CN111089093B - Energy-saving electro-hydraulic proportional direction valve with improved structure and control method thereof - Google Patents
Energy-saving electro-hydraulic proportional direction valve with improved structure and control method thereof Download PDFInfo
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- CN111089093B CN111089093B CN201911349019.6A CN201911349019A CN111089093B CN 111089093 B CN111089093 B CN 111089093B CN 201911349019 A CN201911349019 A CN 201911349019A CN 111089093 B CN111089093 B CN 111089093B
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
- F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
- F15B13/00—Details of servomotor systems ; Valves for servomotor systems
- F15B13/02—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors
Abstract
The invention discloses an energy-saving electro-hydraulic proportional direction valve with an improved structure, which comprises a hydraulic valve body, a valve core, a proportional electromagnet and a displacement sensor, wherein the valve core is arranged in the hydraulic valve body; the proportional electromagnet is used for driving the valve core to move in the hydraulic valve body; the displacement sensor is used for measuring the position of the valve core; by designing the sizes and the position relations of the two load throttling ports, the two oil supply ports, the oil return port and the four shoulders, when the opening degree of one load throttling port is controlled by one shoulder, the other load throttling port is in a fully opened state. The invention also provides a control method of the energy-saving electro-hydraulic proportional directional valve, and specific control design is carried out aiming at four working conditions. By using the hydraulic control system, under some working states, high-pressure energy does not need to be provided for the hydraulic control system at all, and only low-pressure energy needs to be provided for the hydraulic control system to prevent one cavity of the actuating mechanism from being sucked to be empty.
Description
Technical Field
The invention relates to the technical field of hydraulic valves and control thereof, in particular to an energy-saving electro-hydraulic proportional direction valve with an improved structure.
Background
Hydraulic valves, especially electrically-powered hydraulic proportional directional valves, have very wide application in electro-hydraulic servo systems. The energy utilization ratio of a system formed by the electro-hydraulic proportional directional valves is lower, the main reason is that the hydraulic valve has large throttling loss, the throttling loss is consumed in a thermal form, the working efficiency of the hydraulic valve control system is low, the heating is more severe, and large energy waste exists.
The structural principle of the traditional hydraulic proportional directional valve determines that the opening sizes of 2 load throttling mouths of the traditional hydraulic proportional directional valve are always the same and cannot be independently adjusted, and the working mode determines that the traditional hydraulic proportional directional valve can generate extra energy loss under a plurality of load working conditions (such as exceeding load). In order to overcome the energy loss caused by such a situation, some scholars propose to use two or even a plurality of proportional valves to respectively control the actuators, and to use a plurality of valves to control one actuator, so that the cooperative control of a plurality of valves is required, which increases the complexity of the system on one hand and the cost of the system on the other hand.
Disclosure of Invention
In view of this, the present invention provides an energy-saving hydraulic proportional directional valve with an improved structure, in which the opening adjustment relations of two load orifices are not related, and when the opening degree of one of the load orifices is adjusted, the other load orifice is in a fully open state, and only one of the load orifices needs to be subjected to pressure or flow control. The invention also provides a special control method for the hydraulic valve with the structure.
In order to realize the functions, the invention is realized as follows:
an energy-saving electro-hydraulic proportional direction valve with an improved structure comprises a hydraulic valve body, a valve core, a proportional electromagnet and a displacement sensor; the hydraulic valve body is provided with a first load throttling port, a second load throttling port, a first oil supply port, a second oil supply port and an oil return port; the valve core is sequentially provided with a first shoulder, a second shoulder, a third shoulder and a fourth shoulder; it is characterized in that the preparation method is characterized in that,
setting the distance between the first load throttle orifice and the oil return port as L1, and similarly, setting the distance between the second load throttle orifice and the oil return port as L1; the width of the oil return port is L2, and the widths of the first load throttle port, the second shoulder and the third shoulder are L3; the distance between the second shoulder and the third shoulder is L4; the distance between the first oil supply port and the first load throttle port is L5, and the distance between the second oil supply port and the second load throttle port is L5; the first and second oil supply ports have a width of L6; the distance between the first shoulder and the second shoulder is L7, and similarly, the distance between the third shoulder and the fourth shoulder is L7; the first and fourth shoulders have a width of L8; the distance between the left inner wall of the valve body and the first load throttling opening is L9, and the distance between the right inner wall of the valve body (1) and the second load throttling opening is L9;
L1≥L3；
L4≤2×L1+L2+L5；
L7≤L1+L5+L6；
L9≥L3+L4+L7+L8-2×L1-L2。
preferably, the energy-saving electro-hydraulic proportional directional valve further comprises a return spring, and the return spring is arranged between two end faces of the valve core and the inner wall of the hydraulic valve body.
