 315. Further, the maximum load pressure P1max3 is fed back to the regulator 212 of the main pump 202, the displacement (flow rate) of the main pump 202 increases according to the demanded flow rate (opening area) of the flow control valve 6a, the hydraulic fluid at the flow rate corresponding to the input to the boom control lever is supplied from the third delivery port 202a of the main pump 202 to the bottom side of the boom cylinder 3a, and the boom cylinder 3a is driven in the expanding direction by the hydraulic fluid from the third delivery port 202a.
In contrast, the arm control lever is operated by the full operation or full input. Thus, as explained in the above chapter (e), the opening areas of the meter-in channels of the flow control valves 6b and 6j for the main driving and the assist driving of the arm cylinder 3b reach A1 and A2, respectively. The load pressure of the arm cylinder 3b is detected by the first and second load pressure detection circuits 131 and 132 respectively as the maximum load pressures P1max1 and P1max2 (P1max1=P1max2) via the load ports of the flow control valves 6b and 6j, the hydraulic line for discharging the hydraulic fluid in the first hydraulic fluid supply line 105 to the tank is interrupted by the unloading valve 115, and the hydraulic line for discharging the hydraulic fluid in the second hydraulic fluid supply line 205 to the tank is interrupted by the unloading valve 215. Further, the maximum load pressures P1max1 and P1max2 are fed back to the regulator 112 of the main pump 102, the displacement (flow rate) of the main pump 102 increases according to the demanded flow rates of the flow control valves 6b and 6j, the hydraulic fluid at the flow rate corresponding to the input to the arm control lever is supplied from the first and second delivery ports 102a and 102b of the main pump 102 to the bottom side of the arm cylinder 3b, and the arm cylinder 3b is driven in the expanding direction by the merged hydraulic fluid from the first and second delivery ports 102a and 102b.
In the horizontally leveling work, the load pressure of the arm cylinder 3b is generally low and the load pressure of the boom cylinder 3a is generally high in many cases. In this embodiment, actuators differing in the load pressure are driven by separate pumps, namely, the boom cylinder 3a is driven by the main pump 202 and the arm cylinder 3b is driven by the main pump 102, in the horizontally leveling work. Therefore, the wasteful energy consumption caused by the pressure loss in the pressure compensating valve 7b on the low load side, occurring in the conventional one-pump load sensing system which drives multiple actuators differing in the load pressure by use of one pump, does not occur in the hydraulic drive system of this embodiment.
Since the boom raising is a fine operation in this case, as explained in the chapter (b), the torque feedback circuit 112v takes on the setting indicated by the straight lines Bm and Bp in
With such features, even when the arm control lever is operated by the full operation in the horizontally leveling work, the first torque control section controls the tilting angle of the main pump 102 such that the absorption torque of the main pump 102 does not exceed T12max−T3gs, by which the sum of the absorption torque of the main pump 102 and the absorption torque of the main pump 202 is inhibited from exceeding the maximum torque T12max. Consequently, the stoppage of the prime mover 1 (engine stall) can be prevented.
The load lifting work is a type of work in which a wire is attached to a hook formed on the bucket and a load is lifted with the wire and moved to a different place. Also when the boom raising fine operation is performed in the load lifting work, the hydraulic fluid is supplied from the third delivery port 202a of the main pump 202 to the bottom side of the boom cylinder 3a by the load sensing control performed by the regulator 212 and the boom cylinder 3a is driven in the expanding direction as explained in the chapter (b) or (f). However, the boom raising in the load lifting work is work that needs extreme care, and thus the operation amount of the control lever is extremely small and there are cases where the demanded flow rate of the flow control valve is less than the minimum flow rate obtained by the minimum tilting angle q3min of the main pump 202. In such cases, P1s3>Pgr holds, the LS control valve 212b is positioned at the left-hand position in
Here, the load in the load lifting work is heavy and the delivery pressure P3 of the main pump 202 becomes high like the point H on the straight line An in
If the torque feedback circuit 112v is not equipped with the second pressure dividing circuit 112s in this embodiment, the output pressure of the torque feedback circuit 112v is limited to the pressure Ppa in the hydraulic line 112p as the output pressure of the variable pressure reducing valve 112g as shown in
In this embodiment, the torque feedback circuit 112v is equipped with the second pressure dividing circuit 112s. Thus, even when the delivery pressure P3 of the main pump 202 becomes high like the point H on the straight line An in
Earth removal work for moving earth and sand by operating the blade 106 while traveling is performed by a combined operation driving the travel motors 3f and 3g and the blade cylinder 106 at the same time. When the blade control lever is operated in this case, similarly to the aforementioned boom raising fine operation (b), for example, the displacement (flow rate) of the main pump 202 increases according to the demanded flow rate (opening area) of the flow control valve 6h, the hydraulic fluid at the flow rate corresponding to the input to the blade control lever is supplied from the third delivery port 202a of the main pump 202 to the blade cylinder 3h, and the blade cylinder 3h is driven by the hydraulic fluid from the third delivery port 202a.
