Patent Publication Number: US-3878767-A

Title: High pressure radial piston fluid translating device and cylinder construction therefor

Description:
United States Patent 1 3,878,767 Engel Apr. 22, 1975 1 HIGH PRESSURE RADIAL PISTON FLUID 3.799.034 3/1974 Haglund .9|/49o TRANSLATING DEVICE AND CYLINDER CONSTRUCTION THEREFOR [75] Inventor: William Karl Engel, Peoria. 111.  
 [73] Assignee: Caterpillar Tractor Company,  
 Peoria, 111.  
 [22] Filed: Dec. 17, 1973 [211 Appl. No.: 425,192  
 [52] US. Cl. 91/490 [51] Int. Cl. F0lb 13/06 [58] Field of Search 91/490; 417/273 [56] References Cited UNITED STATES PATENTS 2,461,235 2/1949 Raymond 417/273 2,815,718 12/1951 Avery 91/498 3,274,946 9/1966 Simmons 91/490 Primary Examiner-William L. Freeh Attorney, Agent, or Firm-Phillips, Moore, Weissenberger, Lempio &amp; Strabala [57] ABSTRACT A fluid pump or motor having a rotor with hollow radially directed piston spokes, each extending into a separate one of a plurality of cylinders which ride against an eccentrically positioned circular race, is provided with cylinders having an inside diameter which progressively decreases towards the radially innermost end of the cylinders to compensate for expansion of the cylinder walls in the course of each working stroke when operated under extreme high pressures. Problems with excessive wear or leakage at the zone of contact between the cylinders and pistons are thereby alleviated.  
 8 Claims, 10 Drawing Figures PATENTED APR22 I975 szsmzora HIGH PRESSURE RADIAL PISTON FLL&#39;ID TRANSLATING DEVICE AND CYLINDER CONSTRUCTION THEREFOR BACKGROUND OF THE INYENI&#39;ION This invention relates to fluid pumps and motors. herein referred to as fluid translating devices. of the radial piston form and to a cylinder configuration therefor which facilitates operation of such devices under very high fluid pressure conditions.  
  Radial piston pumps or motors are distinguishable from other types of fluid translating devices by having a rotor provided with hollow radial piston spokes. each of which extends into a separate one of a plurality of cylinders. The cylinders ride against a stationary circular race which is eccentrically positioned relative to the axis of the rotor. As the rotor revolves. each cylinder must reciprocate relative to the associated piston and either a pumping or motor action may be realized depending on the form of the fluid and mechanical connections which are made to the device.  
  If all other parameters are unchanged. the power developed when such a device is operated as a motor or the pumping capacity when the device is operated as a pump is dependent on the size of the device. However. increasing the size of such devices in order to realize higher power output or greater flow rates may not be always possible. The device may have to be fitted into a constricted space within other mechanisms with which the device is associated. In other situations where space requirements do not rule out a larger device. increased size may still be undesirable simply because large. heavy pumps or motors are more difficult to handle and more costly to manufacture.  
  An alternative technique for realizing higher power outputs or greater flow rates is to design the device for operation at higher fluid pressures and higher rotational speeds. One device embodying the present invention. for example. has been designed to operate at speeds of up to 12.000 revolutions per minute with internal fluid pressures which may be as high as 7.500 pounds per square inch. Compactness. together with high operating efficiency may be realized under these conditions but certain other problems are encountered that were not of significant consequence in more conventional devices. Most notably. normally rigid components of the device may be distorted significantly by fluid pressures ofthis general magnitude. This presents a particularly serious problem in connection with the fit between the pistons and the cylinders which are carried thereon and which reciprocate relative thereto.  