The invention also provides a control method of the energy-saving electro-hydraulic proportional directional valve, wherein a first load throttling port of the energy-saving electro-hydraulic proportional directional valve is connected with the cavity A of the actuating mechanism, a second load throttling port is connected with the cavity B of the actuating mechanism, and the first load throttling port is definedThe motion direction generated when a load throttle orifice supplies oil to the actuating mechanism is positive; it is characterized by that according to the set movement trace of actuating mechanism and its external load force FxCalculating the driving force F required by the actuatorhAnd controlling according to the situation:
when the actuating mechanism moves in the positive direction and the actuating mechanism requires the driving force FhLoad force F with actuatorxThe resultant force between the first shoulder and the second shoulder is the same as the movement direction, the proportional electromagnet is controlled to enable the second shoulder to shield a partial opening of the first load throttling port from the right side, a communication relation is formed between the first load throttling port and the first oil supply port, and the second load throttling port is not shielded completely; the motion trail of the actuating mechanism can be controlled by adjusting the overlapping degree between the first shoulder and the first load throttle orifice;
when the actuating mechanism moves in the positive direction and the actuating mechanism requires the driving force FhLoad force F with actuatorxThe resultant force between the first load throttling port and the second load throttling port is opposite to the movement direction, the proportional electromagnet is controlled, the third shoulder shields a partial opening of the second load throttling port from the right side, and a communication relation is formed among the second load throttling port, the first load throttling port and the oil return port; the motion trail of the actuating mechanism can be controlled by adjusting the overlapping degree between the third shoulder and the second load throttle orifice;
when the actuating mechanism moves reversely and the actuating mechanism requires the driving force FhLoad force F with actuatorxThe resultant force between the first shoulder and the second shoulder is the same as the movement direction, the proportional electromagnet is controlled to enable the third shoulder to shield a partial opening of the second load throttling port from the left side, a communication relation is formed between the second load throttling port and the second oil supply port, and the first load throttling port is not shielded completely; the motion trail of the actuating mechanism can be controlled by adjusting the overlapping degree between the second shoulder and the second load throttle orifice;
when the actuator moves in the reverse direction, when the actuator requires a driving force FhLoad force F with actuatorxThe resultant force between the first and second shoulders is opposite to the moving direction, the proportional electromagnet is controlled to enable the second shoulder to shield a part of the opening of the first load throttle orifice, the second load throttle orifice and the oil return orifice from the left sideForm a communicating relation between the two; and the motion trail of the actuating mechanism can be controlled by adjusting the overlapping degree between the second shoulder and the first load throttle orifice.
Has the advantages that:
(1) by utilizing the electro-hydraulic proportional direction valve provided by the invention, under some working states, high-pressure energy does not need to be provided for the electro-hydraulic proportional direction valve at all, and only low-pressure energy needs to be provided for the electro-hydraulic proportional direction valve to prevent an actuating mechanism from being sucked empty.
(2) When one load throttle orifice is subjected to high-pressure oil supply throttling control, the other load throttle orifice is in a fully opened state, and the throttle orifice has no throttling loss and also plays a role in saving hydraulic energy.
(3) According to the electro-hydraulic proportional direction valve provided by the invention, the position of the valve core is continuously adjusted in the same movement direction of the actuating mechanism, and sudden change and step of the position do not exist, so that the control of a hydraulic valve and the actuating mechanism is facilitated.
(4) According to the control method provided by the invention, under two working conditions without providing additional driving force, the control strategy is equivalent to a backpressure throttling control mode, and the traditional control method for throttling the inlet and the outlet simultaneously has the contradiction between outlet backpressure regulation and inlet flow regulation, so that the motion track of an actuating mechanism is not smooth, and the motion acceleration has sudden change. The control method provided by the invention eliminates the contradiction of the traditional control method because the inlet does not have throttling, and greatly reduces the control difficulty of the motion trail of the actuating mechanism.
