GB2389639A

GB2389639A - Sliding and check valves

Info

Publication number: GB2389639A
Application number: GB0213668A
Authority: GB
Inventors: Tony Markham
Original assignee: Water Hydraulics Co Ltd; WATER HYDRAULICS Co Ltd
Current assignee: Water Hydraulics Co Ltd; WATER HYDRAULICS Co Ltd
Priority date: 2002-06-14
Filing date: 2002-06-14
Publication date: 2003-12-17
Anticipated expiration: 2022-06-14
Also published as: GB2389639B; GB0213668D0

Abstract

A valve comprises casing 2 with inner surface 3 enclosing volume 4, a first 5 and second 7 fluid port. Sealing member 9 comprising piston 17 and cylinder 18 is inside the casing having surface 10 sealing against the casing inner surface and having exposure surface 12 within cavity 20 arranged for exposure to fluid at inlet pressure to urge the sealing surface against the inner surface sealing portion. The sealing member being movable between a first position separating the first and second ports, and a second position in which the first and second ports communicate. Fluid with in cavity 20 urges surface 21 of piston 17 against top plate 151 and fluid pocket 22 supplied from cavity 20 provides hydrostatic balance, minimising friction when moving the sealing member. Check valve (50 Fig. 6) comprises piston (53) within a spring loaded (63) sleeve (58) with radial holes (60) and an O-ring seal (54) at the root of the sleeve bore.

Description

- 1 FLUID VALVES

Field of the Invention

This invention relates to valves, and in particular, although not exclusively, to valves for use in water hydraulic systems.

Background to the Invention

The field of water hydraulics is an emerging technology. There are now a num-

ber of world players in this market producing hydraulic equipment that can operate upon tap water as the hydraulic fluid. Concentration towards rotary power i.e. axial piston pumps and axial piston motors by two global companies was the major driving force behind this emergence. However the control of such devices in terms of pressure and flow, to date are limited and restricted to adopting current oil industry standards and modifying the materials to suit. The result is a more expensive less efficient over complicated system.

The higher cost associated with such components restrict the rate of growth of the water hydraulic market. The inflated system cost detracts from the advantage of the environmentally friendly alternative and often results in adoption of the conven-

tional oil system. Only where the risk of contamination of fire hazards are primary con-

siderations will the more expensive system be considered confining water technology to niche applications.

In the early 80's a move to high water based fluids (water with a lubrication ad-

ditive applying a corrosive barrier such as 95% water and 5% Glycol-95/5) were ex-

amined in high fire-risk areas such as Steel, Mining and Offshore markets. Traditional material specifications of oil hydraulic components were modified to accept function

upon the low viscosity fluids but a far more expensive component resulted. Life expec-

tancy of the components varied greatly, manufacturers warranty was seldom available.

Some components designed for the conventional oil (i.e. high viscosity hydraulic industry are appropriate for use on pure water. However, their high price inhibits their entry into the newer emerging field of pure water hydraulics.

A further major difference between 95/5 and pure water is the system operat-

ing pressure. Pure systems are limited to 140/160 bar, 95/S systems can operate up to at 350. Components designed for the 9S/S market are therefore far more expen-

sive due to the design specification.

Oil hydraulic system designers can select numerous valve functions offered by many worldwide manufacturers without compromising safety or increasing system complexity. To date a comparable range of water hydraulic functions are not available.

As a result solutions are usually over complicated which ultimately impact upon the reli-

ability of the system.

It is possible to overcome some of the inherent problems of controlling speed, direction and load of a hydraulic prime mover e.g. cylinder or motor by current stan-

dards. However, offering a valve specifically designed for the fluid is the only method of satisfying systems criteria at a reduced cost. It is an object of embodiments of the pres-

ent invention to provide valves which offer an extensive range of control options for low viscosity and corrosive fluids, including control of pressure and direction of flow.

Previous methods of controlling flow and pressure within a water hydraulic sys-

tem have been based upon two well-established valve designs: the spool valve; and the poppet valve.

In spool valves, the low lubrication properties of water offer little support for load bearing dynamic surfaces. Material combinations of bore and spool have to over-

come the following criteria but still maintaining a respectable efficiency: (a) Fluid abrasion Very high flow velocities upon opening and closing cause rapid deterioration of the ports and spool land. A water jet is far more aggressive than the oil equivalent.

(b) Corrosion Both short and long term corrosion will modify the valves performance and prevent operation. The possible material combinations are drastically reduced due to

( - 3 the corrosion resistance requirement. as well as the galvanic erosion considerations from non-compatible materials.

(c) Efficiency Low viscosity fluids command very close fitting manufacturing tolerances to ok tain acceptable volumetric compliance (i.e. low leakage). Most candidate materials such as ceramics are very difficult to produce to such tight tolerances, and the differential thermal expansion of possible material combinations further restrict appropriate com-

binations. (d) Friction The materials static and dynamic friction dictate operational forces for a given pressure. Flow control requires a very responsive fine incremental motion. Pressure balancing can decrease the required operating forces but such complication on a round component is expensive and will reduce efficiency.

The poppet valve offers potential zero leakage but once again tight manufactur-

ing tolerances are required to achieve this. Direct actuation of a poppet against system pressure i.e. withdrawing the poppet from the seat against system pressure is possible on valves with a maximum operating pressure of 25 bar. Above this level designs be-

come far more complicated relying on mechanical advantage or pilot operation to ok tain the motion. The relatively low operational forces available from a solenoid how-

ever, results in a very small diameter feed holes, usually in the region of less than 1 mm diameter. The small diameters impact on the sealing potential of the valve as well as making the valve more susceptible to failure from fluid contaminants.

The poppet design from a single valve can only offer 2 port. 2 position (V2) or 3 port, 2 position (3/2) control function. Control of a double acting cylinder or rota-

tional direction of a motor would normally incorporate 4 off V2 valves and a compli-

cated manifold or pipe installation.

Both designs of valve have many manufacturers world wide for oil application, and an increasing number are looking to the water market as potential growth area.

