JPH0240913B2 - - Google Patents

Description

[Detailed description of the invention]

The present invention relates to a pressure control valve obtained by adding excellent instantaneous actuation characteristics to a balanced pressure control valve having good pressure override characteristics. Generally, a pressure control valve used in a hydraulic circuit has the function of a relief valve that controls the pressure so that the flow rate that cannot be consumed by the actuator is discharged to the tank and the maximum pressure of the hydraulic circuit is kept constant.
It functions as a safety valve that instantly avoids sudden surge pressure when it occurs in the hydraulic circuit. From this point of view, pressure control valves have static characteristics that include pressure override characteristics (pressure override characteristics are the cracking pressure at which the valve begins to open when the circuit pressure increases, and the total flow rate when the rated flow rate flows when the pressure increases further). pressure) and
The dynamic characteristics include instantaneous operating characteristics (instantaneous operating characteristics are
How quickly does it respond to pressure increases in the hydraulic circuit?
The problem is whether the relief operation can be performed without generating surge pressure (the degree of responsiveness). Therefore, no matter how good the pressure override characteristics in a steady state are for a pressure control valve,
If the instantaneous operation characteristics as a safety valve are poor,
Not only can piping systems and hydraulic equipment be damaged, but human life can also be threatened, and the intended functionality cannot be achieved. From the above points, the instantaneous operating characteristics of the pressure control valve are:
From the state of pressure change over time, Figure 1a
It can be classified into four types shown in ~d. That is, the ideal one shown in Fig. 1a, the quasi-ideal one in which the pressure rise at the beginning of instantaneous actuation is quite close to the set pressure shown in Fig. 1b, and the quasi-ideal one in which the pressure rise at the beginning of instantaneous actuation is shown in Fig. 1c. It can be classified into undesirable valves that may become too high and cause equipment damage, and valves that are desirable as safety valves but are unreliable in that they react sensitively to pressure fluctuations and maintain a constant pressure, as shown in Figure 1d. Most of the instantaneous operating characteristics of commonly used pressure control valves belong to the types shown in Figures 1b and 1c, whereas conventional pressure control valves have ideal characteristics as shown in Figure 1a. can't get it. Therefore, by using a combination of the conventional direct type pressure control valve, which has good instantaneous actuation characteristics but lacks pressure override characteristics, and the balanced type pressure control valve, which has slightly inferior instantaneous actuation characteristics but has good pressure override characteristics, it is possible to Efforts have been made to approximate the ideal type shown in Figure 1a. That is, in Fig. 2, the characteristics of the direct pressure control valve increase proportionally from the cracking pressure Pd,
Among the characteristics of a balanced pressure control valve that continues from the main valve cracking pressure Ps, a direct type pressure control valve with good instantaneous operation characteristics is used during the flow rate Qd up to the intersection point A of both characteristics, and a balance type pressure control valve with good instantaneous operation characteristics is used during the flow rate Qs. The purpose is to use a type pressure control valve to obtain the following characteristics. However, this configuration has the disadvantage that two types of pressure control valves are required, resulting in an increased size of equipment and a complicated structure. There have also been attempts to combine two types of valve springs, large and small, and use the smaller valve spring until a predetermined set pressure, and then use the combination of large and small valve springs when the flow rate exceeds a predetermined flow rate. Similarly, there was a drawback that the structure became complicated. In view of the shortcomings of the pressure control valves according to the prior art, the present invention focuses on the influence of back pressure acting on the sub-valve of a balanced pressure control valve, and by controlling this sub-valve back pressure, the present invention provides no new advantages. By imparting the ideal characteristics shown in Figure 1a to a balanced pressure control valve with the same conventional configuration without adding additional equipment elements,
