CN105740552B - A kind of double electric low pressure reversing valve designs methods of deep-sea production tree electrichydraulic control valve group - Google Patents
A kind of double electric low pressure reversing valve designs methods of deep-sea production tree electrichydraulic control valve group Download PDFInfo
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Abstract
The present invention provides a kind of double electric low pressure reversing valve designs methods of deep-sea production tree electrichydraulic control valve group, specific design method includes that four steps such as design scheme, geometric dimension calculation and check, valve body operation stress calculation and check and review check are determined according to practical service environment.Reasonable design method of the present invention, computational efficiency and precision are high, and design value and actual use value are more close, the reliability for effectively raising the design work of the double electric low pressure reversal valves of deep-sea subsea production tree electrichydraulic control valve group, to help to improve the stability and reliability of valve body in actual use.
Description
Technical field
The present invention relates to a kind of double electric low pressure reversing valve designs methods, are exactly a kind of deep-sea production tree electrichydraulic control
The double electric low pressure reversing valve designs methods of valve group.
Background technique
At present in the development process of deep-sea oil gas resource, deep-sea subsea production tree equipment application is very extensive, and in depth
There is vital effect in extra large petrol resources exploitation, but finds in actual use, the current institute on the production tree of deep-sea
The double electric low pressure reversal valves used often pass through empirical equation, by the underwater double electric low pressure reversal valves of traditional fresh water or shallow sea water
Lower pair of electric low pressure reversal valve is transformed and designs preparation, although can satisfy the need of deep-marine-environment operation to a certain extent
It wants, but the deep-sea production tree that design preparation all obtains in this way the running technology parameter and reality of double electric low pressure reversal valves
There are biggish error between the use environment of border, so as to cause double electric low pressure reversal valves, operation stability is not seriously under deep-marine-environment
Foot, while the experience that passes through of tradition is during carrying out double electric low pressure reversing valve designs, one side computational accuracy wretched insufficiency, separately
One side computational efficiency is also relatively low, while also can not carry out effective check to the design structure by being calculated and test
Card to also cause greatly to perplex to design work, therefore is directed to this status, there is an urgent need to develop a kind of versatile and
Simple and easy double electric low pressure reversing valve designs methods, to meet the needs of actual use.
Summary of the invention
The object of the present invention is to provide the present invention to provide a kind of double electric low pressure reversal valves of deep-sea production tree electrichydraulic control valve group
Design method.
In order to achieve the above object, the present invention provides the following technical solutions：
A kind of double electric low pressure reversing valve designs methods of deep-sea production tree electrichydraulic control valve group, deep-sea production tree electric hydraulic control valve
The double electric low pressure reversing valve designs methods of group include the following steps：
The first step determines design scheme according to practical service environment, runs actual environment feelings under water according to the deep-sea of valve body
Condition and working media situation primarily determine effective working environment adaptation range of valve body, set every running technology index of valve body
And the basic mechanical structure of valve body；
Second step, geometric dimension calculation and check are right according to the valve body technical parameter of first step setting and basic mechanical structure
The geometric dimension of valve body carries out calculation and check, wherein needing to oil inlet and outlet diameter, cue ball valve base inner hole and rod diameter, main valve
Valve port minimum aperture and main valve plug stroke carry out calculation and check；
Third step, valve body run stress calculation and check, according to the running environment and technical parameter of first step setting, tie simultaneously
The specific size for closing second mechanical structure being calculated, carries out calculating check to each stress condition in valve body operating condition,
It wherein needs to carry out school to frictional resistance, the resistance of motion, hydraulic card clamping force, stable state fluid power, spool active force and return spring elastic force
Assess calculation；
4th step, review is checked, and according to the setup parameter range of the first step, selectes at least one set data, and by selected number
According to be brought into second step and specific data that third step is calculated in, check then in conjunction with valve body practical operation situation multiple
Audit is calculated.
