US2272684A - Hydraulically actuated member and speed control therefor - Google Patents

Hydraulically actuated member and speed control therefor Download PDF

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US2272684A
US2272684A US26052639A US2272684A US 2272684 A US2272684 A US 2272684A US 26052639 A US26052639 A US 26052639A US 2272684 A US2272684 A US 2272684A
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pressure
valve
chamber
piston
liquid
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Harry F Vickers
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Vickers Inc
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Vickers Inc
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    • BPERFORMING OPERATIONS; TRANSPORTING
    • B23MACHINE TOOLS; METAL-WORKING NOT OTHERWISE PROVIDED FOR
    • B23QDETAILS, COMPONENTS, OR ACCESSORIES FOR MACHINE TOOLS, e.g. ARRANGEMENTS FOR COPYING OR CONTROLLING; MACHINE TOOLS IN GENERAL CHARACTERISED BY THE CONSTRUCTION OF PARTICULAR DETAILS OR COMPONENTS; COMBINATIONS OR ASSOCIATIONS OF METAL-WORKING MACHINES, NOT DIRECTED TO A PARTICULAR RESULT
    • B23Q5/00Driving or feeding mechanisms; Control arrangements therefor
    • B23Q5/22Feeding members carrying tools or work
    • B23Q5/26Fluid-pressure drives
    • B23Q5/266Fluid-pressure drives with means to control the feed rate by controlling the fluid flow
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15BSYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B13/00Details of servomotor systems ; Valves for servomotor systems
    • F15B13/02Fluid distribution or supply devices characterised by their adaptation to the control of servomotors
    • F15B13/04Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor
    • F15B13/0416Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor with means or adapted for load sensing
    • F15B13/0417Load sensing elements; Internal fluid connections therefor; Anti-saturation or pressure-compensation valves
    • YGENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y10TECHNICAL SUBJECTS COVERED BY FORMER USPC
    • Y10TTECHNICAL SUBJECTS COVERED BY FORMER US CLASSIFICATION
    • Y10T137/00Fluid handling
    • Y10T137/7722Line condition change responsive valves
    • Y10T137/7781With separate connected fluid reactor surface
    • Y10T137/7784Responsive to change in rate of fluid flow
    • Y10T137/7787Expansible chamber subject to differential pressures
    • Y10T137/7788Pressures across fixed choke
    • YGENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y10TECHNICAL SUBJECTS COVERED BY FORMER USPC
    • Y10TTECHNICAL SUBJECTS COVERED BY FORMER US CLASSIFICATION
    • Y10T137/00Fluid handling
    • Y10T137/7722Line condition change responsive valves
    • Y10T137/7781With separate connected fluid reactor surface
    • Y10T137/7793With opening bias [e.g., pressure regulator]

Description

Feb. 10, 1942. H. F. vlcKERs 2,272,684
HYDRAULICALLY ACTUATED MEMBER AND SPEED CONTROL THEREFOR Original Filed June 12, 1931 3 Sheets-Sheet l -Wary F Vickers INVENTOR.
ATTORNEYS Feb. 10, 1942. H. F. vlcKERs 2,272,684
HYDRAULIALLY ACTUATED MEMBER AND SPEED CONTROL THEREFOR Original Filed June 12, 1931 .'5 Sheets-Sheet 2 INVENTOR.
ATTORNEY S Feb. 10, 1942. H F, wcm-:RS 2,272,684
HYDRAULICALLY ACTUATED MEMBER AND SPEED CONTROL THEREFOR Original Filed June l2, 1951 3 Sheets-Sheet 3 l u R A ATTORNEYS Patented Feb. 10, 1942 HYDBAULICALLY ACTUATED MEMBER AND SPEED CONTROL THEREFOB Harry F. Vickers, Detroit, Mich., assignor to Vickers, Incorporated poration of Michigan June 12, 1931, Serial No. and this application March 8, 260,526
Original application 543,908. Divided 1939, Serial No.
9 Claims.
This invention relates to a hydraulically actuated member and speed control therefor, and has to do particularly with liquid actuated devices which are subject to a variable speed or a. variable load.
This application is a division of my copending application Serial No. 543,908, led June 12, 1931, since abandoned.
Heretofore, in the control of the speed of liquid actuated devices, such as machine tools and the like, it has been the practice to make use of systems embodying either a variable displacement pump or a constant displacement pump. In systems using variable pumps, it is the practice to set the pump corresponding to a certain load and predetermined speed. Theoretically, such variable pump should maintain a constant speed of the liquid actuated element regardless of a varying load, but practically, there is a change in slippage in such pumps which results in a slowing down when an increased load is met or an increase in speed equivalent to the amount of slippage when load is released. In practice, in systems using variable pumps, fairly good results are often obtained when the work is irst set up, but when the oil warms up and thins, and as the tools become dull, there is a noticeable slowing up due to a change in the slippage in the variable pump.
In control systems utilizing constant displacement pumps, present practice is to control the speed of the actuated member by the use of throttle valves placed either in the intake or discharge sides of the actuating cylinder. In such systems, however, it has been impossible to maintain aconstant speed of actuated member due to change in differential pressure across the throttle valve with the result that the speed varies with the change in load on actuated member.