The aim is to improve the instantaneous operation characteristics. Hereinafter, the pressure control valve of the present invention will be explained. First, the operating principle of the present invention will be explained using a conventional pressure control valve shown in FIG. In FIG. 3, an oil passage 2 leading from a pump to a hydraulic device and a tank port 3 leading to a tank are formed in a casing 1, and a valve body 5 of a pressure control valve 4 is attached to the casing 1 by means of screwing, for example. tightly attached. A pump port 6 that opens to the oil passage 2 is formed in the valve body 5, and a main valve outlet 7 that constitutes a part of the tank port 3 is formed. A sub-valve seat member 8 and a sleeve 9 are fixedly attached to the valve body 5, and the sub-valve seat member 8 defines a main valve chamber 10 in the valve body 5 and a sub-valve chamber 11 in the sleeve 9. are doing. Note that in the drawings, the casing 1 and the valve body 5 are shown as separate members for reasons of workmanship and manufacturing, but they may be integrally constructed. In the valve body 5, a valve seat 12 for separating and seating the main valve is formed between the pump port 6 and the main valve outlet 7 and has an inclined shape whose diameter increases toward the right in the figure. A main valve 13 is slidably provided in the chamber 10 and is seated on and off the valve seat 12, and the main valve 13 is always urged in the valve-closing direction by a main valve valve spring 14. Further, the main valve 13 is provided with a throttle passage 15 that serves as an oil passage for constantly communicating the main valve chamber 10 with the pump port 6. On the other hand, a small hole 16 is bored in the auxiliary valve seat member 8 and serves as an oil passage that constantly communicates the main valve chamber 10 and the auxiliary valve chamber 11. On the side of the auxiliary valve chamber 11, a valve for separating and seating the auxiliary valve is formed. A seat 17 is formed, and a sub-valve 18 is provided in the sub-valve chamber 11 and is seated on and off the valve seat 17, and the sub-valve 18 is always biased in the valve-closing direction by a sub-valve valve spring 19. . 20 is a screw for adjusting the spring force of the valve spring 19 for the auxiliary valve. Reference numeral 21 denotes a discharge hole bored in the sleeve 9 to discharge the pressure oil in the auxiliary valve chamber 11. The discharge hole 21 communicates with a discharge hole 22 formed in the valve body 5, and the discharge hole 22 connects an oil passage 23. Tank port 3 of casing 1 through
It communicates with the side discharge hole 24. And oil road 2
An adjustable back pressure valve 25 that can adjust the back pressure of the auxiliary valve 18 is provided in the middle of the valve 3, and the discharge holes 21, 22,
A relief oil passage includes an oil passage 23, a back pressure valve 25, and the like. In such a pressure control valve, the main valve 13 is seated on the valve seat 12 before operation, and the sub valve 18 is also seated on the valve seat 17.
is seated. Now, when the pressure inside the pump port 6 increases, the pressure inside the main valve chamber 10 also increases, and the sub valve 18
is the valve spring 19 due to the pressure acting on its pressure receiving area.
The valve opens against the For this reason, the pressure oil in the main valve chamber 10 flows through the small hole 16, the auxiliary valve chamber 11, the discharge holes 21, 22,
It is relieved to the relief port 3 through the oil passage 23, the back pressure valve 25, and the discharge hole 24 in order, and the main valve 13
