Pneumatic reciprocating motor

A gas-actuated reciprocal drive apparatus has a double-acting piston in a pneumatic cylinder having a chamber at each end. Gas from an area of higher pressure in a compressed gas system flows into a first chamber, while the second chamber is in fluid communication with an area of lower pressure in the gas system. The piston moves toward the second chamber, purging gas therein back to the lower-pressure area in the gas system, without any venting to the atmosphere. A four-way gas valve reverses the piston motion after each stroke, by reversing the chambers' gas connections. The piston has a pair of circumferential seals, plus a differential shuttle valve that allows gas from the lower-pressure chamber to enter the annular space between the seals, such that the pressure differential across the seals always equals the pressure differential between the two chambers, regardless of the actual pressures in the chambers, thus reducing friction forces on the piston seals, increasing the power output of the apparatus, and extending the service life of the seals.

FIELD OF THE INVENTION

The present invention relates to reciprocating drive apparatus actuated by a pressurized gas, and in particular to reciprocating drive apparatus that is actuated by a pressurized gas without exhausting the actuating gas to the atmosphere.

BACKGROUND OF THE INVENTION

In natural gas production facilities, it is often necessary or desirable to periodically or continuously inject liquids into a high pressure gas pipeline. One example is the injection of methanol to prevent any water present in the natural gas from freezing. Such liquids are injected by means of pumps which overcome the pressure of the compressed gas to force the liquid into the pipeline. These injection pumps are often powered by pneumatic devices, particularly in remote locations. In some situations, the compressed gas flowing in the pipeline is used to drive the pump, but usually only after it has been regulated down to a pressure suitable for the pneumatic device (often around 10 pounds per square inch). The exhaust gas from the pneumatic device comes out of the device at a lower pressure than the gas in the pipeline, so it cannot be reinjected into the pipeline unless it is first compressed. Therefore, the exhaust gas is usually vented to atmosphere. In some situations a gas such as propane is brought to the site, stored in a pressure vessel, and used to drive a pneumatic device. This gas is also vented to atmosphere from the pneumatic device.

This venting of the exhaust gas to the atmosphere is a problem, firstly because it is a waste of valuable gas, secondly because it causes environmental contamination. In the case of sour gas wells (i.e., wells producing natural gas with high hydrogen sulphide content), it is generally prohibited, on environmental and health grounds, to use drive apparatus actuated by well gas where the exhaust gas is vented to atmosphere. Accordingly, there is a need for drive apparatus for driving injection pumps and other equipment associated with natural gas wells, using raw gas from the well to actuate the apparatus, but without venting the actuating gas to the atmosphere.

U.S. Pat. No. 6,336,389, issued Jan. 8, 2002 to English et al., discloses one example of prior art apparatus directed to this objective, mobilizing the kinetic energy inherent in the differential pressure between areas of higher and lower pressure in a pressurized gas system such as a pipeline. The English apparatus uses a single-acting piston that reciprocates within an open-ended cylinder inside a pressure vessel, where the interior of the pressure vessel is in fluid communication with the area of lower pressure, such that the bottom end of the piston is always exposed to the lower pressure. A switching valve allows gas from the area of higher pressure to flow into the chamber at the closed end cylinder, thus inducing a pressure differential between the two ends of the piston, causing the piston to move in a downward or power stroke. Linkage mechanism is provided for transferring the energy from the power stroke to an oscillatingly rotating output shaft, which is then connected to an injection pump or other type of equipment to be driven.

At or near the end of the downward stroke, the switching valve opens the piston chamber to the interior of the pressure vessel and closes off flow or higher pressure gas into the chamber, thus equalizing the pressure on each end of the piston. Biasing means such as a spring then moves the piston back to the top of the piston, thus exhausting the gas in the piston chamber into the pressure vessel and, effectively, into the area of lower pressure within the pressurized gas system. At or near the end of this exhaust stroke, the switching valve closes off the piston chamber from the interior of the pressure vessel and opens the chamber once again to the flow of gas from the area of higher pressure, thus readying the apparatus for the next downward power stroke.