It is the object of the present invention to provide means operating independently of the particular type of pump used and independently of the viscosity of the liquid used, to uniformly and positively control the speed of any liquid actuated member; this control of speed being either uniform regardless of the load, or variable relative to any change in load. More specically, the present invention utilizes the principle that if a given pressure diiferential is maintained across a given size orifice the flow through said orifice will remain constant.
An important feature of the present invention resides in the fact that my speed control means is placed in the discharge side of the system in Detroit, Mich., a cororder to prevent undesired movement in case of change in the load from positive to negative, although it will be understood that very efficient results may be obtained if the unit is placed ahead of the liquid actuated device in any other case.
My complete system, as contemplated in the present invention, embodies a liquid actuated member such as a machine tool element which is subjected to a variable load either positive or negative during its movement in any given direction, a source of uid supply under pressure, an orice regulable to control the ow of liquid therethrough, and hence the speed of the liquid actuated element, and means operating in combination with the other elements of the system to positively control the pressure diierential across the orice to maintain or control the predetermined speed. This means embodies a hydrostatic pressure regulating valve, preferably in combination as a single unit with a variable orice, said hydrostatically operating valve being so arranged and proportioned that any change in pressure due to a change in load can be made to increase, decrease or maintain the pressure in a certain part of the valve and hence cause the differential pressure across the orifice to either increase, decrease, or maintain constant the speed of the liquid actuated element, as desired.
Other features, including the lapplication of the balanced pressure regulating valve to and in combination with the general' system and the structural details of the valve itself, will be more clearly brought out in the specification and claims.
In the drawings:
Fig. 1 is a somewhat diagrammatic view of a complete liquid circulating system embodying the present invention, and showing particularly the manner of connecting my speed control valve into the system whereby to control the speed oi the liquid actuated element in one direction.
Fig. 2 is a sectional view of my hydrostatic 0I speed control valve as embodied in a single complete unit, the balanced valve being shown in raised seated position.
Fig. 3 is a front elevation of the complete unit and showing the adjustable indicating lever for regulating the size of the discharge orifice.
Fig. 4 is a fragmentary sectional view of the end of the balanced valve and showing a modified manner of varying the eifective diameter of one end of the valve.v
Fig. 5 is a somewhat diagrammatic assembly view of a circulating system embodying the present invention and showing the manner of connecting my speed control valve into the discharge side of the system, and controlling the speed of the liquid actuated member in opposite directions of movement.
Fig. 6 is a view similar to Fig.,5, but illustrating the manner of connecting the speed control unit into the intake side of the system.
First taking up the description of the speed control or hydrostatic valve as an integral unit of manufacture, and as best shown in Figs. 2 and 3, it will be seen that said valve consists of a housing 2 which is provided with a chamber 3 in communication with the incoming liquid and a chamber 4 for connecting a balanced inlet valve 5 and a variable oriiice outlet valve 6. This outlet valve 6 controls the volume of liquid passing into the outlet 1.
The balanced valve 5 is connected by means of a stem 8 to a piston 9, the diameter A1 of the stem at the point III being predetermined relative to the effective diameter AT of the valve 5 at its valve seat. Under usual conditions, the diameter of the portion I0 of the valve stem will be equal to the effective diameter of the valve 5, but such respective diameters may be varied as will be later described. In Fig. 2, di ameter A2 is shown smaller than diameter A1. A coil spring II of predetermined pressure is so positioned as to normally tend to open the valve 5. The chamber 4 is in direct communication with a'chamber I2 on one side of the piston 9 by means of a passageway I3 so that the strength of the spring II as balanced against the pressure upon the area. of the piston 9 determines the opening of the valve 5.
The space I4 on the other side of the piston 3 is in communication, by means of the passageway I5, with the outlet passageway 1 to compensate for any back pressure under any operating conditions, as, for instance, when a long pipe connected to the outlet 1 would set up a material amount of resistance. The back pressure in the line 1 will vary and one of the varients will be the length of the pipe 1, but as the back pressure in the pipe I5 is effective against the top of piston 9 it will be seen that the combination of the spring II, piston 9, conduit' I3, valve 5, and orice 6 will maintain head in the chamber 4, or in other words, a constant dierential pressure across the orifice regardless of the amount of back pressure in pipe I5. Because of the arrangement of the conduit I5 connected to the top of the piston 9, whatever back pressure there might be in the line 1 will have no effect whatsoever upon the differential across the orifice E but, of course, it would change the actual pressure within the chamber 4. In other words, let us assume that the actual pressure in the chamber 4 is fteen pounds and the back pressure zero, now if a back pressure should be set up in 1, say, twenty pounds, then the actual pressure in the chamber 4 would be thirty-iive pounds, thus leaving a differential head of fifteen pounds the same as before. It will be obvious that the same is true regardless of whether the valve is balanced or unbalanced.
The housing 2 also preferably provides a suitable bearing for supporting a shaft I 6 rigidly connected to the variable orice valve 6 and actuated by means of a suitable lever I1. It will thus be evident that the orifice may be readily adjusted to any predetermined opening by moving the lever I1 to the desired point on the graduated scale at the face of the unit.
