JP2006083990A - Hydraulic driving device - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To optimize output torque of an engine, by setting a pump absorbing torque characteristic desired to be set originally, in a hydraulic driving device for driving an actuator by merging and supplying delivery oil of hydraulic pumps of a variable displacement type and a fixed displacement type. <P>SOLUTION: An absorbing torque control mechanism 9 is arranged in the first hydraulic pump 2 of the variable displacement type, and a main relief valve 11 is arranged in a merging pipe 8. A check valve 12 is arranged in the middle of a pipe 7b being a delivery oil passage of the second hydraulic pump 3 of the fixed displacement type, and a second relief valve 15 is arranged in a pipe 14 by branching off the pipe 14 reaching a drain tank 13 from a pipe part between the second hydraulic pump 3 of the pipe 7b and the check valve 12. Preset pressure of the relief valve 15 is set lower than preset pressure of the main relief valve 11, and is set so that an override characteristic becomes gentle and differential pressure between cracking pressure and relief setting pressure becomes larger than ordinary pressure. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to a hydraulic drive device for a construction machine such as a hydraulic excavator, and in particular, discharges oil from a variable displacement hydraulic pump and discharge oil from a fixed displacement hydraulic pump to supply and drive a hydraulic actuator. The present invention relates to a hydraulic drive device.

  As a hydraulic drive device that joins and supplies the discharge oil of the variable displacement hydraulic pump and the discharge oil of the fixed displacement hydraulic pump to drive the hydraulic actuator, for example, FIGS. 1 and 2 of JP-A-7-187578. And those described in Japanese Patent Laid-Open No. 2003-202001. The variable displacement hydraulic pump is provided with an absorption torque control mechanism that reduces the pump displacement (displacement volume) when the discharge pressure increases, so that the pump absorption torque does not exceed the limit value. In this way, by using a variable displacement hydraulic pump and a fixed displacement hydraulic pump together, the fixed displacement hydraulic pump can replenish the required flow rate with respect to the maximum discharge flow rate of the variable displacement hydraulic pump. Therefore, it is possible to suppress an increase in the size of the variable displacement hydraulic pump more than necessary.

JP-A-7-187578 JP 2003-202001 A

  The horsepower control characteristic of the whole hydraulic power source in the conventional hydraulic drive apparatus is given by the sum of the horsepower control characteristic of the variable displacement hydraulic pump by the absorption torque control mechanism and the horsepower characteristic by the absorption torque of the fixed displacement hydraulic pump. In this case, in order to suppress the absorption torque of the variable displacement hydraulic pump to zero at the rated pressure (set pressure of the main relief valve), the flow rate by the tilt control of the variable displacement hydraulic pump is the minimum value at the rated pressure. It is conceivable to set it to be zero. However, even if set in this way, the absorption horsepower of the whole hydraulic power source at the rated pressure is determined by the discharge flow rate of the fixed displacement hydraulic pump, and the output torque required for the engine is defined from the absorption horsepower at the rated pressure. As a result, it is necessary to require a high engine output torque with respect to the isotorque line to be originally set.

  Thus, in the conventional technology, if the capacity of a fixed displacement hydraulic pump is set high in order to secure a large flow rate at light load, the pump absorption torque at high load (for example, at rated pressure) becomes excessive, and the engine The output torque required for the engine is higher than the engine output torque that is originally suitable for a hydraulic drive device. This increases the size of the engine itself, the cooling function against the increase in the emitted heat, etc., and the peripheral devices also have excessive performance, impairing the appropriateness in terms of space and cost.

  An object of the present invention is to approximate the isotorque line to be originally set in a hydraulic drive device that supplies and supplies the discharge oil of a variable displacement hydraulic pump and the discharge oil of a fixed displacement hydraulic pump to drive a hydraulic actuator. It is an object of the present invention to provide a hydraulic drive device that can set pump absorption torque and can optimize engine output torque, engine specifications, and peripheral device specifications.

  (1) In order to achieve the above object, the present invention provides a variable displacement type first hydraulic pump and a fixed displacement type second hydraulic pump, a direction switching valve, and pressures from the first and second hydraulic pumps. A hydraulic circuit including an actuator where oil joins and is supplied via the direction switching valve, an absorption torque control mechanism of the first hydraulic pump, and discharge oil of the first hydraulic pump is supplied to the direction switching valve A hydraulic drive device having a first relief valve that is provided in the first pipeline and that defines a rated pressure of the hydraulic circuit, and is connected to the first pipeline, and discharges oil from the second hydraulic pump to the first pipeline. A check valve that is provided in the middle of the second pipe to be joined to one pipe and prevents the backflow of pressure oil to the second hydraulic pump; and a second hydraulic pump and a check valve in the second pipe Between the actuators in the first pipeline. And a drain flow rate control means for controlling the discharge amount of the discharge oil of the second hydraulic pump to the drain tank so that the pump absorption torque characteristic corresponding to the flow rate characteristic of the supply flow rate to the compressor approaches the equal torque characteristic And

  Thus, the check valve and the drain flow rate control means are provided, and the discharge oil of the second hydraulic pump is adjusted so that the pump absorption torque characteristic corresponding to the flow rate characteristic of the supply flow rate to the actuator in the first pipeline approaches the equal torque characteristic. By controlling the discharge amount to the drain tank, it is possible to set the pump absorption torque close to the isotorque line that is originally set, and to optimize the engine output torque, engine specifications, and peripheral equipment specifications. .