But still the cost penalty for the use of a water version when compared to the oil by

( - 4 draulic equivalent can easily be in the order of 8 to 10 times and in some instances greater. US2,653,003 discloses a shear seal valve design that uses flat plates and pres-

sure loaded face poppets to seal. The design is rotary and uses a roller bearing to react the poppet load. The lack of lubrication properties from the operating fluid results in high point or line contact stress imposed on the ball or roller races. Premature failure of the bearing and nonfunction of the valve will result when operating on low viscosity fluids. Sealing will lead to complication of the design, high actuation forces and exces-

sive maintenance. The co-efficient of friction within this design if used with low viscosity fluids would also lead to exaggerated actuation loads.

Face valves (usually rotary) are also available for low pressure water systems such as domestic mixer taps, but the maximum working pressure for such valves is typi-

cally1 2 bar.

It is an object of embodiments of the present invention to provide valves which overcome, at least partially, the disadvantages associated with the previous designs.

It is an object of certain embodiments to provide low cost valves for water hy-

draulic applications.

It is a further object of embodiments to provide valves for use with low viscosity fluids with improved erosion and/or corrosion resistance.

It is a yet a further object of embodiments to provide valves for switching high pressure fluids, with low actuation forces.

Summary of Invention

A first aspect of the invention provides a valve as defined by Claim 1. The seal-

ing force (i.e. the force with which the sealing surface is urged against the sealing por-

tion of the inner surface to form a seal) is determined by the area of the exposure sur-

face and the pressure of the fluid. These can be selected to give a desired actuation force for moving the sealing member between the first and second positions.

- 5 Preferably, the exposure surface is exposed to the fluid pressure in the first port. Thus, the sealing force is provided by pressurised fluid already supplied to the valve, i.e. it is provided by the fluid which the valve is arranged to control. Further-

more, this arrangement provides the advantage that the sealing force is proportional to the fluid pressure in the first port.

The valve may include one or more additional ports Into the casing, and the sealing member may be moveable between a plurality of positions in which different combinations of ports are isolated, blocked or connected.

In a preferred embodiment, when the sealing member is in the first position it forms a seal around just the mouth of the first port, i.e. isolates the first port from all others. Preferably, in the second position the sealing surface of the sealing member forms a seal around the mouths of both the first and second ports.

Preferably, the sealing surface and sealing portion are flat. This provides the ad-

vantage that the mating surfaces are easy to machine, even with materials such as can ramics which offer high cavitation erosion resistance performance and high resistance to corrosion. Advantageously, the sealing surface may be annular, and, where the sealing sur-

face and sealing portion are flat, this results in a sliding circular seal which can be posi-

tioned over one or more mouths of ports in a flat side wall of the valve casing.

The casing may comprise a distribution plate having a flat inner surface, and one or more of the ports may have a mouth opening at the flat inner surface. Advanta-

geously, the sealing member may be arranged to seal against the flat inner surface of the distribution plate.

Advantageously, the sealing member may comprise a sealing bearing, and the sealing surface may be a surface of the bearing. The bearing may be formed of a differ-

ent material from the remaining portion of the sealing member.

Preferably, the seal formed between the seating surface and sealing portion is a sliding seal. and the sealing member is slideable between the first and second positions.

A seal may be maintained during movement from the first to the second position.

( - 6 Advantageously, the valve may comprise an assembly of a piston means and cylinder means arranged inside the casing. The sealing member in such arrangements may be the piston means or the cylinder means.

Preferably the assembly includes a chamber, between the piston and cylinder, this chamber being arranged to be pressurised by a fluid when the valve is in use, such that the piston and cylinder are urged apart and the sealing surface of one of them is urged to seal against the inner surface of the casing.

In addition to the sealing surface on either the piston or cylinder, the assembly preferably includes a further surface on the other component of the assembly, this fur-

ther surface being urged away from the sealing surface by the presence of fluid at pres-

sure in the chamber. In this way, as the sealing surface is urged to seal against the seal-

ing portion of the inner surface, the further surface is urged against another portion of the inner surface of the casing, this other portion being generally opposite the sealing portion. Preferably, the sealing portion and further portion are opposing flat parallel por-

tions of the inner surface. The sealing portion may, for example, be on the inner sur-

face of a flat distribution plate, and the further portion may be on the inner surface of an opposing top plate.

Preferably, the further surface is arranged to form a sliding seal against the fur-

ther portion, and the further surface and further portion may be flat. Advantageously, the seal formed between the further surface and further portion may enclose (i.e. it may bound) a fluid pocket between the assembly and the inner surface. Such a seal may, for example, be annular.

Advantageously, the area of this pocket and/or the pressure of fluid contained within it may be set so as to balance to some degree the affect of fluid at pressure in the cavity (i.e. chamber) urging the further surface against the further portion.

Advantageously, the assembly of piston and cylinder may include a fluid passage connecting the pocket to the chamber, such that the same fluid pressure is present in the pocket and chamber. The presence of the pressurised fluid in the pocket counter

( - 7 acts the affect of the pressurised fluid in the chamber urging the further surface against the further portion, and so gives a desired sealing force.

The fluid passage may be a bore in the piston or cylinder means.

In one preferred arrangement, the sealing surface is a surface of the cylinder and the further surface is a surface of the piston.

Preferably, the chamber enclosed by the piston and cylinder means is pressur-

ised by the fluid from the first port. In this way, the fluid from the first port urges the piston and cylinder assembly to expand inside the casing, forming seals against opposing sides. Advantageously, the cylinder means may be the sealing member, and the sealing surface may form a seal around the mouth of the first port in the first and second posi-

tions, the cylinder means further including a cylinder fluid passage having a mouth inside the sealing surface (i.e. the mouth is bounded by the sealing surface) and communicat-

ing with the cavity. Thus, the mouth of the cylinder fluid passage is generally over (i.e. in fluid communication with) Me mouth of the first port in both the first and second po-

sitions so that the fluid from the first port continually pressurises the chamber to urge the piston and cylinder apart to form seals against the inner walls of the casing.

Alternatively, in arrangements where the piston means is the sealing member, the piston means may include a piston fluid passage having a mouth for communicating with the mouth of the first port in the same way.

The fluid passage mouth should have an area smaller than that of the chamber in a plane parallel to the plane of the sealing surface, such that the sealing member is urged against the sealing portion of the inner surface.

Preferably, the relative areas of the fluid passage mouth and chamber are se-

lected to provide a desired sealing force between the sealing surface and sealing portion for a predetermined fluid pressure in the first port.