The operation of the detailed part of the hydrostatic valve as shown in Fig. 2, and the utilization of the same in controlling the speed of the liquid actuated element may best be understo/od by describing the operation of the complete liquid circulating system. Referring to Fig. l, for instance, wherein a simple circuit is shown as comprising a constant displacement type pump I8, a relief valve I9, a four-way control valve 20, a cylinder 2I containing a liquid actuated element or piston 22, a check valve 23, a feed control valve 2, and a tank 24; the liquid flows through the pump I8, relief valve I9 and four-way valve 2U, to one end of the cylinder 2|, thus exerting pressure on the head of the piston as indicated by the arrows. This pressure tends to move the actuated element or piston 22 in one direction and to force the liquid on the other side of the piston out through the pipe 25. As the check valve 23 will prevent flow therethrough in this direction, the liquid will be forced to enter the chamber 3 of the valve housing 2. The spring II will normally hold the valve 5 open and liquid entering from the pipe 25 will iill the chamber 3 and chamber 4 and enter the passageway I3 and fill the chamber I2 thus creating a pressure under the piston 9 tending to move it towards valve closing position. It will be seen that if this pressure should be sufiiciently high, it would completely overcome the spring II and close the valve 5 against the removable seat 26.
Upon moving the valve 6 by the handle I1 to any selected point, the liquid will flow through this valve and through the pipes and control valve 20 back to the tank. Now, when this valve 6 is open and the liquid allowed to escape from the chamber 4, a pressure drop occurs which will result in the spring II opening the Valve 5 and admitting additional liquid to the chamber 4 whereby to instantly build up the pressure in the chambers 4 and I2 to thus in turn move the piston 9 upwardly'towards valve closing position. As the orifice control valve 6 has been set at a certain position, the outlet of liquid from the chamber 4 will be continuous and therefore the inlet of liquid past the valve 5 would have to be continuous to maintain a pressure in the chamber 4. It will thus be seen that the piston 9 is automatically positioned so that just sufficient liquid is admitted past the valve 5 to the chambers 4 and I2 to maintain pressure necessary to partially overcome the spring II.
Disregarding the back pressure in line I5, and with the effective areas of the portion I0 and the valve 5 equal, the pressure in the chamber 4 will be maintained constant, for any given spring, regardless of the opening of the orifice provided with a valve 6. Any tendency to increase the pressure in the line and in the chamber 3 will be immediately reflected in the chamber I2 to raise the piston and valve 5 towards valve closing position. Likewise, any pressure drop in the chamber 3 will be immediately reected in the chamber I2 and the spring II will immediately open or enlarge the opening past the valve 5 and admit more liquid to the chamber 4 to maintain pressure therein. Thus, it is seen that uniform pressure is maintained in the chambers 4 and I2 regardless of the pressures in the chamber 3 and since a uniform pressure is maintained in the chamber 4, the flow past the orifice valve 6 must always be at the same rate for a given position, any viscosity change due to temperature being negligible.
Taking into consideration the varying back pressure in the line I5, the pressure in the chamber 4 will vary in accordance with any variation in the ,back pressure, but the differential between the pressure in the chamber 4 and that in the line I will be maintained regardless of any change in back pressure. Thus it will be seen that with a uniform pressure in chamber-4 or with a uniform differential pressure across the orifice, the ow past such orifice will, of necessity, be constant for any given opening of the orifice 6; if the rate of ow is constant on the exhaust side of the motor, then regardless of the load upon the piston 22 of the motor and up to the capacity of the setting of the relief valve I9, the speed of the piston 22 will be maintained at the rate set above the orifice opening. The relief valve IS is set to equal the maximum desired load on the piston 22. The speed control unit 2 will, of course, have less capacity, when operating to control the speed of the piston 22, than the capacity of the pump I8. Thus if the pump I8 has a iive gallon per minute capacity, and a certain setting of the speed control unit 2 a one-half gallon per minute capacity, and if the maximum load on the piston 22 and the maximum setting of the relief valve I9 should be say two thousand pounds, then if the pump is started up and the liquid directed into the inlet side of the cylinder 2|, it will be seen that the back pressure against the piston 22 caused by the small capacity of the speed control unit 2 will cause a maximum pressure to build up on Vthe inlet side of the cylinder 2| and against the relief valve I9, with the result that the excess liquid from the pump will be spilled over the relief valve. It will be obvious that there would be no function in having a hydrostatic Valve or ow control unit 2 unless its capacity were less than the pump volume; this excess of the pump over the iiow control unit is taken care of by the relief valve.
When the valve 20 is reversed by means of its actuating lever, vliquid will pass around the speed control valve assembly 2 and through the check valve 23 allowing a quick return and a rate equivalent to the full volume of the pump I8. During the power stroke of the piston 22, any liquid pumped in excess of the required amount to ll the cylinder 2| at the rate it is allowed to move by the speed regulator valve, is allowed to escape through the relief valve I9 back to the tank.
