JPH0544704A - Flow dividing device for hydraulic operating oil - Google Patents

Abstract

PURPOSE:To change flow dividing rates so as to constitute efficient hydraulic systems, in a flow dividing device with which hydraulic operating oil discharged from an oil pump is divided into two or more hydraulic systems. CONSTITUTION:This flow dividing device for hydraulic operating oil is constituted so that area of control orifices from an inflow port 1 to a discharge port 1c are changed so as to change rates of respective discharge hydraulic operating oil flows from discharge ports 1b, 1c, by means of a flow dividing control valve 2 provided with control orifices 2a, 2c and an auxiliary valve 35 of which the control orifice 40 is changeable in opening area.

Description

DETAILED DESCRIPTION OF THE INVENTION (Industrial field of application) The present invention relates to a shunt of hydraulic oil for use in vehicles or the like, in which hydraulic oil discharged from one oil pump is used in two or more hydraulic systems. Regarding the device. (Prior Art) Conventionally, a plurality of constant flow rate hydraulic systems used in a vehicle or the like usually have independent hydraulic sources (oil pumps). On the other hand, even in a system using the same hydraulic pressure source, a flow dividing valve as shown in FIG. 1 is used so that the respective hydraulic systems can control their required pressures independently of each other. The operation is as follows. The hydraulic oil discharged from the oil pump flows into the chamber 6 from the port 1a, and a part of the hydraulic oil flows through the control orifice 2a into the chamber 7,
The oil passage 2b, the variable throttle 18, the annular groove 9 and the port 1b flow, and the remaining hydraulic oil flows through the variable throttle 19 and the annular groove 8 to the port 1c. At this time, if the reaction force of the spring 3 is R, the pressure of the chamber 6 is P6, and the pressure of the chamber 7 is P7, it acts on the control valve 2 fitted to the housing 1 so as to be able to slide liquid-tightly and smoothly. The force balance is P7 × A + R = P6 × A, where A is the pressure receiving area of the spool. On the other hand, the flow rate Qb passing through the control orifice 2a is generally: Qb = C × S × (P6−P7) ^{1/2 where} C is the coefficient S because it has a relation of the opening area of the control orifice 2a. Now, if the flow rate Qb is less than the set value, the pressure at P7 becomes relatively large with respect to P6. At this time, the control valve 2 slides to the right in FIG. 1 to open the variable throttle 18 forming the throttle pair of the variable throttle 19. As a result, Qb increases to a constant value, and the control valve 2 has a pressure and a spring force. Settle into a mechanically balanced position.
When Qb exceeds the set value, conversely to the above, the control valve 2 slides to the left in FIG. 1 to decrease Qb to the set value in the same manner as described above. In any case, since the control valve 2 operates if Qb is always kept constant, the hydraulic oil discharged from the oil pump is diverted to each hydraulic system at a constant rate regardless of the pressures of the ports 1b and 1c. Will be done. The cap 4 is integrally incorporated in the housing 1 with screws or the like to support the spring 3 and at the same time prevent the hydraulic oil from leaking to the outside together with the seal 5. (Problems to be Solved by the Invention) When an oil pump is provided for each hydraulic system, not only the drive structure of the pump becomes complicated, but also the installation space of the oil pump is greatly restricted. .. On the other hand, when the conventional flow dividing valve is used, the drive structure for the above problems can be solved, but there is no difference in that a large capacity oil pump is required. Now, assuming that the flow rates required for the respective hydraulic systems are Qb and Qc, naturally the required discharge flow rate Qa of the oil pump is Qa = Qb + Qc, and the required discharge flow rate and the energy required to drive the oil pump are independent. This is the same as when a pump is provided. By the way, when considering the operating state of the hydraulic system used in a vehicle or the like, it is noticed that the timings of the maximum load states of the respective hydraulic systems do not necessarily match. That is, for example, one hydraulic system is in a maximum load state during low speed running and the other hydraulic system is in a maximum load state during high speed running, or the timing of this maximum load state is divided into low engine speed and high engine speed. .. Therefore, when combining hydraulic systems with