Further, oil outlet diameter, cue ball valve base inner hole and rod diameter, valve port of main valve be most in the second step
It is as follows that small guide vane and main valve plug stroke carry out calculation and check formula：
Oil outlet diameter calculation formula：
Wherein：D-hydraulic fluid port diameter
Q-- metered flow (l/min)；
V-oil inlet and outlet diameter d goes out oil flow, and pressure is bigger, and speed is higher；
Cue ball valve base diameter of bore and rod diameter and steel ball size calculation formula:
d1≥1/2D1
Flow formula by circular passage between valve port and push rod is
Above formula flow Q is brought into metered flow, the oil flow V in circular passage, becauseThen
Wherein：D1-rod diameter
D1-rod diameter cue ball valve base diameter of bore
Oil flow in v-circular passage
Valve port of main valve minimum aperture calculation formula：
It is according to by the flow rate calculation formula of valve port：
In formula：
Q --- pass through valve port fluid flow (m3/s)；
A --- valve port area of passage (m2)；
Δ p --- valve port both ends pressure difference (Pa)；
Ρ --- fluid density (kg/m3)；
Cd --- it is valve port flow coefficient；
Valve port area of passage A calculation formula is：
Wherein,
WhenWhen,
Above formula can be changed to
So the opening minimum aperture formula of valve can be dissolved is：
Main valve plug climb displacement formula：
X1 is had to be larger than by the stroke S of spool to obtain：
S>X1
Wherein：S-spool stroke；
X1-valve port of main valve minimum aperture；
Further, frictional resistance, the resistance of motion, hydraulic card clamping force, stable state fluid power, spool effect in the third step
Power and the calculation formula of return spring elastic force are as follows：
Frictional resistance calculation formula：
Due to bt=0.55d；
D0 is the end face diameter of O-ring seals.It is assumed that O-ring and control piston are only in contact shifting under the action of Pb
It is dynamic, and contact width is constant.
O-ring to control piston clamping force be
Then
Fm=fN=0.275 π fPbdtd0
Wherein：Fm--- frictional resistance；
F --- coefficient of friction can use f=0.1；
Dt --- diameter of piston rod；
D0 --- O-ring end face diameter；
Pb --- allow back pressure；
Resistance of motion calculation formula：
Wherein：FV--- the resistance of motion；
D --- control piston diameter；
L --- the contact length of control piston and valve body hole；
V --- valve core movement speed；
μ --- oil liquid dynamic viscosity；
Between Δ r --- spool and the unilateral cooperation of valve body hole；
When hydraulic card clamping force calculates, the generation of hydraulic card clamping force is because of fluid matching between hydraulic valve core and valve body
When closing the flowing in gap, due to spool and valve body hole is tapered and eccentricity, makes the presence of pressure at the different gap of circumferencial direction
Power changes in distribution, and a radial imbalance force is produced to spool, simultaneously because using O shape between the spool and valve body of the design
Sealing ring sealing, leakage rate is almost nil, and the design ignores hydraulic card clamping force；
Stable state fluid power calculation formula：
Fw=CdπDδΔpsinα (2-17)
Simultaneously as the structure of spool, in spool commutation two mouthfuls all in open state, the liquid stream on one side is at lower manifold
Formula, the liquid stream on one side is at upper form, but the steady-state fluid force that spool is subject to is all in one direction, all with liquid flow path direction phase
Instead, so the steady-state fluid force of two ball valves need to be calculated in calculating formula.And spool is either in left position or right position, situation
All be it is identical, calculate when, only with calculate spool one side steady-state fluid force, spool steady-state fluid force calculating need to calculate
Steady-state fluid force in the case of two kinds, one is small openings, i.e. the δ max of δ=1/3, and one is when spool standard-sized sheet；
Thus it obtains, steady-state fluid force summation is：
ΣFw=Fw1+Fw2=CdπDδΔpsinα1+CdπD(S-δ)Δpsinα2
In formula：
Cd --- it is valve port flow coefficient；
D --- valve port diameter；
The amount of opening of δ --- valve；
Δ p --- valve port both ends pressure difference (Pa)；
α --- fluid flow angle
Spool Calculation of the force formula：
Spool Calculation of the force formula：
The maximum working load of spring can be calculated as the following formula
Ft2> Σ Fw+Fp1=8N
ΣFw--- the stable state steady-state fluid force on spool；
Fp1--- the hydraulic coupling on spool
Take Ft2=15N
So the rigidity of spring：
In formula：
ΣFm--- frictional resistance summation；
Fp--- the pressure that spool generates；
Ft1--- spring minimum workload.