In the construction of the speed control valve assembly, it is important that the diameter A1 of the stem .at the point I bear a certain predetermined relation to the effective diameter A2 of the valve 5. Normally, these two diameters are preferably equal so that pressure in the chamber 3 will not impart either a downward er upward force on the piston 9. However, in many instances, it might be desired to increase the speed of the liquid actuated element as the load increases, and in this case the effective diameter of the valve l will preferably be made smaller than the effective diameter at the point lil (as shown in Fig. 2). In this case, any drop in pressurc in the line 25 due to a sudden load will cause an increase of the pressure in the chamber 4, thus maintaining increased speed during the effective period of said increase in load, as will hereinafter be explained in greater detail.
In `order to change my valve assembly to vary thus be seen that by proportioning the relative areas of the valve 5 and valve stem at IU, any decrease in pressure due to load upon the member 22 can be made vto increase or decrease pressure in chamber 4 and thus increase or decrease the speed of the member 22.
In Fig. 5 I have shown my hydraulic system as applied to a member to be actuated in whichv the speed thereof is controlled in both directions. In this circuit the four-way valve 20 controls the ow of liquid under pressure alternately to opposite sides of the piston 22. As shown, this piston 22 may control the actuation of a turret head, or any other machine tool element. The liquid from thedischarging side of the cylinder 2| will be directed by the four-Way valve through the pipe 24 tothe speed control unit, and this four-way valve will control the application of the pressure to the orice 6, regardless of the pressure in the chamber 3, in the same manner as described in connection with the operation of the circuit shown in Fig. 1. Inasmuch as both sides of the cylinder 2| discharge through the valve unit 2, it will be seen that thespeed of the piston 22 will be accurately controlled in both directions regardless of a negative or positive load against the piston 22. A rapid traverse valve 30 may be inserted as a shunt around the speed control unit 2 to bypass the speed control unit 2 when desired.
I have also indicated in Fig. 5 approximate pressures that might exist in the valve system in a certain condition of operation. With a resistance or load of approximately one thousand pounds and with a piston diameter of four inches and a piston rod diameter of one inch, the pump I8 could be regulated by setting the relief valve I9 to deliver seven hundred and fty pounds per square inch, which pressure will be directed by means of the valve to the left hand side of the piston 22. To obtain the proper movement ofthe piston 22 the pressure on'the discharge side will be approximately 640 pounds per square inch and approximately this same pressure will be present in the chamber 3 of the unit 2. With a certain Spring Il in the unit 2 and a certain diameter of the piston 9, a pressure of ten pounds per square inch will be maintained in the chamber 4; in this particular case the area of the piston plus the valve stem would be approximately two square inches, and the spring strength twenty pounds. vIt will be evident that in this arrangement the variable valve 6 may be set to obtain any speed desired of the work element, regardless of a negative or positive load in a single direction of movement and in every such set-up the pressure in the chamber 4 will remain constant at ten pounds per square inch.
In Fig. 6 I have shown a slightly modified hydraulic hook-up wherein the speed control unit 2 is placed in the intake side of the system. This particular hydraulic hook-up may be utilized with many machine tools and other elements and has practically all the advantages of the of a sudden release of the load as in case wherein the drill has protruded through the work and g the inertia force of the weight of the drill head would be greater than the force exerted by the vacuum behind the piston.
To more specifically express the operation of the system, assuming that there is no pressure and the entire system, as shown in Fig. 1, is at rest, the lever l1 may be placed in any position to give the required speed. The pump I8 may then be started and the liquid under pressure flows through the four-way valve 20 into the upper end of the cylinder 2|. This causes the piston 22 to move. driving the oil on the other side thereof into the control valve 2, the piston of which is at that time held wide open by the spring. As soon as the pressure in chamber 4 builds up due to restriction of the orifice member 5, the pressure on the lower side of the piston 9 overcomes the spring Il closing the valve 5 to the point where flow through this valve just supplies sufiicient volume to maintain a predetermined pressure in the chamber 4.
For example, using a thirty pound spring Il and a piston 9 of two square inches effective area, it will require a pressure of over fifteen pounds in the chamber 4 to overcome the spring and close valve 5 to shut on inflowing liquid. Assuming the control valve 5 to be open to such a position as to pass five gallons of liquid per minute at fifteen pounds per square inch, the pressure in chamber 4 will drop just sufllcient to allow the spring Il to open the valve 5 and in turn allow liquid to flow into the chamber 4 to maintain the pressure in that chamber at a point which will balance the pressure at fiff teen pounds per square inch and, of course, this in turn will maintain the desired pressure differential across the orice valve 5.
As the valve 6 is opened to obtain a higher speed the resulting expansion of the spring II for this greater liquid iiow will result in a slight Vlowering of the maintained pressure in the chamber 4 due to the change of rate of the spring. Citing another example, if there is a thousand pounds pressure above the piston 22 and five hundred pounds below, and the load should suddenly be released, the piston 9 would move an almost infinitesimal amount because the oil in the control valve 2 is practically incompressible; in other words, a change of only one pound pres-- sure will operate the piston 9 so that even if there would be five thousand pounds pressure in the line the pressure in the chamber 4 would still be constant because normally the effective diameters 5 and l0 of the balanced valve are equal.
