Abstract:
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&#39; 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.

Description:
FIELD OF THE INVENTION  
       [0001]     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  
       [0002]     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.  
         [0003]     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.  
         [0004]     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.  
         [0005]     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.  
         [0006]     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.  
         [0007]     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.  
         [0008]     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.  
         [0009]     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.  
         [0010]     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  
       [0011]     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.  
         [0012]     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.  
         [0013]     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.  
         [0014]     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.  
         [0015]     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.        
 
         [0023]     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.       
 
     
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0039]      FIG. 1A  is a side view of a pneumatic reciprocating motor apparatus in accordance with the preferred embodiment of the present invention, configured to power a pair of pumps (illustrated for exemplary purposes as a piston pump and a plunger pump).  
         [0040]      FIG. 1B  is a plan view of the pneumatic reciprocating motor of  FIG. 1A .  
         [0041]      FIG. 2  is a schematic drawing of the pneumatic reciprocating motor in accordance with the preferred embodiment of the invention.  
         [0042]      FIG. 3  is a partial section through the cylinder and piston of the apparatus, particularly illustrating the differential shuttle valve of the invention.  
         [0043]      FIG. 3A  is a cross-sectional detail of a spring-type circumferential seal as preferably used with the piston of the apparatus, conceptually illustrating the normal forces and friction forces associated with the seal.  
         [0044]      FIG. 3B  is an exploded view of the differential shuttle valve of the invention, in accordance with the embodiment illustrated in  FIG. 3 .  
         [0045]      FIG. 3C  is an elevational view (with cross-sectional detail) of the shuttle member of the differential shuttle valve shown in  FIG. 3 .  
         [0046]      FIG. 3D  illustrates an alternative embodiment of the differential shuttle valve, with frustoconical cap members.  
         [0047]      FIG. 3E  illustrates a further alternative embodiment of the differential shuttle valve, with frustoconical cap members and an alternatively configured shuttle member.  
         [0048]      FIG. 4A  is an elevational view of a rotary valve in accordance with a first embodiment.  
         [0049]      FIG. 4B  is a plan view of the rotary valve of  FIG. 4A .  
         [0050]      FIG. 4C  is a cross-sectional view of the rotary valve of  FIG. 4A .  
         [0051]      FIG. 4D  is a cross-section through the rotor of the rotary valve of  FIG. 4A , configured in a first position such that the first chamber of the cylinder of the apparatus is in fluid communication with an area of higher pressure in a pressurized gas system, and the second cylinder chamber is in fluid communication with an area of lower pressure in the pressurized gas system.  
         [0052]      FIG. 4E  is a cross-section through the rotor of the rotary valve of  FIG. 4A , configured in a second position such that the second cylinder chamber is in fluid communication with the area of higher pressure, and the first cylinder chamber is in fluid communication with the area of lower pressure.  
         [0053]      FIG. 5A  is an elevational view of a rotary valve in accordance with a second embodiment of the invention.  
         [0054]      FIG. 5B  is a plan view of the rotary valve of  FIG. 5A .  
         [0055]      FIG. 5C  is a cross-sectional view of the rotary valve of  FIG. 5A .  
         [0056]      FIG. 5D  is a cross-section through of the rotor of the rotary valve of  FIG. 5A , configured in a first position such that the first chamber of the cylinder of the apparatus is in fluid communication with an area of higher pressure in a pressurized gas system, and the second cylinder chamber is in fluid communication with an area of lower pressure in the pressurized gas system.  
         [0057]      FIG. 5E  is a cross-section through of the rotor of the rotary valve of  FIG. 5A , configured in a second position such that the second cylinder chamber is in fluid communication with the area of higher pressure, and the first cylinder chamber is in fluid communication with the area of lower pressure.  
         [0058]      FIG. 5F  is a sectional detail of the resistance adjustment mechanism of the rotary valve shown in  FIGS. 5B and 5C .  