ΣFw--- steady-state fluid force when spool circle is opened；
Ft2--- little spring maximum working load；
The rigidity of K1 --- spring；
Thus：
Open control piston
Control piston spool to be opened is opened, the following conditions need to be met
FOpen 1> Σ Fm+Fp+Ft1
FOpen 1--- open state spool active force；
FOpen 2--- closed state spool active force；
ΣFm--- frictional resistance summation；
Fp--- the pressure that spool generates；
Ft1--- little spring minimum workload；
When spool in place after, the following conditions need to be met:
FOpen 2> Σ Fw+Fp+Ft2-FFrom；
Close control piston
When the valves are closed, closure piston pushes main valve plug open, and needing the power for overcoming self-locking piston and frictional force at this time, there are also stable states
Hydraulic power.P, R, C are communicated when valve is closed, and valve is in unloading condition, and the liquid in self-locking piston loses pressure, close control piston
Only needing the power of very little can close, so only closing power of the control piston when valve is begun to shut off with calculating.
The following conditions need to be met by closing control piston
FIt closes> Σ Fm+ΣFw+FFrom-K(Xt+S)-Fp
KXt> FFrom 1
When spool in place after, the following conditions need to be met, quickly due to spool movement speed, general < 0.01s, thus in order to
Spool it is reliable, it is assumed that control piston in liquid have enough time flowing out not yet, so need herein calculate control piston generate work
Firmly；
KXt> FFrom 1
When spool is to small aperture position, the following conditions need to be met, need to equally calculate the active force that control piston generates；
K(Xt+S-1/3δmax) > FFrom 1+ΣFW1 is small+ΣFm1
ΣFW1 is small--- steady-state fluid force of the spool in the small opening of 69bar；
Reasonable design method of the present invention, computational efficiency and precision are high, and design value and actual use value are more closely, effectively
The design work for improving the double electric low pressure reversal valves of deep-sea subsea production tree electrichydraulic control valve group reliability, to facilitate
Improve the stability and reliability of valve body in actual use.
Detailed description of the invention
In order to more clearly explain the embodiment of the invention or the technical proposal in the existing technology, to embodiment or will show below
There is attached drawing needed in technical description to be briefly described, it should be apparent that, the accompanying drawings in the following description is only this
Some embodiments of invention for those of ordinary skill in the art without creative efforts, can be with
It obtains other drawings based on these drawings.
Fig. 1 is the method for the present invention flow chart.
Specific embodiment
Technical solution of the present invention is clearly and completely described below in conjunction with attached drawing of the invention, it is clear that retouched
The embodiment stated is only a part of the embodiment of the present invention, instead of all the embodiments.Based on the embodiments of the present invention, originally
Field those of ordinary skill every other embodiment obtained without creative efforts, belongs to the present invention
The range of protection.
Embodiment 1：
A kind of double electric low pressure reversing valve designs methods of deep-sea production tree electrichydraulic control valve group as shown in Figure 1, deep-sea is recovered the oil
The double electric low pressure reversing valve designs methods of tree electrichydraulic control valve group include the following steps：
The first step determines design scheme according to practical service environment, runs actual environment feelings under water according to the deep-sea of valve body
Condition and working media situation primarily determine effective working environment adaptation range of valve body, set every running technology index of valve body
And the basic mechanical structure of valve body：
Working environment：
Above-mentioned hydraulic valve is all mounted in the sealing container of the 3000 meters of depth of water in ocean
1.2 temperature requirement：
1) storage temperature range：-18℃—+50℃；
2) operating temperature range：-5℃—+40℃.
3) working media：Water base fluid, such as HW443.
Control principle：When oil circuit is normal, valve is electric within the scope of PP=15.5--37.9Mp at charge oil pressure P=56.9Mpa
Magnet valve can be by the opening of control valve, locking and closing, and providing operating pressure for actuator is 10.3-56.9Mpa hydraulic
Oil.