On the other hand, it is often desirable to obtain an increase of speed in a hydraulic motor coincident with an increase in load and to accomplish this, referring particularly to Figs. 2 and 4, a sleeve 26 or a valve 5a is selected with a diameter smaller than the diameter of the stem I0. With the smaller area on 5 or 5a any drop in the pressure in chamber I will reduce the reaction against the spring Il and result in a proportionate increase in pressure in 4; in other words, with the control valve balanced the presany unequalness in the sure in 4 will always remain constant regardless of conditions in the system, b ut with the valve unbalanced the pressure in 4 will be proportional to the unbalancing pressure. More specifically, areas 5 and l0 will add to or subtract from the effective spring load on the piston and will effect a proportional change in the pressure in chamber 4. An increase or decrease of the pressure in chamber 4- increases or decreases the flow through the valve 8. It will therefore be obvious that with areas 5 andv Il unbalanced a change in pressure in chamber I causes a proportional change in chamber 4 which is either in direct or inverse relation to -the change in 3 depending upon whether the effective area of 5 is larger or smaller than Il. The main use for this feature is4 duel to the fact that all hydraulic motors have some slip or leakage and an increase in load causes an increase in slip due to the increased pressure differential across the piston and it is therefore possible to select a speed control valve 5 so that a change in load will cause no change in piston speed.
It will be understood that the limited line of contact such as shown in Fig. 4, will be more efficient in the operation of the valve and the maintaining of the contact pressure in the chamber 4. This is because the seat of the sleeve 26a has a fixed taper. It is easy to fabricate discs 5a of various diameters. The sharp edge of the disc 5a contacting with the tapered seat on a line contact makes for an accurate eiiicient seal; in changing the valve area it is only necessary to change discs 5a. The Valve shown in Fig. 2 constitutes merely a reversal of parts, and is less desirable mechanically because of the difficulty in matching two tapered surfaces and because both sleeve 26 and valve 5 must be reversed in changing the setting. There is no difference in the functions or operations of either valve assembly and either may be used equally well in the control units shown in Figs. l, 5 and 6.
It will be understood here that where the word maintaining is used in the specification and claims I am referring to substantial maintenance of pressure or pressure differential, as it will be understood that the speed of some types of machines or units may vary somewhat and still be considered a uniform speed within the meaning of the present invention,
What I claim is:
l. A liquid controlling system comprising in combination a liquid operated motor, a source of liquid under pressure and aspeed control device all operatively connected, said device including a control valve having opposed pressure surfaces and a regulable orifice member, said member and valve being operatively connected through a pressure chamber, the pressure in said chamber controlling the flow through the orifice member to control the speed of the motor, said opposed pressure surfaces being unbalanced to maintain a pressure in the pressure chamber in proportion to the unbalancing of said pressure surfaces whereby to increase or decrease the speed of the motor upon a variation in load.
2. In combination, a hydraulic motor, a source of liquid supply operatively connected with the motor, a speed control unit therefor embodying a chamber for receiving the liquid supply, a variable discharge orifice member, a chamber on the intake side of said orifice member, a valve between said two chambers having opposed pressure surfaces, elements forming the effective size of one of said pressure surfaces, one of said elements being replaceable to vary the size of said one surface, and a spring pressed piston operatively connected with said valve for controlling the pressure in said second named chamber to control the pressure differential across the orifice member.
3. In combination, a hydraulic motor, a speed control unit therefor including a chamber for receiving the normal liquid flow from the motor, a secondchamber leading to a discharge orifice member of predetermined size, a piston valve having opposed pressure surfaces for controlling the flow between said two chambers, a piston connected to said valve, a conduit for connecting the second chamber with one side of the piston, one of said opposed pressure surfaces being of different relative effective diameter to vary the pressure in the second chamber in proportion with a change in load in the motor.
4. In combination, a hydraulic motor having a piston and cylinder, a hydraulic circuit including a pump directly connected with said motor, a relief valve in said'circuit for determining the maximum. preloading of one side of said piston, a directional valve for the motor, and a speed control unit comprising a regulable orifice valve, a pressure chamber on one side of said orifice valve and a discharge conduit leading from the unit on the other side, a spring pressed piston having one face exposed to the pressure in said chamber and the other face exposed to the pressure in said discharge conduit, a second chamber for receiving liquid from the motor, the flow of which is to be controlled, and ar piston valve formed as a part of said piston and extending into said second chamber, said piston valve having opposed pressure surfaces with different effective areas exposed to the pressure in said second chamber, the relative size of said different effective areas being predetermined to insure a predetermined relationship between the pressure in the first chamber and any change in pressure in the second chamber, and a passageway between said two chambers forming a valve seat for one of said pressure surfaces, said spring tending to separate said pressure surface and valve seat.
5. In combination, a hydraulic motor having a piston and cylinder, a hydraulic circuit including a. pump directly connect-ed with said motor, a directional valve for the motor, and a speed control unit comprising a regulable orifice valve, a pressure chamber on one side of said orice valve and a discharge conduit leading from the unit on the other side, a spring pressed piston having one face exposed to the pressure in said chamber and the other face exposed to the pressure in said discharge conduit, a second chamber for receiving liquid from the motor, the flow of which is to be controlled, and a piston valve formed as a part of said piston and extending into said second chamber, said piston valve having opposed pressure surfaces with different effective areas exposed to the pressure in said second chamber, the relative size of said different effective areas being predetermined to insure a predetermined relationship between the pressure in the first chamber and any change in pressure in the second chamber, and a passageway between said two chambers forming a valve seat for one of said pressure surfaces, said spring tending to separate said pressure surface and valve seat.