Preferably the valve includes an actuator for moving the sealing member be-

tween the different positions. Advantageously, the different areas, (i.e. the area of the exposure surface, the chamber, the pocket and the fluid passage mouth) area arranged such that when they are all exposed to a fluid at a predetermined pressure, the fric

- 8 tional force between the sealing surfaces does not exceed the maximum actuation force that the actuator can provide.

In certain preferred arrangements the first and second ports have mouths on generally opposing sides of the inner surface, In such arrangements a piston and cylin-

der assembly may slide to form seals around the respective mouths. Preferably the cylinder and piston assembly includes a chamber which is in fluid communication with the first port in all positions so that the chamber is constantly pressurised by the fluid from the first port. The cylinder and piston assembly is moveable to the second posi-

tion in which the first and second ports are connected through the cylinder and piston assembly (by means of the chamber and connecting bores running through the assem-

bly). In other arrangements, a common first port may be arranged so as to constantly pressurise the chamber, whilst sliding of the cylinder and piston assembly connects the first port to one of a number of ports in the opposing side of the inner surface. In such arrangements the numerous ports may be arranged in a top distribution plate for ex-

ample. Preferably the sealing surfaces may be generally annular, and may include at least one straight section. Alternatively, the sealing surfaces may be generally rectan-

gular. When this shape of sealing surface is used in conjunction with ports having slit shaped (i.e. elongate) mouths, this can reduce the distance which needs to be travelled by the sealing member in order to switch the connections between ports.

Advantageously, spring means may be provided to bias the sealing member in the first position.

The actuator for moving the sealing member from the first to the second posi-

tion may advantageously comprise piston means having a surface exposed to fluid pres-

sure from the first port. This piston means may be arranged to be driven, for example, against part of the inner surface of the valve casing in response to exposure to fluid pressure, so tending to move the sealing member.

A second aspect of the present invention provides a valve as defined by Claim 38. The annular sealing member may, for example, be an O ring. It can offer a

( 100% seal duty, as it combines a static sealing arrangement in a dynamic application.

The pressure relief fluid passage connecting the seating to the outlet passage ensures that the annular sealing member is held against its seat by the fluid pressure as the pis-

ton means is withdrawn, i.e. as the piston means moves away from the seat from the first to the second position. The pressure relief passage exposes the "under side" of the annular sealing member (i.e. the portion of the sealing member trapped against the seat) to relatively low pressure.

The pressure relief passage may thus connect a volume underneath the annular sealing member, between the sealing member and the seat, to low pressure. If the sealing member is sufficiently compressed in its seat. however, then this volume may be reduced to zero.

The seating means may, advantageously, comprise a sleeve having a bore ar-

ranged to receive the piston. The seating may be arranged at a root of the bore, or alternatively may be arranged at a shoulder of the bore.

Preferably, the outlet passage may include at least one radial hole in the sleeve at a position axially displaced from the root (i.e. the radial hole and root are spaced apart along the longitudinal axis of the sleeve) and/or from the seating.

Preferably, the piston means may comprise an axially extending bore, and the inlet passage may include that bore.

The piston means may have a bevelled, or angled sealing surface, such as a frus-

toconical sealing face, for compressing the annular sealing member in its seating.

Preferably, the valve may further comprise a spring arranged to bias the piston means and seating means in the first position so that the valve is normally closed. The valve may be used as a pressure relief valve, or a one way valve for example.

Brief description of the drawings

Embodiments of the present invention will now be described, with reference to the accompanying drawings in which:

( - 1 0 Figure 1 is a cross-section of a valve embodying the first aspect of the invention, with the sealing member in a first position; Figure 2 is a cross-section of the valve of Figure 1, but with the sealing member in a second position; Figure 3 is a cross-section of the valve of figures 1 and 2, but with the sealing member in a third position;; Figure 4 is a schematic representation of various valve types to which embodi-

ments of the first aspect of the invention may be applied; Figure 5 is a graph showing erosion resistance properties of examples of mate-

rial suitable for use in embodiments of the present invention; Figure 6 is a cross-section of a valve embodying the second aspect of the present invention; Figure 7 is a cross-section of the valve of Figure 6, shown resisting fluid flow in first direction; Figure 8 is a cross-section of the valve of Figures 6 and 7, allowing flow in a sec-

ond director; Figure 9 is a cross-section of another valve embodying the second aspect of the present invention; Figure 10 is a cross-section of the valve of Figure 9, in the open position; Figure 11 is a schematic representation of various valve types to which em-

bodiments of the second aspect of the invention may be applied; Figures 12 to 15 show schematic, side and plan views of various valves em-

bodying the invention; Figure 16 shows a plan view of another embodiment; Figure 17 is a cross-section of the valve shown in Figure 16; and Figure 18 is a cross-section of a further embodiment.

- 1 1 Detailed description of the preferred embodiments

Referring now to Figure 1, this shows a valve 1 having a casing 2 with an inner surface 3 enclosing a volume 4. A first fluid port 5 (Port P) provides a fluid passage into the casing, and has a mouth 6 at the inner surface 3. A second fluid port 7 also has a mouth 8 at the inner surface. The valve includes a third port 13 and a fourth port 14, which in this example is a tank drain T. An assembly of a piston 17 and cylinder 18 is arranged inside the casing, and the cylinder 18 forms a sealing member 9 which includes a bearing 16. The bearing 16 has a sealing surface 10 which forms a sliding seal against a corresponding sealing portion 11 of the inner surface 3 of the casing. The sealing member 9 has an exposure surface 12 which is exposed to fluid at pressure from the first port 5. This exerts a force on the sealing member 9 which tends to urge the seal-

ing surface 10 against the sealing portion 11 of the inner surface. The piston and cylin-

der assembly is shown in a first position in which the first port 5 is isolated, i.e. it is not connected to any of the other ports. The exposure surface 12 of the cylinder 9 is part of a cavity 20 between the piston 17 and cylinder 18. The cylinder 18 includes a fluid passage having a mouth 24 which is arranged over the mouth 6 of the fluid port 5. This fluid passage keeps the chamber 20 in fluid communication with the fluid in the port 5.

The area of the mouth 24 is less than the area of the cavity 20 in a plane parallel to the sealing surface 11 so that the presence of the fluid in the cavity (chamber) and fluid pas-

sage urges the sealing member 9 against the sealing portion 11 of the inner surface.