         [0059]      FIG. 6A  is an elevational view of a rotary valve in accordance with a third embodiment of the invention.  
         [0060]      FIG. 6B  is a plan view of the rotary valve of  FIG. 6A .  
         [0061]      FIG. 6C  is a cross-sectional view of the rotary valve of  FIG. 6A .  
         [0062]      FIG. 6D  is a cross-section through of the rotor of the rotary valve of  FIG. 6A , configured in a first position such that the first chamber of the cylinder of the apparatus is in fluid communication with an area of higher pressure in a pressurized gas system, and the second cylinder chamber is in fluid communication with an area of lower pressure in the pressurized gas system.  
         [0063]      FIG. 6E  is a cross-section through of the rotor of the rotary valve of  FIG. 6A , configured in a second position such that the second cylinder chamber is in fluid communication with the area of higher pressure, and the first cylinder chamber is in fluid communication with the area of lower pressure.  
         [0064]      FIG. 6F  is a sectional detail of the resistance adjustment mechanism of the rotary valve shown in  FIGS. 6B and 6C .  
         [0065]      FIG. 7  is an elevational view of the switch mechanism of the rotary valve in accordance with a preferred embodiment of the invention.  
         [0066]      FIG. 8A  is an elevational view of a pneumatic filter in accordance with a preferred embodiment of the invention.  
         [0067]      FIG. 8B  is a cross-sectional view through the pneumatic filter shown in  FIG. 8A .  
         [0068]      FIG. 8C  is a detail of an optional gravitational check valve of the pneumatic filter.  
         [0069]      FIG. 9A  is an elevational view of a differential magnetic gauge in accordance with a preferred embodiment of the invention.  
         [0070]      FIG. 9B  is a cross-sectional view of the gauge of  FIG. 9A .  
         [0071]      FIG. 10  is a graph plotting measured output pressures for a plunger pump driven by a pneumatic reciprocating motor in accordance with an embodiment of the present invention, with and without the differential shuttle valve. 
     
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT  
       [0072]     Referring in particular to  FIGS. 1A, 1B ,  2 , and  3 , the pneumatic motor of the present invention (generally designated by reference number  10 ), comprises a pneumatic cylinder  20  and a double-acting piston  30  that is reciprocatingly and coaxially movable within pneumatic cylinder  20 . The pneumatic cylinder  20  has a cylindrical inner wall  22  and is capped at each end by cylinder heads  24 A and  24 B. The piston  30  has circular piston faces  32 A and  32 B and a circumferential side surface  34  extending between piston faces  32 A and  32 B. A piston rod  36 , having ends  36 A and  36 B, is rigidly and coaxially fixed to piston  30 , with rod ends  36 A and  36 B extending through rod openings  26 A and  26 B in cylinder heads  24 A and  24 B respectively. Piston rod seals  26 C are provided in association with rod openings  26 A and  26 B such that piston rod  36  is reciprocatingly movable through rod openings  26 A and  26 B in substantially pressure-tight fashion. In the preferred embodiment, piston rod seals  26 C are dynamic seals similar to the piston seals  38  described elsewhere in this specification.  
         [0073]     Pneumatic cylinder  20  defines an annular cylinder chamber  28 A bounded by cylinder wall  22 , cylinder head  24 A, and piston face  32 A, and an annular cylinder chamber  28 B bounded by cylinder wall  22 , cylinder head  24 B, and piston face  32 B. The length and volume of cylinder chambers  28 A and  28 B varying according to the position of piston  30  within cylinder  20 . For purposes to be explained further herein, cylinder head  24 A has cylinder head gas port  25 A in fluid communication with cylinder chamber  28 A, and cylinder head  24 B has cylinder head gas port  25 B in fluid communication with cylinder chamber  28 B.  