Drawings
Fig. 1 is a schematic view of a valve core structure of a conventional proportional directional valve.
Fig. 2 is a composition diagram of an energy-saving electro-hydraulic proportional direction valve with an improved structure.
Fig. 3 is a dimensional relation diagram of an energy-saving electro-hydraulic proportional direction valve with an improved structure.
Fig. 4 is a schematic diagram of the structural size relationships between L1, L4 and L2, L3, L5.
Fig. 5 is a schematic diagram of the structural size relationship between L7 and L1, L5, and L6.
Fig. 6 is a schematic diagram of structural size relationships between L9 and L1, L2, L3, L4, L7, L8.
FIG. 7 is the hydraulic spool position under condition 1.
FIG. 8 shows the hydraulic spool position under condition 2.
FIG. 9 shows the hydraulic spool position under condition 3.
FIG. 10 is the hydraulic spool position under condition 4.
The hydraulic control valve comprises a hydraulic valve body 1, a valve core 2, a first load throttling port 3, a second load throttling port 4, a return spring 5, a proportional electromagnet 6, a displacement sensor sliding rod 7, a displacement sensor shell 8, a first oil supply port 9, an oil return port 10, a second oil supply port 11, a first shoulder 12, a second shoulder 13, a third shoulder 14, a fourth shoulder 15, a hydraulic cylinder 16, a cavity 17B, a cavity 18A, a piston rod 19, a piston 20 and a load 21.
Detailed Description
The invention is described in detail below by way of example with reference to the accompanying drawings.
The structure of the electro-hydraulic proportional directional valve with an improved structure, which is provided by the invention, is shown in the attached figure 2 and is realized according to the valve core structure of the conventional proportional directional valve shown in the attached figure 1. The electro-hydraulic proportional direction valve with the improved structure is composed of a hydraulic valve body 1, a valve core 2, a return spring 5, a proportional electromagnet 6, a displacement sensor sliding rod 7 and a displacement sensor shell 8.
The proportional electromagnet 6 is used for driving the valve core to move. As shown in the figure, the housing of the proportional electromagnet 6 is fixed at one end outside the hydraulic valve body 1, and one end of the valve core 2 extends out of the hydraulic valve body 1 and is inserted into the cavity of the proportional electromagnet 6. When a certain current is applied to the proportional electromagnet 6, a certain electromagnetic force can be generated to drive the valve core 2 to reciprocate in the hydraulic valve body 1.
The displacement sensor is used for sensing the position of the valve core. And a displacement sensor shell 8 is fixed at one end of the shell of the proportional electromagnet 6, and a displacement sensor sliding rod 7 of the displacement sensor shell is connected with the valve core 2. When the valve core 2 slides in the hydraulic valve body 1, the displacement sensor sliding rod 7 is driven to move together, so that the position of the valve core 2 in the hydraulic valve body 1 can be detected.
Two ends of the valve core 2 are abutted against the inner wall of the hydraulic valve body 1 through a return spring 5.
The hydraulic valve body 1 is provided with a first load throttle orifice 3, a second load throttle orifice 4, a first oil supply port 9, a second oil supply port 11 and an oil return port 10. As shown, the first load throttle 3 is on the left and the second load throttle is on the right. The first fuel feed port 9 is on the left and the second fuel feed port 11 is on the right. Wherein the first load throttle 3 and the second load throttle 4 are used for connecting with an actuator (such as a hydraulic cylinder or a hydraulic motor) to control the movement of the actuator; external hydraulic oil supplies hydraulic energy to the hydraulic valve through the first oil supply port 9 and the second oil supply port 11; the return oil of the hydraulic valve flows back to the oil tank of the hydraulic energy source through an oil return opening 10.
A first shoulder 12, a second shoulder 13, a third shoulder 14 and a fourth shoulder 15 are formed on the valve slide 2 from left to right. When the valve spool 2 moves in the hydraulic valve body 1, the second shoulder 13 and the third shoulder 14 form a certain shielding relationship with the first load throttle 3 and the second load throttle 4, and the first oil supply port 9, the second oil supply port 11 and the oil return port 10. Different valve core positions can form different shielding relations, so that different communication relations are formed.