6. In combination, a hydraulic motor, a hy` draulic circuit including a pump directly connected with said motor, a directional valve for the motor, and a speed control unit comprising a regulable orifice valve, a pressure chamber on one side of said orifice valve and a discharge conduit leading from the unit on the other side, a spring pressed piston having one face exposed to the pressure in said chamber and the other face exposed to the pressure in said discharge conduit, a second chamber for vreceiving liquid from the motor, the ow of which is to be controlled, and a piston valve formed as a part of said piston and extending into said second chamber, said piston valve having opposed pressure surfaces with different effective areas, than the area of the faces of said piston, exposed to the pressure in said second chamber, the relative size of said different effective areas being predetermined to insure a predetermined relationship between the pressure in the rst chamber and any change in pressure in the second chamber, and a passageway between said two chambers forming a valve seat for one of said pressure surfaces, said spring tending to separate said pressure surface and valve seat.
7. A liquid controlling system comprising in combination a liquid operated motor, a source of liquid under pressure constantly preloading one side of said motor during a power stroke, and a speed control device all operatively connected, said device including a control valve having opposed pressure surfaces of different effective areas and a regulable orifice member, said member and valve being operatively connected through a pressure chamber, a spring pressed piston attached to said valve and subject to the pressure in said chamber, the pressure in said chamber controlling the flow through the orifice member to control the speed of the motor, the effective area of one of said pressure surfaces l being predetermined and so arranged relative to the effective area of the other pressure surface that any change in pressure in liquid leading from the motor due to change in load on the motor will predetermine the pressure in said chamber in accordance with said relative effective diameters.
8. A flow control unit, for use with a hydraulic motor circuit of the type having a piston and cylinder motor unit, and pump for applying positive pressure to one side of the piston during its entire working stroke, comprising a regulable orifice valve, a pressure chamber on one side of said valve and a discharge conduit leading from the other side, a spring pressed piston having one face exposed to the pressure in said chamber and the other face exposed to the pressure in said discharge conduit, a second chamber for receiving liquid from the other side of said motor piston, and a piston valve formed as an extension of said piston and extending into said second chamber, said piston valve having opposed pressure surfaces of appreciably smaller size than the effective area of said piston, said pressure surfaces being so proportioned relative toy each other as to predetermine the effect of the pressure in said second chamber upon the pressure in said first chamber, and elements cooperating to form the effective area of one of said surfaces, one of said elements being replaceable to change the relation between the effective areas of said opposed surfaces.
9. A ow control'unit, for use with a hydraulic motor circuit of the type having a piston and cylinder motor unit, and pump for applyingV positive pressure to one side of the piston during its entire Working stroke, comprising a regulable orifice valve, a pressure chamber on one side of said valve and a discharge conduit leading from the other side, a spring pressed piston having one face exposed to the pressure in said chamber and the other face exposed to the pressure in said discharge conduit, a. second chamber for receiving li/quid from the other side of said motor piston, and a piston valve formed as an extension of said piston and extending inhq said second chamber, and connecting the same with said rst chamber, said piston valve havmg opposed v pressure surfaces of appreciably smaller size than' the eiective area. of said piston. said pressure surfaces being so proportioned relative to each other as to predehexmine the effect o! the pressureinsaid second chamberuponthe pressure msadrstchambex-,andmeanstorvaryingthe relative effective area of one of said pressure surfaces.
` HARRY F. VICKERS.