The main valve 13 opens against the valve spring 14 due to the pressure difference between the front and rear, and the pressure oil from the pump port 6 is relieved from the main valve outlet 7 to the tank port 3. Hereinafter, regarding such a pressure control valve 4, the movements of the sub valve 18 and the main valve 13, the flow rate characteristics of the throttle passage 15, etc. will be considered. () About the sub-valve 18 Relief flow rate flowing through the sub-valve (=Q _p ) Balance of forces acting on the sub-valve However, C _P ...Flow coefficient of sub-valve g...Gravity acceleration r...Specific gravity of oil k _p ...Spring constant of valve spring 19 x _S ...Deflection of valve spring 19 when installed x...Sub-valve Displacement of the valve d _P ... Valve seat diameter of the valve seat 17 θ _P ... Half apex angle of the sub-valve P ₂ ... Pressure inside the main valve chamber 10 P _TP ... Sub-valve back pressure C _PV ...... of the sub-valve part Speed coefficient A _P ...Pressure receiving area of the sub-valve (A _P = π/4d _P ² ) α _P ...Jet direction of the outflow jet from the sub-valve () About the throttle passage 15 Flow rate flowing through the throttle passage (=Q _s ) However, C _S ... Flow coefficient of the throttle passage 15 A _S ... Cross-sectional area of the throttle passage 15 P ₁ ... Circuit pressure on the pump port 6 side () Regarding the main valve 13 Relief flow rate flowing through the main valve section (=Q _n ) Balance of forces acting on the main valve However, C _n ...flow coefficient of the main valve section d _n1 ... valve diameter of the main valve on the pump port 6 side when the main valve 13 is seated on the valve seat 12 d _n2 ... within the main valve chamber 10 Valve diameter of the main valve y _S ... Deflection of the valve spring 14 when installed y ... Displacement of the main valve θ _n ... Valve seat angle of the valve seat 12 P _Tn ... Main valve back pressure A _1n ... Pressure receiving area (A _1n = π/4d _n1 ² ) A _2n ... Pressure receiving area (A _2n = π/4d _n2 ² ) A _3n ... Pressure receiving area (A _3n = A _2n −A _1n ) C _nv ... of the main valve section Velocity coefficient α _n ...Jet direction of main valve outflow jet Next, rearranging the relationships in equations (1) to (5) above, the displacement x of the sub valve 18 can be calculated by substituting equation (1) into equation (2). Then, x=A _P (P ₂ −P _TP )−k _P x _S ／2C _P πd _P sinθ _P (
P ₂ −P _TP )cosα _P +k _P Therefore, if the deflection x _S of the valve spring 19 of the sub-valve 18 is determined, x is unique for each value of (P ₂ −P _TP ). Determined. Therefore, it can be determined by inserting the relationship between (P ₂ −P _TP ) and x into the subvalve flow rate Q _P in equation (1). On the other hand, in a steady state where the sub-valve 18 is open,
Since the relationship Q _P = Q _S , the relationship between P ₁ and P ₂ is (3)
Transforming the formula, becomes. Considering the above relationship, the deflection x _S of the auxiliary valve spring 19 is kept at the same value using the adjustment screw 20,
For example, P ₂ = 180Kg/cm ² when back pressure P _TP = 0, and P _TP
P ₂ when = 5Kg/cm ² = 185Kg/cm ² is P ₂ -
Since P _TP =180Kg/cm ² , the sub-valve displacement x and the sub-valve flow rate Q _P have the same value. Also, P ₁ is also P _TP = 0
If P ₁ = 230Kg/cm ² then P _TP = 5
When Kg/cm ² , P ₁ =235Kg/cm ² . On the other hand, for the main valve 13, substituting equation (4) into equation (5) yields y=P ₁ A _1n +P _Tn A _3n −P ₂ A _2n −K _n y _S /2C _n πd _n sinθ _n (P
₁ −P _Tn )cosα _n +k _n , the main valve flow rate Q _n can be expressed as (P ₁ −
It can be determined by entering the relationship between P _Tn ) and y. In order to know the influence of back pressure acting on the sub-valve 18,
The above relationship, which has been analyzed analytically, will be explained with reference to FIG. First, in FIG. 4a, the back pressure valve 25 is opened and the sub valve 1 is opened.
8 shows the relationship between the auxiliary valve relief flow rate Q _P , the circuit pressure P ₁ , and the internal pressure P ₂ of the main valve chamber 10 when the back pressure P _TP acting on the valve 8 is controlled to be 0. Next, in FIG. 4b, the back pressure valve 25 is adjusted with the adjusting screw 20 fixed, and the auxiliary valve back pressure P _TP is applied to the auxiliary valve 18.
=ΔP The relationship between the relief flow rate Q _P , the circuit pressure P ₁ , and the main valve chamber pressure P ₂ is shown when _TP is applied. Fig. 4b is compared with Fig. 4a,
The relationship is such that the pressures P ₁ and P ₂ are shifted in parallel by ΔP _TP with respect to the relief flow rate Q _P. Next, FIG. 4c shows the relationship between the main valve relief flow rate Q _n and the circuit pressure P ₁ in the same way as FIGS. 4a and b, and as can be seen from FIG. By adjusting the back pressure P _TP to be generated using the back pressure valve 25, compared to the case where P _TP = 0, even when P _TP = ΔP _TP is increased, the circuit pressure P ₁ becomes ΔP _TP for the same flow rate Q _n .