The difference between the hydraulic valve provided by the invention and the existing hydraulic valve is that the four shoulders 12, 13, 14 and 15 on the valve core 2 of the hydraulic valve provided by the invention are different from the structural size relationship between the first load throttle orifice 3 and the second load throttle orifice 4, and when the first load throttle orifice 3 or the second load throttle orifice 4 of the hydraulic valve provided by the invention is controlled, the other load throttle orifice is in a fully opened state.
For the sake of convenience and intuition in the following description, fig. 2 is simplified to the form of fig. 3, and the return spring 5, the proportional electromagnet 6, the displacement sensor slide rod 7, and the displacement sensor housing 8 are omitted in fig. 3. The key structure size of the energy-saving hydraulic proportional direction valve provided by the invention is marked as follows:
the distance between the oil return port 10 and the first and second load chokes 3, 4 is designated L1, the width of the oil return port 10 is designated L2, the width of the load chokes 3, 4 and the shoulders 13, 14 is designated L3, the distance between the shoulders 13, 14 is designated L4, the distance between the first oil feed port 9 and the first load chokes 3 is designated L5, the distance between the second oil feed port 11 and the second load chokes 4 is also designated L5, the width of the first and second oil feed ports 9, 11 is designated L6, the distance between the shoulder 12 and the shoulder 13 is designated L7, the distance between the third shoulder 14 and the fourth shoulder 15 is designated L7, and the width of the shoulder 12 and the fourth shoulder 15 is designated L8; the distance of the left inner wall of the valve body 1 from the load throttle 3 is marked L9, and likewise the distance of the right inner wall of the valve body 1 from the load throttle 4 is also marked L9.
To ensure that the first land 13 or the third land 14 does not block the return port 10 when the land controls the opening degree of the load throttle, the minimum dimension of L1 should be no less than L3. As shown in fig. 4, the first land 13 starts moving to the right from the left side of the first load restriction 3, gradually blocking the first load restriction 3 until the first land 13 is located at the right side of the first load restriction 3, and the relationship between L1 and L3 is L1 equal to L3, at which time the first land 13 does not block the oil return port 10.
In order to ensure that when the opening degree of one load throttle orifice is controlled, the other load throttle orifice is only communicated with the oil return port 10, but not communicated with the oil supply port on the same side, the L4 is not larger than the L1-L3+ L2+ L1+ L3+ L5 which is 2 xL 1+ L2+ L5. As shown in fig. 4, the first shoulder 13 moves from the left side of the first load restriction 3 to the right until the first shoulder 13 is located at the right side of the first load restriction 3, and the mutual positional relationship between L4 and L1, L2 and L5 is L4 — 2 × L1+ L2+ L5, at which time the third shoulder 14 just blocks the second oil supply port 11.
In order to ensure that when one shoulder controls the opening degree of the load throttle orifice, the shoulder at the tail end of the valve core 2 on the same side does not shield the oil supply orifice, L7 is not larger than L1-L3+ L3+ L5+ L6, namely L1+ L5+ L6. As shown in fig. 5, the first shoulder 13 moves from the left side of the first load restriction 3 to the right until the first shoulder 13 is located at the right side of the first load restriction 3, and the mutual positional relationship between L7 and L1, L5 and L6 is L7 — L1+ L5+ L6, at which time the first shoulder 12 does not just block the first fuel supply port 9.
In order to ensure that the valve core 2 has enough movement space when the opening degree of the load throttle is controlled, L9 is not less than L3+ L4+ L7+ L8-2 XL 1-L2. As shown in fig. 6, the first land 13 moves to the right starting from the left side of the first load restriction 3 until the first land 13 is located on the right side of the first load restriction 3, L9 is exactly equal to L3+ L4+ L7+ L8-2 xl 1-L2, at which time the fourth land 15 moves just to the right side of the hydraulic valve body 1.
The dimensional relationship ensures that when the first shoulder 13 controls the opening degree of the first load throttle orifice 3, the second load throttle orifice 4 is in a fully open state, and similarly when the third shoulder 14 controls the opening degree of the second load throttle orifice 4, the first load throttle orifice 3 is in the fully open state, and the load throttle orifice in the fully open state is only communicated with the oil return orifice 10, and the shoulders 12 and 15 do not shield the oil supply orifice.