  (2) In order to achieve the above object, the present invention includes a variable displacement type first hydraulic pump and a fixed displacement type second hydraulic pump, a direction switching valve, and the first and second hydraulic pumps. A hydraulic circuit including an actuator supplied with the pressure oil and supplied via the direction switching valve, an absorption torque control mechanism of the first hydraulic pump, and a discharge oil from the first hydraulic pump as the direction switching valve. A hydraulic drive device having a first relief valve that is provided in a first pipeline that supplies a rated pressure of the hydraulic circuit, and that is connected to the first pipeline and discharges oil from the second hydraulic pump. A check valve that is provided in the middle of the second pipe that joins the first pipe and prevents the backflow of pressure oil to the second hydraulic pump; and the second hydraulic pump and the check valve of the second pipe Between the two hydraulic pumps, When the pressure reaches the first pressure, discharge of the discharge oil of the second hydraulic pump to the drain tank is started, and when the discharge pressure of the second hydraulic pump reaches a second pressure higher than the first pressure. The entire amount of discharge oil of the second hydraulic pump is discharged to the drain tank, and while the discharge pressure of the second hydraulic pump is at the first pressure and the second pressure, the discharge pressure of the second hydraulic pump is reduced. It is assumed that a drain flow rate control means for controlling the discharge amount to the drain tank so as to increase the discharge amount in accordance with the rise, and the second pressure is set lower than the set pressure of the first relief valve.

  Thus, by providing the check valve and the drain flow rate control means and controlling the discharge amount of the discharge oil of the second hydraulic pump to the drain tank, the first pipe line is between the first pressure and the second pressure. As the supply pressure (discharge pressure of the second hydraulic pump) increases, the discharge amount to the drain tank increases, so the supply flow rate to the actuator in the first pipe decreases accordingly, and the second pressure and the second pressure Since the entire amount of oil discharged from the second hydraulic pump is discharged to the drain tank between the set pressures of one relief valve, the supply flow rate is only the discharge flow rate of the first hydraulic pump, and the supply flow rate is the absorption torque of the first hydraulic pump. Decreases according to the horsepower control characteristics by the control mechanism. For this reason, in the pressure range higher than the first pressure, the supply flow rate decreases smoothly as the supply pressure increases, and in the pressure range higher than the second pressure, the absorption horsepower of the supply flow rate is determined only by the discharge flow rate of the first hydraulic pump. It becomes possible to decide. As a result, it is possible to set the pump absorption torque close to the isotorque line that is originally set, and to optimize the engine output torque, engine specifications, and peripheral equipment specifications.

  (3) In the above (2), preferably, the drain flow rate control means is provided in a third pipeline branched from a pipeline portion between the check valve of the second pipeline and the second hydraulic pump. , Having a second relief valve in which the differential pressure between the set pressure and the cracking pressure is greater than 10% of the set pressure, setting the set pressure of the second relief valve to the second pressure, and setting the cracking pressure to the first pressure Set to 1 pressure.

  As a result, the drain flow rate control means can be configured using one relief valve, and a simple and inexpensive system can be obtained.

  (4) In the above (2), preferably, the drain flow rate control means includes a third pipeline branched from a pipeline portion between the check valve of the second pipeline and the second hydraulic pump, and A second relief valve and a third relief valve provided in parallel with each other in the fourth pipeline, and a fixed throttle provided on the upstream side of the third relief valve in the fourth pipeline, The set pressure of the relief valve is set to the second pressure, and the set pressure of the third relief valve is set to the first pressure.

  As a result, the flow rate characteristic of the fixed throttle has a large rising rate and then gradually decreases. Therefore, the pump absorption torque characteristic corresponding to the supply flow characteristic of the fixed displacement type second hydraulic pump simulates an isotorque line. Therefore, discontinuities are unlikely to occur in the merged discharge oils of the first and second hydraulic pumps, and a characteristic close to that of an equal torque control by a single ordinary variable displacement hydraulic pump can be obtained.

  (5) Further, in the above (2), preferably, the drain flow rate control means is provided in a third pipeline branched from a pipeline portion between the check valve of the second pipeline and the second hydraulic pump. And a variable throttle valve device provided on the upstream side of the second relief valve in the third pipe line, and the variable throttle valve device has a differential pressure across itself. If the preset pressure is not reached, the minimum opening area is maintained, and when the variable throttle valve device increases its flow rate, the opening area is increased so that the set value is maintained when the pressure difference across the valve reaches the set pressure. The set pressure of the variable throttle device is set equal to the differential pressure between the first pressure and the second pressure.

  As a result, the flow rate characteristic of the variable throttle valve device has a large rate of change of rising and then gradually decreases. Therefore, the pump absorption torque characteristic corresponding to the supply flow rate characteristic of the fixed displacement type second hydraulic pump is equal torque. Since it becomes a locus simulating a line, it is difficult for discontinuities to occur in the merged discharge oil of the first and second hydraulic pumps, and it is possible to obtain characteristics close to equal torque control by a single ordinary variable displacement hydraulic pump it can.

  According to the present invention, in the hydraulic drive device that drives the hydraulic actuator by supplying the discharge oil of the variable displacement type hydraulic pump and the discharge oil of the fixed displacement type hydraulic pump, it is close to the isotorque line to be originally set. Pump absorption torque can be set, and engine output torque, engine specifications, and peripheral equipment specifications can be optimized. As a result, an optimal device configuration including pump selection becomes possible.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings.

  A hydraulic drive apparatus according to a first embodiment of the present invention will be described with reference to FIGS.

  FIG. 1 is a diagram showing a basic circuit of a hydraulic drive apparatus according to the present embodiment. The basic circuit of this hydraulic drive device is mainly used for a hydraulic excavator.