The English apparatus effectively provides means for gas-driven actuation of injection pumps or other equipment without venting of the actuating gas. The English apparatus can operate with pressure differentials as low as 25 psi, so the internal mechanisms of the apparatus are not exposed to high pressures, even though the pressure in the gas system that drives it may be 1,000 psi or higher. However, the output of this apparatus is limited to an oscillating rotary drive. Commonly-used chemical injection pumps, on the other hand, require a reciprocating drive. Accordingly, the use of the English apparatus to drive a reciprocating-drive pump entails some kind of motion-converting mechanism to convert the oscillating rotary output motion to a reciprocating motion. This adds to the overall cost and mechanical complexity of the apparatus used to drive the pump, and reduces the overall mechanical efficiency of the apparatus.

Since the English apparatus uses a single-acting piston, and thus produces power only on half of the piston strokes, its mechanical efficiency is less than would be the case for apparatus using a double-acting piston and producing power on each piston stroke. An additional drawback of the English apparatus is that the spring or other biasing means (for returning the piston to the top of the cylinder after each power stroke) must be compressed during each power stroke, thus consuming part of the energy inherent in the pressure differential and thereby reducing the power output of the apparatus.

U.S. Pat. No. 6,694,858, issued Feb. 24, 2004 to Grimes, discloses a gas-driven reciprocating drive unit that uses a double-acting piston within a closed cylinder, in association with a pressurized gas system such as a gas pipeline. A switching valve directs gas from area of higher and lower pressure to opposite sides of the piston. The pressure differential between the two ends of the double-acting piston causes the piston to move toward a first end of the cylinder, simultaneously exhausting the gas in the first end of the cylinder back into the pressurized gas system. A drive link connected to the piston is used to transfer the power generated by the movement of the piston to a pump or other piece of equipment. At or near the end of each piston stroke, the switching valve reverses the connections to the areas of higher and lower pressure in the pressurized gas system, thus inducing a pressure differential that causes the piston to move in the direction opposite to the previous stroke and thereby exhausting the gas in the second end of the cylinder back into the pressurized gas system.

One of the significant drawbacks and disadvantages of the Grimes apparatus is the susceptibility of the piston seals to wear and deterioration. In order to maintain a pressure differential between the ends of the cylinder, the double-acting piston requires circumferential seals of some suitable type to prevent the flow of gas between the two ends of the cylinder via the annular space between the piston and cylinder. The ambient pressure within the annular space between the seals is constant, and typically atmospheric (i.e., approximately 15 psi). In contrast, the gas pressure within each end of the cylinder may be 1,000 psi or greater. As a result (and unlike the piston seals in the English apparatus), both of the seals in the Grimes apparatus are continuously working against a very large pressure differential, notwithstanding the fact that the piston itself is exposed to only a small pressure differential. The high differential pressure acting across the seals induces proportionately higher friction forces at the cylinder interface. These friction forces must be overcome in order for the piston move, and the power required to do this directly reduces the available power output from the apparatus. If the friction forces become too high, the piston may be susceptible to seizing or stalling (“stiction”). In addition, the high friction forces promote wear on the seals, thus making seal replacement necessary more often than would be the case in absence of high differential pressures across the seals.

For the foregoing reasons, there remains a need for reciprocating drive apparatus that not only may be actuated by raw pressurized gas from a natural gas well without venting the actuating gas to the atmosphere, but that also provides a direct reciprocating final drive output without need for motion-converting mechanisms. There is a further need for reciprocating pneumatic drive apparatus in which the seals between the piston and cylinder of the apparatus are exposed to a low pressure differential, therefore being less susceptible friction-induced power output losses, and less susceptible to wear and deterioration, than in prior art pneumatic drive apparatus. The present invention is directed to these needs.

BRIEF SUMMARY OF THE INVENTION

In general terms, the present invention is a closed-loop, gas-actuated reciprocal drive apparatus that utilizes the potential energy inherent in the pressure differential between an area of higher pressure and an area of lower pressure in a compressed gas system, such as a natural gas pipe line, to enable the pressurized gas to actuate the apparatus while exhausting the actuating gas back into the compressed gas system, without exhausting the actuating gas to atmosphere. The apparatus converts the potential energy from the pressure differential into linear reciprocating motion, using a double-acting, double-rod piston moving within a pneumatic cylinder. The cylinder defines a pneumatic chamber at each end, with the linear length of the chamber varying as the piston moves within the cylinder. Operation of the apparatus is initiated by allowing gas from an area of higher pressure to flow into one chamber, while the other chamber is in fluid communication with an area of lower pressure. This induces a pressure differential that causes the piston to move toward the lower-pressure chamber, and at the same time purging the gas from that chamber. A four-way, two-position gas valve is used in conjunction with an angular incremental switch mechanism to reverse the motion of the piston at the end of each stroke, by reversing the connections of the chambers to the areas of higher and lower pressure in the gas system.