  In fluid translating devices in general. it is recognized that there is some optimum degree of fit between a piston and the associated cylinder at which fluid leakage is minimized without causing excessive friction and wear or risking binding from an overly tight fit. In conventional relatively low pressure systems. the inside diameter of the cylinder and the outside diameter of the piston are simply selected to provide the preferred fit and it is assumed. without serious practical effect. that these predetermined dimensions are maintained in operation. This assumption cannot be relied upon in the case of extreme high pressure systems. Both the piston and the cylinder can be significantly expanded by the extremely high fluid pressures. Moreover. the pressure expansion of the piston and that of the cylinder do not tend to be proportionate. Expansion of the cylinder is minimal at the relatively high strength head end and is progressively greater towards the open, or skirt end of the cylinder. As a consequence. the designer is seemingly faced with a choice of providing an undesirably close fit at certain portions of the working stroke at which friction and wear are excessive and the risk of binding is present or else providing an undesirably loose fit at which fluid leakage will be severe at least at certain portions of the working stroke.  
 SUMMARY OF THE INVENTION The problems arising from differential expansion of different portions of the cylinder of a radial piston fluid translating device under high pressure conditions are alleviated by forming the cylinders with a bore which is slightly conical. under unpressurized conditions. rather than being strictly cylindrical. The inside diameter of the cylinder becomes larger towards the head end thereof and smaller towards the open or skirt end. Accordingly. fluid pressure-induced expansion of the cylinder in the course of each stroke. which expansion is greatest at the open or skirt end of the cylinder and minimal at the head end. tends to maintain a desired fit between the cylinder and associated piston throughout the working stroke.  
  In a preferred form of the invention. the cylinders and pistons are proportioned to exhibit a sizable clearance fit near the head end of the cylinder under unprcssurized conditions. This clearance tapers to a lesser clearance or. in some cases. to a slight interference fit near the opposite end ofthe cylinders. The wall of each cylinder is made as thick as is practical to minimize cylinder expansion to the extent possible but the piston wall is proportioned to undergo a sizable expansion. under pressure. to reduce or even eliminate the clearance of the cylinder under operating conditions in the head region. This allows a substantial degree of taper to be provided in the cylinder bore for the above described purpose without creating a severe interference fit at one end of the cylinder under unpressurized conditions.  
  It is an object of this invention to provide a more compact and efficient construction for radial piston fluid translating devices by facilitating operation thereof at high speeds and under high fluid pressure.  
  It is another object of this invention to reduce friction. wear and leakage at the piston and cylinder interfaces in a radial piston fluid translating device.  
  It is still another object of this invention to provide for a more uniform fit between the pistons and cylinders of a radial piston fluid translating device throughout all portions of the working cycle.  
  The invention together with further objects and advantages thereof will best be understood by reference to the following description of a preferred embodiment taken in conjunction with the accompanying drawings.  
 BRIEF DESCRIPTION OF THE DRAWINGS In the accompanying drawings:  
  FIG. 1 is a cross section view through a radial piston fluid translating device embodying the present invention;  
  FIG. 2 is a partial axial section view of the device of FIG. I taken along line IlII thereof.  
  FIG. 3 is an exterior view of the head end of a cylinder of the device of FIGS. 1 and 2;  
  FIG. 4 is a view of the cylinder of FIG. 3 taken along line IV-IV thereof.  
  FIG. 5 is a view of the cylinder of FIG. 3 taken along line thereof:  
  FIG. 6 is an axial section view of a cylinder similar to that of FIGS. 3 to 5 shown disposed on an associated piston.  
  FIG. 7 is a diagrammatic view illustrating. in exaggerated form. the nonuniform internal diameter of the cylinder of FIGS. 3 to 5:  
  FIG. 8 is a graphical illustration of typical changes in the inside diameters of a prior art bore cylinder and a piston in the course of a working stroke in the presence of very high fluid pressures:  
  FIG. 9 is a graphical illustration of changes in the diameters of a straight bore cylinder and piston in the course of a working stroke in the presence of very high fluid pressure wherein the cyclinder walls have been thickened relative to the cylinders of FIG. 8. and  
  FIG. I0 is a graphical illustration of changes in the diameters of the cylinders and pistons of the present invention in the course of a working stroke in the presence of very high pressures.  
 DESCRIPTION OF A PREFERRED EMBODIMENT Referring initially to FIGS. 1 and 2 of the drawings in conjunction. a fluid translating device 11 embodying the invention may have a general construction essentially similar to that of conventional devices except insofar as components may have a more massive highstrcngth construction to adapt to the extremely high speed and high pressure conditions and except insofar as modifications are present to accommodate to the novel configuration of the cylinders 12. Accordingly. the overall construction of the device 11. aside from the cylinder configuration. will be only briefly described to provide a background for an understanding of the present invention.  
