US3053195A - High pressure hydraulic pump - Google Patents

Description

Sept. 11, 1962 L. R. WILLIAMSON .HIGH PRESSURE HYDRAULIC PUMP 4 Sheets-Sheet 1 Filed April 14. 1959 INWNTO? LAR/(Ml E. M//LLMMSON ffm.

Sept- 11, 1962 l.. R. WILLIAMSON 3,053,195

HIGH PRESSURE HYDRAULIC PUMP 24 25 f@ f Il mrllll E7 '6 INVENTQP 4e/WN e. Ma/AMso/v B1 Sept. ll, 1962 L. R. WILLIAMSON HIGH PRESSURE HYDRAULIC PUMP 4 Sheets-Sheet 3 Filed April 14, 1959 Sept. 11, 1962 L. R. WILLIAMSON 3,053,195'

HIGH PRESSURE HYDRAULIC PUMP Filed April 14, 1959 4 sheets-sheet 4 NVENTOR 12T; .4, 4e/0N IMLL/AMSON j Y 6 j@ mavfgs United States Patent Oilice 3,053,195 HIGH PRESSURE HYDRAULIC PUMP Larkin R. Williamson, 304 S. Main St., Red Springs, N.C. substituted for abandoned application Ser. No. 269,917, Feb. 5, 1952. This application Apr. 14, 1959, Ser. No.

8 Claims. (Cl. 10S-153) The present invention is concerned with the problem of the generation and delivery, to the location of consumption, of extremely high iiuid pressures.

In the art of providing means for operating machinery, such as, for example, hydraulic presses, various systems have been developed and utilized for the generating of the fluid pressure and the utilization of the same in the press. Such systems usually comprise either variable displacement pumps of substantially constant speed in the form of either radially arranged or axially arranged piston and cylinder units or over-powered pumps wherein the displacement of the pump piston is always uniform, but the speed of operation of the pump is varied.

Apart from the drawback of the extreme expense of electrical power equipment to drive a pump of the last type referred to, the pressure developing capacity of such pumps has been limited by the bearing capacity of the connecting means between the piston and a rotary element for operating the pump piston.

A second limitation has been the volume of the leakage inevitable in such pumps and present in such mechanisms even when new and relatively tight, and ever increasing with continued use and wear. Should such pumps be designed to increase the operating pressures delivered to the press, the resulting bearing structures required would necessitate an extraordinarily large and impractical pump structure. This limitation is also true of both the radially and axially adjustable piston cylinder types of pumps, i.e., the maximum pressure capacity of such pumps is limited by the practical limits of the pressure or fatigue capacity of the bearing surface of whatever type is used. In view of these limitations as to the pressure withstanding capacity of bearings within practical dimensional limits and the leakage factor, the maximum operating pressure capacities of pumps of whatever type are correspondingly limited when the overall dimensions of such pumps remain within practical limits.

The results of the foregoing as related to the design and construction of hydraulically operated machines, for example, hydraulic presses, has necessitated the provision of large cylinder and piston dimensions in the press to attain high tonnage pressure, thus requiring the use of large steel castings in order to obtain an accumulation of sufficient overall pressure on the press piston to reach such high press capacities. The size of such castings is further increased due to the factor of safety practice of over-dimensioning or over-designing the steel castings in order to prevent leakage or seepage of the oil or other power transmitting medium through the castings, due to the probability of porosity characteristic of large steel castings. At times very substantial sums of money become invested in such large castings, in the form of expensive foundry and machine shop labor before blow holes are encountered in the machining of the casting and if the casting has not been purposely over-designed, all the expended work is lost.

It is to be seen, therefore, that the design and proportioning of the cylinder and piston mechanisms of a heavy duty press is determined to a considerable extent by the expedient of fluid pressure limitation from the source thereof, to-wit, the uid pressure pump.

Should it be possible to obtain in a practical way a uid pressure at the source of pressure generation, several times greater than the currently known practical pressure a 3,953,195'A Patented Sept. l1, 1962 sources, it would be possible to decrease the dimensions of the piston and cylinder of the hydraulic press and when thus decreased to bring within the realm of practicability the utilization of forgings instead of steel castings in the press structure. Accordingly, the general object of the present invention is to provide a fluid pressure generating medium of practical dimensions but of extremely high pressure capacity and utilizable as a source of uid power to operate a hydraulic press of total high pressure capacity while using a comparatively small piston and cylinder mechanism.