In the energy-saving hydraulic valve provided by the invention, the sizes L2, L3, L6, L8 and other structural sizes can be determined according to specific applications by referring to the structural size of the existing hydraulic valve, and no special requirement exists.
The energy-saving electro-hydraulic proportional direction valve with the improved structure is structurally different from the traditional hydraulic valve in that the position relation between a convex shoulder of a valve core and a load throttling opening is changed. Based on this change, the control manner of the hydraulic valve is different from the conventional control manner. The traditional hydraulic valve control only considers the movement direction of the actuating mechanism, and then controls the position of the valve core, and the invention needs to consider the movement direction of the actuating mechanism and the driving force direction required by the actuating mechanism. The present embodiment is described by taking a linear hydraulic cylinder as an example. A first load throttling port 3 of the energy-saving electro-hydraulic proportional directional valve is connected with a cavity A of the hydraulic cylinder, and a second load throttling port 4 of the energy-saving electro-hydraulic proportional directional valve is connected with a cavity B of the hydraulic cylinder.
Consider the force balance equation for an actuator:
Fh＝ma+Fx(1)
where m is the equivalent load mass of the actuator, a is the desired actuator motion acceleration, FxFor application to actuatorsThe load force of the structure. FhAt the driving force F for the driving force expected to be generatedhCan make the driving mechanism generate the expected motion acceleration a.
When in control, the control is firstly carried out according to the set motion trail of the actuating mechanism and the external load force F thereofxCalculating the driving force F required by the actuatorh(ii) a Then, according to the moving direction and the driving force direction required by the actuating mechanism, the control of the invention can be divided into 4 working conditions: (in the present embodiment, the piston rod is defined to project upward positively)
When the piston rod 19 is extended upward, the pressure difference between the chamber a 18 and the chamber B17 generates a driving force F on the piston 20hIf driving force FhGreater than the load force FxThe resultant force direction is the same as the movement direction, and the piston rod moves in an accelerated extension mode and belongs to the working condition 1; if driving force FhLess than the load force FxThe resultant force direction is opposite to the movement direction, and the piston rod performs deceleration extension movement, belonging to the working condition 2. When the piston rod 19 moves in retraction, the pressure difference between the chamber a 18 and the chamber B17 generates a driving force F on the piston 20hIf driving force FhGreater than the load force FxThe resultant force direction is the same as the movement direction, and the piston rod performs accelerated retraction movement, belonging to the working condition 3; if driving force FhLess than the load force FxAnd the resultant force direction is opposite to the movement direction, and the piston rod performs deceleration retraction movement, belonging to the working condition 4.
In each of the four working conditions of the invention, the hydraulic valve core 2 needs to be controlled to a corresponding working area, and then the size of the throttling opening is adjusted in the working area according to position feedback. The specific control method under each working condition is as follows:
(1) working condition 1
When the piston rod 19 of the hydraulic cylinder 16 is accelerated to move in an extending way, the driving force is required to be the same as the moving direction of the piston rod 19 so as to overcome the resistance to the piston rod.
In this state, the proportional solenoid 6 is controlled to move the spool 2 until the first land 13 and the first load restriction 3 are brought into a positional relationship as shown in fig. 7. At this time, the first land 13 blocks a partial opening of the first load restriction 3 from the right side, that is, a communicating relationship is formed between the first load restriction 3 and the first oil feed port 9. In this state, high pressure oil is supplied to the hydraulic valve through the first supply port 9, and after entering the hydraulic valve, the high pressure oil enters the chamber a 18 of the hydraulic cylinder 16 through the first load throttle 3. The high pressure oil will create a pressure drop after passing through the first load restriction 3, the magnitude of the pressure drop depending on the opening degree of the first load restriction 3.
The hydraulic oil in the cavity B17 enters the hydraulic valve through the second load throttle 4 and then flows back to the oil tank through the oil return port 10. At this time, the second load throttle 4 is not blocked at all, so that no throttling action is generated and no pressure loss is caused.
By adjusting the overlap between the first shoulder 13 and the first load restriction 3, i.e. by adjusting the opening of the first load restriction 3, the pressure of the hydraulic oil flowing into the a chamber 18 can be controlled, so that the pressure difference between the a chamber 18 and the B chamber 17 can be controlled, and thus the driving force applied to the piston 20, and finally the trajectory of the extension movement of the piston rod 19 can be controlled.