         [0074]     As particularly illustrated in  FIGS. 3 and 3 A, piston  30  is provided with two circumferential piston seals  38 , each disposed in a circumferential chase  39  formed into side surface  34  of piston  30  near one end of piston  30 . Piston seals  38  are at all times sealingly engaged against cylinder wall  22 , so as to substantially prevent leakage of gas from either of the cylinder chambers  28 A and  28 B. In the preferred embodiment, as shown in  FIGS. 3 and 3 A, piston seals  38  are 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 chases  39 . The elastic energy or spring force thus induced in the piston seals  38  biases them radially outward and into contact with cylinder wall  22 . As conceptually illustrated in  FIG. 3A , this outward biasing force manifests as a normal force F n  acting against cylinder wall  22 . The friction force F f  required to overcome normal force F n  (in order for piston  30  to move) is directly proportional to normal force F n . Accordingly, piston seals  38  are ideally designed or selected so as to induce a normal force F n  that is as low as possible in order to minimize friction force F f , while being high enough to ensure a vapor-tight seal against cylinder wall  22 .  
         [0075]     Referring to  FIGS. 2 and 3 , piston  30  incorporates a shuttle valve  40  whereby pressurized gas can be introduced into the annular space  29  radially bounded by piston  30  and cylinder  20 , and longitudinally bounded by piston seals  38 . A transverse passage  41  extends through piston  30  at a selected location, with said passage  41  configured to include a central bore  41 A and a concentric and larger diameter recess  41 B adjacent to each of piston faces  32 A and  32 B, such that an annular shoulder  42  is formed between central bore  41 A and each recess  41 B. A radial passage  43  extends through piston  30  between central bore  41 A and annular space  29 . Shuttle valve  40  includes a shuttle member  44  with cap members  45  at each end, with the clear distance between the cap members  45  being greater than the length of central bore  41 A between recesses  41 B. Each cap member has an outer face  45 A and an inner face  45 B and an annular groove  45 C is formed in each inner face  45 B for receiving an O-ring  46  or similar sealing member.  
         [0076]     The cross-sectional geometry of shuttle member  44  is configured such that shuttle member  44  can slide freely within central bore  41 A but with fairly close tolerances so that it slides substantially coaxially within central bore  41 A, while at the same time defining at least one longitudinal channel between shuttle member  44  and the walls of central bore  41 A. In one embodiment, this feature is provided by forming shuttle member  44  from initially round stock into which one or more longitudinal flattened surfaces are formed. This creates one or more longitudinal channels  47  which in cross section resemble a circular segment. This and alternative embodiments of the shuttle member  44  are illustrated in  FIGS. 3B through 3E  (described in further detail below).  
         [0077]     As shown in  FIG. 3 , shuttle valve  40  is assembled with shuttle member  44  disposed within central bore  41 A, and with each cap member  45  disposed within a corresponding recess  41 B. Accordingly, each  0 -ring  46  directly faces and is substantially parallel to a corresponding shoulder  42 . Because the length of shuttle member  44  is greater than the length of central bore  41 A, one of cap members  45  will always be separated slightly away from their corresponding shoulders  42 . When one cap member  45  is separated from its corresponding shoulder  42  (such as the lefthand cap member  45  in  FIG. 3 ), a pathway is created whereby gas present in cylinder chamber  28 A can pass around cap member  45 , through longitudinal channel(s)  47 , through radial passage  43 , and into annular space  29 . As may be seen from  FIG. 3 , if the righthand cap member  45  is being pressed against its corresponding shoulder  42 , the corresponding O-ring  46  will seal the righthand cap member  45  against its corresponding shoulder  42 , thus preventing any flow of gas between cylinder chamber  28 B and the shuttle valve  40 .  
         [0078]      FIGS. 3B and 3C  illustrate one alternative construction of the shuttle valve  40 . As shown in  FIG. 3B , each cap member  45  has a threaded stem  45 D that is matingly engageable with threaded bore  44 A of shuttle member  44 . To assemble the shuttle valve  40 , one cap member  45  is screwed into one end of shuttle member  44 , and this subassembly is inserted into central bore  41 A of piston  30 . The other cap member  45  may then be screwed into the other end of shuttle member  44 .  