  Device ll may typically have a stationary cylindrical pintle 13 which in this example is provided with a flange 14 that forms one endwall of a rotor chamber 16. A circular opposite endwall 17 is spaced from flange l4 and and annular race retainer 18 extends between the outermost portions ofthe two endwalls in coaxial relation therewith. To transmit drive to or from the device 11 depending on whether it is operated as a pump or motor. a rotatable shaft 19 extends through an opening 21 at the center ofendplate l7 and into a bore 22 in the adjacent end of pintle 13. the shaft beingjournaled for rotation by a bearing 23 and an additional bearing 24 situated in an opening 21 and in bore 22 respectively.  
  An annular rotor 26 is disposed coaxially on the end of pintle 13 within rotor chamber 16 and is provided with equiangularly spaced radially directed hollow spokes 27&#39; of which the outer ends constitute the pistons 27 of device 11. there being five such pistons in this example. An end portion 28 of rotor 26 extends radially inward between pintle l3 and endplate l7 and carries splines 29. which engage splines 31 on shaft 19 to couple the rotor and shaft for synchronous rotation. Each ofthe spokes 27&#39; has a radially extending internal passage 32. A piston ring 33. having a spherical outer surface 34. is disposed coaxially on the outermost end of each piston 27.  
  Cylinders 12 each have a substantially cylindrical skirt or open end portion 36 into which an associated individual one of the pistons 27 and piston rings 33 is received. Each cylinder has a substantially closed outer or head end 37 which rides against a cylindrical inner surface 38 of a race 39 that encircles the cylinders and which is contained between endwalls l4 and 17 within race retainer 18.  
  As best seen in FIG. 1 in particular. the race 39 is positioned eccentric-ally relative to the axis of rotation of rotor 26 and thus each cylinder 12 is constrained to reciprocate relative to the associated piston 27 as the rotor turns. In this example of the invention. the race 39 is of the form which is shiftable along a diameter of the device 11 within race retainer 18 to shift the direction of the eccentricity and thereby reverse rotor rotation when the device is being operated as a motor or to reverse flow direction when the device is being operated as a pump. For this purpose. the race retainer 18 has flat sections 41 at diametrically opposite points and the race 39 has matching flat outer surfaces 42 disposed thereagainst. Accordingly the race 39 may be shifted sidewardly within retainer 18 and may beheld at a selected position by admitting pressurized fluid to the region 43 between the race and retainer at one side while draining fluid from the similar region 43&#39; at the other side of the device. It should be understood that devices 11 may also be constructed with a fixed race if this particular technique of reversing operation is not utilized.  
  To provide for admission of fluid to those of the cylinders 12 that are moving radially outward on the associated postons 27 at any given time. a first group of four passages 44 extend longitudinally within pintle l3 and each such passage terminates at an arcuate slot 46 in the cylindrical surface of the pintle. Slot 46 is positioned to communicate with the internal passages 32 of these pistons 27 at which cylinders 12 are moving towards the rotational axis of rotor 26. A second group of four passages 47 within the pintle l3 terminate at a similar but oppositely facing slot 48 which communicates with pistons carrying cylinders that are moving outward thereon. Thus passages 47 constitute the fluid outlet for device 11 and passages 44 constitute the fluid inlet. If race 39 is shifted as described above. passages 47 become the inlet and passages 44 become the outlet. If fluid under pressure is forced into the inlet passages. then the device 11 may function as a motor with shaft 19 being coupled to mechanism which is to be driven. Conversely. if the shaft 19 is forcibly turned by external drive means. the device 11 may function as a pump in that fluid will be drawn into the inlet passages and will be forcibly discharged from the outlet passages. the position of the race 39 determining the direction of the pumping action.  