The presence of fluid in the chamber 20 also urges the piston 17 against a top plate 151 of the casing. The piston 17 includes a further surface 21 (i.e. a further sealing surface) which forms a seal against the inner surface of the top plate 151. Inside this seal there is a fluid pocket 22 which is in fluid communication with the cavity 20 via a bore 23.

The presence of fluid in the pocket 22 at pressure exerts a force on the piston 17 tending to urge it away from the inner surface of the top plate 151, i.e. it opposes the effect of the pressurised fluid in the cavity 20 urging the piston towards the upper plate 151. Thus, the area of the pocket 22 is selected to provide a hydrostatic balance, i.e. it reduces the reactive force at the interface between the piston sealing surface 21 and the

- 1 2 top plate 151 to an acceptable level such that the frictional force does not exceed the maximum actuation force which can be provided to slide the piston and cylinder assem-

bly. The piston includes an annular sealing member 25 to improve the seal between the piston and cylinder walls.

It will be appreciated that the correct arrangement of sizes and shapes of fluid passage mouth, chamber 20 and pocket 22 result in a piston and cylinder assembly which tends to expand when exposed to fluid at pressure from the first port S. The total expansive force for a given fluid pressure is determined by the relative areas. This expansive force will also be termed a clamping force, and hydrostatic balancing of forces on the components is employed to produce a clamping force which provides adequate sealing but which does not result in the valve sticking.

Moving on to Figure 2, this shows the valve of Figure 1, but with the piston 17 and cylinder 18 assembly, which includes the sealing member 9, in a second position. In this position, the bearing 16 forms a seal around the mouths 6, 8 of the first and second ports. At the same time, the third port 13 is connected to the tank drain 14.

In Figure 3, the cylinder and piston assembly is shown in a third position, in which the first port 5 is in fluid communication with the third port 13, whilst the second port 7 is connected to the tank drain 14. Movement of the cylinder and piston assem-

bly between the first, second and third positions is achieved by linear sliding. At all times, the piston and cylinder components form seals with the opposing flat surfaces inside the casing 2. Thus, the cylinder and piston slide between parallel flat plates 151 and 1 5.

It will be apparent that the valve shown in figures 1, 2 and 3 offers fluid direc-

tional control under controlled leakage conditions. Although this design lends itself predominantly for directional control, sealing interfaces may comprise ceramics, and the ceramic interface option will be resilient and offer excellent durability to flow erosion experienced during pressure control applications. The valve may be activated by hand, pilot (air or fluid) or electrical operation, and proportional or servo control may also be employed.

- 13 Returning to Figure 1, this shows the sectional view through a hydrostatically balanced face valve. The central Port P is the main pressure supply and is isolated in the distribution plate 15 from the ports A and B by the bearing surface 10 of the control piston. In this position the ports are also isolated from the Tank return line T (the case drain) in this instance, and for valves with different centre positions the port Grillings can be repositioned to suit. Leakage across the bearing surface 10 is minimised by the flat-

ness of the mating surfaces 10, 11 and the axial load imposed by the piston 17 and cyl-

inder 18. The resultant force imposed on the piston's bearing surface is hydrostatically balanced to minimise the force required to move the assembly. Actuation of the cylin-

der assembly to the left will connect port P and A, and simultaneously port B is con-

nected to the case drain T. Reversal of the control piston will reverse the porting cons sections. In the valve shown in Figures 1, 2 and 3, it is possible to define a clamp ratio, which is the ratio between the clamping piston/cylinder force and opposing forces in-

duced by the leakage across the sealing faces. The clamp ratio can be altered by alter-

ing the dimensions (in particular the areas) of the chamber 20, the pressure pocket 22, and the mouth 24. With a high clamp ratio, a very high volumetric efficiency can be achieved but the actuation force required is greater. A high actuation force may be un-

acceptable for electrical actuation.

One of the main advantages to the design shown in Figures 1, 2 and 3 is the relative ease of processing hard and uncompromising materials to obtain a wear and flow erosion resistant component. Achieving the final form in both spool and poppet designs requires many machine operations and working to micron tolerances. One of the main costs for such valves is the finishing process and the expensive machinery and measuring equipment required to meet the design specification. The major require-

ment on the face valve shown in figures 1, 2 and 3 is flatness, and this can be achieved with relative ease.

In the valve on Figures 1, 2 and 3, the design is based on achieving a sealed sur-

face through flatness and hydrostatic balancing of mating components. The opening and closing of the main port 5 is achieved by linear motion of the pressure balanced pis

- 1 4 ton/cylinder assembly 17, 18. The material combination satisfyingthe criteria of resis-

tance to flow erosion and corrosion resistance may exhibit very high fictional charac-

teristics. Opffmisation of the hydrostatic balance is important The use of high grade engineering polymers can be considered as the differenffal thermal expansion rates are less important with the face valve design. Figure 5 shows the results of cavitation tests on a variety of materials suitable for use in valves embodying the present invention. All of the components of the valves embodying the invention may be formed from a single one of these materials, or more typically, a combination of materials can be chosen.

In embodiments of the present invention, where the actuation forces are re-

duced to an appropriate level it is possible to position the cylinder/piston assembly di-

rectly, rather than via a mechanical linkage. This may especially apply to valves operat-

ing at the low pressure range. For on/off style valves this will reduce the cost and sim-

plify the valve. For a more sophisticated level of control, i.e. for proportional or servo options, a linkage will offer a greater deal of control (i.e. smaller actuation forces with incremental movements will yield finer motion at the control orifice). In embodiments in which solenoids are used in the actuators, there is a trade off between actuation force and operating distance. Optimisation of available designs will dictate the most appropri-

ate function ratio.

The level of over clamp and ultimate position of the clamp cylinder are impor-

tant factors in the correct functioning of the valve. The area projected to direct system pressures varies under the opening and closing of the port. This is not considered to be a problem on "bang, bang" (either fully open or fully closed) functions, but under servo or proportional operation it may affect the efficiency of the valve. The moment op-

posed on the control face under the flow control condition can also be affected by the position of the opposing clamp force, and it is desirable to design the valve to overcome this moment.