The balance will be balanced at a higher value. As is clear from Fig. 4 a to c, even when the sub-valve back pressure P _TP is increased by P _TP =ΔP _TP , for the same sub-valve relief flow rate Q _P and main valve relief flow rate Q _n ,
Since the circuit pressure P ₁ is simply moved in parallel to a position higher by ΔP _TP , the back pressure acting on the sub-valve 18 is
By controlling P _TP , Fig. 4 d and Fig. 4
The characteristics shown in Figure e are obtained. That is, in FIG. 4d, Q _n0 indicates the main valve cracking position, and after the main valve 13 cracks at this Q _n0 , the back pressure P _TP that acts on the sub valve 18 until the main valve flow rate Q _n1 is reached. This shows the characteristics when controlled to increase continuously from 0 to ΔP _TP and keep P _TP =ΔP _TP constant after reaching Q _n1 . in this way,
By varying the back pressure P _TP , characteristics similar to those shown in FIG. 2 can be obtained. However, although the diagram shown in Figure 2 or Figure 4 d is preferable when considering instantaneous actuation characteristics, in terms of relief performance, the line _diagram shown in Figure 2 or 4 is preferable.
In the section _{n0 n1} in Figure d, the main valve 13 opens rapidly and cannot relieve a large amount of pressure oil into the tank, sacrificing the pressure overwrite characteristic, which is not a desirable performance. . On the other hand, Fig. 4e shows the characteristics with the drawbacks as shown in Fig. 4d removed. At the position of flow rate Q _n0 , the main valve 13 is cracked with the auxiliary valve back pressure P _TP = 0, and when the flow rate is _{n0 n1} . The auxiliary valve back pressure P _TP is maintained at 0 in a small section, and then the auxiliary valve back pressure P _TP is rapidly increased in a small section of the flow rate _{n1 n2} , and after the flow rate Q _n2 , the auxiliary valve back pressure P _TP = ΔP It is controlled to keep _TP at a constant value. As described above, the balance type pressure control valve shown in FIG. 3 and the back pressure acting on the sub-valve 18 of the pressure control valve
By combining it with a back pressure valve 25 that controls P _TP , a pressure control valve with good instantaneous actuation characteristics as shown in FIG. 4e can be obtained without sacrificing pressure override characteristics. However _, in the case shown in Fig. 3, _the principle of Although it is technically superior, it is an extremely complicated configuration and is not practical. An object of the present invention is to provide a pressure control valve that can improve instantaneous actuation characteristics without using the back pressure valve according to the prior art described above. In order to achieve this object, the feature of the means adopted in the present invention is that the valve seat for separating and seating the main valve is formed as an inclined valve seat with a predetermined valve seat angle that expands in diameter toward the outlet side. , a relief oil passage is bored in the valve body as an oil hole, and the passage outlet portion of the relief passage is opened slightly outward in the radial direction from an extension line of the inclined shape of the main valve seating/seating valve seat; Near cracking of the main valve, the main valve jet does not collide with the passage outlet of the relief oil passage, but as the main valve jet increases, it collides with the passage outlet, increasing the auxiliary valve back pressure, and maintaining the predetermined relief flow rate. Thereafter, the configuration is such that the back pressure of the sub-valve is maintained at a predetermined high pressure. With this configuration, the circuit pressure can have the characteristics shown in Figure 4e in response to changes in the relief flow rate without adding new equipment elements such as a back pressure valve, and the pressure override characteristics can be sacrificed. Instantaneous operating characteristics having the ideal characteristics shown in FIG. Embodiments of the present invention will be described below with reference to FIGS. 5 and 6. Components that are the same as those of the prior art pressure control valve shown in FIG. It is formed as one or more oil holes located and drilled in the axial direction, and one end of the oil hole communicates with a discharge hole 21 that communicates with the sub-valve chamber 11, and the other end of the oil hole is a passage outlet. 32 and opens to the main valve outlet 7. Here, looking at the relationship between the valve seat 12 and the passage outlet part 32, the extension line of the valve seat angle θ _n of the valve seat 12 is a point 33 ( (Actually it is annular) and the main valve outlet 7
It is formed so that it intersects with the inner wall of. Therefore, immediately after the main valve 13 cracks, the valve seat 12 and the main valve 1
The main valve jet flow ejected from the gap with the point 3 in the direction indicated by the solid line arrow X in FIG. 6 collides with the point 33. Next, the main valve 13 opens to the position shown by the imaginary line in FIG. The arrangement is such that the passage outlet portion 32 is closed. In this case, as a result of various experiments, by rounding the tip 12A of the valve seat 12, the main valve jet flow indicated by the imaginary line Y
I learned that it definitely blocks the. The pressure control valve of the present invention is constructed as described above.Next, its operation will be explained with reference to the explanatory diagram shown in FIG. 4e. First, when the main valve 13 is still closed and the auxiliary valve 18 has cracked, there is no jet flow from the main valve 13, and the internal oil in the auxiliary valve chamber 19 flows through the discharge hole 21.