I have discovered that it is feasible to construct a4 pump, in which comparatively small pistons are used, which operate at a relatively low piston feet per minute, but `which operate at relatively high reversal speeds per minute and have such a short pumping stroke that the speed of the piston during such working stroke is greater than the leakage speed of the fluid being pumped. This pumping piston speed is high enough together with certain structural features to prevent the fluid of predetermined viscosity which is being pumped from reversing its flow and passing out between the cylinder and piston walls.

The structural features which aid in controlling the leakage past the piston are, lirst, to operate the working end of the piston in a cavity which is larger than the cylinder bore and, second, it is necessary to provide an initial piston clearance less than the normal run-off clearance of the fluid being pumped. In addition, I have provided means for maintaining the piston clearance below the run-olf clearance of the uid as this run-off clearance may decrease due to the increase of pressure throughout the pumping stroke.

For any predetermined maximum pressure desired, the volume which can be obtained from any single cylinder is limited by the bearings which can be provided on the crankshaft. As the pump piston speed during its working stroke must be above a predetermined piston speed per minute, the rpm. of the crankshaft must provide this piston speed.

A further object of the invention is the provision of hydraulic pump mechanism of high pressure capacity but of comparatively small construction whereby the pressure of the pump chambers may be formed in steel forgings of comparatively small dimensions.

Other objects and advantages of the invention will be apparent from the following detailed description of a preferred form of embodiment of the invention, reference being made to the accompanying drawings wherein- FIG. 1 is a side view of my improved pump;

FIG. 2 is a longitudinal center cross-section of the same;

fFIG. 3 is a plan view of my ing four cylinders;

FIG. 4 is a fragmentary view similar to FIG. 2 but showing the bearing plates in the second and different relation;

FIG. 5 is a cross-sectional view of one of the connections leading from the pump header;

FIG. 6 is a fragmentary sectional View on an enlarged scale, showing the end of the piston disposed within the end of the piston chamber;

FIG. 7 is a similar view showing the piston extending into the end of the piston chamber and deformed by the pressure generated by the pumping action; and

FIG. 8 is a fragmentary similar view upon a still more enlarged scale, showing the form of the piston under compression in the cylinder.

The essentials of the present pump construction are shown in FIGS. l and 2, which show only a single cylinder in detail but which may, of course, be duplicated or multiplied as desired, depending upon the volume recomplete pump comprisquired for a given use". The pump consists of a cylinder of a block 1 having ak cylinder bore 2 formed therein in which is mounted a piston 3 which in turn is operated by a crosshead member 4 actuated to reciprocate by means of a shoe 5 mounted upon yand driven by a crankshaft 6, whieh in turn is mounted in 4suitable bearings 8 in a frame 7. Connections leading to and from the pump cylinder bore 2 consist of an inlet conduit 10 leading to an intake header 11 secured to the end of the block 1 and provided with passages connecting with an inlet passage 12 leading to the cylinder 2. Between the header and the cylinder block passage 12 is a Y connecting block 16 with passages 17 removably connecting the intake header 11 to the inlet passage 12.

The discharge passages from the cylinder 2 for carrying away the fluid pressure consist of the passage 13 formed in the block 1, leading to a second larger passage 14, ialso formed therein, and in turn connected with these suitable passages is a discharge header block 15 secured to the block 1.

. The driving mechanism for the -shoe 5 and its piston 3 consists of the crankshaft 6 which is mounted in the bearings 8 (see FIG. 3) mounted lin the side plates 21 of the pump housing 22. This crankshaft 6 is provided with an integrally formed eccentric 23 on which is mounted a roller bearing 24 which has its outer race 25 securely pressed into the bearing bore 18 of the shoe 5 so that as the crankshaft turns the shoe is moved in a constrained movement with the face 26 of the shoe maintained in a vertical position (see FIG. 2) during its constrained travel back and forth in a horizontal direction. That is, in the arrangement shown in FIG. 2, the shoe is constrained to reciprocate in a direction aligned with the movement of the piston, but it also has vertical movement at right angles to the movement of the piston during its movement in the horizontal direction.