US26052639 1931-06-12 1939-03-08 Hydraulically actuated member and speed control therefor Expired - Lifetime US2272684A (en)

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Cited By (44)

* Cited by examiner, † Cited by third party
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US2421219A (en) * 1942-03-04 1947-05-27 Landis Tool Co Balancing valve
US2462796A (en) * 1944-12-11 1949-02-22 Bendix Aviat Corp Regulator
US2487520A (en) * 1944-12-26 1949-11-08 Vickers Inc Hydraulic power transmission with bypass flow control valve
US2495785A (en) * 1945-01-11 1950-01-31 Hydraulic Equipment Company Lowering valve
US2501483A (en) * 1948-04-03 1950-03-21 Warner Swasey Co Hydraulic power system
US2526835A (en) * 1946-10-18 1950-10-24 Hpm Dev Corp Hydraulic motor control
US2536558A (en) * 1946-01-04 1951-01-02 Keelavite Co Ltd Pump and motor hydraulic transmission system
US2546579A (en) * 1946-08-13 1951-03-27 Denison Eng Co Fluid motor control mechanism
US2587449A (en) * 1945-02-10 1952-02-26 Farmingdale Corp Hydraulic feed for machine tools
US2603235A (en) * 1952-07-15 Kirkham
US2618932A (en) * 1949-09-09 1952-11-25 Vickers Inc Pump and motor hydraulic system, including multiple pumps
US2620679A (en) * 1949-04-13 1952-12-09 Morris Motors Ltd Power transmission for motor vehicles
US2655902A (en) * 1949-12-22 1953-10-20 Askania Regulator Co System for proportioning fluid flow to control signal magnitude
US2657906A (en) * 1947-09-18 1953-11-03 Buda Co Earth drill
US2752895A (en) * 1951-03-09 1956-07-03 Bendix Aviat Corp Hydraulic motor and control therefor
US2825358A (en) * 1954-03-01 1958-03-04 Oilgear Co Pressure regulator
US2833374A (en) * 1954-07-07 1958-05-06 Sidney P Glasser Constant flow lube system
US2855752A (en) * 1955-10-21 1958-10-14 Brusque Rene Le Hydraulic device for controlling the feed and stop position of a machine element in cutting, sawing and slicing machines
US2862523A (en) * 1955-04-18 1958-12-02 Dole Valve Co Thermostatic fluid restrictor
US2903007A (en) * 1950-04-28 1959-09-08 Gpe Controls Inc Compensation of back pressure variation in discharge type regulators
US2951501A (en) * 1954-09-20 1960-09-06 Separator Ab Regulating device for a flow medium
US2953152A (en) * 1956-08-17 1960-09-20 Thompson Ramo Wooldridge Inc Pressure regulating valve
US2982258A (en) * 1957-06-04 1961-05-02 United Aircraft Corp Pressure ratio device utilizing a free piston valve for pressure ratio regulation and a servo mechanism coacting therewith to amplify pressure ratio error correction
US3015350A (en) * 1957-10-14 1962-01-02 Swift & Co Bacon slicer having adjustable control of group size
US3046950A (en) * 1958-01-22 1962-07-31 Whiting Corp Constant mechanical advantage rotary hydraulic device
US3088688A (en) * 1958-09-25 1963-05-07 H G Weber And Company Inc Hydraulic system
US3115892A (en) * 1954-10-11 1963-12-31 Fischer & Porter Co Flow controller
US3119306A (en) * 1960-08-01 1964-01-28 Onsrud Machine Works Inc Contouring and profiling machines
US3132485A (en) * 1961-03-31 1964-05-12 Blackhawk Mfg Co Hydraulic motor control
US3212525A (en) * 1962-10-18 1965-10-19 Henderson Hallie Valves for refrigeration apparatus having cooling and/or heating cycles
US3437012A (en) * 1965-12-28 1969-04-08 Asea Ab Valve system for hydraulic elevators
US3596677A (en) * 1969-01-13 1971-08-03 Rex Chainbelt Inc Remotely operable pressure compensated flow control valve
US3678952A (en) * 1968-10-22 1972-07-25 Honda Motor Co Ltd Pressure fluid circuit in automatic transmission apparatus
US3795260A (en) * 1972-09-27 1974-03-05 G Bergson Three way valve for flow regulator connected to moisture analyzer
US4147179A (en) * 1976-02-24 1979-04-03 Shoketsu Kinzoku Kogyo Co., Ltd. Pressure governor valve equipped with flow control valve
US4175473A (en) * 1976-06-08 1979-11-27 Shoketsu Kinzoku Kogyo Kabushiki Kaisha Fluid circuit
DE2826613A1 (en) * 1978-06-19 1979-12-20 Werner & Pfleiderer Flow control valve for pressurised liq. - has spring-loaded pressure balancing piston between pressures up and downstream of throttle
US4271864A (en) * 1980-03-31 1981-06-09 Mac Valves, Inc. Pressure regulating valve
US4420014A (en) * 1980-04-21 1983-12-13 The Bendix Corporation Pressure regulator for a fluid motor
US4779419A (en) * 1985-11-12 1988-10-25 Caterpillar Inc. Adjustable flow limiting pressure compensated flow control
US4811649A (en) * 1987-02-18 1989-03-14 Heilmeier & Weinlein, Fabrik Fur Oelhydraulik Gmbh & Co. Kg Hydraulic control apparatus
US5107886A (en) * 1991-02-15 1992-04-28 Taylor Julian S Constant flow orifice valve
US5282490A (en) * 1989-12-18 1994-02-01 Higgs Robert E Flow metering injection controller
US5460199A (en) * 1992-07-13 1995-10-24 Sumitomo Electric Industries, Ltd. Flow control valve and control method therefor

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* Cited by examiner, † Cited by third party