Note：The state being wherein often used is：P=56.9Mpa, PP=34.5Mpa provide operating pressure for actuator
10.3-56.9Mpa hydraulic oil.
When in-line decompression, i.e. when P≤6.9Mpa, valve is automatically closed, and is performed simultaneously device reset.
1) when circuit breaks down, inlet pressure is let out by platform, carries out closing actuator, resets double electric low pressure valves.When
When inlet pressure P≤6.9Mpa, valve automatically resets, and is performed simultaneously device closing.
2) valve is in charge oil pressure P=56.9Mpa, and when controlling oil circuit PP=6.9Mpa, solenoid valve cannot open and close valve,
The state of actuator is unrelated with control oil circuit, only related with oil-feed port pressure.Actuator is maintained at some position.
3) valve inner spring starts to reset in 3.45Mpa-6.9Mpa, resets pressure and is controlled by HPU waterborne.(spring
Power when initial pressure is 3.45Mpa, power when being 6.9Mpa after compression)
When test pressure：It can be opened under P=85.3Mpa, PP=56.9Mpa T=31Mpa, self-locking and reset valve.
Second step, geometric dimension calculation and check are right according to the valve body technical parameter of first step setting and basic mechanical structure
The geometric dimension of valve body carries out calculation and check, wherein needing to oil inlet and outlet diameter, cue ball valve base inner hole and rod diameter, main valve
Valve port minimum aperture and main valve plug stroke carry out calculation and check：
Oil outlet diameter calculation formula：
Wherein：D-hydraulic fluid port diameter
Q-- metered flow (l/min)；
V-oil inlet and outlet diameter d goes out oil flow, and pressure is bigger, and speed is higher, selects 10m/s herein
So：
Rounding takes d=6mm；
Cue ball valve base diameter of bore and rod diameter and steel ball size calculation formula:
d1≥1/2D1
Flow formula by circular passage between valve port and push rod is
Above formula flow Q is brought into metered flow, the oil flow V in circular passage, becauseThen
D1=6mm, d1=4.5mm are taken after rounding,
Steel ball size 8mm；
Wherein：D1-rod diameter
D1-rod diameter cue ball valve base diameter of bore
Oil flow in v-circular passage
Valve port of main valve minimum aperture calculation formula：
It is according to by the flow rate calculation formula of valve port：
In formula：
Q --- pass through valve port fluid flow (m3/s)；
A --- valve port area of passage (m2)；
Δ p --- valve port both ends pressure difference (Pa)；
Ρ --- fluid density (kg/m3)；
Cd --- it is valve port flow coefficient；
Valve port area of passage A calculation formula is：
Wherein,
WhenWhen,
Above formula can be changed to
So the opening minimum aperture formula of valve can be dissolved is：
Δ P=1MPa, Cd=1 are taken, so
X1=0.38mm；
Main valve plug climb displacement formula：
X1 is had to be larger than by the stroke S of spool to obtain：S>X1 takes
S=1mm；
Wherein：S-spool stroke；
X1-valve port of main valve minimum aperture；
Third step, valve body run stress calculation and check, according to the running environment and technical parameter of first step setting, tie simultaneously
The specific size for closing second mechanical structure being calculated, carries out calculating check to each stress condition in valve body operating condition,
It wherein needs to carry out school to frictional resistance, the resistance of motion, hydraulic card clamping force, stable state fluid power, spool active force and return spring elastic force
Assess calculation：
Frictional resistance calculation formula：
Due to bt=0.55d；
D0 is the end face diameter of O-ring seals.It is assumed that O-ring and control piston are only in contact shifting under the action of Pb
It is dynamic, and contact width is constant.