Extremely high flow velocities will result when opening and closing the valve, and erosion of the sealing surfaces will affect the performance and ultimate life of the valve. Material combinations and maximum operating pressure should be selected to give the best performance. Geometry of the port orifices and also the resultant jet di

i - 1 5 rection will also affect performance. By projecting the jet into an area of little influence, the erosion will not be a deciding factor on the ultimate life of the component. The ge-

ometry of the ports will also impact on the valves capacities to control flow and pres-

sure. When used with low viscosity fluids, such as water, relatively large variation in performance will occur from minor motions of the control piston. A restricting opening to the ports may be used, and can give a desirable affect for the control aspect of the valve. Actuation force is also very important if accurate flow control is to be gained by the face valve. Under electrical actuation the actuation force will be set at a very low target. To incorporate proportional control within the valve, mechanical assistance to the mechanism may be more appropriate. It is important that the following criteria be established and optimised to obtain the full range of control.

Firstly, efficiency. A poppet style valve will generate 100% efficiency, i.e. zero leakage when produced at high tolerances. The alternative spool arrangement for wa-

ter hydraulics can yield less than a 90% rating. It is important that maximum efficiency be gained, and a target of 98% of maximum flow rate be transferred in useable volume on the face valve design.

Secondly, hydrostatic balance. To obtain the 98% efficiency specified above, the clamp ratio of the hydrostatic balance will be a critical parameter. The balance is also very important regarding the actuation of the valve. There will be a balance between obtaining acceptable efficiency and appropriate actuation force for all modes of opera-

tion. An alternative approach, used in certain embodiments of the invention, is to of-

fer the efficiency level in operation, but reduce this level when actuating the valve. This will result in minimised actuation load but will not greatly reduce the overall system effi-

ciency. This effect could be produced with a pressure pocket that is energised during control piston motion, or by temporarily unloading the clamp cylinder.

The valves embodying the present invention may be operated in three different modes, namely manual, air or fluid pilot, and electrical.

- 1 6 The actuation of the valve via an electrical input is the most demanding of the three due to the relatively low forces available from a solenoid. Actuation via a motor would be too slow for accurate flow control. The static friction co-effcient is much greater than the dynamic condition, and a high starting force can be applied by using some form of electrical amplification.

It will be further appreciated that the valve shown in Figures 1, 2 and 3 is a 4/3 valve. To optimise the valve performance, it is possible to design the valve first of all with a high degree of over clamp (i.e. with a very high clamping force being generated by the fluid at pressure supplied to the piston and cylinder assembly.) To optimise the design, the balance area of the pocket 22 can be gradually increased. Thus the balance area can be increased by the introduction of a pressure pocket on the working face of

the control piston 17. The pocket dimensions can be incrementally increased, and per formance variation and actuation load recorded.

The valve shown in Figure 1 is a directional 4/3 valve with a closed centre posi-

tion. The ports may include inserts.

Referring now to Figure 6, this shows a cross-section through a check valve em-

bodying the second aspect of the present invention. The male piston 53 carries a sleeve 58 that has a number of radial holes 60. The sleeve 58 is part of a seating member 55 which includes a seat 56 for an O ring 54. With the O ring seal 54 located in the root of the bore of the sleeve 58 it is possible to achieve a 100% seal with flow from the di-

rection of port A 52, as shown also in Figure 7. With flow from this direction, the fluid pressure in the port 52 urges the seating means 55 against the piston 53 and so holds the O ring in compression between the angled sealing surface 62 at the end of the pis-

ton and the seating 56. In this way, the O ring 54 forms an annular seal between the piston and seating means. The spring 63 biases the piston 53 and seating means 55 ton "ether to form the seal with the O ring 54. The spring 63 initially positions the sleeve but the differential area of the sleeve open to pressure results in an axial force adequate to obtain and maintain the seal. With the flow in the opposite direction, from port B 51 the sleeve may be displaced by the flow if the fluid supply has sufficient pressure to compress the spring, and so gives a free path from port 51 to port 52.

( - 1 7 Thus, the valve can act as a pressure relief valve, or a one way valve.

One of the major benefits of the valve shown in Figures 6, 7 and 8 is that when the sleeve is withdrawn (i.e. when the piston 53 and seating means 55 are separated, in the second position) the seal is no longer functioning. Flow control is achieved by the radial holes 60 in the sleeve 58 and the fit on the pin (i.e. piston member 53) as in a spool valve arrangement. By positioning the radial holes of the sleeve to offer a delay when opening to the cavity, the seal is away from the immediate area, thus reducing the possibility of flow erosion. The fit of the sleeve 58 to the pin (piston 63) is, however, not as critical as with the standard spool arrangement. This delay when opening the valve from the first to the second position is thus achieved by positioning the radial hole 60 at a location axially displaced from the seating 56. The greater this axial separation, the greater the delay.

A prototype valve constructed in accordance with Figure 6 proved successful, holding pressure for a number of days without decay. This function is especially desir-

able for lifting equipment and safety functions, which are therefore particular applica-

tions for the valve. With the flow in the opposite direction however, without the pres-

sure relief fluid passage 57 the seal became displaced causing the check not to seal when the flow was again reversed. Pressure held the sealing member 54 against the angular sealing face 62 of the piston 53 as was separated from the sleeve member 58, rather than maintaining contact with the root of the bore (i.e. the O ring seating). The pres-

sure relief fluid passage 57 is a small pressure relieving hole which was drilled in the closed side of the sealed cavity (i.e. O ring seating) which results in positive pressure holding the sealing member into the root of the bore (i.e. against the seating) as the valve is opened. With the fluid passage 57 incorporated in the valve, the valve func-

tioned correctly in both forward and reverse directions without affecting its sealing ca-

pacity. It will be apparent that the valve of Figure 6 incorporates the dual functionality of poppet and spool style sealing' and provides the possibility of total sealing being achieved with minimal actuation forces.

( - 18 A direct acting, electrically energised cartridge valve is shown in Figures 9 and 10. This valve also embodies the second aspect of the present invention. The cartridge valve design offers a cost reduction in multi-function applications, although complicated manifold designs usually warrant higher volumes to withstand the increased design cost.