         [0079]     As shown in  FIG. 3C , shuttle member  44  is made from round stock that has been milled flat on four sides  44 B, leaving four longitudinal surfaces  44 C which retain the radius of the round stock. The radius of the round stock is slightly less than the radius of central bore  41 A, such that shuttle member  44  can slide freely within central bore  41 A but without significant “play”. When the shuttle valve  40  is installed in central bore  41 A of piston  30 , the space between the surface of central bore  41 A and each flattened side surface  44 B forms a longitudinal channel  47 . Centrally-located portions of the longitudinal surfaces  44 C of shuttle member  44  are milled to create recessed areas  44 D that permit fluid communication between adjacent longitudinal channels  47 . The length of the recessed areas  44 D is such that at least a portion of the length will coincide with the opening from central bore  41 A into radial passage  43  regardless of the position of shuttle valve  40  within central bore  41 A. This arrangement ensures that gas flowing into the longitudinal channels  47  from cylinder chamber  28 A or cylinder chamber  28 B will pass through longitudinal channels  47  into radial passage  43  and thence into annular space  29 .  
         [0080]      FIG. 3D  illustrates an alternative construction of shuttle valve  40  largely similar to that shown in  FIGS. 3, 3B , and  3 C except that cap members  45  are of frustoconical configuration and recesses  41 B are correspondingly shaped.  FIG. 3E  illustrates an alternative construction of shuttle valve  40  having frustoconical cap members  45  as in  FIG. 3D  but with a differently-configured shuttle member  44 . As conceptually indicated, the stems  45 C of cap members  45  are internally threaded and mate with externally-threaded ends of shuttle member  44 . The frustoconical cap members  45  are effectively self-centering within central bore  41 A, so the diameter of stems  45 C can be sufficiently smaller than that of central bore  41 A so as to form a substantially longitudinal channel  47  therebetween. The diameter of shuttle member  44  is less than that of stems  45 C, so as to form an annular recessed area  44 D. Alternatively, stems  45 C may be fabricated with flattened surfaces similar to the flattened side surface  44 B of the shuttle member  44  in  FIG. 3C , with corresponding longitudinal surfaces  44 C, such that stems  45 C can slide freely but without play within central bore  41 A.  
         [0081]     It can be readily seen that if the gas pressure in cylinder chamber  28 B exceeds the gas pressure in cylinder chamber  28 A, the shuttle valve assembly  40  will move to the left, into the position shown in  FIG. 3 , with gas free to flow from cylinder chamber  28 A to annular space  29  as described above. If the pressure in cylinder chamber  28 A is then made to exceed the gas pressure in cylinder chamber  28 B, the shuttle valve assembly  40  will move to the right, sealing the lefthand cap member  45  against its corresponding shoulder  42  and preventing any flow of gas between cylinder chamber  28 A and the shuttle valve  40 , while at the same time allowing gas to flow from cylinder chamber  28 B to annular space  29 .  
         [0082]     Other configurations of shuttle valve  40 , functioning substantially as described above, may be devised without departing from the principles and scope of the present invention.  
         [0083]     The pneumatic motor  10  also includes a multi-position gas valve  50  having valve ports  52 A,  52 B,  52 C, and  52 D. By means of suitable conduits, valve port  52 A is in fluid communication with cylinder head port  25 A and valve port  52 B is in fluid communication with cylinder head port  25 B. Valve port  52 C is in fluid communication with an area HP in a pressurized gas system (such as a gas pipeline), and valve port  52 D is in fluid communication with an area LP in the gas system, said area LP being at a pressure lower than area HP. Gas valve  50  is operable between: 
        a first position in which valve ports  52 A and  52 C are in fluid communication, putting cylinder chamber  28 A in fluid communication with area HP, while valve ports  52 B and  52 D are in fluid communication, putting cylinder chamber  28 B in fluid communication with area LP; and     a second position in which valve ports  52 A and  52 D are in fluid communication, putting cylinder chamber  28 A in fluid communication with area LP, while valve ports  52 B and  52 C are in fluid communication, putting cylinder chamber  28 B in fluid communication with area HP.        