different timings for reaching these maximum load states, a flow dividing device for hydraulic oil is provided that can reduce the total energy required to drive the oil pump and can also downsize the pump body. That is the subject of the present application. (Means for Solving the Problems) In order to solve the above problems, the load pressure in each hydraulic system, the rotation speed of the device driving the oil pump, the traveling speed of the vehicle, etc. are detected, and It suffices if the split ratio of the hydraulic oil can be set according to the values such as. Therefore, the first embodiment which realizes the object of the present application by detecting the load pressure of the hydraulic stem will be described in detail with reference to FIG. In FIG. 2, port 1a is an inflow port from the pump, and 1b and 1c are discharge ports to the hydraulic systems A and B, respectively. The control valve 2 is supported by the cap 4 via a spring seat 14. On the other hand, an auxiliary valve 10 having a large diameter portion and a small diameter portion is incorporated in the control valve 2 so as to be able to slide smoothly, one end is incorporated into the control valve 2 via a spring 11, and the other end is incorporated in the control valve 2. It is supported by the snap ring 13. Since the chamber 12 communicates with the annular groove 8 via the oil passage 2d, the pressure of the chamber 12 is the discharge port 1c.
Pressure, that is, the pressure of the hydraulic system B. The chamber 6 communicates with the chamber 7 through the control orifice 2a and the oil passage 10a, but the control orifice 2c is closed by the auxiliary valve 10 in the illustrated state. (Operation) When a constant flow rate of hydraulic oil flows into the port 1a from the oil pump in such a state, the chamber 6 is generated according to the flow rate passing through the control orifice 2a as described above.
Due to the pressure difference between # 1 and # 7, the flow rate discharged from the ports 1b and 1c has a constant ratio. When the load pressure of the hydraulic system A increases in this state, that is, the pressure of the chamber 7 increases, the pressure of the chamber 12 is also low when the load pressure of the hydraulic system B is low. In FIG. 2, the control orifice 2c that has been closed by sliding to the right in FIG. 2 is opened. Then chamber 6
Since the hydraulic oil flow rate from No. 7 to No. 7 increases, it is possible to change the diversion ratio according to the load pressure of the system A. Since the spring 11 can be installed by applying compression in advance, it is possible to easily realize a setting in which the diversion ratio is gradually changed when the pressure of the system A exceeds a certain value and the diversion ratio is exceeded. it can. Control orifice 2
If a plurality of c's are provided while being shifted in the axial direction of the control valve 2,
The rate of change of the flow rate with respect to the load pressure of the system A can be changed very easily. With such a device, when the load pressure of one hydraulic system changes, the supply flow rate can be increased or decreased as needed, and at the same time, the intended purpose can be achieved. Needless to say, such a flow rate increase / decrease can be performed without delay because the control is performed using a hydraulic valve. (Embodiment) Hereinafter, other embodiments will be described in detail from FIG. 3 to FIG. Second Embodiment ... See FIG. 3 The difference from the first embodiment is that the auxiliary valve 10 is provided with a communication hole 10b and the right end of the control valve 2 is closed. As a result, the pressure of the chamber 7 also acts on the right end of the auxiliary valve 10, so that the auxiliary valve 10 slides purely according to the pressure difference between the chamber 7 and the discharge port 1c, and the opening area of the ports 2a and 2c. The hydraulic oil discharged from the pump is divided in proportion to the sum of the above. Third embodiment: See FIG. 4. A part of the hydraulic oil flowing from the port 1a is discharged from the port 1b through the control orifice 2a, the chamber 7, the oil passage 2b, the variable throttle 18, and the annular groove 9. To be done. The remaining hydraulic oil flows into the chamber 25 through the control orifice 2c and the oil passage 10a, and further the oil passage 2e, the variable throttle 19,
It is discharged from the port 1c via the annular groove 8. One end of the control valve 2 is supported by the cap 4 via the snap ring 13, the spring seat 21, and the spring 3, and the other end is supported by the housing 1 via the spring 20. Now set the pressure of chambers 7 and 25 to D ₇ and P, respectively.