In the earth removal work, it is when the LS drive pressure Px3 is at an intermediate pressure between the tank pressure and the pilot primary pressure Ppilot that the main pump 202 operates at the point X3 (P3c, q3c) in
With such features, the first torque control section controls the tilting angle of the main pump 102 such that the absorption torque of the main pump 102 does not exceed T12max−T3hs, by which the sum of the absorption torque of the main pump 102 and the absorption torque of the main pump 202 is inhibited from exceeding the maximum torque T12max. Consequently, the stoppage of the prime mover 1 (engine stall) can be prevented.
In this embodiment configured as above, not only when the main pump 202 (second hydraulic pump) is in the operational state of undergoing the limitation by the torque control and operating at the maximum torque T3max of the torque control but also when the main pump 202 is in the operational state of not undergoing the limitation by the torque control and performing the displacement control by means of the load sensing control, the delivery pressure P3 of the main pump 202 is modified by the torque feedback circuit 112v to achieve a characteristic simulating the absorption torque of the main pump 202 and the maximum torque T12max is modified by the torque feedback piston 112f (third torque control actuator) to decrease by an amount corresponding to the modified delivery pressure P3t. As above, the absorption torque of the main pump 202 is detected precisely by use of a purely hydraulic structure (torque feedback circuit 112v). By feeding back the absorption torque to the main pump 102's side, the total torque control can be performed precisely and the rated output torque Terate of the prime mover 1 can be utilized efficiently.
In the comparative example shown in
In this comparative example, when the main pump 202 is operating at the point X1 (P3a, q3a) on the curve 602 of the maximum torque T3max in
However, when the main pump 202 is operating at the point X2 (P3a, q3b) in
The comparative example cannot achieve the effects of this embodiment also when the main pump 202 is operating at the point X3 (P3c, q3c) in
As mentioned above, in this embodiment, when the main pump 202 is operating at the point X2 (P3a, q3b) in
Further, when the main pump 202 is operating at the point X3 (P3c, q3c) in
As above, in this embodiment, the total horsepower control for preventing the stoppage of the prime mover 1 (engine stall) can be performed precisely and the output torque Terate of the prime mover 1 can be utilized efficiently by having the torque feedback circuit 112v precisely feed back the absorption torque T3max, T3g or T3h of the main pump 202 to the main pump 102's side.
Further, in this embodiment in which the torque feedback circuit 112v is equipped with the second pressure dividing circuit 112s, even when the delivery pressure P3 of the main pump 202 becomes high like the point H on the straight line An in
In
Specifically, the torque feedback circuit 112Av in this embodiment includes a variable pressure reducing valve 112g, a pressure dividing circuit 112s, and a shuttle valve (higher pressure selection valve) 112j. The variable pressure reducing valve 112g is supplied with the delivery pressure P3 of the main pump 202 (the pressure in the third hydraulic fluid supply line 305), outputs the delivery pressure P3 of the main pump 202 without change when the delivery pressure P3 of the main pump 202 is lower than or equal to a set pressure, and reduces the delivery pressure P3 of the main pump 202 to the set pressure and outputs the reduced pressure when the delivery pressure P3 of the main pump 202 is higher than the set pressure. The pressure dividing circuit 112s includes a second fixed restrictor 112k to which the delivery pressure P3 of the main pump 202 is led and a third fixed restrictor 1121 situated downstream of the second fixed restrictor 112k and connected to the tank on the downstream side. The pressure dividing circuit 112s outputs the pressure in the hydraulic line 112n between the second fixed restrictor 112k and the third fixed restrictor 112l. The shuttle valve (higher pressure selection valve) 112j selects a pressure on the high pressure side from the output pressure of the variable pressure reducing valve 112g and the output pressure of the pressure dividing circuit 112s and outputs the selected higher pressure.
In
When the LS drive pressure Px3 is at an intermediate pressure between the tank pressure and the pilot primary pressure Ppilot, the set pressure Pp of the variable pressure reducing valve 112g drops from the initial value Ppf to Ppc. Thus, when the delivery pressure P3 of the main pump 202 rises, the output pressure Pp of the variable pressure reducing valve 112g changes like the straight lines Cm1 and Bp. Specifically, the output pressure Pp of the variable pressure reducing valve 112g increases linearly and proportionally like the straight line Cm1 (Pp=P3) until the delivery pressure P3 of the main pump 202 rises to Ppc. After the delivery pressure P3 reaches Ppc, the output pressure Pp does not increase further and is limited to Ppc lower than the pressure Ppf of the straight line Cp like the straight line Bp.