 
         [0086]      FIGS. 4A  to  4 E illustrate a multi-position gas valve  50  in accordance with a preferred embodiment of the present invention. As best seen in  FIGS. 4D and 4E , the gas valve  50  in this embodiment is a rotary valve having a cylindrical interior cavity  54 , with valve ports  52 A,  52 B,  52 C, and  52 D all in communication therewith. Cavity  54  is circumferentially bounded by cylindrical surface  53 . A rotor  56  is coaxially rotatable within cavity  54  about rotational axis A, and is geometrically configured such that particular valve ports will be in fluid communication, via segmental sub-cavities  54 A on either side of rotor  56 , when the valve  50  is in the first and second positions, as described above. Rotor  56  is fixed to valve shaft  67  so as to be coaxially rotatable about rotational axis A. Rotor  56  has rotor ends  58  that engage cylindrical surface  55  as rotor  56  cycles between operational positions, in substantially vapor-tight fashion such that there is no leakage of gas between segmental sub-cavities  54 A.  
         [0087]     Preferably, the vapor-tight engagement of rotor ends  58  with cylindrical surface  55  is facilitated by use of a separate sealing means, an example of which is illustrated in  FIGS. 4D and 4E . In the illustrated embodiment, a longitudinal slot  62  is formed in each rotor end  58 , and a resilient biasing means  64  is disposed along the base of each slot  62 . A selected pressure seal material  66  (such as, for instance, Teflon™ lamella) is then inserted into each slot  62 , with the dimensions of the pressure seal  66  being such that it will project slightly beyond the face of rotor end  58  when not subject to compressive force urging it radially into slot  62 . Thus, when rotor  58  is positioned within cavity  54 , pressure seal  66  will at all times be in contact with cylindrical surface  55 , with resilient biasing means  64  constantly urging pressure seal  66  radially outward against cylindrical surface  55 .  
         [0088]     In  FIGS. 4D and 4E , rotor  56  is shown having straight or flat side portions, but this is not critical. The rotor  56  may have curvilinear or other geometric contours without substantively affecting the functioning of valve  50 , so long as the stated operational interrelation of valve ports  52 A,  52 B,  52 C, and  52 D is maintained when valve  50  is in the first and second operational positions.  
         [0089]     Gas valve  50  is actuated between its first and second operational positions by means of a switch mechanism  70  which cycles the valve  50  at the end of each stroke of piston  30  and piston rod  36 . 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 mechanism  70  in accordance with the operational mode described above.  FIG. 7  illustrates one example of such a mechanism, as used in a preferred embodiment of the invention. Switch mechanism  70  is disposed within switch housing  71 . A sleeve  74  is slidingly disposed around the portion of piston rod  36 B extending from cylinder  20 . Piston rod  36 B is reciprocatingly movable relative to switch housing  71  as piston  30  reciprocates within cylinder  20 . Suitable collars  73 A and  73 B are positioned at a desired spacing on either side of the sleeve  74  so as to limit the range of sliding movement of sleeve  74  on piston rod  36 B. A bracket  74 A fixed to sleeve  74  has a spring-retaining pin  74 B for receiving the first end of a tension spring  76 . A lever arm  72  is mounted at one end to valve shaft  67 , which projects into switch housing  71 . The other end of lever arm  72  has a spring-retaining pin  72 A which receives the second end of a tension spring  76  (shown in discontinuous fashion in  FIG. 7  for purposes of clarity) Lever arm bumpers  78 A and  78 B are mounted to switch housing  71  to limit the travel of lever arm  72 . Lever arm  72  is offset from piston rod  36 B so as not to impede its reciprocating movement.  