₂ balance equation of ₅ Tosureba control valve Becomes In this embodiment, since there are two springs such as 3 and 20, it is possible to set R≈O near the neutral position of the control valve 2. At this time, P ₇ ≈P ₂₅ from the above equation. .. On the other hand, since the flow rates passing through the control orifices 2a and 2c are proportional to the pressure differences between the chambers 6 and 7 and the chambers 6 and 25, respectively, the diversion ratio is
a. It will be determined by the area of 2c. For example, when the flow rate passing through the control orifice 2a is higher than the set flow dividing ratio, the pressure at P ₇ becomes relatively lower than the pressure at P ₂₅ . Then, the control valve 2 moves to the left in FIG. 4 and throttles the variable throttle 18. As a result, the pressure of P ₇ increases and the flow rate passing through the control orifice 2a decreases. On the contrary, when the flow rate passing through the control orifice 2a is small, the control valve 2 slides in the opposite direction to the variable throttle 19, and the flow rate passing through the control orifice 2a increases. Therefore, the diversion ratio is always kept at the set value. The feature of this structure is that the hole diameters of the control orifices 2a and 2c can be set to be the same. It is generally known that the flow coefficient changes according to the Reynolds number determined by the flow velocity and kinematic viscosity of the hydraulic oil passing through the orifice, and the dimension represented by the diameter of the orifice. By making the diameters of the holes the same and changing only the number of holes, the flow coefficient of each hole can be made constant. This indicates that according to this structure, the working oil can be diverted with extremely high accuracy regardless of changes in the discharge flow rate from the pump and changes in the oil temperature.
(In the conventional example of FIG. 1, the control orifice 2a and the variable throttle 1 are
It is structurally impossible to configure 9 and 9 with the same hole diameter. By the way, since the large diameter portion of the spring seat 21 is incorporated in the inner diameter of the control valve 2 with a gap, when the pressure of the discharge port 1b and the chamber 7 becomes high, this pressure acts on the left end of the auxiliary valve 10. Meanwhile chamber 12
Is communicated with the annular groove 8 by the oil passage 2d, so when the pressure of the discharge port 1b becomes relatively higher than the pressure of the discharge port 1c, the auxiliary valve 10 slides to the right in FIG. The area of 2c is reduced and the amount of hydraulic oil discharged from the discharge port 1b is increased. The spring 11
Is incorporated between the auxiliary valve 10 and the control valve 2 as in the first embodiment. Fourth Embodiment ... See FIG. 5 The basic operation is the same as that of the third embodiment, but the auxiliary valve 1
Chambers 26 and 12 for exerting hydraulic pressure on 0
Are connected to the ports 1d and 1e, respectively, and are set so that the auxiliary valve can be operated by the pressure from the outside to change the diversion ratio. The spring seat 21 is set so as to keep the chamber 26 liquid-tight by providing a seal 29 on the large diameter portion or press-fitting into the control valve 2, and the snap ring 13 keeps the pressure in the chamber 26. Prevents the spring seat from coming off when loaded. Further, the spring seat 21 is provided with a slight step in the inner diameter portion of the control valve 2 so as not to move to the right in FIG. 5 due to the reaction force of the spring 3 and the hydraulic pressure of the chamber 7. Fifth embodiment: See FIG. 6 In this embodiment, an auxiliary valve is not incorporated in the control valve 2, but an auxiliary valve provided separately from the control valve 2 by a solenoid valve 30 controlled by a controller. 35 is being driven. The auxiliary valve 35 has a negative end supported by the housing 1 via a spring 36, and the other end is in contact with the rod 32 of the solenoid valve 30 so that the auxiliary valve 35 can slide smoothly in the housing. For example, FIG. 6 shows a state where the solenoid current is 0. By energizing this solenoid, the rod 32 pops out in response to this current and the prosthetic flap valve 3