The bearing face of the shoe is provided with a suitable flat bearing or Wear plate 27 which maintains contact throughout its movement with the end of the crosshead 4, and which has a relatively large surface area compared to the stressed `bearing area of the roller bearing. Also, as seen in FIGS. 2, 3, and 4, for example, the magnitude of the contact area between the face 26 of the shoe and the wear plate 27 is several times that of the cross-sectional area of the piston 3. The shoe is provided near its upper end with a pair of guide members 28 to prevent rocking motion of the shoe but to allow both its vertical and horizontal movement, thus maintaining the face of the shoe and the contacting face of -the crosshead in complete surface contact during the entire instroke of the piston. The piston is returned from its inward `or pressure position by means of a coil spring 29 which surrounds the rear end of the piston in the crosshead and has a bearing against the rear end of the cylinder block. The crosshead is likewise returned from the inward or pressure position of the piston by meansof a pair of coil springs 30 mounted between the rear face of the cylinder block and the front face of the crosshead in suitable cylindrical openings 31 yformed in the cylinder block.

To eliminate the tendency to tip the crosshead, due to.-

friction between the bearing plate of the shoe and the ehd of the crosshead, I have lowered the center line of the crankshaft below the horizontal center line of the piston so that at the maximum vertical height of the shoe during the travel of the eccentric, the center line of the .eccentric will be on the centerline of the piston. It

was found that unless the conditions stated were provided that there was a considerable tendency on the part of p the shoe to tip 4or cock the crosshead, which produced a serious binding action between these parts and dilhculty in the operation of the piston in its guides. Y

'It was found by calculation that the offset required in the example shown in FIGS. 1 and 2 was approximately one-half of the ,stroke of the piston, i.e., and one-half the eccentricity. This amount of offset is correct where the coefficient of friction, which can be determined by calculation in any given case, approximates 0.2 and will, of course, vary as the coefficient of friction varies in various structures.

In the present pump, -as shown in FIGS. 2 and 3, the pump housing 22 consists of a base plate 35 on which are mounted a pair of side plates 21, these plates clamping four separate cylinder blocks therebetween, as best shown in FIG. 3, the blocks and the plates being held together by means of a series Iof bolts 36 and 37. The bolts 36 extend through the rear end of the plates and the bolts 3'7 extend through the side plates and the cylinder blocks to hold the entire structure rigidly together, in order to obtain satisfactory bearing support for the crankshaft and a pair of sleeves 38, which are in turn secured to the side plates 21 by means of bolts 39. In this construction, as shown in FIG. 3, the pairs of cylinders adjacent the end plates are operated together and oppose each other during reciprocation.

Such intake header 11 is a continuous bar, as shown in FIG. 3, and attached to the several cylinder blocks by means of bolts 40 over each intake opening in the cylinder, and each cylinder is provided with its connecting block 16 having a passage 61 communicating with the inlet passage 12 of the cylinder passage, and a second passage to a header opening. This exhaust header is a similar longitudinal block 50 having four openings connecting to cylinder conduits 45 from the four cylinders and having a vertical opening connected through these passages so as to provide a single outlet passage 63 from the header to an 'outlet pipe, which in turn leads to the desired working mechanism. In turn, the head 50 is provided with the single outlet pipe 64 which is connected by a header block 56 to the header block 50. This outlet 64 is suitable for connection to any operating mechanism by means of conduit 55.

On the top of the exhaust header block is a safety block having a vertical passage 81 connected to a passage 83 in the header block. This safety block has a thin wall left between the passage 81 and a concentric opening 84 thereabove. A cover plate 85 is mounted above the opening 84 and several small outlet openings 86 extend from the concentric opening to the sides of the block. The thickness of the wall 86 in the passage is used as `a safety blow-out seal, the pressure needed to blow out the wall being predetermined to provide the needed safety factor for the pump. Preferably this Wall will be reduced to allow it to blow out or rupture at pressures between 25,000 lbs. to 35,000 lbs., the exact limit being determined by the Working pressure to be top to bottom by the amount of the eccentricity and in the horizontal direction also varies by the same amount. The horizontal direction or movement of the shoe is transmitted to the piston by the means already described and in one full revolution of the eccentric the piston is moved from its retracted'position, shown in'full lines in FIG. 2, into thedotted line position or pumping position, also there shown, and returned to its retracted position. As the piston is kforced inwardly lby the action of the eccentric and the connections between the eccentric and the piston, which have been described, it 4forces fluid from the piston chamber 2, which is larger than the piston, through the passage 13, lifting the Aball valve S2 which is seated against an integral portionV ofthe block at the end of the small passage 13 and against the pressure of the spring 14. This liquid which represents the displacement of the piston, is thus forced outwardly through the discharge passage 13, through the connecting block 15, conduit 45 and int-o the exhaust header block 60, into passages 70, entering the header block and leading to a common vertical passage 7'1 from which the iluid is discharged through the conduit 55 into the desired mechanism, not shown, which is to be operated.