         [0090]     The operation of switch mechanism  70  may be understood from  FIG. 7 , in which sleeve  74  (shown cross-hatched for clarity) is at its leftmost limit of travel relative to piston rod  36 B. For purposes of illustration, valve  50  may be considered to be in its first position when the switch mechanism is as shown in solid outline in  FIG. 7 . As piston rod  36 B moves to the right (indicated by arrow R in  FIG. 7 ), sleeve  74  will be pushed to the right as well by collar  73 A. The rightward movement of sleeve  74  causes tension spring  76  to stretch, but this initially has no effect on lever arm  72 , which remains in position against the left bumper  78 A. However, as the center of spring-retaining pin  74 B moves rightward past rotational axis A of valve shaft  67 , the tensile force in tension spring  76 , acting downward and to the right against spring-retaining pin  72 A, applies a clockwise moment on lever arm  72 , around rotational axis A. The magnitude of this moment increases as the rightward movement of sleeve  74  progresses, until it overcomes the resistant moment acting on valve shaft  67  (e.g., due to friction forces within the valve  50 ). At that point, lever arm  72  will swing clockwise to the position shown in phantom outline. Since lever arm  72  is fixed to valve shaft  67 , this has the effect of switching valve  50  from its first position to its second position. Piston rod  36 B will reach the end of its rightward stroke soon after this happens; at this point, sleeve  74  will be abutting collar  73 B. Piston rod  36 B will then begin its leftward stroke, ultimately causing lever arm  72  will swing counterclockwise, thus switching valve  50  from the second position back to the first position.  
         [0091]     The positions of collars  73 A and  73 B relative to piston rod  36 B may be adjusted so as to regulate the lag between the swing of lever arm  72  and the end of the piston rod stroke.  
         [0092]     The operation of the pneumatic motor of the present invention may now be easily understood having reference to  FIGS. 2, 3 ,  4 D, and  4 E in particular. With gas valve  50  in the first position, higher-pressure gas from area HP flows into cylinder chamber  28 A while lower-pressure gas from area LP flows into cylinder chamber  28 B. The pressure differential between the two chambers causes piston  30  to move to the right, into the position shown in  FIG. 2 . This causes piston rod  36  to move in a rightward power stroke. At the same time, the pressure differential causes differential valve  40  to move to the right such that the lefthand cap member  45  (of shuttle valve  40 ) and its associated O-ring  46  are urged against their corresponding shoulder  42 , while the righthand cap member  45  and its associated O-ring  46  are moved away from their corresponding shoulder  42 . In this configuration, gas is prevented from escaping from cylinder chamber  28 A into central bore  41 A of piston  30 , while gas is free to flow from cylinder chamber  28 B into annular space  29 , thus eliminating or greatly reducing the pressure differential across the piston seals  38 .  
         [0093]     As piston  30  reaches or nears the end of its rightward power stroke, switching mechanism  70  cycles gas valve  50  to the second position. Now, higher-pressure gas from area HP flows into cylinder chamber  28 B while lower-pressure gas from area LP flows into cylinder chamber  28 A. The pressure differential between the two chambers causes piston  30  to move to the left, into the position shown in  FIG. 3 . This causes piston rod  36  to move in a lefttward power stroke. At the same time, the pressure differential causes differential valve  40  to move to the left such that the righthand cap member  45  and its associated  0 -ring  46  are urged against their corresponding shoulder  42 , while the lefthand cap member  45  and its associated  0 -ring  46  are moved away from their corresponding shoulder  42 . In this configuration, gas is prevented from escaping from cylinder chamber  28 B into central bore  41 A of piston  30 , while gas is free to flow from cylinder chamber  28 A into annular space  29 , thus once again eliminating or greatly reducing the pressure differential across the piston seals  38 . As piston  30  reaches or nears the end of its leftward power stroke, switching mechanism  70  cycles gas valve  50  back to the first position, and the alternating cycles continue as long as valve ports  52 C and  52 D remain in fluid communication with areas HP and LP respectively in a pressurized gas system.  