Each end of the piston has a piston rod reciprocatingly extending through a corresponding end the cylinder, for providing linear drive force to a plunger pump or piston pump (or other devices). The apparatus is thus capable of driving two pumps at the same time. Moreover, the apparatus is capable of doing so in conditions where the differential between the areas of higher and lower pressure is as low as 10 psi.

The gas used to actuate the apparatus is always returned to the pressurized gas system from which it was supplied. Accordingly, the apparatus is a fully-closed system that vents no gas to atmosphere, and therefore is readily usable in conjunction with sour gas wells.

The piston has a circumferential piston seal near each end, and further incorporates a differential shuttle valve that allows gas from the low-pressure chamber of the cylinder to enter the annular space between the seals. The pressure differential across the seals is thus equal to the differential between the two chambers of the cylinder, regardless of the magnitude of the gas pressures in the chambers. As a result, the friction forces between the piston seals and the cylinder walls remain substantially constant, and of substantially lesser magnitude than in prior art apparatus having double-acting cylinders, thereby increasing the power output of the apparatus and extending the service life of the seals.

Accordingly, in one aspect the present invention is a reciprocating pneumatic drive apparatus for use in association with a compressed gas system having an area of higher pressure and an area of lower pressure, said apparatus comprising:(a) a cylinder having a cylindrical sidewall extending between a pair of cylinder heads, each of which has a piston rod opening;(b) a piston having first and second piston faces plus first and second piston rods, each projecting from a corresponding piston face, said piston being reciprocatingly slidable within the cylinder, with each piston rod being sealingly slidable through the piston rod opening of a corresponding one of the cylinder heads, said piston demarcating first and second variable-length cylinder chambers, one at each end of the cylinder;(c) a pair of spaced-apart piston seals disposed circumferentially around the piston, for sealing between the piston and the sidewall, said piston seals defining the ends of an annular space;(d) valve means operable between a first position in which the first and second cylinder chambers are in fluid communication with the areas of higher and lower pressure respectively, and a second position in which the first and second cylinder chambers are in fluid communication with the areas of lower and higher pressure respectively, so as to induce reciprocating movement of the piston within the cylinder; and(e) switch means operable to switch the position of the gas flow valve at or near the end of each stroke of the piston;
wherein:(f) the piston has a transverse passage extending between the piston faces, and a radial passage extending between the transverse passage and the annular space; and(g) the apparatus further comprises shuttle valve means retainingly disposed within the transverse passage, for enabling gas from whichever cylinder chamber is under lower pressure to flow through the transverse and radial passages into the annular space, while preventing the flow of gas from the cylinder chamber under higher pressure into the transverse passage.

In a second aspect, the invention is a reciprocating pneumatic drive apparatus for use in association with a compressed gas system having an area of higher pressure and an area of lower pressure, said apparatus comprising:(a) a cylinder having a cylindrical sidewall and first and second cylinder heads, each cylinder head having a piston rod opening;(b) a piston reciprocatingly disposed within the cylinder, said piston having first and second piston faces, and having a circumferential side face extending between said first and second piston faces;(c) a first cylinder chamber defined by said sidewall, first cylinder head, and first piston face, the size of said first cylinder chamber varying according to the position of the piston within the cylinder;(d) a second cylinder chamber defined by said sidewall, second cylinder head, and second piston face;(e) a first piston rod rigidly fixed to the piston and extending from the first piston face, and being reciprocatingly and sealingly movable through the piston rod opening of the first cylinder head;(f) a second piston rod rigidly fixed to the piston and extending from the second piston face, and being reciprocatingly and sealingly movable through the piston rod opening of the second cylinder head;(g) first piston sealing means, for sealing between the sidewall and the side face of the piston, adjacent to the first piston face;(h) second piston sealing means, for sealing between the sidewall and the side face of the piston, adjacent to the second piston face;(i) first cylinder head port, in fluid communication with the first cylinder chamber;(j) second cylinder head port, in fluid communication with the second cylinder chamber;(k) a gas flow control valve alternatingly operable between a first position in which the first and second cylinder head ports are in fluid communication with the areas of higher and lower pressure respectively, and a second position in which the first and second cylinder head ports are in fluid communication with the areas of lower and higher pressure respectively, so as to induce reciprocating movement of the piston within the cylinder; and(l) switch means operable to switch the position of the gas flow valve at or near the end of each stroke of the piston;
wherein:(m) the cylinder sidewall, the piston side face, and the first and second piston sealing means define an annular space;(n) the piston has a transverse passage extending between the piston faces, and a radial passage extending between the transverse passage and the annular space; and(o) the apparatus further comprises shuttle valve means retainingly disposed within the transverse passage, for enabling gas from whichever cylinder chamber is under lower pressure to flow through the transverse and radial passages into the annular space, while preventing the flow of gas from the cylinder chamber under higher pressure into the transverse passage.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