  Referring now to FIGS. 3. 4 and Sin conjunction. the head portions 36 of the cylinders 12 have an outer surface 49 with a curvature conforming substantially to that of the inner surface of the race 39 although the cylinder surface curvature may be slightly greater than the race curvature to entrap fluid between the surfaces 49 and the race to establish a hydrodynamic bearing action as is understood in-the art. A hydrostatic bearing effect may also be provided by forming a central recess 51 in the cylinder surfaces 49 which recess is communicated with the interior of the cylinder by a small aperture 52. This causes the high fluid pressures within the cylinders 12 to act between the cylinders and race 39 in such a manner as to prevent extreme direct pressure of the cylinders against the race from the combined ef fects of internal fluid pressure and centrifugal force. Cylinders 12 are further provided with pads 53 on the innermost surface of head portions 36. Pads 53 ride on annular retainer rings 54 disposed in rotor chamber 16 as shown in FIG. 2 to assure that the cylinders remain adjacent the race under all conditions.  
  As previously pointed out. the device ll has been designed to be extremely compact relative to the output power when operated as a motor or to the pumping rate when operated as a pump. and this dictates operation at extremely high speeds and under extremely high fluid pressures. One unit of the device 11 has a race 39 with an inside diameter of only 6 inches while providing a maximum piston displacement of 6.28 cubic inches per revolution. and may operate at speeds up to 12.000 RPM with internal fluid pressures of up to 7.500 psi. Under&#39;these extreme conditions. effects which are of minor concern in more conventional devices become very significant design considerations. For example. if the cylinders 12 each weigh 0.8 pounds at rest. centrifugal force at 12.000 RPM within a six inch race causes each single cylinder to exert 9.821 pounds of force against the race 39. Thus. the race and supporting structure must have an extremely high strength construction.  
  The extreme high pressures which may be present in the device 11 also create unique problems. In the conventional design of pumps. motors and the like. it is customary to assume that components formed of nominally rigid material such as metal have constant dimensions in operation and under normal circumstances. this is a reasonably valid assumption if components are made sufficiently massive. Given the conditions which may exist within the present device 11 as discussed above. the conventional practice of providing for rigidity of components by providing adequate thickness or massiveness may not be possible with respect to certain components and most notably with respect to the piston 27 including piston rings 33 and the associated cylinders 12. Given the objective of compactness of the device as a whole. these elements cannot be made sufficiently thick and massive to avoid significant distortion in the presence of extreme high fluid pressures. In instances where dimensional limits do not prevent a very massive construction of the pistons and cylinders. this may still be undesirable for other reasons. such as added cost. increased inertia. momentum and the severe centrifugal force effects.  
  Referring now to FIG. 6. for efficient operation without excessive wear or fluid leakage. there exists some optimum degree of fit which should exit between the piston ring 33 and the inside wall 56 of the cylinder 12. Conventionally. a cylinder has a uniform bore and the outer diameter of the piston ring and the inner diameter of the cylinder are simply selected to provide the desired fit. If the device is then operated under extremely high fluid pressure conditions. this desired clearance will not in fact be present except possibly at one specific instant in each working stroke of the cylinder and piston. The reason is that both the piston and the cylinder expand appreciably when the very high fluid pressures are present. but not by the same amount. Still further. the expansion of the cylinder in particular is not uniform at all portions thereof. The  
 open skirt end 36. Thus. the radial enlargement of the inside diameter of the cylinder in response to high fluid pressure is minimal when the piston is at the top dead center position and becomes increasingly larger as the piston ring 33 recedes therefrom towards the bottom dead center position. Thus the cylinder 12 will cyclically swell and shrink in the course of each piston stroke as the extent of this radial movement is a function of the position of the piston within the cylinder.  