The oil is relieved from the passage outlet part 32 to the main valve outlet 7 via the oil passage 31 and returned to the tank from the tank port 3. However, the tank port 3 is a passage where the return oil from each actuator joins, and if there is a large pressure loss, the energy of the hydraulic circuit will be wasted, so the pressure inside the tank port 3 is small, and therefore, Sub-valve back pressure near passage outlet 32
P _TP ≒0. Next, immediately after the main valve 13 cracks due to the differential pressure acting before and after the main valve 13 (see Fig. 4 e),
At position Q _n0 ), the main valve jet from the gap between the valve seat 12 and the main valve 13 is ejected in the direction of the solid line arrow X in FIG.
The passage outlet portion 32 will not be obstructed by the collision. Therefore, the auxiliary valve back pressure at the passage outlet portion 32 becomes P _TP =0, as described above. When the valve opening degree (relief flow rate) of the main valve 13 increases and the main valve jet flow increases, this main valve jet flow changes from the solid line arrow X toward the imaginary line Y and reaches the opening end of the passage outlet portion 32. do. This position is the relief flow rate in Figure 4 e.
Q is the position of _n1 . Therefore, in the section of the flow rate _{n0 n1} , the auxiliary valve back pressure P _TP = 0, and as mentioned above, no surge pressure is generated in the circuit pressure P ₁ on the pump port 6 side, that is, if the instantaneous operation characteristics are good. The main valve 13 can be opened in the closed state. Next, when the opening degree of the main valve 13 further increases,
The main valve jet moves to the state shown by the virtual line Y, and the passage outlet portion 32 is sequentially closed by the dynamic pressure of this main valve jet, gradually increasing the auxiliary valve back pressure in the oil passage 31 to a predetermined level. When the relief flow rate Q _n2 is reached, the passage outlet portion 32 is completely closed. That is, this is the fourth
Figure e shows the section where the flow rate is _{n1 n2} , and the auxiliary valve back pressure P _TP =ΔP _TP is given at the flow rate Q _n2 . Furthermore, after the relief flow rate reaches Q _n2 , the flow velocity of the main valve jet is as follows in equation (5):

Since it is proportional to [Formula], the auxiliary valve back pressure ΔP _TP can be kept constant, so a predetermined relief performance can be maintained. In the embodiments of the present invention, the balanced pressure control valve is illustrated as a balanced piston type pressure control valve in which the main valve 13 is provided with the throttle passage 15, but the present invention is not limited to this; The main valve chamber 10 may be connected to the oil passage 2 via a throttle passage. Also, the fourth
In order to obtain the characteristics shown in Figure e, valve seat 1 for main valve separation and seating.