Referring in particular toFIGS. 1A,1B,2, and3, the pneumatic motor of the present invention (generally designated by reference number10), comprises a pneumatic cylinder20and a double-acting piston30that is reciprocatingly and coaxially movable within pneumatic cylinder20. The pneumatic cylinder20has a cylindrical inner wall22and is capped at each end by cylinder heads24A and24B. The piston30has circular piston faces32A and32B and a circumferential side surface34extending between piston faces32A and32B. A piston rod36, having ends36A and36B, is rigidly and coaxially fixed to piston30, with rod ends36A and36B extending through rod openings26A and26B in cylinder heads24A and24B respectively. Piston rod seals26C are provided in association with rod openings26A and26B such that piston rod36is reciprocatingly movable through rod openings26A and26B in substantially pressure-tight fashion. In the preferred embodiment, piston rod seals26C are dynamic seals similar to the piston seals38described elsewhere in this specification.

Pneumatic cylinder20defines an annular cylinder chamber28A bounded by cylinder wall22, cylinder head24A, and piston face32A, and an annular cylinder chamber28B bounded by cylinder wall22, cylinder head24B, and piston face32B. The length and volume of cylinder chambers28A and28B varying according to the position of piston30within cylinder20. For purposes to be explained further herein, cylinder head24A has cylinder head gas port25A in fluid communication with cylinder chamber28A, and cylinder head24B has cylinder head gas port25B in fluid communication with cylinder chamber28B.

As particularly illustrated inFIGS. 3 and 3A, piston30is provided with two circumferential piston seals38, each disposed in a circumferential chase39formed into side surface34of piston30near one end of piston30. Piston seals38are at all times sealingly engaged against cylinder wall22, so as to substantially prevent leakage of gas from either of the cylinder chambers28A and28B. In the preferred embodiment, as shown inFIGS. 3 and 3A, piston seals38are dynamic seals that include a core element made from an elastic material and formed with a “horseshoe” cross-section, such that they need to be radially compressed for insertion into their respective chases39. The elastic energy or spring force thus induced in the piston seals38biases them radially outward and into contact with cylinder wall22. As conceptually illustrated inFIG. 3A, this outward biasing force manifests as a normal force Fnacting against cylinder wall22. The friction force Ffrequired to overcome normal force Fn(in order for piston30to move) is directly proportional to normal force Fn. Accordingly, piston seals38are ideally designed or selected so as to induce a normal force Fnthat is as low as possible in order to minimize friction force Ff, while being high enough to ensure a vapor-tight seal against cylinder wall22.

Referring toFIGS. 2 and 3, piston30incorporates a shuttle valve40whereby pressurized gas can be introduced into the annular space29radially bounded by piston30and cylinder20, and longitudinally bounded by piston seals38. A transverse passage41extends through piston30at a selected location, with said passage41configured to include a central bore41A and a concentric and larger diameter recess41B adjacent to each of piston faces32A and32B, such that an annular shoulder42is formed between central bore41A and each recess41B. A radial passage43extends through piston30between central bore41A and annular space29. Shuttle valve40includes a shuttle member44with cap members45at each end, with the clear distance between the cap members45being greater than the length of central bore41A between recesses41B. Each cap member has an outer face45A and an inner face45B and an annular groove45C is formed in each inner face45B for receiving an O-ring46or similar sealing member.