(2) Working condition 2
When the piston rod of the hydraulic cylinder extends and moves in a decelerating way, the driving force required at the moment is opposite to the moving direction of the piston rod so as to overcome the action of the external force.
In this state, the proportional solenoid 6 is controlled to move the spool 2 until the third land 14 and the second load restriction 4 are brought into a positional relationship as shown in fig. 8. At this time, the third land 14 shields a partial opening of the second load throttle 4 from the right side, that is, a communication relationship is formed among the second load throttle 4, the first load throttle 3, and the return port 10. In this state, the hydraulic oil in the B chamber 17 flows into the hydraulic valve through the second load throttle 4.
Meanwhile, hydraulic oil needs to be supplied to the cavity A18 to prevent the cavity A from being sucked empty, if the hydraulic cylinder is a symmetrical hydraulic cylinder (namely, as shown in figure 8, piston rods are arranged in the two cavities and the diameters of the piston rods are the same), the hydraulic oil flowing out of the cavity B17 is exactly equal to the hydraulic oil needed by the cavity A18, and at the moment, additional hydraulic oil does not need to be supplied to the cavity A; if the hydraulic cylinder is asymmetric (i.e. only one of the chambers has a piston rod, and the other chamber does not have a piston rod), when the B chamber 17 has a piston rod, the effective area of the piston 20 in the a chamber 18 is larger than that in the B chamber 17, so the hydraulic oil flowing out of the B chamber 17 is not enough to meet the requirement of the a chamber 18, and therefore, the oil needs to be sucked from the hydraulic oil tank in a self-priming manner to provide a part of the hydraulic oil for the a chamber 18, whereas when the piston rod is in the a chamber 18, the hydraulic oil flowing out of the B chamber 17 is more than that required by the a chamber 18, and the surplus part of the hydraulic oil flows back to the oil tank, i.e. at this time, the extra hydraulic.
By adjusting the overlapping degree between the third shoulder 14 and the second load throttle 4, that is, the opening degree of the second load throttle 4, the pressure of the hydraulic oil in the B cavity 17 can be adjusted, so that the pressure difference between the a cavity 18 and the B cavity 17 can be controlled, the driving force applied to the piston 20 can be further controlled, and finally the trajectory of the extending motion of the piston rod 19 can be controlled. It can be seen that the hydraulic valve does not need to be supplied with high-pressure hydraulic oil at all, and the purpose of saving energy is achieved.
(3) Working condition 3
When the piston rod 19 of the hydraulic cylinder 16 accelerates the retraction movement, the driving force is required to be the same as the movement direction of the piston rod 19 so as to overcome the resistance force.
In this state, the proportional solenoid 6 is controlled to move the spool 2 until the third land 14 and the second load restriction 4 are brought into a positional relationship as shown in fig. 9. At this time, the third land 14 blocks a partial opening of the second load restriction 4 from the left side, that is, a communication relationship is formed between the second load restriction 4 and the second oil feed port 11. In this state, high pressure oil is supplied to the hydraulic valve through the second oil supply port 11, and after entering the hydraulic valve, the high pressure oil enters the B chamber 17 of the hydraulic cylinder 16 through the second load throttle port 4. The high pressure oil will create a pressure drop after passing through the second load restriction 4, the magnitude of the pressure drop depending on the degree of opening of the second load restriction 4.
The hydraulic oil in the chamber a 18 enters the hydraulic valve through the first load throttle 3 and flows back to the oil tank through the oil return port 10. At this time, the first load throttle 3 is not shielded at all, and therefore, no throttling action is generated, and no pressure loss is caused.
By adjusting the overlap between the third shoulder 14 and the second load restriction 4, i.e. the opening of the second load restriction 4, the pressure of the hydraulic oil flowing into the B chamber 17 can be controlled, so that the pressure difference between the a chamber 18 and the B chamber 17, and thus the driving force applied to the piston 20, and finally the retraction path of the piston rod 19, can be controlled.
(4) Working condition 4
When the piston rod of the hydraulic cylinder performs the deceleration retraction movement, the required driving force F is generated at the momenthShould be opposite to the direction of movement of the piston rod to overcome the effect of this external force.
