WO1995034756A1

WO1995034756A1 - Tandem positive displacement pump

Info

Publication number: WO1995034756A1
Application number: PCT/US1995/007455
Authority: WO
Inventors: Douglas C. Hicks
Original assignee: Fmc Corporation
Priority date: 1994-06-15
Filing date: 1995-06-13
Publication date: 1995-12-21
Also published as: AU2772295A

Abstract

A high pressure, tandem positive displacement piston pump (10) for pumping a corrosive fluid is disclosed. The pump includes two axially opposed pump bodies each having an inlet (26A, 26A') and outlet (28A, 28A') as well as a plurality of pistons (54, 54') arranged in axially opposed cylinders (51, 51'). A two sided cam device (62) moves each pair of opposed pistons (54, 54') reciprocally and has two opposed first camming surfaces (131, 131') which cyclically rotate. At the end of each piston (54, 54'), a second camming surface is provided which engages the first camming surface. A dynamic bearing system (75, 77, 81) supports the cam device axially and overcomes any imbalance between the opposed pistons. A shaft (66), with suitable journaling device (79, 79'), is preferably used for rotating the two sided cam device. A cooling and lubricating system is also provided for cooling the camming surfaces, the dynamic bearing system, and the journaling device. The coolant liquid can be the corrosive liquid being pumped.

Description

TANDEM POSITIVE DISPLACEMENT PUMP

The present invention relates generally to pumps, and more particularly to high pressure, positive displacement piston pumps for pumping corrosive fluids.

When faced with a need to increase the volumetric discharge capacity of positive displacement style pumps one may generally increase the size of the pistons/plungers by increasing the diameter and/or stroke length of the pistons/plungers, increasing the operating speed, or increasing the total number of pistons/plungers used in the pump. Physical limitations imposed by the laws of physics and fluid mechanics may limit the success when any of these modifications are made separately or in concert. For example, larger diameter pistons will increase the fluid displacement of the pump, but the associated increase in mechanical component weight and inertia may require that the pump's operating speed be reduced, resulting in a lesser gain in overall discharge capacity. Inertial and viscous forces also limit how fast a given size piston/plunger can be operated. A larger-sized pump piston may not be able to refill quickly enough on an intake stroke. In general, pumps for aqueous solutions with discharge capabilities above ten gallons per minute (37.8 liters per minute) normally cannot be operated at rotational speeds above about 1000 rpm. There are also practical limitations to the maximum stroke length and number of pistons that can be engineered into a reasonable-sized pump frame. Similarly, the load forces associated with containing pressure within a larger pump with increased capacity can exceed the limits of the available materials of construction.

Commercially available high pressure pumps used in pumping corrosive fluids, such as seawater, have their fluid ends constructed from a combination of expensive metal alloys to withstand the corrosive effects of the working fluid. Positive displacement type pumps require a transmission to convert the rotational drive input, such as from an electric motor, into the linear motion to pump the working fluid. Conventional pumps rely on an oil bath to cool and lubricate the transmission, with dynamic seals employed to isolate the transmission oil from the working fluid in the fluid-end. These designs require periodic oil changes (approximately every 300-500 hours) and frequent adjustment or replacement of the oil/working fluid seals. In addition, the combination of metal alloys commonly employed in the fluid end often result in electrolysis and premature failure of components such as valve springs, seats, and seals.

A positive displacement piston pump for pumping a corrosive fluid is disclosed in U.S. Patent No. 5,013,219. The assignee of this application is the exclusive licensee of U.S. Patent No. 5,013,219. The pump disclosed in that patent has a cam device with a rotating member (swash plate) having a single camming surface which is cyclically rotated adjacent the ends of a plurality of pistons. The piston ends are provided with camming surfaces to engage the single camming surface of the rotating member. Rotation of the rotating member causes each of the pistons to reciprocate axially within piston chambers. The camming surfaces of the pump may be cooled by the fluid being pumped rather than a separate lubricant or coolant fluid. The pump is constructed primarily of non-corrodible materials and has a capacity of about 0.1 to 120 liters per minute. This patent does not describe how to increase pumping capacity without substantially increasing pump size (the dimensions of the outer housing of the pump). Nor does this patent teach use of a double-sided centered swash plate cam device that controls a plurality of opposing pistons. Nor does this patent teach improved means for centering and aligning the rotating shaft axially.

A two-stage pump is disclosed in U.S. Patent No. 4,105,369. The two stage pump has low and high pressure stages formed by a pair of pumps positioned in a casing. Each pump has a fluid inlet and a fluid outlet, with the fluid outlet of the low stage pump being in fluid flow relation with the fluid inlet of the high stage pump and with an outlet port from the casing. As shown in the patent, only the high stage pump is an axial piston pump and it has only a single plurality of reciprocal pistons 54 within a plurality of axially-extending cylinders 53. The strokes of the pistons are controlled by only a single side of a rotatable angle plate 55. This patent does not describe how to increase pumping capacity without substantially increasing pump size (the dimensions of the outer housing of the pump).

Another two-stage pumping system is disclosed in U.S. Patent No. 2,709,339. The pumping system relates to high pressure piston pumps for delivering hydraulic fluid to a hydraulic motor, primarily for aircraft. The pistons are made of metal alloys that expand at a faster rate than the metal of the cylinder block. The pump has two separate outer cams 31 and 31' that rotate with shaft 16. Axially opposed aligned pistons are mounted in cylinder bores. As the thrust bearings or washers 33 and 33' rotate with the cams, grooves 35 in the washers distribute lubricant and the washers compress to form wedge shaped pockets, loaded with hydraulic fluid to support the end thrust load on the cams. The rotating shaft is journaled in roller bearings 18 and 18'. This patent does not teach the use of corrosion- resistant materials for the pump body, cylinders, pistons and other pump parts. This patent further does not teach use of a double-sided centered swash plate cam device to balance the thrust loads from a pair of opposed pistons, and, therefore, requires two sets of thrust bearings and has a reduced overall efficiency. Moreover, because the pump has two separate outer cams and a central fluid intake, it cannot be used to pump two different fluids simultaneously.