O-ring to control piston clamping force be
Then
Fm=fN=0.275 π fPbdtd0
Wherein：Fm--- frictional resistance；
F --- coefficient of friction can use f=0.1；
Dt --- diameter of piston rod；
D0 --- O-ring end face diameter；
Pb --- allow back pressure；
So
Fm=0.0864Pbdtd0
If there is 5 O-ring seals on main valve plug, for safety, according to the maximal friction of each O-ring seals when calculating
I.e. maximum back pressure when calculate, so respective frictional force is：
From lock control piston
Control piston maximum back pressure is Pb1=569bar, O-ring diameter of section d01=1.8mm, piston diameter dt1
Fm1=0.0864Pb1dt1d01=4916dt1
Open control piston
Control piston maximum back pressure is Pb2=569bar, O-ring diameter of section d02=1.8mm, piston diameter dt2
Fm2=0.0864Pb2dt2d02=4916dt2
Close control piston push rod
Push rod maximum back pressure is Pb3=569bar, O-ring diameter of section d03=1.8mm, piston diameter dt3
Fm3=0.0864Pb3dt3d03=4916dt3
Take dt3=6mm
Fm3=4916 × 0.006=29.5N
Close control piston
Control piston maximum back pressure is Pb4=379bar, O-ring diameter of section d02=1mm, piston diameter dt4
Fm2=0.0864Pb2dt2d02=4916dt2
Resistance of motion calculation formula：
Wherein：FV--- the resistance of motion；
D --- control piston diameter；
L --- the contact length of control piston and valve body hole；
V --- valve core movement speed, can with spool actuation time by 0.01s when average speed replace；So spool
μ --- oil liquid dynamic viscosity, the kinematic viscosity of HW443 are 1.9mm2/s；
The unilateral fit clearance of Δ r --- spool and valve body hole, takes 0.05mm here.
So
By formula (2-17) it is found that since hydraulic oil viscosity is low, and the fit clearance of valve is relatively large, so the resistance of motion is non-
It is often small, so being ignored in calculating formula.
When hydraulic card clamping force calculates, the generation of hydraulic card clamping force is because of fluid matching between hydraulic valve core and valve body
When closing the flowing in gap, due to spool and valve body hole is tapered and eccentricity, makes the presence of pressure at the different gap of circumferencial direction
Power changes in distribution, and a radial imbalance force is produced to spool, simultaneously because using O shape between the spool and valve body of the design
Sealing ring sealing, leakage rate is almost nil, and the design ignores hydraulic card clamping force；
Stable state fluid power calculation formula：
Fw=CdπDδΔpsinα (2-17)
Simultaneously as the structure of spool, in spool commutation two mouthfuls all in open state, the liquid stream on one side is at lower manifold
Formula, the liquid stream on one side is at upper form, but the steady-state fluid force that spool is subject to is all in one direction, all with liquid flow path direction phase
Instead, so the steady-state fluid force of two ball valves need to be calculated in calculating formula.And spool is either in left position or right position, situation
All be it is identical, calculate when, only with calculate spool one side steady-state fluid force, spool steady-state fluid force calculating need to calculate
Steady-state fluid force in the case of two kinds, one is small openings, i.e. the δ max of δ=1/3, and one is when spool standard-sized sheet；
Thus it obtains, steady-state fluid force summation is：
ΣFw=Fw1+Fw2=CdπDδΔpsinα1+CdπD(S-δ)Δpsinα2
In formula：
Cd --- it is valve port flow coefficient, takes Cd=1；
D --- valve port diameter；
The amount of opening of δ --- valve；
Δ p --- valve port both ends pressure difference (Pa), when small opening, take 1MPa, and when big opening takes 0.1MPa；
α --- fluid flow angle, when big opening, take α=69 °, and when small opening takes α=21 °；
Due to the structure of spool, in spool commutation two mouthfuls all in open state, the liquid stream on one side is at dirty form, one
The liquid stream on side is at upper form, but the steady-state fluid force that spool is subject to is all in one direction, all with liquid flow path direction on the contrary, so
The steady-state fluid force of two ball valves need to be calculated in calculating formula.And spool is either in left position or right position, and situation is all phase
With, when calculating, only uses and calculate spool in the steady-state fluid force on one side.
The steady-state fluid force of spool calculates the steady-state fluid force that need to be calculated in the case of two kinds, and one is small openings, i.e. δ=1/
3 δ max, one is when spool standard-sized sheet.