  In FIG. 1, the basic circuit of the hydraulic drive apparatus according to the present embodiment includes an engine 1, a variable displacement type first hydraulic pump 2 and a fixed displacement type second hydraulic pump 3 driven by the engine 1, and directions. A switching valve 4 and a hydraulic actuator 5 are provided. The first hydraulic pump 2 is connected to the direction switching valve 4 via a pipeline 7a and a pipeline 8, and the second hydraulic pump 3 is connected to the direction switching valve 4 via a pipeline 7b and a pipeline 8. 7 a and 7 b are connected to the pipe line 8 in parallel. The pipe lines 7a and 7b constitute discharge oil paths of the first and second hydraulic pumps 2 and 3, respectively, and the pipe line 8 constitutes a merged pipe line of discharge oil of the first and second hydraulic pumps 2 and 3. The oil discharged from the first and second hydraulic pumps 2 and 3 is supplied via pipes 7a and 7b, respectively, and merges in the pipe 8. The oil discharged from the first and second hydraulic pumps 2 and 3 merged in the pipe 8 is supplied to the hydraulic actuator 5 via the direction switching valve 4. By supplementing the flow rate required by the second hydraulic pump (fixed capacity pump) 3 with respect to the maximum discharge flow rate of the first hydraulic pump (variable capacity pump) 2, the size of the first hydraulic pump 2 can be increased more than necessary. It becomes possible to suppress.

  The variable displacement first hydraulic pump 2 is provided with an absorption torque control mechanism 9 that reduces the pump displacement (displacement volume) when the discharge pressure rises so that the pump absorption torque does not exceed the limit value. The pipe 8 is provided with a main relief valve 11 that defines a rated pressure (maximum circuit pressure) of a hydraulic circuit including the first and second hydraulic pumps 2 and 3, the direction switching valve 4, and the hydraulic actuator 5.

  A check valve 12 that allows a flow of pressure oil from the second hydraulic pump 3 to the pipe 8 and prevents a reverse flow in the middle of the pipe 7b that is a discharge oil path of the fixed displacement type second hydraulic pump 3. Similarly, in the middle of the pipe line 7b, a pipe line 14 from the pipe line part between the second hydraulic pump 3 and the check valve 12 to the drain tank 13 branches, and the second pipe line 14 The relief valve 15 is provided.

  The second relief valve 15 sets the set pressure (relief set pressure) lower than the set pressure of the main relief valve 11 (rated pressure of the hydraulic circuit). Also, in a normal relief valve, the differential pressure between the cracking pressure and the relief setting pressure is set to be as small as possible so that the override characteristic becomes steep, and usually the differential pressure is 10% or less of the relief setting pressure. is there. On the other hand, in the relief valve 15 according to the present embodiment, the differential pressure between the cracking pressure and the relief set pressure is set to be larger than usual so that the override characteristic becomes gentle, and the differential pressure is more than 10% of the relief set pressure. Is also big.

  The relief valve 15 configured in this manner releases the oil discharged from the second hydraulic pump 3 to the drain tank 13 when the discharge pressure of the second hydraulic pump 3 reaches the cracking pressure (first pressure) of the relief valve 15. When the discharge pressure of the second hydraulic pump 3 reaches the set pressure of the relief valve 15 (second pressure higher than the first pressure), the entire amount of oil discharged from the second hydraulic pump 3 is released to the drain tank 13 While the discharge pressure of the second hydraulic pump 3 is between the first pressure and the second pressure, the drain tank 13 is configured to continuously increase the discharge amount in accordance with the increase of the discharge pressure of the second hydraulic pump 3. It functions as a drain flow rate control means for controlling the discharge amount to the water.

  That is, when the pressure in the pipe line portion between the check valve 12 of the pipe line 7 b and the second hydraulic pump 3 rises above the cracking pressure of the relief valve 15, the discharge oil of the second hydraulic pump 3 is discharged from the relief valve 15. When the pressure is released to the drain tank 13 and the pressure reaches the set pressure of the relief valve 15, the entire amount of oil discharged from the second hydraulic pump 2 is released to the drain tank 13. Under this condition, the discharge oil of the first hydraulic pump 2 does not flow back to the pipe line 7b on the second hydraulic pump 3 side by the check valve 12. For this reason, the pressure in the pipe line 7 b (the discharge pressure of the second hydraulic pump 3) is held at the set pressure of the relief valve 15. As a result, the absorption torque of the second hydraulic pump 3 can be set regardless of the rated pressure of the hydraulic circuit defined by the main relief valve 11.

  In addition, since the override characteristic of the relief valve 15 is set more gradual than that of a normal relief valve, it is possible to suppress a discontinuous change in the supply flow rate when the relief valve 15 is operated.

  As a result, the second relief valve 15 discharges oil from the fixed displacement type second hydraulic pump 3 so that the pump absorption torque characteristic corresponding to the flow characteristic of the supply flow rate to the actuator 5 in the conduit 8 approaches the equal torque characteristic. The amount of discharge to the drain tank is controlled.

  FIG. 2 is a diagram showing absorption torque characteristics and horsepower control characteristics by the absorption torque control mechanism 9 of the first hydraulic pump 2. In FIG. 2, for the sake of convenience, the pump absorption torque characteristic and the horsepower control characteristic are expressed by the same characteristic line on the assumption that the rotational speed of the engine 1 is constant at the rated rotational speed. That is, the pump absorption torque is represented by the relationship between pressure and pump capacity (displacement volume), and the horsepower control characteristic is represented by the relationship between pressure and flow rate. The flow rate is a linear function of the pump capacity (displacement volume). When the rotational speed of the engine 1 is constant at the rated rotational speed, the pump flow rate is constant and the horsepower control characteristics are also constant. Therefore, the pump absorption torque characteristic and the horsepower control characteristic can be expressed by the same characteristic line. This is the same in the following figures (FIGS. 5, 7, 12, and 17).

  In FIG. 2, the first hydraulic pump 2 reduces the pump capacity (displacement volume) when the discharge pressure of the first hydraulic pump 2 is increased by the absorption torque control mechanism 9 so that the pump absorption torque does not exceed the limit value A. Be controlled. As a result, the second hydraulic pump 2 is controlled so that the pump discharge flow rate is reduced and the pump absorption horsepower does not exceed the limit value B when the discharge pressure of the first hydraulic pump 2 increases at the rated speed of the engine 1. .

  FIG. 3 is a diagram showing an override characteristic of the relief valve 15, and FIG. 4 is a diagram showing a supply flow rate characteristic of the second hydraulic pump (fixed capacity pump) 3.