In accordance with the present invention, a high pressure, tandem positive displacement piston pump for pumping fluids, preferably corrosive fluids, is provided. The tandem pump includes two axially opposed pump bodies each having a plurality of cylinders therein. Each pump body is provided with an inlet and outlet through the pump body to the cylinders. An inlet one-way valve means and an outlet one-way valve means are disposed, respectively in the inlets and outlets for allowing pumped fluid flow into and out of each cylinder. A piston is disposed in each cylinder for reciprocal movement therein in order to pump the fluid from the inlet to the outlet. A centrally located and centered cam means is provided for moving the pistons reciprocally in each cylinder. The cam means includes a double sided rotating member or swash plate having at least two first camming surfaces. The swash plate is cyclically rotated adjacent an end of each piston. A second camming surface at the end of each piston engages one of the first camming surfaces to move the piston reciprocally in its respective cylinder.

The first camming surfaces are preferably formed from a corrosion resistant metal alloy such as stainless steel, monel, titanium, etc.. Other suitable materials include ceramics, epoxies, polyamide-imides, polyetheretherketones, composites, and highly polymerized organic materials. The second camming surface is preferably formed of an organic material preferably selected from the polymer group consisting of epoxies, polyvinyl chloride, acetal, polyester, polyimide, polyamide, polyamide-imide, polyetheretherketone, Teflon or polytetrafluoroethylene (PTFE), ultra high molecular weight polyethylene, and polyurethane. These materials are considered to include also those materials with internal lubricants, such as PTFE, molydisulfide, etc., and reinforcing from fibers as desired. In low duty applications, both the first and second camming surfaces may be formed from two different organic materials selected from those listed above.

A cooling means is further provided for cooling and lubricating the first and second camming surfaces. The cooling means includes a liquid coolant which contacts the first and second camming surfaces. In one embodiment of the present invention, the liquid being pumped is also used as the coolant. Where the pump of the present invention is used for pumping seawater, for example, or any other aqueous solutions, the coolant portion of the solution can be conducted from the pressurized inlet of the pump to the camming and bearing surfaces to be cooled and lubricated. In other situations where the liquid being pumped is suitable for use as the coolant and lubricant, a portion of the pressurized pump liquid from the inlet or outlet can be channeled and used for cooling and lubrication. When the pumped fluid is used also as the coolant it is an advantage of the present invention that seals between the pumped fluid and the cooling fluid are not required to completely isolate the two fluids. Some mixing of the two fluids by leakage between various components within the pump is easily tolerated. Alternatively, a separate fluid stream, such as fresh water, could be used as the coolant.

In one preferred embodiment of the invention, the rotating cam member includes a shaft and a means for journaling the shaft for rotation in the pump bodies. The shaft may preferably be made from a non-corrodible metal alloy and preferably be joumaled by bearings made of a polymer material, such as described with respect to the camming surfaces. In this preferred embodiment, the rotating member is a double sided centered swash plate having two first camming surfaces. While the double sided swash plate could be provided with first camming surfaces of various geometries, preferably the ramped surface on the first of the first camming surfaces has a slope angle of equal magnitude but opposite sign of the angle of the ramped surface on the second of the first camming surfaces so as to better balance the load on the swash plate. The swash plate is firmly attached to the shaft in a central position between the two pump bodies and rotates about the center axis of the shaft. The second camming surfaces of the pistons from each of the pump bodies engage or contact the first camming surfaces. Preferably, each pump body is provided with an equal number of cylinders and pistons so that the pistons may be positioned on opposite sides of the swash plate and axially opposite a respective piston of the other pump body. Most preferably, the cylinders and pistons are spaced annularly at an equal radial distance from the shaft and separated by equal angles. As the shaft and swash plate rotate within the pump cavity, the first camming surfaces of the swash plate continually contact the second camming surfaces of the pistons, causing the pistons to reciprocate within their respective cylinders to pump the fluid. The reaction force on the first of the first camming surfaces when the first of the first camming surfaces causes a piston to reciprocate upward into its cylinder is substantially balanced by the reaction force on the second of the first camming surfaces when the second of the first camming surfaces simultaneously causes the opposed piston from the other pump body to reciprocate upward into its cylinder.

The tandem positive displacement piston pump of the preferred embodiment of the present invention, which has a double sided centered swash plate with opposed pistons from two pump bodies acting thereon, is more efficient than two separate piston pumps each having only a single sided swash plate. The double sided swash plate supports working loads on both sides in a tandem pump, whereas greater frictional losses would be generated using the single pumps with single sided cams. The tandem positive displacement piston pump of the invention allows for a pump of double the discharge capacity while utilizing components of nominally the same size and operating at nominally the same speed.