The cross-sectional geometry of shuttle member44is configured such that shuttle member44can slide freely within central bore41A but with fairly close tolerances so that it slides substantially coaxially within central bore41A, while at the same time defining at least one longitudinal channel between shuttle member44and the walls of central bore41A. In one embodiment, this feature is provided by forming shuttle member44from initially round stock into which one or more longitudinal flattened surfaces are formed. This creates one or more longitudinal channels47which in cross section resemble a circular segment. This and alternative embodiments of the shuttle member44are illustrated inFIGS. 3B through 3E(described in further detail below).

As shown inFIG. 3, shuttle valve40is assembled with shuttle member44disposed within central bore41A, and with each cap member45disposed within a corresponding recess41B. Accordingly, each O-ring46directly faces and is substantially parallel to a corresponding shoulder42. Because the length of shuttle member44is greater than the length of central bore41A, one of cap members45will always be separated slightly away from their corresponding shoulders42. When one cap member45is separated from its corresponding shoulder42(such as the lefthand cap member45inFIG. 3), a pathway is created whereby gas present in cylinder chamber28A can pass around cap member45, through longitudinal channel(s)47, through radial passage43, and into annular space29. As may be seen fromFIG. 3, if the righthand cap member45is being pressed against its corresponding shoulder42, the corresponding O-ring46will seal the righthand cap member45against its corresponding shoulder42, thus preventing any flow of gas between cylinder chamber28B and the shuttle valve40.

FIGS. 3B and 3Cillustrate one alternative construction of the shuttle valve40. As shown inFIG. 3B, each cap member45has a threaded stem45D that is matingly engageable with threaded bore44A of shuttle member44. To assemble the shuttle valve40, one cap member45is screwed into one end of shuttle member44, and this subassembly is inserted into central bore41A of piston30. The other cap member45may then be screwed into the other end of shuttle member44.

As shown inFIG. 3C, shuttle member44is made from round stock that has been milled flat on four sides44B, leaving four longitudinal surfaces44C which retain the radius of the round stock. The radius of the round stock is slightly less than the radius of central bore41A, such that shuttle member44can slide freely within central bore41A but without significant “play”. When the shuttle valve40is installed in central bore41A of piston30, the space between the surface of central bore41A and each flattened side surface44B forms a longitudinal channel47. Centrally-located portions of the longitudinal surfaces44C of shuttle member44are milled to create recessed areas44D that permit fluid communication between adjacent longitudinal channels47. The length of the recessed areas44D is such that at least a portion of the length will coincide with the opening from central bore41A into radial passage43regardless of the position of shuttle valve40within central bore41A. This arrangement ensures that gas flowing into the longitudinal channels47from cylinder chamber28A or cylinder chamber28B will pass through longitudinal channels47into radial passage43and thence into annular space29.

FIG. 3Dillustrates an alternative construction of shuttle valve40largely similar to that shown inFIGS. 3,3B, and3C except that cap members45are of frustoconical configuration and recesses41B are correspondingly shaped.FIG. 3Eillustrates an alternative construction of shuttle valve40having frustoconical cap members45as inFIG. 3Dbut with a differently-configured shuttle member44. As conceptually indicated, the stems45C of cap members45are internally threaded and mate with externally-threaded ends of shuttle member44. The frustoconical cap members45are effectively self-centering within central bore41A, so the diameter of stems45C can be sufficiently smaller than that of central bore41A so as to form a substantially longitudinal channel47therebetween. The diameter of shuttle member44is less than that of stems45C, so as to form an annular recessed area44D. Alternatively, stems45C may be fabricated with flattened surfaces similar to the flattened side surface44B of the shuttle member44inFIG. 3C, with corresponding longitudinal surfaces44C, such that stems45C can slide freely but without play within central bore41A.

It can be readily seen that if the gas pressure in cylinder chamber28B exceeds the gas pressure in cylinder chamber28A, the shuttle valve assembly40will move to the left, into the position shown inFIG. 3, with gas free to flow from cylinder chamber28A to annular space29as described above. If the pressure in cylinder chamber28A is then made to exceed the gas pressure in cylinder chamber28B, the shuttle valve assembly40will move to the right, sealing the lefthand cap member45against its corresponding shoulder42and preventing any flow of gas between cylinder chamber28A and the shuttle valve40, while at the same time allowing gas to flow from cylinder chamber28B to annular space29.