  The cracking pressure of the relief valve 15 is set to P1, and the set pressure (relief set pressure) of the relief valve 15 is set to P2. P2 is a pressure at which the absorption torque of the second hydraulic pump 3 is desired to be limited. In this example, the cracking pressure P1 is set to about 50% of the relief setting pressure P2. Due to this characteristic, the supply flow rate of the pressure oil that joins the pipe 8 from the second hydraulic pump 3 is constant until the discharge pressure of the second hydraulic pump 3 reaches the cracking pressure P1 of the relief valve 15, and the second hydraulic pressure After the discharge pressure of the pump 3 reaches the cracking pressure P1, when the pump discharge pressure further increases, it decreases according to the override characteristics of the relief valve 15, and the supply flow rate becomes 0 at the relief set pressure P2. .

  FIG. 5 is a diagram for explaining the system function of the present embodiment. The horsepower control characteristic of the first hydraulic pump 2 is shown on the upper right side of FIG. 5, and the supply flow rate of the second hydraulic pump 3 is shown on the upper left side of FIG. 5 shows the relationship between the supply flow rate obtained by synthesizing them and the supply pressure (the discharge pressure of the first hydraulic pump 2 or the pressure of the pipe line 8) on the lower side of FIG. This relationship corresponds to the pump horsepower control characteristics of the whole hydraulic power source including the first and second hydraulic pumps 2 and 3, the absorption torque control mechanism 8, and the relief valve 15. In the figure, P3 is the set pressure of the main relief valve 11, that is, the rated pressure of the hydraulic circuit. C is the supply flow rate characteristic of the second hydraulic pump 3, and the horsepower control characteristic (right side in FIG. 5) by the absorption torque control mechanism 8 of the first hydraulic pump 2 is set according to the supply flow rate characteristic of the second hydraulic pump 3. Has been.

  The supply flow rate and the supply pressure given by the sum of the horsepower control characteristic by the absorption torque control mechanism 8 of the first hydraulic pump 2 and the supply flow rate characteristic of the second hydraulic pump 3 (horsepower characteristic by the absorption torque of the second hydraulic pump 3). The relationship is set as D on the lower side of FIG. Between the cracking pressure P1 of the relief valve 15 and the set pressure P2, when the supply pressure increases, the discharge amount to the drain tank 13 increases according to the override characteristics of the relief valve 15, so the supply flow rate decreases accordingly, and the relief valve Between the set pressure P2 of 15 and the rated pressure P3, the entire amount of oil discharged from the second hydraulic pump 3 is discharged to the drain tank, so the supply flow rate is only the discharge flow rate of the first hydraulic pump 2, and the supply flow rate is 1 Decrease according to the horsepower control characteristic by the absorption torque control mechanism of the hydraulic pump 2. As a result, the absorption torque characteristics (shown by the same curve D for convenience) of the first and second hydraulic pumps 2 and 3 corresponding to the flow characteristic D are set close to the equal torque line F that is originally set. In the lower side of FIG. 5, E is the relationship between the combined flow rate of the first and second hydraulic pumps 2 and 3 of the following comparative example (conventional device) and the pressure of the combined flow rate.

  FIG. 6 is a diagram showing a basic circuit of a conventional hydraulic drive apparatus as a comparative example. Similar to the basic circuit of the hydraulic drive apparatus according to the present embodiment, the variable displacement type first hydraulic pump 2 and the fixed displacement type second hydraulic pump 3 are connected to the direction switching valve 4 via the pipelines 7 a, 7 b, 8. It is connected. A check valve and a relief valve are not provided in the pipe line 4b which is a discharge oil path of the second hydraulic pump.

  FIG. 7 shows the relationship between the supply flow rate and the supply pressure, which is the combined flow rate of the first and second hydraulic pumps 2 and 3 (the pump of the entire hydraulic source including the first and second hydraulic pumps 2 and 3 and the absorption torque control mechanism 8 It is a figure which shows the characteristic equivalent to a horsepower control characteristic. This relationship is given by the sum of the horsepower control characteristics of the first hydraulic pump 2 by the absorption torque control mechanism 8 and the supply flow characteristics of the second hydraulic pump 3 (horsepower characteristics by the absorption torque of the second hydraulic pump 3). In the figure, G is the supply flow rate characteristic of the second hydraulic pump 3, and E is the relationship between the combined flow rate of the first and second hydraulic pumps 2 and 3, and the pressure of the combined flow rate.

  In the illustrated example, in order to suppress the absorption torque of the variable displacement type first hydraulic pump 2 to 0 at the rated pressure P3, the flow rate by the tilt control of the first hydraulic pump 2 becomes 0 at the minimum value at the rated pressure P3. It is set as follows. However, even if this setting is made, the absorption horsepower of the entire hydraulic power source at the rated pressure P3 is determined by the discharge flow rate of the second hydraulic pump (fixed capacity pump) 3, and is required for the engine 1 from the absorption horsepower at the rated pressure P3. Output torque is defined. As a result, a high engine output torque H is inevitably required for the equal torque line F that is originally set.

  Thus, in the prior art, if the capacity of the second hydraulic pump (fixed capacity pump) 3 is set high in order to ensure a large flow rate at light load, the pump absorption torque at high load (for example, at the rated pressure P3). Becomes excessive, and the output torque required for the engine becomes higher than the engine output torque that is inherently appropriate for a hydraulic drive device. This increases the size of the engine itself, the cooling function against the increase in the emitted heat, etc., and the peripheral devices also have excessive performance, impairing the appropriateness in terms of space and cost.

  On the other hand, in the present embodiment, it is possible to set the pump absorption torque that is originally set, and it is possible to optimize the engine output torque, the engine specifications, and the specifications of peripheral devices. As a result, an optimal device configuration including pump selection becomes possible.