In a further preferred embodiment of the invention, centering means for aligning the position of the axially rotating member or shaft are provided. Because the axial thrust on the double sided swash plate generated by the pistons in the cylinder housing of the first pump body may not always precisely match the thrust forces from the pistons in the cylinder housing of the second pump body, centering means for the shaft or for the shaft and swash plate include an axially rotatable shaft bearing ring, a substantially nonrotatable segmented thrust bearing ring, a relatively fixed, substantially compliant support pad, and a relatively axially movable, substantially compliant support pad. Preferably, the substantially compliant support pads do not contact the shaft or swash plate and are fitted within a recess, preferably an annular recess, in the pump body. One segmented thrust bearing ring is loosely attached to one of the compliant support pads so that a surface of the ring is adjacent to and contacts a surface of the pad and a portion of the segmented thrust bearing ring is within the recess in the pump body. The axially rotatable shaft bearing ring is mounted or attached to the rotatable member or shaft in a position so that it contacts the other surface of the nonrotatable segmented thrust bearing ring. The individual segments of the segmented thrust bearing ring are separated by gaps to allow flow of lubricant and coolant fluid therethrough. The segmented thrust bearing ring preferably is made of a noncorrodible polymer material, such as epoxy, acetal, polyester, polyamide, polyamide-imide, polytetrafluoroethylene (PTFE), ultra high molecular weight polyethylene, polyurethane, polyetheretherketone, polycarbonate, or polysulfone, or other highly polymerized organic materials. The polymer may also be reinforced with suitable filler materials, such as fibers or glass or molydisulfide. The compliant support pads preferably are made of elastomer materials such as polyurethane or Neoprene, or the like.

When the shaft rotates, the axially rotatable shaft bearing ring also rotates. From above the end of the rotating member or shaft, pressurized lubricating fluid flows through one or more bores into the shaft cavity and then through angled bores into the recess, preferably an annular recess. The fluid exerts fluid pressure on the relatively axially movable substantially compliant support pad filling that recess, and that compliant pad flexes or acts against the segmented thrust bearing ring, which in turn acts against the rotatable shaft bearing ring. The axially movable compliant pad moves, slides or flexes back and forth in an axial direction in response to fluid pressure from the pump inlet and any opposing axial loads on the segmented thrust bearing ring. The relatively fixed compliant support pad remains substantially stationary within the recess, but springs or flexes axially under the segmented thrust bearing ring to distribute the thrust and allowing the segments of the segmented thrust bearing ring to incline to form a hydrodynamic lubricating film between the segmented thrust bearing ring and the rotatable shaft bearing ring. The centering means maintains the axial position of the shaft and swash plate within the pump transmission chamber without need for other mechanical means. In other words, the compliant support pads and the segmented thrust bearing ring act on or react from forces on the rotatable shaft bearing ring.

The tandem positive displacement pump herein provides a pump with an increased capacity and efficiency over prior piston pumps of equal dimensions (i.e., size, weight and operating speed). Moreover, different fluids optionally may be pumped by each pump body of the tandem positive displacement pump.

Other features and advantages of the present invention are stated in or apparent from a detailed description of a presently preferred embodiment of the invention provided below. Fig. 1 is a top plan view of a pump according to the present invention;

Fig. 2 is a cross-sectional view in side_. elevation taken along line 2-2 of Fig. 1 ;

Fig. 3 is a fragmental top plan view of the valve housing of the pump shown in Fig. 2;

Fig. 4 is a fragmental top plan view of the cylinder housing of the pump shown in Fig. 2;

Fig. 5 is a bottom plan view of the cylinder housing shown in Fig. 4;

Fig. 6 is a cross-sectional view of the shaft bearing; Fig. 7 is a top plan view of the bearing ring shown in Fig. 2;

Fig. 8 is a side elevational view of the swash plate, drive shaft and bearing rings;

Fig. 9 is a fragmental top plan view of a piston and the flexible ring shown in Fig. 2; and Fig. 10 is a fragmental side elevational view of the piston and flexible ring of Fig. 9.

With reference to the drawings in which like numerals represent like elements throughout the several views, the presently preferred embodiment of a high pressure, positive displacement piston pump 10 are depicted in FIGS. 1 and 2. Pump 10 includes two opposed pump bodies 12 and 12', which are composed of galleries 14 and 14', valve housings 16 and 16', cylinder housings 18 and 18' and a transmission housing 19. Pump bodies 12 and 12' are held together by a plurality of bolt means 22 such as depicted in FIGS. 1 and 2 which extend through bores 24 in pump bodies 12 and 12'. Conveniently, bolt means 22 are also non-corrodible and are made of stainless steel, brass, or the like. Bolts 23 and 23' are also made of non-corrodible metals and serve to align the pump body parts, ensuring that the valves and cylinders are properly positioned relative to one another, and helping to contain the operating pressure within the pump.

Galleries 14 and 14' each include an inlet port 26 and 26', and an outlet port 28 and 28'. Gallery 14 and cylinder housing 18' also include coolant inlet ports 30 and 30', respectively. Ports 26, 28, 30, and 30' can be configured to receive piping in various forms such as pipe thread as is shown on ports 30 and 30', or pipe tubes sealed with O-rings and held in place with split retaining rings as shown on ports 26 and 28. Inlet port 26 is fluidly connected to a annular inlet channel 32 extending circumferentially in galleries 14 and 14' concentric to coolant inlet port 30 and pump shaft 66. Outlet port 28 is similarly connected to a annular outlet channel 34 inside and concentric with inlet channel 32. Coolant inlet port 30' has been shown only in FIG. 2. The exact radial position of this port 30' is not critical to the operation of the present invention.