Other configurations of shuttle valve40, functioning substantially as described above, may be devised without departing from the principles and scope of the present invention.

The pneumatic motor10also includes a multi-position gas valve50having valve ports52A,52B,52C, and52D. By means of suitable conduits, valve port52A is in fluid communication with cylinder head port25A and valve port52B is in fluid communication with cylinder head port25B. Valve port52C is in fluid communication with an area HP in a pressurized gas system (such as a gas pipeline), and valve port52D is in fluid communication with an area LP in the gas system, said area LP being at a pressure lower than area HP. Gas valve50is operable between:a first position in which valve ports52A and52C are in fluid communication, putting cylinder chamber28A in fluid communication with area HP, while valve ports52B and52D are in fluid communication, putting cylinder chamber28B in fluid communication with area LP; anda second position in which valve ports52A and52D are in fluid communication, putting cylinder chamber28A in fluid communication with area LP, while valve ports52B and52C are in fluid communication, putting cylinder chamber28B in fluid communication with area HP.

FIGS. 4A to 4Eillustrate a multi-position gas valve50in accordance with a preferred embodiment of the present invention. As best seen inFIGS. 4D and 4E, the gas valve50in this embodiment is a rotary valve having a cylindrical interior cavity54, with valve ports52A,52B,52C, and52D all in communication therewith. Cavity54is circumferentially bounded by cylindrical surface53. A rotor56is coaxially rotatable within cavity54about rotational axis A, and is geometrically configured such that particular valve ports will be in fluid communication, via segmental sub-cavities54A on either side of rotor56, when the valve50is in the first and second positions, as described above. Rotor56is fixed to valve shaft67so as to be coaxially rotatable about rotational axis A. Rotor56has rotor ends58that engage cylindrical surface55as rotor56cycles between operational positions, in substantially vapor-tight fashion such that there is no leakage of gas between segmental sub-cavities54A.

Preferably, the vapor-tight engagement of rotor ends58with cylindrical surface55is facilitated by use of a separate sealing means, an example of which is illustrated inFIGS. 4D and 4E. In the illustrated embodiment, a longitudinal slot62is formed in each rotor end58, and a resilient biasing means64is disposed along the base of each slot62. A selected pressure seal material66(such as, for instance, Teflon™ lamella) is then inserted into each slot62, with the dimensions of the pressure seal66being such that it will project slightly beyond the face of rotor end58when not subject to compressive force urging it radially into slot62. Thus, when rotor58is positioned within cavity54, pressure seal66will at all times be in contact with cylindrical surface55, with resilient biasing means64constantly urging pressure seal66radially outward against cylindrical surface55.

InFIGS. 4D and 4E, rotor56is shown having straight or flat side portions, but this is not critical. The rotor56may have curvilinear or other geometric contours without substantively affecting the functioning of valve50, so long as the stated operational interrelation of valve ports52A,52B,52C, and52D is maintained when valve50is in the first and second operational positions.

Gas valve50is actuated between its first and second operational positions by means of a switch mechanism70which cycles the valve50at the end of each stroke of piston30and piston rod36. It will be readily apparent to persons skilled in the art of the invention that a variety of mechanisms could be devised to carry out the function of switch mechanism70in accordance with the operational mode described above.FIG. 7illustrates one example of such a mechanism, as used in a preferred embodiment of the invention. Switch mechanism70is disposed within switch housing71. A sleeve74is slidingly disposed around the portion of piston rod36B extending from cylinder20. Piston rod36B is reciprocatingly movable relative to switch housing71as piston30reciprocates within cylinder20. Suitable collars73A and73B are positioned at a desired spacing on either side of the sleeve74so as to limit the range of sliding movement of sleeve74on piston rod36B. A bracket74A fixed to sleeve74has a spring-retaining pin74B for receiving the first end of a tension spring76. A lever arm72is mounted at one end to valve shaft67, which projects into switch housing71. The other end of lever arm72has a spring-retaining pin72A which receives the second end of a tension spring76(shown in discontinuous fashion inFIG. 7for purposes of clarity) Lever arm bumpers78A and78B are mounted to switch housing71to limit the travel of lever arm72. Lever arm72is offset from piston rod36B so as not to impede its reciprocating movement.