In the valve of Figures 9 and 10, operation is dependent upon gaining the seal at the spool diameter rather than an alternative diameter within the sealed cavity. An ac-

tuator 100 comprises a solenoid 101 which, when energised, attracts a pin 102 further into its bore. The pin 102 is coupled to piston means 53. A bias spring 63 biases the piston means 53 in a first position when the solenoid is not energised. In this first posi-

tion, a seating 531 at an end of the piston means 53 is urged against an O ring 54 seated in a seating 56 in a sleeve member 58. This results in an annular seal being formed between the piston means and sleeve member 58, which prevents fluid 51 sup-

plied to radial hole 60 in the sleeve member 58 from passing into a bore 52 of the sleeve member 58. The piston 53 includes a second O ring seating 67 carrying a sec-

ond O ring 64. This second seating is axially displaced from the end seating 531.

When the piston 53 is in the first position, this second O ring forms a further annular seal between the piston 53 and sleeve member 58. The two annular seals formed when the piston is in the first position acts to further isolate the inlet port 60. The sleeve member 58 includes a further seating 66 to receive the second O ring 64 when in the first position. In this example, the first and second scalings 56, 66 are both lo-

cated on shoulders of the bore 52 running axially through the sleeve 58. This is in con-

trast to the arrangement shown in Figure 6, in which the seating 56 was at a closed root of the bore.

A portion 59 of the sleeve member bore receives the piston 53.

Figure 10 shows the valve of Figure 9, but with the solenoid energised and the valve open. The piston 53 has been withdrawn, against the spring 63, such that the O rings 54 and 64 no longer form seals between the piston and sleeve, and the inlet ports 60 are connected to the outlet port 52.

It will be appreciated that the design of the valve shown in Figures 9 and 10 is dependent upon gaining the seal at the spool diameter rather than an alternative diame

- 1 9 ter within the seal cavity. Failure to achieve this will result in an axial load that the spring and solenoid will need to overcome for actuation. By achieving an appropriate design, little or possible zero axial load will be applied to the moving section of the valve. Actuation forces via spring and electrical solenoid will be possible for minimum energy input allowing direct, as opposed to pilot, operation. The force required to achieve the seal against a given pressure is an area of development of the valve. The size of the valve will also determine the required actuation forces, and it is proposed that 6mm ports are used for prototypes appropriate for flow rates up to 30 litres per minutes. Although not shown in Figures 9 and 10, this cartridge style valve will also in-

clude a pressure relieving passage connecting the underside of the O ring 54 to the outlet port 52 to ensure that the O ring is retained in its seating 56 as the piston is withdrawn. Material combinations are less important with respect to frictional characteristics with the design of valve shown in Figures 6 to 10. Flow erosion resistance also has less of an affect on the performance and ultimate life of the components owing to the two different positions of sealing and the flow control. It will be apparent that the valves shown in Figures 6 to 10 offer 100% sealed performance when in the closed position but, due to the sealing technique, mimic a seal-less spool actuation performance. Op-

eration by hand, pilot or electrical means are possible directly rather than through pilot or mechanical advantage. The valves shown in Figures 6 to 10 can also provide other possible functions such as check, priority flow control valve, pilot operated check, and sealed V2 valves. Thus, valves embodying the second aspect of the present invention can be used for a variety of applications, as shown in Figure 11.

Figure 1 2(a) is a schematic representation of a further valve embodying the pre-

sent invention. Figure 1 2(b) shows side and plan views of part of this valve. Ports 8 and 10 are provided in a top distribution plate. A piston 17 carries a bearing member 161 which forms a sliding seal against the inner surface of the top distribution plate. In Figure 12(b) the bearing 161 is shown in a first position in which the port 8 is con-

nected to a bore 23 running through the piston 17. The sealing surface of the bearing

( - 20 161 is generally annular 21, although it has straight side portions. Ports 8 and 10 are elongate, and generally resemble slits arranged parallel to the straight sides of the seal-

ing surface 21. This reduces the distance which the piston 17 has to travel to switch from connecting port 8 to the bore 23 to port 10. Figure 1 2(c) shows the piston 17 in a central position in which the sealing surface of the bearing 161 blocks both ports 8 and 10. Figure 12(d) shows the piston 17 in a third position in which it connects port 10 to the bore 23, but isolates port 8.

Figure 1 3(a) is a schematic view of a further embodiment. Figures 13 (b), (c) and (d) show side and plan views of this embodiment with the piston 17 in first, second and third positions. As with the embodiment shown in Figure 12, pressure may not build in the centre position, and may only increase when the cylinder is positioned over a port. Although not shown in Figures 12 to 15, the valve may include a valve feed port, which may be on a lower surface of the casing. The valve feed port and distribu-

tion ports may be in opposing surfaces. This provides the advantage that the ports can be drilled closer together, thereby reducing the overall valve size and minimising the required stroke length to open and close the ports.

The distribution hydrostatic pockets may vary in shape to minimise the stroke length for opening and closing. The ports may be kidney shaped, slot or rounded, de-

pending on the opening conditions and desired stroke of the cylinder.

Moving on to Figure 14, Figure 14(a) shows a further embodiment in schematic form. Figures 1 4(b), (c) and (d) show side and plan views of the embodiment with the piston 17 in three different positions. In the central position shown in Figure 14(c) the ports 8 and 10 are in fluid communication with another. In such an arrangement, if there is a central valve feed port providing pressurised fluid to the cylinder and piston assembly then that feed port is in a fluid communication with both ports 8 and 10 in this central position.

Figure 1 S(a) shows a schematic view of a further embodiment. As can be seen from Figures 15 (b) and (d), the sealing surface 21 in this case is generally rectangular.