  A second embodiment of the present invention will be described with reference to FIGS.

  FIG. 8 is a diagram showing a basic circuit of the hydraulic drive apparatus according to the present embodiment. In the figure, the same components as those shown in FIG. 1 are denoted by the same reference numerals, and description thereof will be omitted as appropriate.

  In FIG. 8, in the middle of the pipe line 7 b that is a discharge oil path of the fixed capacity type second hydraulic pump 3, a common pipe is connected from the pipe line part between the second hydraulic pump 3 and the check valve 12. The pipelines 14b and 14c leading to the drain tank 13 are branched through the passage 14a, the second relief valve 21 is provided in the pipeline 14b, and the fixed throttle 23 and the third relief valve 24 are provided in the pipeline 14c. It has been. The fixed throttle 23 is located on the upstream side of the third relief valve 24, and the second relief valve 21 and the third relief valve 25 are arranged in parallel to each other.

  The set pressure (relief set pressure) of the second relief valve 21 is set lower than the set pressure (rated pressure value of the hydraulic circuit) of the main relief valve 11, and the set pressure (relief set pressure) of the second relief valve 24 is set. Is set lower than the set pressure of the second relief valve 21. Both the second and third relief valves 21 and 24 are normal relief valves whose cracking pressure is 10% or less of the relief set pressure.

  FIG. 9 is a diagram showing the flow rate characteristics of the fixed throttle 23, FIG. 10 is a diagram showing the flow rate characteristics of the drain circuit including the fixed throttle 23 and the second and third relief valves 21, 24, and FIG. FIG. 3 is a diagram showing supply flow rate characteristics of a fixed displacement type second hydraulic pump 3.

  The flow characteristics of the second relief valve 21 and the third relief valve 24 are shown by broken lines in FIG. As can be seen from this figure, the set pressure of the third relief valve 24 is set in the vicinity of P1, which is a pressure corresponding to the cracking pressure of the relief valve 15 in the first embodiment, and the set pressure of the second relief valve 21. Is set to P2 (> P1), which is a pressure corresponding to the set pressure of the relief valve 15 in the first embodiment. P2 is a pressure at which the absorption torque of the second hydraulic pump 3 is desired to be limited.

  Assuming that the opening area of the fixed throttle 23 is a, the relationship between the differential pressure ΔP across the fixed throttle 23 and the passage flow rate Q is as follows.

Q = c * a * (2 / ρ * ΔP) 1/2
Here, c is a flow coefficient, and ρ is the viscosity of the fluid (oil).

  The horizontal axis in FIG. 9 is the front-rear differential pressure ΔP of the fixed throttle 23, and the vertical axis is the passage flow rate Q of the fixed throttle 23. When the front-rear differential pressure ΔP of the fixed throttle is 0, the passage flow rate Q is 0, and the passage flow rate Q increases as the front-rear differential pressure ΔP increases. Further, the flow rate characteristic is different from the override characteristic of FIG. 3 and has a maximum rate of change of rising and then gradually decreases. In the present embodiment, the opening area a of the fixed throttle 23 is such that the differential pressure ΔP is P2−P1 (the differential pressure between the set pressure P2 of the second relief valve 21 and the set differential pressure P1 of the third relief valve 24). ), The passage flow rate Q is set to be the discharge flow rate of the fixed displacement hydraulic pump 3.

  Pressure oil discharged to the drain tank 13 through the fixed throttle 23 and the third relief valve 24 by setting the relief pressure of the second and third relief valves 21 and 24 and setting the opening area of the fixed throttle 23 as described above. The flow rate characteristic is as shown by the solid line in FIG. That is, the flow rate of the pressure oil discharged from the pipe line 7b to the drain tank 13 starts to flow from the set pressure P1 of the third relief valve 24 with the characteristics shown in FIG. The total discharge flow rate of the fixed displacement hydraulic pump 3 is discharged to the drain tank 13 at the set pressure P2. With this characteristic, the supply flow rate of the pressure oil that joins the fixed displacement type hydraulic pump 3 to the pipe line 8 is until the pump discharge pressure reaches the set pressure P1 of the third relief valve 24 as shown in FIG. After the discharge pressure of the second hydraulic pump 3 reaches the set pressure pressure P1 of the third relief valve 24 after the discharge pressure of the second hydraulic pump 3 reaches the set pressure pressure P1, the pump discharge pressure is further increased. It decreases according to the flow rate characteristic of the throttle 23 and becomes 0 at the set pressure P2 of the second relief valve 21. In this case, as a result of the flow characteristic of the fixed throttle 23 shown in FIG. 9, the pump absorption torque characteristic corresponding to the supply flow characteristic of the second hydraulic pump 3 becomes a locus simulating an isotorque line.

  FIG. 12 is a view similar to FIG. 5 for explaining the system function of the present embodiment, showing the horsepower control characteristics of the first hydraulic pump 2 on the upper right side of FIG. 12, and the second hydraulic pump on the upper left side of FIG. 3, the combined flow rate of the first and second hydraulic pumps 2 and 3 obtained by combining them on the lower side of FIG. 12 and the pressure of the combined flow rate (the discharge pressure of the first hydraulic pump 2 or The relationship with the pressure of the pipe line 8) is shown. This relationship corresponds to the pump horsepower control characteristics of the entire hydraulic power source including the first and second hydraulic pumps 2 and 3, the absorption torque control mechanism 8, the relief valves 21 and 24, and the fixed throttle 23.