Move 5 to the left in the figure. On the other hand, the housing 1 was provided with a control orifice 40 communicating with the chamber 6, and the chamber 41 was connected to the chamber 25 via an oil passage 42. Therefore, when the solenoid valve 30 is energized and the auxiliary valve 35 is moved to the left and the control orifice 40 is opened to the chamber 41, the working oil in the chamber 6 passes through the control orifice 40, the chamber 41 and the oil passage 42 to the chamber 25. Pour in. That is, control orifice 4
When 0 is closed, the hydraulic oil is divided according to the areas of the control orifices 2a and 2c, but the control orifice 4
When 0 is opened in the chamber 41, the diversion ratio is determined by the sum of the area of the control orifice 2a and the area of the control orifices 2c and 40. Therefore, it becomes possible to change the diversion ratio according to the solenoid current. FIG. 6 shows a configuration in which the area of the control orifice 40 changes, but this oil passage is always opened in the chamber 41, and instead, the opening area changes in the oil passage 42 and the auxiliary valve 35. It goes without saying that the same function can be achieved even if the control orifice is configured. In FIG. 6, both ends of the auxiliary valve 35 are communicated with each other through the oil passage 35a to balance the oil pressure acting on the auxiliary valve 35. However, the oil passage 35a is eliminated and hydraulic pressure is applied to the chambers 33 and 36 from the outside. Thus, the area of the control orifice 40, that is, the diversion ratio can be changed not only by the output of the solenoid but also by the hydraulic pressure. Reference numeral 31 is a seal, and various sensors are, for example, a vehicle speed sensor, an engine speed sensor, a hydraulic pressure sensor, and the like. (Effect of the Invention) According to the present invention, not only can a plurality of hydraulic systems be driven simultaneously by one oil pump, but also the discharge flow rate of the oil pump can be efficiently adjusted according to the load state of each hydraulic system or the required timing. Since the flow can be divided, the drive energy of the oil pump can be significantly reduced as compared with the conventional system. Also, since the discharge amount per unit rotation of the oil pump can be reduced, downsizing of the oil pump body can be easily realized and the weight of the entire system can be reduced. Further, by adopting the structure of the control orifice as shown in FIGS. 4 to 6, it is possible to divide the working oil very stably and highly accurately.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 shows a structural example of a conventional flow dividing device and an overall configuration of a system, and FIGS. 2 to 6 show an embodiment of the flow dividing device according to the present application. 1 Housing 1a Inflow port 1b Discharge port 1c Discharge port 2 Control valve 2a Control orifice 2b Oil passage 2c Control orifice 3 Spring 6,7 Chamber 8,9 Annular groove 10 Auxiliary valve 11 Spring 12 Chamber 18 Variable throttle 19 Variable throttle 20 Spring 21 Spring seat 25 Chamber 30 Solenoid valve 32 Rod 35 Auxiliary valve 36 Spring 40 Control orifice 41 Chamber 42 Oil passage

Claims (1)

Claims: (1) In a device for dividing hydraulic oil discharged from one oil pump into two or more hydraulic systems, which is used in a vehicle or the like, a control orifice and a control orifice that determine a split ratio of the hydraulic oil. A split flow of hydraulic oil, which is equipped with a pair of variable throttles and is configured to change the split flow ratio by changing the area of the control orifice that determines the split flow ratio according to pressure or electric signals. apparatus. (2) In the item (1), the control valve provided with the control orifice has an inner hole, an auxiliary valve slidably fitted in the inner hole, and an anti-reaction depending on the relative movement between the control valve and the auxiliary valve. A hydraulic oil flow dividing device, comprising: a spring for generating a force, and configured to change an area of a control orifice in accordance with an oil pressure acting on an auxiliary valve.