The valve housings 16 and 16' (refer to FIG. 3) are located immediately adjacent the galleries 14 and 14'. In one preferred embodiment of pump 10, valve housings 16 and 16' include eleven bores 36 located equidistant from one another and underneath a respective portion of outlet channel 34. Similarly, pump 10 could be configured with a greater or lesser number of bores 36. Immediately adjacent each bore 36 is a bore 38 located underneath a respective portion of inlet channel 32. Disposed in each bore 38 is an inlet one-way valve means 40. Located in each bore 36 is an outlet one-way valve means 42. One-way valve means 40 and 42 are similar in appearance to conventional ball valves typically having three apertures at the sealing end and seven apertures at the opposite retainer end. As shown in FIGS. 2 and 4, cylinder housings 18 and 18' include cylinders 50 each provided with a liner 51 located immediately below each respective pair of inlet and outlet one-way valve means 40 and 42. Liner 51 is held into place in cylinder 50 by either a press-fit and/or adhesive bonding to secure and seal said liner 51. Although not shown, cylinder 50, could be held in place mechanically with an O-ring or the like used for sealing purposes. Disposed in each cylinder 50 and associated liner 51 is a piston 52 having a suitable sealing means 54 with a respective cylinder liner 51. On the end of each piston 52 opposite the one-way valve means is the second camming surface formed as a hemispherical ball and socket joint 130 between each piston 52 and a slipper bearing 131. The slipper bearings are designed to engage a double sided swash plate 62 rotating within cavity 64 provided in transmission housing 19. Double sided swash plate 62 is mounted for rotation on shaft 66 which is rotated by a suitable motor, engine, or the like. The slipper bearings 131 can be held in proper alignment beneath each piston 52 by means of a flexible ring such as 132 (refer to FIGS. 9 and 10), however other provisions such as pins or sleeves could also be employed. Vent holes 93 may optionally be provided in the flexible ring 132 to allow lubricant and cooling liquid to pass through the ring. To reduce side wear and friction on piston 52, a wear sleeve 160 can be used (FIG. 10).

In cylinder housing 18, shaft 66 is journaled for rotation by a suitable journaling means 68 which includes a shaft bearing 72. Similarly, in cylinder housing 18', shaft 66 is journaled for rotation by a suitable journaling means 68' which includes a shaft bearing 69. Coolant inlet port 30 is connected by a bore 74 to journaling means 68. Likewise, coolant inlet port 30' is connected by a bore 74' to journaling means 68'. As shown in FIG. 6, shaft bearings 72 and 69 include channels 82 along the interior surface thereof between which bearing surfaces 84 for shaft 66 are located. Shaft bearing 69 contains a cross bore 71 to form a fluid path between bore 74' and a channel 82. The channels 82 along the interior of shaft bearings 72 and 69 open into cavity 64 within the transmission housing 19 which leads to a coolant outlet port 94 as shown in FIG. 2. In cylinder housing 18', below journaling means 68' is a sealing ring 88 to contain the fluid within the pump. Mounted into annular recesses 73 and 73' in cylinder housings 18 and 18', are rigid, movable, multi-piece segmented thrust bearing rings 75 and 75' and elastic supporting pad rings 77 and 77' (FIGS. 2, 7 and 8). The annular recesses 73 and 73' are in fluid communication with the cavity 64 within the transmission housing 19. The segmented thrust bearing rings 75 and 75' have gaps 91 between the individual segments for lubricating fluid to flow therethrough forming a fluid path between channel 82 in shaft bearing 72 and 69 and cavity 64. The gaps 91 also serve to supply fluid to segmented thrust bearing rings 75 and 75' for purposes of cooling and lubricating. Perpendicular to the axis of shaft 66 and spaced above the double sided swash plate 62 are mounted a pair of flat, polished shaft bearing rings 79 and 79', which rotate along with shaft 66. As shown in FIGS. 2 and 7, segmented thrust bearing rings 75 and 75' as well as support pads 77 and 77' have bores 95 perpendicular to their radial plane that serve to loosely key these components to pins 81 and 81 '. Said pins are pressed, bonded, or fastened into the cylinder housings 18 and 18'. Keyed in place by the pins 81 and 81', the segmented thrust bearing rings 75 and 75' are prevented from rotating along with the rings 79 and 79'. Loosely keyed in this way and mounted on the compliant elastic supporting pads 77 and 77', the segmented thrust bearing rings 75 and 75' align themselves in an efficient inclined manner to form a thin wedge shaped hydrodynamic lubricating fluid film layer that effectively lifts and separates them from the rotating shaft bearing rings 79 and 79'. By virtue of the present invention, this fluid film will form whether the shaft 66 rotates in either a clockwise or counter-clockwise direction. Segmented thrust bearing rings 75 and 75' preferably are fabricated from polymer materials such as polyamide-imide or polyetheretherketone, or other highly polymerized organic materials. Elastic support pad rings 77 and 77' are preferably fabricated from elastomers such as polyurethanes or Neoprene, or the like. To seal along the mating faces of galleries 14 and 14', valve housings 16 and 16', cylinder housings 18 and 18', and transmission housing 19, sealing means 100 are provided. Typically, each sealing means

100 is a suitable O-ring provided in a circular channel in one of the mating faces.

Double sided swash plate 62 includes a pair of camming surfaces in the form of a circumferential ramp surfaces 102 and 102' which extend from lower-most surface portions 104 and 104' to upper-most surface portions 106 and 106'. In the preferred embodiments, ramp surfaces 102 and 102' are at the same opposite angles relative the axis of the shaft 66. Thus, as swash plate 62 is rotated, each piston 52 or 52' is raised by contact with ramp surfaces 102 and 102' to provide a pumping action for the pumped fluid. The fluid coming into the pump 10 under pressure refills and lowers the pistons. Likewise, springs or other mechanical means could be employed to push the pistons downwards to effect their refill. Surfaces 102 and 102' can either be an integral part of swash plate 62, separate disks bonded in place, or applied surface coatings.