  This effect is shown graphically in FlG. 8 which illustrates the radial growth of a typical conventional piston and straight bored cylinder. in the course of a single piston stroke. in the presence of an unusually high internal fluid pressure. The solid vertical line 57 at the left edge of FIG. 8 represents a piston and cylinder having identical diameters operated at zero pressure. Under such a condition. neither element would expand in the course of a piston stroke between top dead center and bottom dead center. While this condition can probably never be precisely realized in practice. it is reasonably closely approximated in conventional devices which are not subjected to extreme high pressures. Considering now what actually occurs ifthis conventional structure is subjected to high pressure such as 7.500 psi. dashed line 58 represents the radial growth of the piston in response to such pressure which in this example is a radial enlargement of 0.0003 inches. This piston growth would be constant throughout the stroke if it were not affected by the encircling cylinder. Dashed line 59 indicates the radial growth of the cylinder if unaffected by that of the piston. and it should be observed that this varies substantially according to the position of the piston in the cylinder. At the top dead center position. the radial growth of the cylinder is only 0.0001 inches but at the bottom dead center position. the radial growth of the piston is 0.0020 inches. It should be understood that the interference between the piston and cylinder which is depicted above the intersection point 61 of the dashed lines 58 and 59 does not exist in practice since the piston ring cannot actually have a greater diameter than the adjacent inner surface of the cylinder and the actual condition above intersection 61 is that both have some common diameter depicted by dotted line 62 which is intermediate between lines 58 and 59 in the interference fit region.  
  The situation depicted in FIG. 8 is highly undesirable in a fluid translating device. While a possibly acceptable degree of interference fit exists for a small portion of each stroke near top dead center. an increasingly excessive clearance appears each time the piston moves from point 61 towards the bottom dead center position. An unacceptable amount of leakage will occur throughout a major portion of each stroke.  
  FIG. 9 is a graph comparable to FIG. 8 but illustrating what occurs if an attempt is made to counteract the leakage problem exhibited in FIG. 8 simply by thickening the walls of the cylinder within practical limits. The variable cylinder growth represented by dashed line 59a has been reduced substantially and the zone of acceptable fit above point 61a has been extended but not enough to resolve the problem. The radial growth of the cylinder at the bottom dead center position has been decreased to 0.0009 inches. This degree of clearance near the bottom dead center region. 0.0006 inches. is still undesirably high and severe fluid leakage occurs. Thus. mere thickening of the cylinder walls.  
 given size and weight limits. does not necessarily resolve the problem.  
  Considering now an important aspect of the present invention with reference again to FIG. 6. the problems discussed above are alleviated by preferably forming the wall of the cylinder to be as thick as possible within practical limitations and then. in addition. by forming the internal surface 56 of the cylinder to have a tapered or conical configuration wherein the diameter of the cylinder bore is greatest near the head end 37 and becomes progressively smaller towards the open skirt end 36. The degree of taper may be slight. typically of the order of 0.024 inch per foot of cylinder length. and thus. is not readily apparent in a drawing of the scale of FIG. 6. Accordingly. the direction of the taper is illustrated in exaggerated form in FIG. 7.  
  Referring again to FIG. 6. the most advantageous results are obtained if the cylinder bore and piston ring are proportioned so that a sizable clearance. typically ofthe order of(l.00(l-l inches exists at the top dead center position in the absence of fluid pressure. A smaller clearance or an exact fit or. as in this example. a small interference fit. typically 0.000l inches. exists at the bottom dead center position in the unpressurized condition, This is illustrated graphically in FIG. 10. wherein lines 57b and 57h. respectively. indicate the outer diameter of the piston and the inner diameter of the cylinder along the length thereof under unpressurized conditions. It should again be understood that the interference depicted in FIG. 10 at the lower portions of lines 57b and 57h would not actually exist if the cylinder were reciprocated on the piston at unpressurized conditions. but instead both elements are forced to have a single compromise diameter 570 intermediate between the diameters represented by the lines 57!: and 57/).  
  Referring still to FIG. I0. the zone of acceptable interference fit 62b has been extended to cover almost all of the working strokev While some clearance exists in the immediate vicinity ofbottom dead center in this example. it is far smaller than in the previous cases and is tolerable. Although the radial growth of the cylinder under pressure at the bottom dead center position remains the same as that depicted in FIG. 9 in absolute terms. this radial growth relative to that of the piston as depicted by dashed line 58b has been drastically reduced.  
  It may be observed in FIG. 10 that the radial expansion of the piston. 0.0007 inches in this example as indicated by line 58b. is substantially greater than was the case in FIGS. 8 and 9. This is deliberately provided for by utilizing a relatively thinner walled piston or a more elastic piston material. in order to take full advantage of the effects ofthe tapered cylinder bore. If the piston expansion 58h remained as it was in the earlier discussed situations. that is. a radial expansion of 0.0003 inches. then the tapered bore would still provide a more uniform fit of the piston and cylinder throughout the stroke. but would also allow a degree of clearance to exist at portions of the entire stroke. which might be excessive in some devices.  