In this state, the proportional solenoid 6 is controlled to move the spool 2 until the first land 13 and the first load restriction 3 are brought into a positional relationship as shown in fig. 10. At this time, the first land 13 blocks a partial opening of the first load restriction 3 from the left side, that is, a communication relationship is formed between the first load restriction 3, the second load restriction 4, and the return port 10. In this state, the hydraulic oil in the a chamber 18 flows into the hydraulic valve through the first load throttle 3.
Meanwhile, hydraulic oil needs to be supplied to the cavity B17 to prevent the cavity B from being sucked empty, if the hydraulic cylinder is a symmetrical hydraulic cylinder (namely, as shown in the attached drawing 10, piston rods are arranged in the two cavities and the diameters of the piston rods are the same), the hydraulic oil flowing out of the cavity A18 is exactly equal to the hydraulic oil needed by the cavity B17, and at the moment, additional hydraulic oil does not need to be supplied to the cavity B; if the hydraulic cylinder is asymmetrical (i.e. only one of the chambers has a piston rod, the other chamber does not have a piston rod), when the B chamber 17 has a piston rod, the effective area of the piston 20 in the a chamber 18 is larger than that in the B chamber 17, so the hydraulic oil flowing out of the a chamber 18 exceeds the hydraulic oil required by the B chamber 17, and therefore the surplus hydraulic oil flows back into the oil tank, whereas when the piston rod is in the a chamber 18, the hydraulic oil flowing out of the a chamber 18 is less than the hydraulic oil required by the B chamber 17, and the oil needs to be sucked from the hydraulic oil tank in a self-priming manner to provide a part of the hydraulic oil for the B chamber 18.
By adjusting the overlap between the first shoulder 13 and the first load restriction 3, i.e. adjusting the opening of the first load restriction 3, the pressure of the hydraulic oil in the a-chamber 18 can be adjusted, so that the pressure difference between the a-chamber 18 and the B-chamber 17 can be controlled, and further the driving force applied to the piston 20, and finally the trajectory of the retraction movement of the piston rod 19 can be controlled. It can be seen that the hydraulic valve does not need to be supplied with high-pressure hydraulic oil at all, and the purpose of saving energy is achieved.
The situation that the hydraulic valve controls the hydraulic motor is similar to the situation that the hydraulic valve controls the hydraulic cylinder, and the situation can also be divided into the 4 working states. Only the linear motion of the extension and retraction of the piston rod of the hydraulic cylinder is changed into the clockwise and anticlockwise rotation of the rotating shaft of the hydraulic motor.
In summary, the above description is only a preferred embodiment of the present invention, and is not intended to limit the scope of the present invention. Any modification, equivalent replacement, or improvement made within the spirit and principle of the present invention should be included in the protection scope of the present invention.
Claims (3)
1. An energy-saving electro-hydraulic proportional direction valve with an improved structure comprises a hydraulic valve body (1), a valve core (2), a proportional electromagnet (6) and a displacement sensor; the hydraulic valve body (1) is provided with a first load throttling port (3), a second load throttling port (4), a first oil supply port (9), a second oil supply port (11) and an oil return port (10); the valve core (2) is sequentially provided with a first shoulder (12), a second shoulder (13), a third shoulder (14) and a fourth shoulder (15); it is characterized in that the preparation method is characterized in that,
setting the distance between the first load throttle (3) and the oil return (10) as L1, and similarly, setting the distance between the second load throttle (4) and the oil return (10) as L1; the width of the oil return port (10) is L2, and the widths of the first load throttle port (3), the second load throttle port (4), the second shoulder (13) and the third shoulder (14) are L3; the distance between the second shoulder (13) and the third shoulder (14) is L4; the distance between the first oil supply port (9) and the first load throttle (3) is L5, and the distance between the second oil supply port (11) and the second load throttle (4) is L5; the widths of the first oil supply port (9) and the second oil supply port (11) are L6; the distance between the first shoulder (12) and the second shoulder (13) is L7, and similarly, the distance between the third shoulder (14) and the fourth shoulder (15) is L7; the first shoulder (12) and the fourth shoulder (15) have a width L8; the distance between the left inner wall of the valve body (1) and the first load throttling port (3) is L9, and the distance between the right inner wall of the valve body (1) and the second load throttling port (4) is L9;
L1≥L3；
L4≤2×L1+L2+L5；
L7≤L1+L5+L6；
L9≥L3+L4+L7+L8-2×L1-L2。
2. the improved energy-saving electrohydraulic proportional direction valve of claim 1, further comprising a return spring (5) disposed between two end faces of said spool (2) and an inner wall of said hydraulic valve body (1).