When spool is in small opening, δ=0.127mm, steady-state fluid force summation at this time is：
ΣFw=Fw1+Fw2=CdπDδΔpsinα1+CdπD(S-δ)Δpsinα2
ΣFw=4.8N
When valve is in standard-sized sheet, δ=S=2mm, steady-state fluid force summation at this time is
ΣFw=CdπDδΔpsinα1=4.1N
Spool Calculation of the force formula：
The maximum working load of spring can be calculated as the following formula
Ft2> Σ Fw+Fp1=8N
ΣFw--- the stable state steady-state fluid force on spool；
Fp1--- the hydraulic coupling on spool
Take Ft2=15N
So the rigidity of spring：
In formula：
ΣFm--- frictional resistance summation；
Fp--- the pressure that spool generates；
Ft1--- spring minimum workload.
ΣFw--- steady-state fluid force when spool circle is opened；
Ft2--- little spring maximum working load；
The rigidity of K1 --- spring；
Thus：
Open control piston
Control piston spool to be opened is opened, the following conditions need to be met
FOpen 1> Σ Fm+Fp+Ft1
FOpen 1--- open state spool active force；
FOpen 2--- closed state spool active force；
ΣFm--- frictional resistance summation；
Fp--- the pressure that spool generates；
Ft1--- little spring minimum workload；
When spool in place after, the following conditions need to be met:
FOpen 2> Σ Fw+Fp+Ft2-FFrom；
Close control piston
When the valves are closed, closure piston pushes main valve plug open, and needing the power for overcoming self-locking piston and frictional force at this time, there are also stable states
Hydraulic power.P, R, C are communicated when valve is closed, and valve is in unloading condition, and the liquid in self-locking piston loses pressure, close control piston
Only needing the power of very little can close, so only closing power of the control piston when valve is begun to shut off with calculating.
The following conditions need to be met by closing control piston
FIt closes> Σ Fm+ΣFw+FFrom-K(Xt+S)-Fp
Return spring elastic force calculation formula：
K1(XT+ S)=Σ Fw1+ΣFm1+Fp1
In formula：
ΣFw1--- steady-state fluid force of the spool in 569bar open；
ΣFm1--- spool the sum of frictional resistance in 569bar；
Fp1--- the fluid pressure that ball valve core is subject in 569bar；
When spool in place after, the following conditions need to be met, quickly due to spool movement speed, general < 0.01s, thus in order to
Spool it is reliable, it is assumed that control piston in liquid have enough time flowing out not yet, so need herein calculate control piston generate work
Firmly.
KXt> FFrom 1
When spool is to small aperture position, the following conditions need to be met, need to equally calculate the active force that control piston generates.
K(Xt+S-1/3δmax) > FFrom 1+ΣFW1 is small+ΣFm1
ΣFW1 is small--- steady-state fluid force of the spool in the small opening of 69bar.
4th step, review is checked, and according to the setup parameter range of the first step, selectes at least one set data, and by selected number
According to be brought into second step and specific data that third step is calculated in, check then in conjunction with valve body practical operation situation multiple
Audit is calculated.
Reasonable design method of the present invention, computational efficiency and precision are high, and design value and actual use value are more closely, effectively
The design work for improving the double electric low pressure reversal valves of deep-sea subsea production tree electrichydraulic control valve group reliability, to facilitate
Improve the stability and reliability of valve body in actual use.
The above description is merely a specific embodiment, but scope of protection of the present invention is not limited thereto, any
Those familiar with the art in the technical scope disclosed by the present invention, can easily think of the change or the replacement, and should all contain
Lid is within protection scope of the present invention.Therefore, the protection scope of the present invention shall be subject to the protection scope of the claims.