As soon as the inlet stroke of the piston is completed, the spring returns the ball into engagement with its seat, closing the discharge line from the cylinder and permitting the inlet line to be opened by the suction created by the piston in its return stroke, which lifts the intake ball valve from its seat and draws a fresh supply of fluid into the conjoint chamber, consisting of the passage 12 and the chamber at right angles thereto which constitute the enlarged piston chamber proper.

The very high pressure created by the operation of the present pump is due to, or rather limited by, the ratio between the charge of uid moved on a given inward stroke of the piston and the total volume of the chamber into which this fluid is yforced through the pump or inward stroke of the piston. The amount or the volume of each charge forced through the system by one inward stroke of the piston is the volume indicated between the end of the piston and the dotted lines to the right thereof.

The present pump is capable of producing pressures of from 20,000 lbs. per square inch up, and the production of this pressure represents a direct force acting ngitudinally on the end of the piston of a very considerable amount in total pounds. This total force acts longitudinally on the piston and represents the load which must be absorbed through the bearing plate and crosshead, and in turn between the crosshead, the shoe and the roller bearing actuated by the eccentric. It will be understood that the force per square inch of this load is reduced very greatly by the construction of the crosshead and bearing plate which are many times the area of the piston itself. Similarly, the load per square inch on the roller bearing is again reduced by reason of the size of the roller bearing to a fraction of that of the loading per square inch on the piston.

It will be possible, of course, to operate the presentI pump as described to this point and if made by ordinary methods and tolerances to produce a very considerable pressure, there would obviously, with ordinary tolerances, be quite a considerable leakage past the piston due to the high pressure produced within the cylinder cavity. This I have been able to eliminate substantially completely by the construction of piston and cylinder, which is shown in FIGS. 7 and 8. In an ordinary pump employed to produce high pressures of the order of 3000 to 5000 pounds, it is common to pack the pistons and to renew the packing at intervals, and even with such packing a certain amount ot leakage results. The present piston and cylinder have no packing of the ordinary types interposed between them. The cylinder consists, so far as the pump in which the piston operates is concerned, of a truly cylindrical, bored-out space, nished by ordinary machining and to a xed tolerance.

The piston constructed in the form indicated, is slightly longer than the cylinder bore in which it is guided so that it projects at all times into the enlarged portion of the cylinder or cavity and is machined and ground, by ordinary methods, to a tolerance of approximately .0022 inch with respect to the cylinder in the size herein indicated, which is fi of an inch. As a result the iluid forces -acting on the piston produce the eifect which is illustrated in FIG. 7, when the piston is in its inward or pumping position.

The piston is loaded longitudinally as a column by reason of the pressure exerted against the inward end face thereof. It is also loaded radially through that portion which extends into the enlarged portion of the cylinder or cavity, producing a slight contraction of the end of the piston, as indicated in FIG. 7. The end loading of the piston, however, produces a bulging or barrelling action of the piston within the cylinder bore su'icient to stop all practical leakage. This barrelling action increases as the pressure increases and reduces the clearance between piston and cylinder bore directly Iwith the pressure pumped and maintains the clearance therebetween always at a clearance under the leakage factor of the uid pumped.