3. A control method of an energy-saving electro-hydraulic proportional directional valve according to any one of claims 1-2, wherein a first load throttle (3) of the energy-saving electro-hydraulic proportional directional valve is connected with a cavity A of an actuator, a second load throttle (4) is connected with a cavity B of the actuator, and the movement direction generated when the first load throttle supplies oil to the actuator is defined as a forward direction; it is characterized by that according to the set movement locus of actuating mechanism and load force F of actuating mechanismxCalculating the driving force F required by the actuatorhAnd controlling according to the situation:
when the actuating mechanism moves in the positive direction and the actuating mechanism requires the driving force FhLoad force F with actuatorxThe resultant force between the first and second load throttling mouths is the same as the movement direction, the proportional electromagnet (6) is controlled to enable the second shoulder (13) to shield a partial opening of the first load throttling mouth (3) from the right side, a communication relation is formed between the first load throttling mouth (3) and the first oil supply mouth (9), and the second load throttling mouth (4) is not shielded at all; the motion trail of the actuating mechanism can be controlled by adjusting the overlapping degree between the second shoulder (13) and the first load throttle orifice (3);
when the actuating mechanism moves in the positive direction and the actuating mechanism requires the driving force FhLoad force F with actuatorxThe resultant force between the first load throttling orifice and the second load throttling orifice is opposite to the movement direction, the proportional electromagnet (6) is controlled, the third shoulder (14) shields a partial opening of the second load throttling orifice (4) from the right side, and the second load throttling orifice (4), the first load throttling orifice (3) and the oil return port (10) form a communication relation; adjusting the third shoulder(14) The degree of overlap between the second load throttle orifice (4) and the second load throttle orifice can control the motion track of the actuating mechanism;
when the actuating mechanism moves reversely and the actuating mechanism requires the driving force FhLoad force F with actuatorxThe resultant force between the first load throttling orifice and the second load throttling orifice is the same as the movement direction, the proportional electromagnet (6) is controlled to enable the third shoulder (14) to shield a partial opening of the second load throttling orifice (4) from the left side, a communication relation is formed between the second load throttling orifice (4) and the second oil supply orifice (11), and the first load throttling orifice (3) is not shielded at all; the motion trail of the actuating mechanism can be controlled by adjusting the overlapping degree of the third shoulder (14) and the second load throttle orifice (4);
when the actuator moves in the reverse direction, when the actuator requires a driving force FhLoad force F with actuatorxThe resultant force between the first and second load throttle orifices (3, 4) is opposite to the movement direction, the proportional electromagnet (6) is controlled to enable the second shoulder (13) to shield a partial opening of the first load throttle orifice (3) from the left side, and a communication relation is formed among the first load throttle orifice (3), the second load throttle orifice (4) and the oil return port (10); the motion trail of the actuating mechanism can be controlled by adjusting the overlapping degree between the second shoulder (13) and the first load throttle orifice (3).
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| DE3605312A1 (en) * | 1985-02-22 | 1986-08-28 | Linde Ag, 6200 Wiesbaden | Spool valve |
| CN2833201Y (en) * | 2005-06-08 | 2006-11-01 | 佟世儒 | Hydraulic reversing valve capable of avoiding external leakage |
| CN201554709U (en) * | 2009-10-13 | 2010-08-18 | 宁波汉商液压有限公司 | Multi-loop selection valve |
| CN103133448B (en) * | 2013-02-28 | 2015-04-15 | 北京理工大学 | Energy-saving hydraulic valve with independently-controllable load throttle mouths and hydraulic valve control system |
| CN106704622B (en) * | 2016-11-22 | 2019-08-02 | 浙江大学 | Proportional reversing valve lightweight spool and proportional reversing valve based on SLM technology |
| CN108797696A (en) * | 2018-07-06 | 2018-11-13 | 佛山信卓派思机械科技有限公司 | A kind of excavator operation mode switching control valve |
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