  The discharge pressure of the second hydraulic pump (fixed capacity pump) 3 is held at the set pressure P2 of the second relief valve 21, and the absorption torque of the second hydraulic pump 3 is suppressed. The absorption torque of the second hydraulic pump 3 can be set by the set pressure P2 of the second relief valve 21. Further, due to the flow rate characteristic of the fixed throttle 23, the pump absorption torque characteristic corresponding to the supply flow rate characteristic of the second hydraulic pump 3 becomes a locus simulating an isotorque line. As a result, the combined flow rate of the first and second hydraulic pumps 2 and 3, which combines the horsepower control characteristic by the absorption torque control mechanism 8 of the first hydraulic pump 2 and the supply flow rate characteristic of the second hydraulic pump 3, and the pressure of the combined flow rate. And the absorption torque characteristics (shown by the same curve J for convenience) of the first and second hydraulic pumps 2 and 3 corresponding to the relationship are set to the equal torque to be originally set. The setting is close to the line F. In FIG. 12, I is a supply flow rate characteristic of the second hydraulic pump 3.

  Further, since the pump absorption torque characteristic corresponding to the supply flow rate characteristic of the fixed displacement type second hydraulic pump 3 becomes a locus simulating an isotorque line (FIG. 11), the discharge of the first and second hydraulic pumps 2 and 3 is performed. Discontinuous points are less likely to occur due to oil merging, and characteristics close to those of equal torque control using a single variable displacement hydraulic pump can be obtained.

  Therefore, according to the present embodiment, it is possible to set the pump absorption torque that is originally set, and it is possible to optimize the engine output torque, the engine specifications, and the specifications of peripheral devices. As a result, an optimal device configuration including pump selection becomes possible.

  A third embodiment of the present invention will be described with reference to FIGS.

  FIG. 13 is a diagram showing a basic circuit of the hydraulic drive apparatus according to the present embodiment. In the figure, the same components as those shown in FIG. 1 are denoted by the same reference numerals, and description thereof will be omitted as appropriate.

  In FIG. 13, in the middle of a pipe line 7 b that is a discharge oil path of the second hydraulic pump 3, a pipe line 14 extending from the pipe line part between the second hydraulic pump 3 and the check valve 12 to the drain tank 13. And a variable throttle valve device 31 and a second relief valve 32 are provided in the pipeline 14. The variable throttle valve device 31 is located upstream of the second relief valve 32.

  The throttle valve device 31 includes a variable throttle portion 31b that changes the opening area by the movement of the valve body 31a, pressure receiving portions 31c and 31d that bias the valve body 31a, and a spring 31e. The variable throttle portion 31 b has a certain minimum opening area, and the discharge pressure of the second hydraulic pump 3 always acts on the second relief valve 32. When the discharge pressure of the second hydraulic pump 3 exceeds the set pressure of the second relief valve 32, the second relief valve 32 opens, and the discharge oil of the second hydraulic pump 3 via the variable throttle 31b is drained. It is discharged into the tank 13. At this time, a differential pressure is generated in the variable throttle portion 31b. The pressure receiving sections 31c and 31d are respectively guided to the upstream and downstream pressures of the variable throttle section 31b, and the differential pressure across the variable throttle section 31b acts on the side that increases the opening area of the variable throttle section 31b. The spring 31e is provided on the side facing the differential pressure before and after it and holding the minimum opening area of the variable throttle portion 31b.

  When the differential pressure across the variable throttle 31b does not reach the differential pressure set by the spring 31e, the variable throttle 31b maintains a minimum opening area and functions as a fixed throttle. When the differential pressure across the variable throttle 31b exceeds the set pressure of the spring 31e, the variable throttle 31b increases the opening area and holds the differential pressure across the variable throttle 31b at the set value of the spring 31e.

  The set pressure (relief set pressure) of the second relief valve 32 is set lower than the set pressure (rated pressure value of the hydraulic circuit) of the main relief valve 11. The second relief valve 32 is a normal relief valve whose cracking pressure is 10% or less of the relief set pressure. The set value of the spring 31 e of the variable throttle valve device 31 is set so that the sum of the set pressure of the second relief valve 32 and the set value of the spring 31 e is lower than the set pressure of the main relief valve 11.

  FIG. 14 is a diagram showing the flow characteristics of the variable throttle unit 31b of the variable throttle device 31, and FIG. 15 is a diagram showing the flow characteristics of the drain circuit including the variable throttle device 31 and the second relief valve 32. FIG. 16 is a graph showing supply flow rate characteristics of the fixed displacement type second hydraulic pump 3.

  The flow rate characteristics of the second relief valve 32 are indicated by broken lines in FIG. As can be seen from this figure, the set pressure of the second relief valve 32 is set in the vicinity of P1, which is a pressure corresponding to the cracking pressure of the relief valve 15 in the first embodiment.

  The horizontal axis in FIG. 14 is the differential pressure ΔP across the variable throttle 31b, and the vertical axis is the flow rate Q through the variable throttle 31b. P2-P1 on the horizontal axis is a set pressure of the spring 31e, and P2 is a pressure at which the absorption torque of the second hydraulic pump 3 is desired to be limited. The sum of the set pressure P1 of the second relief valve 32 and the set pressure (P2-P1) of the spring 31e is P2. This P2 is set to be lower than the set pressure of the main relief valve 11.

  In FIG. 14, before the second relief valve 32 is opened, the flow rate through the variable throttle 31b is 0, and the differential pressure ΔP across the variable throttle 31b is 0. When the second relief valve 32 is opened at the pressure P1, pressure oil flows through the variable throttle 31b, and a front-rear differential pressure ΔP is generated. When the differential pressure across the variable throttle portion 31b is equal to or lower than the set pressure (P2-P1) of the spring 31e, the variable throttle portion 31b functions as a fixed throttle as described above. This is the flow rate characteristic of the minimum opening area of the throttle part 31b. This flow rate characteristic is set in the same manner as the flow rate characteristic of the fixed throttle 23 shown in FIG. 9 in the second embodiment. That is, when the front-rear differential pressure ΔP of the variable restrictor 31b functioning as a fixed throttle is 0, the passing flow rate Q is 0, and as the front-rear differential pressure ΔP increases, the passing flow rate Q increases and the front-rear differential pressure ΔP. When the pressure reaches the set pressure (P2-P1) of the spring 31e, the minimum opening area of the variable throttle portion 31b is set so that the passing flow rate Q becomes the total amount of the discharge flow rate of the second hydraulic pump 3. Further, the flow rate characteristic is a characteristic in which the rate of change of rising is the maximum and then gradually decreases, like the flow rate characteristic of the fixed throttle in FIG. Since the set pressure of the spring 31e is (P2-P1), the differential pressure across the variable throttle 31b does not exceed (P2-P1).