  This clearance may be eliminated or reduced by using a piston having a greater diameter under unpressurized conditions. in other words. by shifting line 57b to the right in FIG. 10. However. this results in a severe degree of interference fit near the bottom dead center position when pressure is absent. thereby increasing friction and wear and complicating assembly and disassembly. The alternate technique described above. that of forming the piston for a greater expansion under pressure. does not have those disadvantages.  
  Accordingly. to realize the benefits of the invention in its optimum form. the cylinder is provided with a conical bore and the hoop strength of the piston is chosen to provide for a piston expansion under pressure which reduces any excessive clearance which might otherwise be present because of the taper of the cylinder bore.  
  While the invention has been disclosed with respect to a single embodiment. it well be apparent that many variations are possible and it is not intended to limit the invention except as defined in the following claims.  
 What is claimed is:  
  1. In a radial piston fluid translating device for operation with high internal fluid pressure. the combination comprising an annular rotor having at least one radially extending tubular piston. and a cylinder fitted on said piston for reciprocation relative thereto. said cylinder having a head end and having a skirt end with a bore into which said piston extends. at least a portion of said bore having a tapered configuration and being of progressively smaller internal diameter from said head end towards said skirt end 2. The combination defined in claim I wherein said cylinder reciprocates relative to said piston between a top dead center position at which said head end is closest to said piston and a bottom dead center position at which said head end is furthest from said piston. said cylinder bore having an internal diameter at said top dead center position which exceeds the external diameter of said piston in the absence of said high operating fluid pressure.  
  3. The combination defined in claim 2 wherein said cylinder bore has an internal diameter at said bottom dead center position which is less than said external diameter of said piston in the absence of fluid pressure.  
  4. The combination defined in claim 2 wherein said piston is proportioned to expand in response to said high internal fluid operating pressure by an amount which exceeds said clearance between said piston and said cylinder at said top dead center position.  
 5. A fluid translating device comprising:  
 an annular rotor having a plurality of radially extending tubular piston spokes thereon.  
 means for admitting fluid to said piston spokes at one side of said rotor and for releasing fluid from said piston spokes at the other side thereof.  
 an annular race encircling said rotor including said spokes thereof in radially spaced relation therefrom and in eccentric relationship thereto. and  
 a plurality of cylinders each being fitted on a separate one of said piston spokes and having a head end bearing against the inner surface of said race and having a skirt end with a bore into which the associated one of said spokes is received whereby each of said pistons reciprocates on said associated one of said spokes as said rotor turns within said eccentric race. each of said cylinder bores having a substantially conical configuration at a region thereof in which the associated one of said pistons reciprocates with the larger diameter portion of said bores being closest to said head ends of said cylinders.  
  6. A fluid translating device as defined in claim 5 wherein each of said piston spokes has a piston ring disposed coaxially thereon and wherein said larger diameter portions of said cylinder bores are of greater diameter than said piston rings when fluid pressure is absent from said device and wherein said smaller diameter portions of said cylinder bores are of less diameter than said piston rings when said fluid pressure is absent from said device.  
  7. A fluid translating device as defined in claim 5 wherein said piston spokes are proportioned to expand radially in response to operating fluid pressure an amount which is greater than the clearance which ex- UNITED STATES PATENT OFFICE CERTIFICATE OF CORRECTION PATENT NO. 3 878 767 DATED I April 22, 1975 INVE M 1 William Karl Engel ?t is certified that error appears in he above-menhfied patent and that said Letters Patent are heleby corrected as shown below:  
 On the Title Page, Item [73],, change the spelling of the assignee&#39;s corporate name from Caterpillar Tractor Company&#34; to Caterpillar Tractor Co.  
 Signed and Scaled this twenty-second Day of July 1975 [SEAL] A tlesr:  
 I RUTH c. MASON c. MARSHALL DANN Alreslmg Officer Commixsioner of Parcnls and Trademarks i