         [0094]     The foregoing discussion has been in the context of a pneumatic motor using the rotary valve illustrated in  FIGS. 4A  to  4 D. However, various other forms of gas valve  50  may be used without departing from the principles and scope of the present invention.  FIGS. 5A  to  5 E illustrate a second embodiment of gas valve  50 , which may be alternatively described as a planar valve. The valve body has valve ports  52 A,  52 B,  52 C, and  52 D as previously described in connection with the valve in  FIGS. 4A  to  4 D. These ports are in fluid communication, respectively, with internal horizontal passages  55 A,  55 B,  55 C, and  55 D, which terminate at a common planar surface  51 . As best seen in  FIG. 5C , a valve disc  57 , preferably made of Teflon™ (or an alternative material with good sealing and abrasion-resistance characteristics) is co-rotatably fixed to valve shaft  67 . Valve disc  57  interfaces tightly against planar surface  51  as shown, and is retained by retainer plate  57 D. Valve disc  57  has arcuate channels  57 A and  57 B, the configuration of which can best be seen in  FIGS. 5D and 5E . Arcuate channels  57 A and  57 B, which extend only partly through the thickness of valve disc  57 , are configured so as to align with horizontal passages  55 A,  55 B,  55 C, and  55 D, as schematically shown in  FIGS. 5D and 5E , which show the valve  50  in its first and second positions respectively.  
         [0095]     In the first position ( FIG. 5D ), higher-pressure gas flows through port  52 C, horizontal passage  55 C, and channel  57 A into horizontal passage  55 A, and thence to cylinder chamber  28 A. At the same time, spent gas from cylinder chamber  28 B flows from horizontal passage  55 B into channel  57 B, and thence through horizontal passage  55 D and port  52 D to the area of lower pressure. In the second position ( FIG. 5E ), higher-pressure gas flows through port  52 C, horizontal passage  55 C, and channel  57 A into horizontal passage  55 B, and thence to cylinder chamber  28 B, while spent gas from cylinder chamber  28 A flows from horizontal passage  55 A into channel  57 B, and thence through horizontal passage  55 D and port  52 D to the area of lower pressure.  
         [0096]     As shown in  FIG. 5C , gas valve  50  may have a pressure chamber  59 . In this configuration, and as may been seen in  FIGS. 5C  to  5 E, valve disc  57  has an auxiliary passage  55 C centered within channel  57 A and passing through the full thickness of valve disc  57 . Retainer plate  57 D has a corresponding opening such that gas can flow from channel  57 A into pressure chamber  59 . This has beneficial effect of pressurizing pressure chamber  59  so as to assist in maintaining valve disc  57  in close sealing contact against planar surface  51 . As illustrated in  FIGS. 5B, 5C , and  5 F, gas valve  50  in this embodiment may have a spring-loaded resistance-adjustment mechanism with adjustment screw  58 , for adjusting the interfacial pressure between the valve disc  57  and planar surface  51 . This in turn adjusts the resisting moment acting on valve shaft  67 , thus providing additional means of controlling or fine-tuning the operation of switching means  70 .  
         [0097]      FIGS. 6A  to  6 E illustrate a third embodiment of gas valve  50 , and  FIG. 6F  illustrates 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.  6 A to  6 E will be readily comprehended by persons skilled in the art, without need of detailed discussion.  
         [0098]     In preferred embodiments, the pneumatic motor also incorporates a pneumatic filter as illustrated in  FIGS. 1A, 1B ,  8 A, and  8 B, 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 in  FIG. 8C . Also in the preferred embodiment, the pneumatic motor incorporates a combined relief valve and differential magnetic gauge, as illustrated in  FIGS. 1A, 1B ,  9 A and  9 B, for indicating the pressure differential between the higher and lower pressure areas, and for maintaining the pressure differential within desired limits.  
         [0099]      FIG. 10  provides 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 on  FIG. 10 , indicate a large increase in the pump&#39;s output pressure when driven by the motor having the differential shuttle valve.  
         [0100]     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.  
         [0101]     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.