2 and the relief passage 31 are not limited to the illustrated shapes, and can adopt various shapes. However, as shown in FIG. 7, the extension line X' from the valve seat 12 intersects with the inner wall of the main valve outlet 7 at a point 33' radially outside the opening end of the passage outlet 32. In other words, the passage outlet 32
Of course, if the relationship is such that the main valve jet flow is not affected, the desired performance cannot be obtained. Further, as described above, by rounding the tip 12A of the valve seat 12, it is possible to provide a degree of freedom in design regarding the relationship between the valve seat 12 and the passage outlet portion 32. The pressure control valve according to the present invention has been described in detail above, and has many effects listed in the following sections. In the vicinity of the main valve cracking, the main valve jet does not collide with the passage outlet of the oil hole forming the relief oil passage, and as the relief flow rate increases, the passage outlet is closed by the dynamic pressure of the main valve jet. Since the sub-valve back pressure is increased, the instantaneous actuation characteristics can be improved without sacrificing the pressure override characteristics. According to the present invention, pressure override is small and relief performance is reduced without using a combination of a direct type pressure control valve and a balanced type pressure control valve or using two types of large and small valve springs as in the prior art. The excellent balance type pressure control valve can be used in its conventional shape. In relation to the above item, it is sufficient to change the relationship between the passage outlet of the relief oil passage and the main valve seating/separation valve seat without adding a back pressure valve with variable set pressure. , there is no need to drastically change the design dimensions, and it can be manufactured at low cost. Since a direct type poppet can be used as the sub-valve, the response is good, and since no surge pressure is generated, it can fully function as a safety valve.

[Brief explanation of drawings]

Figures 1 a to d are explanatory diagrams showing the instantaneous actuation characteristics of the pressure control valve, and Figure 2 shows the performance when a direct type pressure control valve and a balance type pressure control valve are combined to improve the instantaneous actuation characteristics. An explanatory diagram, FIG. 3 is a vertical cross-sectional view of a pressure control valve according to the prior art, and FIG. 4 a
~e is an explanatory diagram of the characteristics analyzed using the pressure control valve shown in Fig. 3, Fig. 5 is a vertical sectional view showing the pressure control valve according to the present invention, and Fig. 6 is a partially enlarged view of Fig. 5. , FIG. 7 is a partially enlarged view similar to FIG. 6 showing the shape when the desired performance cannot be obtained. 1...Casing, 3...Tank port, 4...
...Pressure control valve, 5...Valve body, 6...Pump port, 7...Main valve outlet, 8...Sub-valve seat member, 10
...Main valve chamber, 11...Sub-valve chamber, 12...Valve seat, 1
3... Main valve, 15... Throttle passage, 16... Small hole,
18...Sub-valve, 31...Relief oil passage, 32
...Aisle exit.

Claims (1)

[Claims] 1. A valve body whose interior is divided into a main valve chamber and a sub-valve chamber, and which has a pump port and a tank port with a valve seat for separating and seating the main valve on the side of the main valve chamber, and a main valve chamber of the valve body. a main valve provided in a valve chamber and seated and unseated on the valve seat for separating and seating the main valve; an oil passage provided in the main valve and communicating the pump port with the main valve chamber;
Another oil passage is provided in the valve body so as to communicate the main valve chamber and the auxiliary valve chamber, and has a valve seat for separating and seating the auxiliary valve on the side of the auxiliary valve chamber; A sub-valve is provided in the auxiliary valve, which is disposed in the valve seat for separating and seating the sub-valve, and the sub-valve chamber is communicated with the tank port, and when the sub-valve is opened, the first oil passage and the other oil passage are connected to each other. and a relief oil passage for relieving pressure oil from the pump port to a tank port via a predetermined valve seat angle that expands the diameter of the main valve separation/seating valve seat toward the outlet side. The relief oil passage is formed as an oil hole in the valve body, and the passage outlet portion of the relief passage is formed as an inclined valve seat with an inclined shape of the valve seat for separating and seating the main valve. The opening is made slightly outward in the radial direction from the extension line, so that the main valve jet does not collide with the passage outlet of the relief oil passage near the cracking of the main valve, and as the main valve jet increases, the passage exit 1. A pressure control valve characterized by being configured to increase sub-valve back pressure upon collision and maintain the sub-valve back pressure at a predetermined high pressure after a predetermined relief flow rate.