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DE1151713B (en) * 1954-09-11 1963-07-18 Klopp Werke G M B H Control device for planing machines, especially high-speed planers

Cited By (46)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2603235A (en) * 1952-07-15 Kirkham
US2421219A (en) * 1942-03-04 1947-05-27 Landis Tool Co Balancing valve
US2462796A (en) * 1944-12-11 1949-02-22 Bendix Aviat Corp Regulator
US2487520A (en) * 1944-12-26 1949-11-08 Vickers Inc Hydraulic power transmission with bypass flow control valve
US2495785A (en) * 1945-01-11 1950-01-31 Hydraulic Equipment Company Lowering valve
US2587449A (en) * 1945-02-10 1952-02-26 Farmingdale Corp Hydraulic feed for machine tools
US2536558A (en) * 1946-01-04 1951-01-02 Keelavite Co Ltd Pump and motor hydraulic transmission system
US2546579A (en) * 1946-08-13 1951-03-27 Denison Eng Co Fluid motor control mechanism
US2526835A (en) * 1946-10-18 1950-10-24 Hpm Dev Corp Hydraulic motor control
US2657906A (en) * 1947-09-18 1953-11-03 Buda Co Earth drill
US2501483A (en) * 1948-04-03 1950-03-21 Warner Swasey Co Hydraulic power system
US2620679A (en) * 1949-04-13 1952-12-09 Morris Motors Ltd Power transmission for motor vehicles
US2618932A (en) * 1949-09-09 1952-11-25 Vickers Inc Pump and motor hydraulic system, including multiple pumps
US2655902A (en) * 1949-12-22 1953-10-20 Askania Regulator Co System for proportioning fluid flow to control signal magnitude
US2903007A (en) * 1950-04-28 1959-09-08 Gpe Controls Inc Compensation of back pressure variation in discharge type regulators
US2752895A (en) * 1951-03-09 1956-07-03 Bendix Aviat Corp Hydraulic motor and control therefor
US2825358A (en) * 1954-03-01 1958-03-04 Oilgear Co Pressure regulator
US2833374A (en) * 1954-07-07 1958-05-06 Sidney P Glasser Constant flow lube system
US2951501A (en) * 1954-09-20 1960-09-06 Separator Ab Regulating device for a flow medium
US3115892A (en) * 1954-10-11 1963-12-31 Fischer & Porter Co Flow controller
US2862523A (en) * 1955-04-18 1958-12-02 Dole Valve Co Thermostatic fluid restrictor
US2855752A (en) * 1955-10-21 1958-10-14 Brusque Rene Le Hydraulic device for controlling the feed and stop position of a machine element in cutting, sawing and slicing machines
US2953152A (en) * 1956-08-17 1960-09-20 Thompson Ramo Wooldridge Inc Pressure regulating valve
US2982258A (en) * 1957-06-04 1961-05-02 United Aircraft Corp Pressure ratio device utilizing a free piston valve for pressure ratio regulation and a servo mechanism coacting therewith to amplify pressure ratio error correction
US3015350A (en) * 1957-10-14 1962-01-02 Swift & Co Bacon slicer having adjustable control of group size
US3046950A (en) * 1958-01-22 1962-07-31 Whiting Corp Constant mechanical advantage rotary hydraulic device
US3088688A (en) * 1958-09-25 1963-05-07 H G Weber And Company Inc Hydraulic system
US3119306A (en) * 1960-08-01 1964-01-28 Onsrud Machine Works Inc Contouring and profiling machines
US3132485A (en) * 1961-03-31 1964-05-12 Blackhawk Mfg Co Hydraulic motor control
US3212525A (en) * 1962-10-18 1965-10-19 Henderson Hallie Valves for refrigeration apparatus having cooling and/or heating cycles
US3437012A (en) * 1965-12-28 1969-04-08 Asea Ab Valve system for hydraulic elevators
US3678952A (en) * 1968-10-22 1972-07-25 Honda Motor Co Ltd Pressure fluid circuit in automatic transmission apparatus
US3596677A (en) * 1969-01-13 1971-08-03 Rex Chainbelt Inc Remotely operable pressure compensated flow control valve
US3795260A (en) * 1972-09-27 1974-03-05 G Bergson Three way valve for flow regulator connected to moisture analyzer
US4147179A (en) * 1976-02-24 1979-04-03 Shoketsu Kinzoku Kogyo Co., Ltd. Pressure governor valve equipped with flow control valve
US4175473A (en) * 1976-06-08 1979-11-27 Shoketsu Kinzoku Kogyo Kabushiki Kaisha Fluid circuit
DE2826613A1 (en) * 1978-06-19 1979-12-20 Werner & Pfleiderer Flow control valve for pressurised liq. - has spring-loaded pressure balancing piston between pressures up and downstream of throttle
US4271864A (en) * 1980-03-31 1981-06-09 Mac Valves, Inc. Pressure regulating valve
US4420014A (en) * 1980-04-21 1983-12-13 The Bendix Corporation Pressure regulator for a fluid motor
US4779419A (en) * 1985-11-12 1988-10-25 Caterpillar Inc. Adjustable flow limiting pressure compensated flow control
US4811649A (en) * 1987-02-18 1989-03-14 Heilmeier & Weinlein, Fabrik Fur Oelhydraulik Gmbh & Co. Kg Hydraulic control apparatus
US5282490A (en) * 1989-12-18 1994-02-01 Higgs Robert E Flow metering injection controller
US5427149A (en) * 1989-12-18 1995-06-27 Higgs; Robert E. Flow metering injection controller
US5107886A (en) * 1991-02-15 1992-04-28 Taylor Julian S Constant flow orifice valve
US5460199A (en) * 1992-07-13 1995-10-24 Sumitomo Electric Industries, Ltd. Flow control valve and control method therefor
US5524659A (en) * 1992-07-13 1996-06-11 Sumitomo Electric Industries, Ltd. Flow control valve and control method therefor

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