Claims (1)
1. a kind of double electric low pressure reversing valve designs methods of deep-sea production tree electrichydraulic control valve group, it is characterised in that：The deep-sea
The double electric low pressure reversing valve designs methods of production tree electrichydraulic control valve group include the following steps：
The first step determines design scheme according to practical service environment, runs actual environment situation under water according to the deep-sea of valve body, and
Working media situation, the every running technology index and valve for primarily determining effective working environment adaptation range of valve body, setting valve body
The basic mechanical structure of body；
Second step, geometric dimension calculation and check, according to valve body technical indicator and basic mechanical structure that the first step is set, to valve body
Geometric dimension carry out calculation and check, wherein needing to oil inlet and outlet diameter, cue ball valve base inner hole and rod diameter, valve port of main valve
Minimum aperture and main valve plug stroke carry out calculation and check；
Third step, valve body run stress calculation and check, according to running environment and technical indicator that the first step is set, in combination with the
The specific size for the mechanical structure that two steps are calculated carries out calculating check to each stress condition in valve body operating condition, wherein
Check meter need to be carried out to frictional resistance, the resistance of motion, hydraulic card clamping force, stable state fluid power, spool active force and return spring elastic force
It calculates；
4th step, review is checked, and according to the setting technique indication range of the first step, selectes at least one set data, and by selected number
According to be brought into second step and specific data that third step is calculated in, check then in conjunction with valve body practical operation situation multiple
Audit is calculated；
Oil outlet diameter, cue ball valve base inner hole and rod diameter, valve port of main valve minimum aperture and main valve in the second step
It is as follows that core stroke carries out calculation and check formula：
Oil outlet diameter calculation formula： QUOTE
Wherein：D-hydraulic fluid port diameter
Q-- metered flow l/min；
- oil inlet and outlet diameter d goes out oil flow, and pressure is bigger, and speed is higher；
Cue ball valve base diameter of bore and rod diameter and steel ball size calculation formula:
d1≥1/2 D1
Flow formula by circular passage between valve port and push rod is
Above formula flow Q is brought into metered flow, the oil flow V in circular passage, because of QUOTE , then
Wherein： QUOTE - rod diameter
QUOTE - rod diameter cue ball valve base diameter of bore
QUOTE Oil flow in-circular passage
Valve port of main valve minimum aperture calculation formula：
It is according to by the flow rate calculation formula of valve port：
In formula：
QUOTE --- pass through valve port fluid flow m3/s；
QUOTE --- valve port area of passage m2；
QUOTE --- valve port both ends pressure difference Pa；
QUOTE --- fluid density kg/m3;
QUOTE --- it is valve port flow coefficient；
Valve port area of passage A calculation formula is：
QUOTE ；
Wherein, QUOTE ；D --- valve port diameter；
Work as QUOTE When, QUOTE ,
Above formula becomes
QUOTE ；
So the opening minimum aperture formula neutralizing of valve is：
QUOTE ；
Main valve plug climb displacement formula：
QUOTE is had to be larger than by the stroke S of spool ?：
S> QUOTE
Wherein：S-spool stroke；
QUOTE - valve port of main valve minimum aperture；
Frictional resistance, the resistance of motion, hydraulic card clamping force, stable state fluid power, spool active force and return spring in the third step
The calculation formula of elastic force is as follows：
Frictional resistance calculation formula：
Due to QUOTE =0.55 QUOTE ；
QUOTE For the end face diameter of O-ring seals, it is assumed that in QUOTE Under the action of, O-ring and control are lived
Plug is only in contact movement, and contact width is constant；
O-ring to control piston clamping force be
QUOTE ；
Then
Wherein： QUOTE --- frictional resistance；
QUOTE --- coefficient of friction takes QUOTE =0.1；
QUOTE --- diameter of piston rod；
QUOTE --- O-ring end face diameter；
QUOTE --- allow back pressure；
Resistance of motion calculation formula：
Wherein： QUOTE --- the resistance of motion；
D --- control piston diameter；
L --- the contact length of control piston and valve body hole；
V --- valve core movement speed；
μ --- oil liquid dynamic viscosity；
QUOTE --- the unilateral cooperation spacing of spool and valve body hole；
When hydraulic card clamping force calculates, the generation of hydraulic card clamping force is because between cooperation of the fluid between hydraulic valve core and valve body
When flowing in gap, due to spool and valve body hole is tapered and eccentricity, makes at the different gap of circumferencial direction that there are pressure point