  By setting the relief pressure of the second relief valve 32, setting the minimum opening area of the variable throttle portion 31b of the throttle valve device 31, and setting the spring 31e, the variable throttle portion 31b and the second relief valve 32 are used. The flow rate characteristic of the pressure oil discharged to the drain tank 13 is as shown by the solid line in FIG. That is, the flow rate of the pressure oil discharged from the pipe line 7b to the drain tank 13 starts to flow from the set pressure P1 of the second relief valve 32 with the characteristics shown in FIG. 14 of the variable throttle portion 31b functioning as a fixed throttle. 2 When the discharge pressure of the hydraulic pump 3 becomes P2, and the differential pressure ΔP across the variable throttle 31b reaches the set differential pressure (P2-P1) of the spring 31e, the total discharge flow rate of the second hydraulic pump 3 is transferred to the drain tank 13. Released. Due to this characteristic, the supply flow rate of the pressure oil joined from the second hydraulic pump 3 to the pipe line 8 is constant until the pump discharge pressure reaches the set pressure P1 of the second relief valve 32 as shown in FIG. The second hydraulic pump flow rate), and after the discharge pressure of the second hydraulic pump 3 reaches the set pressure pressure P1 of the second relief valve 32, the variable throttle part It decreases according to the flow rate characteristic of 31b and becomes 0 at a pressure P2 that is higher by the set pressure (P2-P1) of the spring 31e. In this case, the pump absorption torque characteristic corresponding to the supply flow characteristic of the second hydraulic pump 3 is the same as that of the second embodiment shown in FIG. 11 as a result of the flow characteristic of the variable throttle 31b shown in FIG. This is a locus simulating an isotorque line.

  FIG. 17 is a view similar to FIGS. 5 and 12 for explaining the system function of the present embodiment. The horsepower control characteristic of the first hydraulic pump 2 is shown on the upper right side of FIG. 2 shows the supply flow rate characteristics of the hydraulic pump 3, and the combined flow rate of the first and second hydraulic pumps 2, 3 obtained by combining them on the lower side of FIG. 17 and the pressure of the combined flow rate (of the first hydraulic pump 2) The relationship with the discharge pressure or the pressure of the pipe line 8) is shown. This relationship corresponds to the pump horsepower control characteristics of the entire hydraulic power source including the first and second hydraulic pumps 2 and 3, the absorption torque control mechanism 8, the relief valve 32, and the variable throttle valve device 31.

  The discharge pressure of the second hydraulic pump (fixed capacity pump) 3 is held at P2, which is the sum of the set pressure P1 of the second relief valve 32 and the set pressure (P2-P1) of the spring 31e of the variable throttle valve device 31. The absorption torque of the second hydraulic pump 3 is suppressed. The absorption torque of the second hydraulic pump 3 can be set by the set pressure P1 of the second relief valve 32 and the set pressure of the spring 31e. Further, the pump absorption torque characteristic corresponding to the supply flow rate characteristic of the second hydraulic pump 3 becomes a locus simulating an isotorque line due to the flow rate characteristic when the variable throttle part 31b of the variable throttle valve device 31 functions as a fixed throttle. . As a result, the combined flow rate of the first and second hydraulic pumps 2 and 3, which combines the horsepower control characteristic by the absorption torque control mechanism 8 of the first hydraulic pump 2 and the supply flow rate characteristic of the second hydraulic pump 3, and the pressure of the combined flow rate. And the absorption torque characteristics (indicated by the same curve L for convenience) of the first and second hydraulic pumps 2 and 3 corresponding to the relationship are set to the equal torque originally desired to be set. The setting is close to the line F. In FIG. 17, K is a supply flow rate characteristic of the second hydraulic pump 3.

  Further, since the pump absorption torque characteristic corresponding to the supply flow rate characteristic of the fixed displacement type second hydraulic pump 3 becomes a locus simulating an isotorque line (FIG. 16), the discharge of the first and second hydraulic pumps 2 and 3 is performed. Discontinuous points are less likely to occur due to oil merging, and characteristics close to those of equal torque control using a single variable displacement hydraulic pump can be obtained.

  Therefore, according to the present embodiment, it is possible to set the pump absorption torque that is originally set, and it is possible to optimize the engine output torque, the engine specifications, and the specifications of peripheral devices. As a result, an optimal device configuration including pump selection becomes possible.

  In the above embodiment, the hydraulic drive device of the present invention is described with reference to a basic circuit. However, the actual hydraulic drive device is provided with various hydraulic elements (not shown), and various types of hydraulic elements are correspondingly provided. Deformation is possible. For example, in the basic circuit shown in FIG. 1 and the like, only the check valve 12 is arranged in the pipeline 7b through which the discharge oil from the fixed displacement type second hydraulic pump 3 flows, but before and after that, a merging switching valve and other actuators are arranged. A direction switching valve or the like may be arranged. In addition, although one directional switching valve 4 is disposed downstream of the pipeline 8 through which the discharge oils of the first and second hydraulic pumps 2 and 3 merge, a plurality of directional switching valves may be disposed.