Pump 10 is specifically designed for the pumping of corrosive liquids such as seawater or other aqueous solutions. For this reason, the elements of pump 10 are specially constructed to be non-corrodible while still operating effectively without significant wear. It should also be appreciated that these materials are usable in a pump according to the present invention due to the cooling and lubrication of the coolant liquid conducted through pump 10. In general, with the exception of the sealing means (which are generally elastomers) and shaft 66, double sided swash plate 62, shaft bearing rings 79 and 79', and cylinder liners 51 , which are currently fabricated from a corrosion resistant metal (such as stainless steel) due to the high forces generated (it should be noted that items 66, 62, 79, 79'and 51 could also be of a material covered by a plastic selected from the below identified polymer group or possibly of a suitable plastic, composite, or ceramic material with fiber reinforcing), the remaining elements of pump 10 are made of organic materials which are preferably selected from the polymer group consisting of epoxies, polyvinyl chloride, acetal, polyester, polyamide, polyamide-imide, teflon, ultra high molecular weight polyethylene, polyurethane, polyetheretherketone, polycarbonate, 12

polysulfone (including such materials also having fillers to increase strength or reduce friction).

In particular, the preferred material for galleries, 14 and 14', valve housings 16 and 16', cylinder housing 18 and 18', and transmission housing 19 is a glass reinforced epoxy resin. Polyacetal is advantageously used for constructing inlet and outlet one-way valve means 40 and 42 seats and retainers with the valve ball constructed from polycarbonate. Teflon filled acetal is the preferred material for pistons 52. Polyurethane, Neoprene, or the like, are preferred for constructing the flexible ring 132 and elastic supporting pad rings 77 and 77'. Finally, graphite, teflon and carbon fiber filled polyamide-imide or polyetheretherketone are the preferred materials for shaft bearing 68, shaft bearing 69, piston wear ring

160, slipper bearing 131 , and segmented thrust bearing rings 75 and 75'.

In operation, pump 10 functions in the following manner. Initially, shaft 66 is connected to a suitable motor of the like in order to drive shaft 66 in rotation about its longitudinal axis. In addition, a suitable connection using piping 26A and 26A¹ is made between inlet ports 26 and 26' and the fluid to be pumped, which is supplied under low pressure in this preferred embodiment. Similarly, a suitable connection using outlet piping 28A and 28A' Is also made between outlet ports 28 and 28" and the area to which the fluid at high pressure is to be pumped. Piping 26A, 26A', 28A, and 28A' may be swivelably connected to the inlet and outlet ports respectively. Finally, coolant inlet ports 30 and 30' are connected via piping 30A and 30A' to a suitable source of coolant, such as the fluid under low pressure in inlet piping 26A and 26A'.

In certain applications where seawater is being pumped, such as reverse osmosis desalination systems, the seawater must first be filtered so that the seawater is pressurized to push the seawater through the filters. Typically, the inlet seawater pressure is about 15-150 psi. This pressure must be reduced before delivery of the seawater to cavity 64, therefore the bores 74 and 74' as well as channels 82 in shaft bearings 72 and 69 are appropriately sized so as to restrict the flow. Alternately, a pressurized coolant such as tap water or the like which would be compatible with the pumped fluid can be supplied to ports 30 and 30' to serve as the coolant and lubricant for the pump 10. After the desired plumbing connections are made, shaft 66 is rotated by the motor or the like to cause double sided swash plate 62 to rotate within cavity 64. As double sided swash plate 62 rotates, ramp surfaces 102 and 102' continually contact each piston slipper bearing 131, 131' of a respective piston 52 and 52'. Thus, for example, when piston slipper bearing 131 contacts the lower-most surface portions 104, the associated piston 52 is at the lowest point of its stroke. Then, as ramp surface 102 rotates past a particular piston slipper bearing 131 , the piston 52 and the associated slipper bearing 131 are raised to the uppermost point of the stroke of the piston at the location of upper-most surface portions 106 as depicted in FIG. 2. As ramp surface 102 contacts each piston slipper bearing 131 during the upward movement of the associated piston 52, the reaction force is balanced by the piston 52' and piston slipper bearing 131 ' simultaneously being contacted by ramp surface 102'. Continued rotation of ramp surfaces 102 and 102' allows pistons 52,

52' to complete a downward stroke to the lower-most points at the location of lower-most surface portions 104 and 104'. For example, piston 52 is forced downwards by the pressure of the liquid in inlet piping 26A as the liquid flows past inlet one-way valve means 40. It should be appreciated that the pressure of the liquid inlet piping 26A must be greater than the pressure on the opposite side of piston 52 in cavity 64. As the pressure in cavity 64 is created by the coolant liquid flowing in piping 30A which comes from inlet piping 26A, bores 74', as well as channel 82 in shaft bearing 72, reduce the pressure before delivery to cavity 64. Typically, where the inlet pressure is 15-150 psi, and, preferably 15-60 psi, the pressure in cavity 64 is reduced to about 2 to 10 psi.