The operation of switch mechanism70may be understood fromFIG. 7, in which sleeve74(shown cross-hatched for clarity) is at its leftmost limit of travel relative to piston rod36B. For purposes of illustration, valve50may be considered to be in its first position when the switch mechanism is as shown in solid outline inFIG. 7. As piston rod36B moves to the right (indicated by arrow R inFIG. 7), sleeve74will be pushed to the right as well by collar73A. The rightward movement of sleeve74causes tension spring76to stretch, but this initially has no effect on lever arm72, which remains in position against the left bumper78A. However, as the center of spring-retaining pin74B moves rightward past rotational axis A of valve shaft67, the tensile force in tension spring76, acting downward and to the right against spring-retaining pin72A, applies a clockwise moment on lever arm72, around rotational axis A. The magnitude of this moment increases as the rightward movement of sleeve74progresses, until it overcomes the resistant moment acting on valve shaft67(e.g., due to friction forces within the valve50). At that point, lever arm72will swing clockwise to the position shown in phantom outline. Since lever arm72is fixed to valve shaft67, this has the effect of switching valve50from its first position to its second position. Piston rod36B will reach the end of its rightward stroke soon after this happens; at this point, sleeve74will be abutting collar73B. Piston rod36B will then begin its leftward stroke, ultimately causing lever arm72will swing counterclockwise, thus switching valve50from the second position back to the first position.

The positions of collars73A and73B relative to piston rod36B may be adjusted so as to regulate the lag between the swing of lever arm72and the end of the piston rod stroke.

The operation of the pneumatic motor of the present invention may now be easily understood having reference toFIGS. 2,3,4D, and4E in particular. With gas valve50in the first position, higher-pressure gas from area HP flows into cylinder chamber28A while lower-pressure gas from area LP flows into cylinder chamber28B. The pressure differential between the two chambers causes piston30to move to the right, into the position shown inFIG. 2. This causes piston rod36to move in a rightward power stroke. At the same time, the pressure differential causes differential valve40to move to the right such that the lefthand cap member45(of shuttle valve40) and its associated O-ring46are urged against their corresponding shoulder42, while the righthand cap member45and its associated O-ring46are moved away from their corresponding shoulder42. In this configuration, gas is prevented from escaping from cylinder chamber28A into central bore41A of piston30, while gas is free to flow from cylinder chamber28B into annular space29, thus eliminating or greatly reducing the pressure differential across the piston seals38.

As piston30reaches or nears the end of its rightward power stroke, switching mechanism70cycles gas valve50to the second position. Now, higher-pressure gas from area HP flows into cylinder chamber28B while lower-pressure gas from area LP flows into cylinder chamber28A. The pressure differential between the two chambers causes piston30to move to the left, into the position shown inFIG. 3. This causes piston rod36to move in a lefttward power stroke. At the same time, the pressure differential causes differential valve40to move to the left such that the righthand cap member45and its associated O-ring46are urged against their corresponding shoulder42, while the lefthand cap member45and its associated O-ring46are moved away from their corresponding shoulder42. In this configuration, gas is prevented from escaping from cylinder chamber28B into central bore41A of piston30, while gas is free to flow from cylinder chamber28A into annular space29, thus once again eliminating or greatly reducing the pressure differential across the piston seals38. As piston30reaches or nears the end of its leftward power stroke, switching mechanism70cycles gas valve50back to the first position, and the alternating cycles continue as long as valve ports52C and52D remain in fluid communication with areas HP and LP respectively in a pressurized gas system.