Moving on to Figure 17, a further embodiment comprises a feed port 5 having a mouth 6 at an inner surface of a bottom plate 15 of the valve casing. Two further

ports, 7 and 13, are provided in a top distribution plate 151. These ports have respec-

tive mouths 8 and 10 at the inner surface of the distribution plate 151. The inner sur-

faces of the distribution plate 151 and bottom plate 15 are flat and parallel to each other. Between these surfaces is located an assembly comprising a piston 17 and a cyl-

inder 18. The piston and cylinder carry respective bearings 16 and 161 which form respective sliding seals against the flat inner surfaces of the top and bottom plates. In-

side the piston and cylinder assembly there is a cavity 20, connected by bores 23 in the piston and cylinder to respective mouths at the bearing surfaces. The piston and cylin-

der assembly is slidable. but in all positions the mouth of the bore 23 at the lower bearing surface 16 is in fluid communication with the mouth 6 of the feed port 5. Thus, the cavity 20 is constantly pressurised by fluid from the feed port 5. Figure 17 shows the cylinder and piston assembly in a central position in which the bore 23 of the cylin-

der is in fluid communication with the right-hand port 10, so connecting port 13 to feed port 5. In this position the mouth 8 of the other port 7 is blocked. Actuation of the valve is achieved by means of fluid supplied through a further port 800. This acts to drive the cylinder and piston assembly towards the left, to connect the port 7 to the feed port S and disconnect port 13. Figure 16 shows the valve of Figure 17 from above. Moving on to Figure 18, this shows a further embodiment. In this arrangement a single port 5 is provided in a bottom plate 15, and a single port 7 is provided in a top plate 151. A spring 200 biases the assembly of cylinder 18 and piston 17 in a first posi-

tion in which the ports are isolated from one another. In this first position the fluid from the port 5 pressurises the chamber 20 and bores 23 in the cylinder and piston assembly. A transverse piston 201 is arranged in a transverse bore of the cylinder 18.

A head surface 202 of the transverse piston 201 is exposed to fluid pressure in the chamber 20 and bores 23 and the piston 201 is thus arranged to drive the piston and cylinder assembly to the left of the figure in response to this exposure. Thus, as fluid pressure in the port 5 increases, the piston 201, drive the piston and cylinder assembly against the biasing spring 200 until eventually the mouth 22 of the piston is brought into fluid communication with the mouth 8 of the port 7. When this happens, the two ports

( - 22 are in fluid communication with one another. It will be apparent that this valve can act as a pressure relief valve or pressure control valve. In this embodiment the piston and cylinder do not carry separate bearing members, but instead the sealing surfaces 21 and 10 are simply generally annular surfaces of the piston and cylinder themselves.

By using the pressure within the cylinder and piston assembly a force acting perpendicular to the clamp force can be applied via the second (i. e. the transverse) pis-

ton 201. By opposing this force with the pre-loaded spring 200 the valve will remain in a closed position until the force is greater than the spring force. When the load is greater than the spring force the cylinder motion will connect the cylinder port to the second port and that will ensure that the maximum pressure predetermined by the spring force is not exceeded.

It will be apparent from the foregoing description that the sealing surfaces in

embodiments of the present invention may be provided by two separate interfaces.

Embodiments of the invention may have an open-centre valve configuration, which re-

sults in pressure not increasing until actuation of the valve.

Overclamp is the term used for the ratio between the positive sealing force within the cylinder and the balance force opposing the clamp force via pressure pockets in one or both bearing surfaces.

An advantage of certain preferred embodiments of the invention is that the sys-

tem pressure is only seen within the cylinder and piston assembly and the ports them-

selves. This gives the advantage that one can use materials such as plastic for the valve housing. In the arrangement shown in Figure 18, displacement of the sliding cylinder and piston assembly is achieved by means of a perpendicular piston which gives pressure control. Valves embodying the present invention can be adapted to offer relief, unload-

ing, pressure compensated flow control, and over-centre operation, some of which functions to date are not commercially available for water service.

Embodiments of the present invention (both the first and second aspects) can be used in three different pressure ranges of operation. Firstly, low cost plastic versions

* -23 can be used for maximum system pressures of 70 bar. Secondly, a plastic/ stainless steel valve embodying the invention can operate on system pressures up to 160 bar.

Thirdly, embodiments employing stainless steel and ceramic materials can be used for 350 bar operation. These three options enable cost effective water systems to be pro-

duced for a broad spectrum of industry requirements.

It will be apparent that the hydrostatic pressure balancing of the poppet forces is an important principle in certain embodiments of the first aspect. Coupled with the lin-

ear actuation motion and the possible mechanical advantage, such valves offer a high potential for control.

The valves embodying the second aspect of the invention utilise a static sealing arrangement in a dynamic application. They offer excellent sealing potential without the high actuation forces associated with dynamic high pressure seals.

Applications for valves embodying the present invention include water hydraulic systems, and control of water flow for traditional fields such as fire fighting, desalination

systems (reverse osmosis), simple water jetting applications and process industry con-

trol methods. Valves embodying the invention may also find particular application in the relatively new field of low pressure water hydraulics, that is valves offering controllabil-

ity and greater efficiency in systems operating at a maximum of 50 bar. Valves em-

bodying the present invention may also be used in pneumatic systems.

Claims

- 24 Claims 1. A valve comprising: a casing having an inner surface

enclosing a volume; a first fluid port into the casing; a second fluid port into the casing; a sealing member inside the casing having a sealing surface for sealing against a corresponding sealing portion of the inner surface of the casing, and an expo sure surface arranged for exposure to a fluid at pressure so as to urge the seal ing surface into sealing engagement with the sealing portion of the inner surface, the sealing member being movable betvveen a first position in which the seal formed between the sealing surface and sealing portion separates the first and second ports, and a second position in which the first and second ports are in fluid communication with one another.
2. A valve in accordance with claim 1, wherein the exposure surface of the sealing member is arranged so as to be exposed, in use, to fluid at pressure from the first port.
3. A valve in accordance with claim 1 or claim 2, further comprising a third fluid port into the casing, and wherein the sealing member blocks the first, second and third ports when in the first position, connects the first and second ports when in the second position, and is further movable to a third position in which the first and third ports are connected.
4. A valve in accordance with claim 3, further comprising a fourth fluid port into the casing, wherein the sealing member in the first position isolates the fourth port from the first, second and third ports, In the second position connects the third and fourth ports, and in the third position connects the second and fourth ports.
5. A valve in accordance with any preceding claim, wherein the first port has a mouth at the inner surface and the sealing surface and sealing portion from a seal around the mouth of the first port in the first position.