During the downward stroke of piston 52 or 52' as fluid is forced into the associated cylinder 50 or 50' from inlet ports 26 and 26' and inlet channel 32 through inlet one-way valve means 40 through inlet bore 46, the pressure of the fluid keeps outlet one-way valve means 42 closed. For example, as soon as piston 52 starts its upward stroke, the fluid contained in cylinder 50 is further pressurized and causes inlet one-way valve means 40 to close and outlet one-way valve means 42 to open. The fluid is then pumped from cylinder 50 through outlet bore 48 and outlet one-way valve means 42 to outlet channel 34 and outlet ports 28 during the upward stroke of piston 52. It should be appreciated that sealing means 54 for piston 52 can allow some leakage without adversely affecting the operation of pump 10 where the fluid being pumped is also used as the coolant. Thus, leakage past piston 52 does not introduce any new or harmful fluid into pump 10, and the fluid in pump 10 already is properly disposed of by a suitable connection to coolant outlet 94.

As shaft 66 rotates, friction is developed between shaft 66 and bearings 72 and 69. The friction is usually low, and is a consequence of the shearing of the fluid films that are held by chemical forces to the opposing solid surfaces. At the same time friction is developed, coolant fluid is conducted through coolant inlet ports 30 and 30' and bores 74 and 74' to journaling means 68 and 68'. This coolant fluid is then conducted through channels 82 of shaft bearings 72 and 69. This coolant liquid serves not only to cool bearings 72 and 69, but due to the materials of construction and operating clearances of shaft 66 and bearings 72 and 69, the coolant further serves to reduce the friction generated between these surfaces by forming a hydrodynamic film that separates the moving surfaces from one another. From channels 82, the coolant enters cavity 64 within the transmission housing 19 by passing through the gaps or cooling grooves 91 formed between segmented thrust bearing rings 75 and 75'. While passing through the gaps 91 in segmented thrust bearing rings 75 and 75', the coolant liquid serves to cool and lubricate the surfaces of the segmented thrust bearing rings 75 and 75' as well as shaft bearing rings 79 and 79'. In cavity 64, the coolant liquid similarly serves to both cool and lubricate ramps 102 and 102', slipper bearings 131, and pistons 52 and 52'.

The slightly pressurized coolant liquid in cavity 64 finally exits pump body 12 through coolant outlet port 94. The fluid leaving pump body 12 via coolant outlet port 94 should not be back-pressured so that substantial pressure is not exerted on the ends of the pistons 52 extending into cavity 64, which pressure would hinder refilling during an intake stroke. Typically, the coolant from coolant outlet port 94 is directed to a drain or to the suction side of a feed pump used to supply pressurized fluid to pump 10. In a case where seawater or brackish water is pumped, coolant outlet port 94 may be connected back to the sea, river, or bay. During the normal course of operation there may and often will be times when the axial thrust associated with the pistons 52 in cylinder housing 18 will not precisely match the thrust from the pistons 52' in cylinder housing 18'. To support this imbalanced axial load, the segmented thrust bearing rings 75 and 75' are dynamically actuated by the fluid supplied to the pump 10 under pressure. During normal operation, a portion of the fluid supplied under pressure to the pump 10 is diverted to port 30 using suitable plumbing means. The fluid is conveyed through bore 74 to the cavity 83 above the end of shaft 66 and then conducted through the angled radial bores 85 to the cavity 87 formed above the elastic supporting pad ring 77 (FIGS. 2 and 5). This cavity allows the fluid under pressure to act across the majority of the surface of supporting pad 77 and not just the area exposed beneath bores 85. The elastic supporting pad ring 77 is configured so as to form a snug fit, sealing against the sides of annular bore 73. The compliant nature of pad ring 77 and the loose fit of pins 81 enables the pad ring 77 to move, flex or slide axially in bore 73 and push the segmented thrust bearing ring 75 against shaft bearing ring 79. Each individual segment of thrust bearing ring 75 is loosely fit into bore 73, enabling them to incline slightly for form an efficient wedge-shaped hydrodynamic lubricating film, with the lubricating fluid being supplied through gaps 91. This thrust in turn pushes the shaft 66 and associated shaft bearing ring 79' against the segmented thrust bearing ring 75'. The elastic support pad ring 77' beneath the segmented thrust bearing ring 75' is stationary, but has sufficient resiliency and clearance on its outer and inner diameters to allow it to spring or flex under ring 75'. This springing action enables the individual segments of the segmented thrust bearing ring to evenly distribute the thrust or load and for the individual segments to incline slightly to form an efficient wedge shaped hydrodynamic lubricating film. When in operation, this dynamic bearing system effectively maintains the axial position of shaft 66 and double sided swash plate 62 within the pump transmission chamber 64 without the need of added springs or mechanical means. Further, the dynamic nature of the design automatically compensates for any unwanted play or backlash introduced by wear, shrinkage, expansion or tolerance variations of the individual components. The tandem positive displacement pump of the present invention is suitable for pumping fluids having inlet pressures in the range of from 15 to 150 psi, preferably 15 to 60 psi. It is anticipated that pump 10 of the present invention can be used to pump approximately 10 to 600 liters per minute of a wide range of corrosive and non-corrosive fluids over a pressure range of approximately 0 to 1 ,200 psi when operated at between about 1000 to 1750 rpm. The lower tensile strengths of plastics, relative to metals, limits the operation of pump 10 shown in FIG. 2 to approximately 1 ,500 psi. However, with proper fiber reinforcement, this limit can be increased to about 1 ,500 to 2,500 psi. The thermoplastic nature of some of the materials used in pump 10 also limits the operating temperature of the fluid being conveyed to approximately 150°F. However, by switching these elements to a thermoset material or a thermoplastic with higher distortion temperatures, this temperature limit could be increased to approximately 200°F.

The satisfactory hydrodynamic lubrication of the bearings within pump 10 depends on a combination of operating pressure, temperature, the viscosity of the fluid pumped, and the operating rotational speed. For particular applications, the clearances and size of the bearing components within pump 10 can be altered during its manufacture to suit specific combinations of pressure, temperature, viscosity and rotational speed.