When the LS drive pressure Px3 rises to the pilot primary pressure Ppilot, the set pressure of the variable pressure reducing valve 112g drops to the minimum value Ppa. Thus, when the delivery pressure P3 of the main pump 202 rises, the output pressure of the variable pressure reducing valve 112g changes like the straight lines Cm2 and Ap. In short, the output pressure Pp of the variable pressure reducing valve 112g is limited to the lowest pressure Ppa like the straight line Ap in the entire range from the minimum delivery pressure of the main pump 202.
The output characteristic of the pressure dividing circuit 112s is identical with that of the second pressure dividing circuit 112s in the first embodiment. The output pressure Pn of the pressure dividing circuit increases linearly and proportionally as the delivery pressure P3 of the main pump 202 increases as indicated by the straight line An in
In
Also in this embodiment configured as above, effects similar to those of the first embodiment can be achieved when the LS drive pressure Px3 is at an intermediate pressure between the tank pressure and the pilot primary pressure Ppilot, except that the setting of the torque feedback circuit 112v indicated by the straight line Bm in
For example, when the main pump 202 is operating at the point X1 (P3a, q3a) on the curve 602 of the maximum torque T3max in
When the main pump 202 is operating at the point X2 (P3a, q3b) in
As above, also in this embodiment, the total horsepower control for preventing the stoppage of the prime mover 1 (engine stall) can be performed precisely and the output torque Terate of the prime mover 1 can be utilized efficiently by having the torque feedback circuit 112Av precisely feed back the absorption torque T3max or T3g of the main pump 202 to the main pump 102′s side.
In
Specifically, the torque feedback circuit 112Bv in this embodiment includes the first pressure dividing circuit 112Br, the variable pressure reducing valve 112g, the second pressure dividing circuit 112s, and the shuttle valve (higher pressure selection valve) 112j.
The first pressure dividing circuit 112Br includes the first fixed restrictor 112i to which the delivery pressure P3 of the main pump 202 (the pressure in the third hydraulic fluid supply line 305) is led and the variable relief valve 112z situated downstream of the first fixed restrictor 112i and connected to the tank on the downstream side. The pressure in the hydraulic line 112m between the first fixed restrictor 112i and the variable relief valve 112z is led to one input port of the shuttle valve 112j.
The LS drive pressure Px3 of the regulator 212 is led to a side of the variable relief valve 112z in the direction for increasing the opening area of the valve. The variable relief valve 112z is configured such that the valve is set at a prescribed relief pressure when the pressure Px3 is at the tank pressure, the relief pressure decreases as the pressure Px3 increases, and the relief pressure becomes zero and the valve has a preset maximum opening area when the pressure Px3 is at the constant pilot primary pressure Ppilot generated in the pilot hydraulic fluid supply line 31b by the pilot relief valve 32.
The structure of the variable pressure reducing valve 112g and the second pressure dividing circuit 112s is the same as that in the first embodiment.
In this embodiment configured as above, the output characteristic of the variable relief valve 112z is equivalent to that of the variable pressure reducing valve 112g in the first embodiment and the output characteristic of the torque feedback circuit 112Bv is equivalent to that of the torque feedback circuit 112v in the first embodiment shown in
While the description of the above embodiments has been given of a case where the first hydraulic pump is the split flow type hydraulic pump 102 having the first and second delivery ports 102a and 102b, the first hydraulic pump can also be a variable displacement hydraulic pump having a single delivery port.
Further, while the first pump control unit has been assumed to be the regulator 112 including the load sensing control section (the low-pressure selection valve 112a, the LS control valve 112b and the LS control piston 112c) and the torque control section (the torque control pistons 112d and 112e and the spring 112u), the load sensing control section in the first pump control unit is not essential. Other types of control methods such as the so-called positive control or negative control may also be employed as long as the displacement of the first hydraulic pump can be controlled according to the operation amount of a control lever (the opening area of a flow control valve - the demanded flow rate).
Furthermore, the load sensing system in the above embodiment is just an example and can be modified in various ways. For example, while a differential pressure reducing valve outputting a pump delivery pressure and a maximum load pressure as absolute pressures is employed, and the target compensation pressure is set by leading the output pressure of the differential pressure reducing valve to a pressure compensating valve, and the target differential pressure of the load sensing control is set by leading the output pressure of the differential pressure reducing valve to an LS control valve in the above embodiment, it is also possible to lead the pump delivery pressure and the maximum load pressure to a pressure control valve or an LS control valve through separate hydraulic lines.
| Number | Date | Country | Kind |
|---|---|---|---|
| 2013-246800 | Nov 2013 | JP | national |
| Filing Document | Filing Date | Country | Kind |
|---|---|---|---|
| PCT/JP2014/081145 | 11/26/2014 | WO | 00 |