- 25
6. A valve in accordance with any preceding claim, wherein the first and second ports have respective mouths at the inner surface, and the sealing surface and sealing portion form a seal around the mouths of the first and second parts in the second position.
7. A valve in accordance with any one of claims 1 to 5, wherein the first and sec ond ports have respective mouths at the inner surface, said mouths being ar ranged on generally opposing sides of said inner surface.
8. A valve in accordance with my preceding claim wherein the sealing surface and sealing portion are flat.
9. A valve in accordance with any preceding claim wherein the sealing surface is generally annular.
10. A valve in accordance with claim 9 wherein the sealing surface comprises at least one straight section.
11. A valve in accordance with any one of claims 1 to 9 wherein the sealing surface is generally rectangular.
12. A valve in accordance with any preceding claim wherein the casing comprises a distribution plate, having a flat inner surface, at least the first and second ports having respective mouths at said flat inner surface, and the sealing member be-

ing arranged to seal against said flat inner surface.
13. A valve in accordance with any preceding claim, wherein the sealing member comprises a sealing bearing, and the sealing surface is a surface of said bearing.
14. A valve in accordance with any preceding claim, wherein the seal formed be-

tween the sealing surface and sealing portion is a sliding seal, and the sealing member is slidable between the first and second positions.
15. A valve in accordance with any preceding claim, comprising a piston means and cylinder means assembly arranged inside the casing, the sealing member being the piston means or the cylinder means.
16. A valve in accordance with claim 15 wherein the assembly includes a chamber, between the piston means and cylinder means, said chamber being arranged so as to be pressurized by said fluid, in use, such that the piston means and cylinder

- 26 means are urged apart and the sealing surface is urged to seal against the sealing portion of the inner surface.
17. A valve in accordance with claim 16 wherein the assembly comprises a further surface, said further surface and sealing surface being urged apart by the pres surisation of the chamber, such that the sealing surface is urged to seal against the sealing portion and the further surface is urged against a further portion of the inner surface, said further portion being generally opposite the sealing por tion.
18. A valve in accordance with claim 17 wherein the sealing portion and further portion are opposing flat parallel portions of the inner surface.
19. A valve in accordance with claim 17 or claim 18 wherein the further surface is arranged to form a sliding seal against the further portion.
20. A valve in accordance with claim 19 wherein the further surface and further portion are flat.
21. A valve in accordance with claim 19 or claim 20 wherein the seal formed be tween the further surface and further portion encloses a fluid pocket between the assembly and inner surface.
22. A valve in accordance with any one of claims 19 to 21 wherein the seal formed between the further surface and further portion form a seal around the mouth of the second port when the sealing member is in the second position.
23. A valve in accordance with claim 21 or claim 22 wherein the further surface is generally annular or rectangular.
24. A valve in accordance with claim 21, or claim 22 or claim 23 wherein the as sembly includes a fluid passage connecting said pocket and said chamber, whereby the presence of pressurised fluid in the pocket counteracts the effect of pressurised fluid in the chamber urging the further surface against the further portion to give a desired sealing force.
25. A valve in accordance with claim 24 wherein said fluid passage is a bore in the piston or cylinder means.

( - 27
26. A valve in accordance with any one of claims 17 to 25 wherein the sealing sur-

face is a surface of the cylinder means and the further surface in a surface of the piston means.
27. A valve in accordance with any one of the claims 16 to 26 wherein said chamber is arranged so as to be in fluid communication with said first port in the first and second positions.
28. A valve in accordance with claim 27 wherein the cylinder means is the sealing member and the sealing surface forms a seal around a mouth of the first port in the first and second positions, the cylinder means further including a cylinder fluid passage having a mouth inside the sealing surface and communicating with the cavity.
29. A valve in accordance with claim 27 wherein the piston means is the sealing member and the sealing surface forms a seal around a mouth of the first port in the first and second positions, the piston means further including a piston fluid passage having a mouth inside the sealing surface and communicating with the cavity.
30. A valve in accordance with claim 28 or claim 29 wherein the sealing surface is generally annular or rectangular.
31. A valve in accordance with any one of claims 28 to 30 wherein the sealing sur-

face is planar and the fluid passage mouth has an area smaller than that of the chamber in a plane parallel to the plane of the sealing surface.
32. A valve in accordance with claim 31 wherein the relative said areas of the fluid passage mouth and chamber are selected to provide a desired sealing force be-

tween the sealing surface and sealing portion for a predetermined fluid pressure in the first port.
33. A valve in accordance with any preceding claim, comprising an actuator for moving the sealing member between the first and second positions.
34. A valve in accordance with claim 33, as dependent on claim 24, wherein the ac-

tuator is operable to provide up to a maximum actuation force, and the pocket and chamber have respective areas selected such that when exposed to fluid at a

( - 28 predetermined pressure, the frictional force between the sealing and further surfaces of the assembly and the inner surface of the casing is less than the maximum actuation force.
35. A valve in accordance with claim 33 or claim 34, comprising spring means ar-

ranged to bias the sealing member in said first position.
36. A valve in accordance with any one of claims 33 to 35 wherein said actuator comprises piston means having a surface exposed to fluid pressure from the first part, the piston means being arranged to urge the sealing member towards said second position in response to the force exerted on said surface by said expo-

sure.
37. A valve in accordance with any preceding claim wherein at least one of said ports and/or said mouths is generally elongate in crosssection.
38. A valve comprising: a fluid inlet passage; a fluid outlet passage; piston means; an annular resilient sealing member; and seating means comprising a seating for the annular member, the piston means and seating means being movable relative to one another between a first posi-

tion, in which the annular member is compressed between the piston and seat-

ing means such that the fluid inlet and outlet passages are not connected, and a second position in which the annular member is not compressed between the piston and seating means and does not form a seal between them such that the inlet and outlet passages are in fluid communication with each other' the seating means further comprising a pressure relief fluid passage connecting the seating to the outlet passage.
39. A valve in accordance with claim 38, wherein the seating means comprises a sleeve having a bore arranged to receive the piston means.
40. A valve in accordance with claim 39 wherein the seating is arranged at a root of the bore.

f -29
41. A valve in accordance with claim 39 or claim 40 wherein said outlet passage in-

cludes at least one radial hole in the sleeve at a position axially displaced from the seating.
42. A valve in accordance with any one of claims 38 to 41 wherein said piston means comprises an axially extending bore, and said inlet passage comprises said bore.
43. A valve in accordance with any one of claims 38 to 42 wherein said piston means comprises a frustoconical sealing face for compressing the annular mem-

ber in the seating.
44. A valve in accordance with any one of claims 38 to 43, further comprising spring means arranged to bias the piston means and seating means in said first position.
45. A valve substantially as hereinbefore described with reference to and/or as shown in the accompanying drawings.