Pump 10 of the present invention provides a light weight, compact, reliable and efficient pump which will operate over an extended period of time with little or no maintenance. This compact design, efficiency and reliability is achieved by use of the tandem configuration with a double- sided swash plate, the unique flow through cooling design, the shaft centering aligning means, and the non-metallic bearing and construction materials. The overall power efficiency of the transmission in the present invention is greater than that of a conventional single ended pump because the second set of plain bearings that normally are provided to support the back side of a conventional one-sided swash plate against downward loads exerted by the pistons are not required. In the present invention, this downward thrust is offset by the opposing set of pistons that are concurrently generating useful work, eliminating the parasitic frictional losses associated with the plain support bearings of a conventional single- ended pump with a one-sided swash plate.

The liquid cooling flow rate for pump 10 is between about 1 to 120 1/min., depending on speed, discharge capacity, pressure, and temperature. These non-metallic bearing materials can be operated at loads and speeds that are 10 to 20 times higher than loads obtainable under dry operating conditions. In addition, the low cost construction and noncorrodible nature of pump 10 make it ideal for use in commercial applications such as reverse osmosis and chemical processing. The light weight and compact nature of pump 10 greatly increase its utility in portable applications such as trailer mounted systems and/or disaster relief water purification equipment. Furthermore, the fluid cooled drive-end could be used in other systems requiring a rotary power source converted into a linear displacement such as hydraulic tool systems and motors. The use of the double-ended configuration of the present invention and the associated high speed (up to about 1750 rpm) enables the discharge capacity of the pump to be increased by a factor of two with only about a 60 to 70 percent increase in the overall size and weight of the pump. Conventional metallic pumps of comparable discharge capacity normally are limited to operational speeds below about 1000 rpm and weigh from about 4 and up to about 10 times more. The higher operating speed of the present invention allows for direct mounting to standard 1750 rpm motors, thereby eliminating the added sheaves, drive belts and/or transmissions, frames and shrouds normally employed with conventional ^" pumps to achieve the slower (below about 1000 rpm) operating speed.

The tandem design of pump 10 could also be easily used to serve as two separate pumps supplying pressured fluid for two applications simultaneously. In such an application, outlet ports 28 and 28' would be connected with suitable plumbing means to the two applications requiring pressurized fluid. There would be a limit on how large a pressure difference between the two applications could be tolerated. To extend the range of pressure difference to be accommodated, cylinder housings 18 and 18' could each be fitted with pistons 52 and 52' of different diameters, with the specific diameters sized to give a specific ratio of piston projected areas to balance the pressure difference in the opposing cylinder housings 18 and 18'. The tandem pump of the invention can do the work of two separate pumps, but at a much reduced overall size and weight.

Furthermore, the double sided swash plate of the present invention reduces approximately by one half the number of bearing surfaces in the transmission, relative to two separate single-ended pumps, thereby significantly reducing the associated frictional losses. For example, the pump body of the tandem pump of the invention does not require wear pads on either side of the double sided swash plate, but such wear pads would be required with a single sided swash plate. It should be appreciated that the ramped surfaces 102 and 102' of double side swash plate 62 could be of various geometries including multiple ramps to give more than one stroke per revolution. The stroke length could also be varied by changing the slope of the ramps, but keeping both ramp faces 102 and 102' at the same angle but of opposite signs relative to one another so as to balance the load. In addition, the number of cylinders, their diameters, and their spacing could be altered in order to change the output capacity of the pump.

A particularly important aspect of the invention is the use of pressurized inlet fluid to refill the cylinders. This feature markedly reduces the complexity of the pump design, eliminating the need for crank arms, wrist pins, refill springs, yokes, ball joints, etc., and increases the pump's resistance to wear induced failure. In particular, the pistons can be considered equivalent to brushes in a motor. The same design advantage is associated with the mechanism to load the segmented thrust bearing rings 75 and 75' against the shaft bearing rings 79 and 79'. Even though this will normally not happen, the pump of this invention could lose 0.15 or more inches from the second camming surface on the pistons or the segmented thrust bearing rings without reducing the pump's volumetric efficiency or introducing any unwanted play or backlash. Thus, while the present invention has been described above with respect to an exemplary embodiment thereof, it will be understood by those of ordinary skill in the art that variations and modifications can be effected within the scope and spirit of the invention.

Claims

What is claimed is:

1. A tandem positive displacement piston pump for pumping a fluid, characterized by: pump bodies including plate means, said plate means comprising an outer plate and an inner valve plate mounted thereto; an axially rotating member in said pump bodies; a swash plate mounted to said rotating member for joint rotation with the axially rotating member, said swash plate having at least two opposed first camming surfaces; a plurality of cylinders mounted in said plate means, said cylinders of one pump body being substantially axially aligned with the cylinders of another pump body and disposed on opposite sides of the first camming surfaces of the swash plate; a piston in each of said cylinders, each of said pistons having a piston head at one end thereof and a second camming surface at its opposite end, said piston head being located in a piston head chamber at one end of its said cylinder, each of said second camming surfaces contacting at least one of said first camming surfaces, with the second camming surfaces of pistons of said first pump body contacting said first of the first camming surfaces of the swash plate and with the second camming surfaces of pistons of another pump body contacting said second of the first camming surfaces of the swash plate so that rotation of said rotating member periodically overcomes the opposing force of the fluid pressure acting against each of said piston heads and thereby causes each of said piston heads to reciprocate axially in its said piston head chamber; inlet valve means including an exposed inlet port in said outer plate and a plurality of inlet one-way valve means in said valve plate corresponding to the number of said cylinders, with each of said inlet valve means being associated with a respective one of said cylinders, inlet channels creating flow communication between said inlet ports and said plurality of said inlet valve means, each of said inlet valve means being in flow communication with a respective piston head chamber; an exposed outlet port in said outer plate and a plurality of outlet valve means in said valve plate corresponding to the number of said cylinders with each of said outlet valve means being associated with a respective one of said cylinders, outlet channels creating flow communication between said outlet ports and said plurality of said outlet valve means, each of said outlet valve means communicating with a respective piston head chamber; fluid paths formed by the elements of said inlet ports and said inlet channels and said inlet valve means and said piston head chambers and said outlet valve means and said outlet channels and said outlet ports, of all of said elements of said pump bodies which comprise said fluid paths being made of a material which is non-corrodible in the pumped fluid; and a lubricating and cooling means for lubricating and cooling said axially rotating member and said first and second camming surfaces, said lubricating and cooling means including a liquid coolant and lubricant in contact with said axially rotating member and said first and second camming surfaces via a coolant passageway communicating with said axially rotating member and with said first and said second camming surfaces, and a coolant outlet passage downstream from said first and said second camming surfaces and exiting from said pump body.

2. The piston pump of claim 1 further characterized by a centering means for aligning the position of the axially rotating member relative to at least one pump body, including an axially rotatable shaft bearing with an upper surface mounted perpendicularly to the axis of the axially rotating member, a substantially nonrotatable segmented thrust bearing with a lower surface movably attached to a relatively movable, substantially compliant support pad positioned within a recess in the pump body so that the upper surface of the segmented thrust bearing is adjacent to the upper surface of said axially rotatable shaft bearing, said segmented thrust bearing having one or more gaps that allow liquid coolant and lubricant to flow therethrough, and the pump body defining a bore through which fluid may flow to the recess, so that when the axially rotatable shaft bearing and the axially rotating member rotate about their axis, the substantially compliant support pad moves axially in response to fluid pressure from fluid that flows through the bore to the recess and the segmented thrust bearing contacts the upper surface of the rotatable shaft bearing to cause axial position adjustment of the axially rotatable shaft.

3. The piston pump of claim 2 characterized in that the substantially compliant support pad is formed from a noncorrodible elastomeric material.

4. The piston pump of claim 2 characterized in that the segmented thrust bearing is formed from a noncorrodible material selected from the polymer group comprising: epoxies, acetal, polyester, polyamide, polyamide-imide, polytetrafluoroethylene, ultra high molecular weight polyethylene, polyurethane, polyetheretherketone, polycarbonate, and polysulfone.

5. The piston pump of claim 2 characterized in that the centering means further comprises a substantially fixed compliant support pad within a recess in the other pump body and a second substantially nonrotatable segmented thrust bearing with an upper and lower surface, said second substantially nonrotatable segmented thrust bearing being movably attached to the substantially fixed compliant support pad and having its lower surface in the recess and its upper surface adjacent to an axially rotatable shaft bearing that is mounted on the shaft, said second segmented thrust bearing having one or more gaps that allow liquid lubricant to flow therethrough, wherein when the axially rotatable shaft bearing rotates about its axis, the upper surface of the second segmented thrust bearing contacts the upper surface of the axially rotatable shaft bearing and the substantially fixed compliant support pad supports the second segmented thrust bearing against axial loads to cause axial position adjustment of the axially rotatable shaft.

6. The piston pump of claim 1 characterized in that a portion of the fluid being pumped is used as the liquid coolant and lubricant.

7. The piston pump of claim 2 characterized in that a portion of the fluid being pumped is used as the liquid coolant and lubricant.

8. The piston pump of claim 1 characterized in that the two first camming surfaces of the swash plate are angled with respect to a plane perpendicular to the axis of the axially rotating member such that the second of the first camming surfaces has a ramp slope angle of equal magnitude, but opposite sign of the ramp slope angle of the first of the first camming surfaces.

9. The piston pump of claim 1 characterized in that the swash plate is substantially centered between the cylinders in the first and second pump bodies and the cylinders of the first and second pump bodies are substantially axially aligned.

10. The piston pump of claim 1 characterized in that the fluid being pumped enters the piston head chambers of the cylinders at a pressure in the range of about 15 to about 150 psi.

11. The piston pump of claim 10 characterized in that the fluid being pumped enters the piston head chambers of the cylinders at a pressure in the range of about 15 to about 60 psi.

12. The piston pump of claim 1 characterized in that the pumped fluid is a corrosive liquid.

13. The piston pump of claim 12 characterized in that the pumped fluid is seawater or an aqueous solution containing salts.

14. The piston pump of claim 2 characterized in that the pumped fluid is a corrosive liquid.

15. The piston pump of claim 14 characterized in that the pumped fluid is seawater or an aqueous solution containing salts.

16. The piston pump of claim 1 characterized in that the inlet port of one pump body is connected by piping means to a first fluid to be pumped and the inlet port of the second pump body is connected by piping means to a second fluid to be pumped.

17. The piston pump of claim 2 characterized in that the inlet port of one pump body is connected by piping means to a first fluid to be pumped and the inlet port of the second pump body is connected by piping means to a second fluid to be pumped.

18. The piston pump of claim 5 characterized in that the inlet port of one pump body is connected by piping means to a first fluid to be pumped and the inlet port of the second pump body is connected by piping means to a second fluid to be pumped.