It will be understood, of course, that while the piston is machined and ground to an extremely smooth finish, the cylinder is machined by ordinary methods and hence its wall surface is not as smooth or as regular as that of the piston. The radial clearance between the piston and cylinder in the size indicated approximates in the finished condition of the piston .0001, exists throughout the area of working engagement between these two members but the points represented by machine tool marks or by ridges found on the surface of the cylinder and extending circumferentially thereof. These ridges, in effect, represent metallic packing elements for sealing the piston yand cylinder against `any longitudinal ow of fluid between the two under the engaged conditions described in FIGS. 7 and 8. In fact, it has been found that the pump can be operated to produce a pressure of 15,000 to 20,000 lbs. which can then be stopped with the piston in its inward position, and this pressure can be maintained `for a matter of many h-ours, which is not true of any of the previous types of pumps in which the relationship between the piston and cylinder has to be maintained by piston rings, packing etc.

Since the ribs on the cylinder, just referred to, in al1 Cases run circumferentially, they produce between them grooves which act as reservoirs of Huid, which fluid serves as a continued source of lubrication between the piston and cylinder. The degree of lubrication, of course, depends upon the particular tluid in question.

I claim:

1. In a high pressure hydraulic pump, a cylinder, a piston therein, and means for reciprocating said piston including a crankshaft, a shoe eccentrically mounted on said crankshaft, a crosshead engaged with said piston and frictionally engaging said shoe over `a relatively flat area having a magnitude considerably greater than the crosssectional area of the piston, at least the portion of the cylinder surrounding said piston being relatively inexpansible with respect to said piston, an initial radial clear ance between said piston `and said cylinder, said initial clearance being of a size to prevent leakage between said piston and lcylinder of fluid to be pumped when said iluid is in an uncompressed state, means for causing said piston during the compression stroke thereof to be radially eX- panded by the pressure of the said fluid :as it is compressed to reduce said initial clearance and thereby prevent leakage of said fluid when the same is in a compressed state.

2. A device as defined in claim 1 in which minute surface irregularities are provided between said piston and cylinder, said irregularities acting as sealing and -lubricating reservoirs.

3. A device as defined in claim 1 in which the external surface of said piston is provided with an extremely smooth finish, and in which said cylinder is provided with a rougher surface finish than said piston, the coaction of said smooth piston surface and said rough cylinder surface forming sealing `and lubricating reservoirs.

4. IIn a high pressure hydraulic pump, a piston and a cylinder, means to reciprocate said piston in said cylinder, said cylinder being relatively inexpansible with respect to said piston, an initial radial clearance provided between said piston Iand said cylinder, said initial clearance being so chosen as to prevent leakage of uid between said piston and said cylinder when said fluid is in an uncompressed state, means 'for causing said piston to be radially 7 expanded during the compression stroke Athereof lby the pressure yof the said uid as it is compressed to reduce said initial clearance 4and thereby prevent leakage of said uid when the same is in a compressed state.

5. A device as defined in claim 4 in which the surface of said cylinder is provided in the region of its association with said piston, with a series of minute circumferentially extending ridges and grooves, said piston being extremely smooth with respect to said cylinder surfaces, said ridges and grooves acting to prevent leakage of fluid between said cylinder land said piston.

6. In a high pressure hydraulic pump, a plurality of cylinders land a plurality of pistons in said cylinders and arranged side-by-side, means for reciprocating said pistons, said means including a crankshaft with a plurality of eocentrices thereon mounted side-by-side, some of said eccentrics being out of phase with respect to the others, a shoe mounted on each eccentric, bearing means between each eccentric and its `associated shoe, and means operatively associating each shoe with at least one piston and including a crosshead engaged with the piston and frictiona-lly engaged with the shoe, `the centenline of each piston being offset from the centerline of the crankshaft to neutralize thrust which would otherwise tend to tip the corresponding crosshead and shoe relative to one another. Y

7. A device as defined in claim 6 characterized further by guide means to act between each crosshead 4and shoe and prevent rocking motion of Iche shoes while at the same time allowing the shoes to move horizontally.

8. A device -as defined in claim 6 in which the frictional engagements between the shoes 4and the crossheads are fiat engagements having areas considerably greater than the cross-sectional areas of the pistons.

References Cited in the le of this patent UNITED STATES PATENTS 2,168,850 Towler et al Aug. 8, 1939 2,253,152 Towler et al. n Aug. 19, 1941 2,324,291 Dodge July 13, 1943 2,345,125 Huber Mar. 28, 1944 2,562,954 Schmdln Aug. 7, 1951 2,730,960 Krause Ian. 17, 1956

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