The foregoing discussion has been in the context of a pneumatic motor using the rotary valve illustrated inFIGS. 4A to 4D. However, various other forms of gas valve50may be used without departing from the principles and scope of the present invention.FIGS. 5A to 5Eillustrate a second embodiment of gas valve50, which may be alternatively described as a planar valve. The valve body has valve ports52A,52B,52C, and52D as previously described in connection with the valve inFIGS. 4A to 4D. These ports are in fluid communication, respectively, with internal horizontal passages55A,55B,55C, and55D, which terminate at a common planar surface51. As best seen inFIG. 5C, a valve disc57, preferably made of Teflon™ (or an alternative material with good sealing and abrasion-resistance characteristics) is co-rotatably fixed to valve shaft67. Valve disc57interfaces tightly against planar surface51as shown, and is retained by retainer plate57D. Valve disc57has arcuate channels57A and57B, the configuration of which can best be seen inFIGS. 5D and 5E. Arcuate channels57A and57B, which extend only partly through the thickness of valve disc57, are configured so as to align with horizontal passages55A,55B,55C, and55D, as schematically shown inFIGS. 5D and 5E, which show the valve50in its first and second positions respectively.

In the first position (FIG. 5D), higher-pressure gas flows through port52C, horizontal passage55C, and channel57A into horizontal passage55A, and thence to cylinder chamber28A. At the same time, spent gas from cylinder chamber28B flows from horizontal passage55B into channel57B, and thence through horizontal passage55D and port52D to the area of lower pressure. In the second position (FIG. 5E), higher-pressure gas flows through port52C, horizontal passage55C, and channel57A into horizontal passage55B, and thence to cylinder chamber28B, while spent gas from cylinder chamber28A flows from horizontal passage55A into channel57B, and thence through horizontal passage55D and port52D to the area of lower pressure.

As shown inFIG. 5C, gas valve50may have a pressure chamber59. In this configuration, and as may been seen inFIGS. 5C to 5E, valve disc57has an auxiliary passage55C centered within channel57A and passing through the full thickness of valve disc57. Retainer plate57D has a corresponding opening such that gas can flow from channel57A into pressure chamber59. This has beneficial effect of pressurizing pressure chamber59so as to assist in maintaining valve disc57in close sealing contact against planar surface51. As illustrated inFIGS. 5B,5C, and5F, gas valve50in this embodiment may have a spring-loaded resistance-adjustment mechanism with adjustment screw58, for adjusting the interfacial pressure between the valve disc57and planar surface51. This in turn adjusts the resisting moment acting on valve shaft67, thus providing additional means of controlling or fine-tuning the operation of switching means70.

FIGS. 6A to 6Eillustrate a third embodiment of gas valve50, andFIG. 6Fillustrates a spring-loaded resistance-adjustment mechanism. Having regard to the preceding explanations of the first and second gas valve embodiments, the configuration and operation of the valve in FIGS.6A to6E will be readily comprehended by persons skilled in the art, without need of detailed discussion.

In preferred embodiments, the pneumatic motor also incorporates a pneumatic filter as illustrated inFIGS. 1A,1B,8A, and8B, to remove impurities from gas flowing into the motor from the area of higher pressure. Even more preferably, the pneumatic filter features a gravitational check valve as shown inFIG. 8C. Also in the preferred embodiment, the pneumatic motor incorporates a combined relief valve and differential magnetic gauge, as illustrated inFIGS. 1A,1B,9A and9B, for indicating the pressure differential between the higher and lower pressure areas, and for maintaining the pressure differential within desired limits.

FIG. 10provides a graphic illustration of the beneficial effectiveness of the differential shuttle valve of the present invention. Tests were performed using two pneumatic reciprocating motors, in accordance with one embodiment of the invention. The two test motors were essentially identical except that one had a differential shuttle valve and the other did not. The piston of each test motor had a diameter of six inches and a stroke of three inches. Each test motor was used to drive a plunger pump under conditions where the input gas pressure to the motor was 100 psi, and the outlet gas pressure from the motor was varied from 90, 80, and 70 psi (i.e., corresponding to differential pressures of 10, 20, and 30 psi.). The maximum oil pressure produced by the plunger pump was read on a pressure gauge having a capacity of 7,000 psi. The results of these tests, plotted onFIG. 10, indicate a large increase in the pump's output pressure when driven by the motor having the differential shuttle valve.

It will be readily seen by those skilled in the art that various modifications of the present invention may be devised without departing from the essential concept of the invention, and all such modifications are intended to be included in the scope of the claims appended hereto.

In this patent document, the word “comprising” is used in its non-limiting sense to mean that items following that word are included, but items not specifically mentioned are not excluded. A reference to an element by the indefinite article “a” does not exclude the possibility that more than one of the element is present, unless the context clearly requires that there be one and only one such element.