Cloth variation, and a radial imbalance force is produced to spool, simultaneously because being sealed between the spool and valve body of the design using O shape
Circle sealing, leakage rate is almost nil, and the design ignores hydraulic card clamping force；
Stable state fluid power calculation formula：
QUOTE （2-17）
Simultaneously as the structure of spool, in spool commutation two mouthfuls all in open state, the liquid stream on one side at dirty form,
The liquid stream on one side is at upper form, but the steady-state fluid force that spool is subject to is all in one direction, all with liquid flow path direction on the contrary, institute
The steady-state fluid force of two ball valves need to be calculated in calculating formula；And spool is either in left position or right position, situation are all
Identical, when calculating, only with spool is calculated in the steady-state fluid force on one side, the steady-state fluid force calculating of spool need to calculate two kinds
In the case of steady-state fluid force, one is small opening, the i.e. δ of δ=1/3max, one is when spool standard-sized sheet；
Thus it obtains, steady-state fluid force summation is：
In formula：
QUOTE --- it is valve port flow coefficient；
QUOTE --- valve port diameter；
The amount of opening of δ --- valve；
QUOTE --- valve port both ends pressure difference Pa；
α --- fluid flow angle
Spool Calculation of the force formula：
At work due to valve body, spool has two kinds of working conditions of unlatching and closure, it is therefore desirable to which spool is being opened and closed
The favorite active force of conjunction state calculates separately,
Therefore open state active force： QUOTE
Closed state active force： QUOTE
So the rigidity of spring：
In formula： QUOTE --- open state spool active force；
QUOTE --- closed state spool active force；
QUOTE --- frictional resistance summation；
QUOTE --- the pressure that spool generates；
QUOTE --- spring minimum workload；
QUOTE --- steady-state fluid force when spool circle is opened；
QUOTE --- little spring maximum working load；
QUOTE --- the rigidity of spring；
Thus：
Open control piston
Control piston spool to be opened is opened, the following conditions need to be met
QUOTE --- frictional resistance summation；
QUOTE --- the pressure that spool generates；
QUOTE --- little spring minimum workload；
When spool in place after, the following conditions need to be met:
QUOTE --- steady-state fluid force when spool circle is opened；
QUOTE --- little spring maximum working load；
QUOTE --- locking piston effect；
For the reliable and stable movement of valve energy, QUOTE is taken When maximum, i.e., system pressure maximum when calculate
From lock control piston
It is that spool is withstood, spool is forced to be constantly in open state after spool opening from lock control piston；
So self-locking piston need to meet the following conditions
QUOTE ；
Close control piston
When the valves are closed, closure piston pushes main valve plug open, and needing the power for overcoming self-locking piston and frictional force at this time, there are also stable states to surge
Power；P, R, C are communicated when valve is closed, and valve is in unloading condition, and the liquid in self-locking piston loses pressure, close control piston, institute
Only to close power of the control piston when valve is begun to shut off with calculating；
The following conditions need to be met by closing control piston
Return spring elastic force calculation formula：
In formula：
QUOTE --- steady-state fluid force of the spool in 569bar open；
QUOTE --- spool the sum of frictional resistance in 569bar；
QUOTE --- the fluid pressure that ball valve core is subject in 569bar;
When spool in place after, the following conditions need to be met, quickly due to spool movement speed, be less than 0.01s, so for spool
Reliably, it is assumed that liquid has enough time flowing out not yet in control piston, so need to calculate the active force that control piston generates herein；
When spool is to small aperture position, the following conditions need to be met, need to equally calculate the active force that control piston generates；
QUOTE --- steady-state fluid force of the spool in the small opening of 69bar.
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| CN105179708A (en) * | 2015-10-14 | 2015-12-23 | 四川华林自控科技有限公司 | Energy-saving balance high-pressure valve and design method |
| CN105201947A (en) * | 2015-10-19 | 2015-12-30 | 中国石油大学(北京) | Electrohydraulic reversing valve capable of maintaining valve position under hydraulic pressure |
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| CN103192815A (en) * | 2013-04-28 | 2013-07-10 | 清华大学 | Method for controlling hydraulic braking force |
| CN105179708A (en) * | 2015-10-14 | 2015-12-23 | 四川华林自控科技有限公司 | Energy-saving balance high-pressure valve and design method |
| CN105201947A (en) * | 2015-10-19 | 2015-12-30 | 中国石油大学(北京) | Electrohydraulic reversing valve capable of maintaining valve position under hydraulic pressure |
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