It is a figure which shows the basic circuit of the hydraulic drive unit concerning the 1st Embodiment of this invention. It is a figure which shows the absorption torque control characteristic and horsepower control characteristic by the absorption torque control mechanism of a 1st hydraulic pump. It is a figure which shows the override characteristic of a relief valve. It is a figure which shows the supply flow volume characteristic of a fixed displacement type 2nd hydraulic pump. It is a figure explaining the system function of this Embodiment. It is a figure which shows the basic circuit of the conventional hydraulic drive device as a comparative example. It is a figure which shows the relationship between the combined flow volume of the variable displacement type hydraulic pump and fixed displacement type hydraulic pump in a prior art, and the pressure of the combined flow rate. It is a figure which shows the basic circuit of the hydraulic drive unit concerning the 2nd Embodiment of this invention. It is a figure which shows the flow volume characteristic of a fixed throttle. It is a figure which shows the flow volume characteristic of the drain circuit containing a fixed throttle and a 2nd and 3rd relief valve. It is a figure which shows the supply flow volume characteristic of a fixed displacement type 2nd hydraulic pump. It is a figure explaining the system function of this Embodiment. It is a figure which shows the basic circuit of the hydraulic drive unit concerning the 3rd Embodiment of this invention. It is a figure which shows the flow volume characteristic of a variable throttle part. It is a figure which shows the flow volume characteristic of the drain circuit containing a variable throttle valve apparatus and a 2nd relief valve. It is a figure which shows the supply flow volume characteristic of a fixed displacement type 2nd hydraulic pump. It is a figure explaining the system function of this Embodiment.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Engine 2 Variable displacement type 1st hydraulic pump 3 Fixed displacement type 2nd hydraulic pump 4 Directional switching valve 5 Hydraulic actuator 7a Pipe line (1st pipe line)
7b Pipeline (second pipe)
8 pipeline (1st pipeline)
9 Absorption torque control mechanism 11 Main relief valve (first relief valve)
12 Check valve 13 Drain tank 14 Pipe lines 14a, 14b, 14c Pipe line 15 Relief valve (second relief valve)
21 Relief valve (second relief valve)
23 Fixed throttle 24 Relief valve (Third relief valve)
31 Variable throttle valve device 31a Valve body 31b Variable throttle parts 31c, 31d Pressure receiving part 31e Spring 32 Relief valve (second relief valve)

Claims (5)

The variable displacement type first hydraulic pump and the fixed displacement type second hydraulic pump, the direction switching valve, and the pressure oil from the first and second hydraulic pumps merge and are supplied via the direction switching valve. A hydraulic circuit including an actuator and an absorption torque control mechanism of the first hydraulic pump; and a first pipe that supplies discharge oil of the first hydraulic pump to the direction switching valve. A hydraulic drive device having a first relief valve that defines
A reverse connected to the first conduit and provided in the middle of a second conduit that joins the discharge oil of the second hydraulic pump to the first conduit, and prevents reverse flow of pressure oil to the second hydraulic pump. A stop valve,
A pump absorption torque characteristic corresponding to a flow rate characteristic of a supply flow rate to the actuator in the first pipe line is provided in the pipe line part between the second hydraulic pump and the check valve in the second pipe line. And a drain flow rate control means for controlling the discharge amount of the discharge oil of the second hydraulic pump to the drain tank so as to approach the hydraulic pressure. The variable displacement type first hydraulic pump and the fixed displacement type second hydraulic pump, the direction switching valve, and the pressure oil from the first and second hydraulic pumps merge and are supplied via the direction switching valve. A hydraulic circuit including an actuator and an absorption torque control mechanism of the first hydraulic pump; and a first pipe that supplies discharge oil of the first hydraulic pump to the direction switching valve. A hydraulic drive device having a first relief valve that defines
A reverse connected to the first conduit and provided in the middle of a second conduit that joins the discharge oil of the second hydraulic pump to the first conduit, and prevents reverse flow of pressure oil to the second hydraulic pump. A stop valve,
It is provided in a pipe line portion between the second hydraulic pump and the check valve in the second pipe line, and when the discharge pressure of the second hydraulic pump reaches the first pressure, the discharge oil of the second hydraulic pump When the discharge to the drain tank is started and the discharge pressure of the second hydraulic pump reaches a second pressure higher than the first pressure, the entire amount of discharge oil of the second hydraulic pump is discharged to the drain tank, While the discharge pressure of the second hydraulic pump is between the first pressure and the second pressure, the discharge amount to the drain tank is controlled so as to increase the discharge amount according to the increase of the discharge pressure of the second hydraulic pump. And a drain flow rate control means for
The hydraulic drive device characterized in that the second pressure is set lower than a set pressure of the first relief valve. The hydraulic drive device according to claim 2, wherein
The drain flow rate control means is provided in a third pipeline branched from a pipeline portion between the check valve of the second pipeline and the second hydraulic pump, and a differential pressure between a set pressure and a cracking pressure is set. Having a second relief valve greater than 10% of the pressure;
A hydraulic drive device characterized in that a set pressure of the second relief valve is set to the second pressure, and a cracking pressure is set to the first pressure. The hydraulic drive device according to claim 2, wherein
The drain flow rate control means includes a second relief provided in parallel with the third and fourth pipelines branched from the pipeline portion between the check valve of the second pipeline and the second hydraulic pump. A valve and a third relief valve, and a fixed throttle provided on the upstream side of the third relief valve of the fourth pipeline,
The hydraulic drive device characterized in that the set pressure of the second relief valve is set to the second pressure, and the set pressure of the third relief valve is set to the first pressure. The hydraulic drive device according to claim 2, wherein
The drain flow rate control means includes a second relief valve provided in a third pipeline branched from a pipeline portion between the check valve of the second pipeline and the second hydraulic pump, and the third pipeline A variable throttle valve device provided on the upstream side of the second relief valve,
The variable throttle valve device maintains a minimum opening area when its own differential pressure before and after reaching a predetermined set pressure, and the flow rate of the variable throttle valve device increases so that the differential pressure across itself becomes the set pressure. When it reaches, it is configured to increase the opening area to maintain the set pressure,
A hydraulic drive device characterized in that a set pressure of the variable throttle device is set equal to a differential pressure between the first pressure and the second pressure.

Images