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Publication numberUS3135210 A
Publication typeGrant
Publication dateJun 2, 1964
Filing dateApr 27, 1961
Priority dateApr 27, 1961
Publication numberUS 3135210 A, US 3135210A, US-A-3135210, US3135210 A, US3135210A
InventorsEnglish Charles L
Original AssigneeEnglish Charles L
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Hydraulic pressure boosting device
US 3135210 A
Abstract  available in
Images(3)
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Claims  available in
Description  (OCR text may contain errors)

June 2, 1964 c. L. ENGLISH HYDRAULIC PRESSURE BOOSTING DEVICE Filed April 27, 1961 3 Sheets-Sheet 1 INVENTOR. CHAEL 55 A. NL/ZTH C. L. ENGLISH HYDRAULIC PRESSURE BOOSTING DEVICE June 2 1964 3 Shecs-Sheet 3 Filed April 27, 1961 diw lrflfi. mwv HLn U T QL o& mm 9% QM HLHHI INVENTOR. cHAElE-S L. Emu/5H A TTDPA/FYS strokereversal and/or crank action.

V r 3,135,210 HYDRAULIC PRESSURE BOOSTING DEVIC Charles L. English, 2204 E. 25th Place, Tulsa, Okla.

Filed Apr. 27, 1961, Ser. No. 105,955

12 Claims. (Cl. 1ll349) invention relates to hydraulic power transmission devices, and more particularly, but not by way of limitation, to a device for boosting the pressure in a second UnitedStates Patent Patented June 2, 196f1 from the pump end of the device without the occurrence of surges or pulsations. The curves of output volume and pressure versus time for the pump are substantially straight lines-an advantage which cannot be claimed for existing duplex, triplex or even quintuplex pumps. Moreover, the cost of construction of the pump of this invention is considerably less than that of the latter types. For

. example, the heavy supporting bedframe which has charpower-driven pumps employing cranks and flywheels,

have, for many years, accountedfor a large share of industrys miscellaneous pumping needs. However, because of certain limitations inherent in previous pump design,- versatility of application has also been limited, so

' that any substantial variation in pumping requirements and conditions necessitates the employment of a different type of pump having pumping characteristics suited to the particular needs at hand. Thus, although reciprocating pumps possess numerous characteristics which make their use attractive in many situations, other features preclude their successful use in other instances. One of the latter features is the pulsating discharge characteristic of such pumps, whether the reciprocating pump be directacting or power-driven. Such pulsations can be reduced in magnitude by increasing the number of acting pistons or plungers and overlapping their strokes, as in duplex, triplex and other multiplex pump systems, or by providing a surge chamber in which the pumped fluid is accumulated at peak delivery and discharged at stroke reversal. However, neither of these devices for'reducing the magnitude of the discharge pressure surges completely eliminate such pulsations, and the cost of the pumping unit is considerably increased when they are utilized.

The present invention contemplates a. rodless multiplex pump which is characterized by its adaptability to many specialized pumping needs, and by its ability to deliver a substantially constant volume of pumped fluid at a high pressure without substantial output surge resulting from The method of stroke control used in the pump is unique in that each stroke reversal is accomplished independently of the move ment of the engine pistons of the pump in their cylinders and is, instead, caused to occur in alternate progressive relationship in each of the cylinders when the combined displacement of the pistons during their stroke is such as to give any desired output of the fluid pumped within the mechanical limits of the total engine cylinder volume available. I Stroke reversal, then, occurs following a predetermined time interval and after a predetermined volume of power fluid has been received in the engine pistons and is not brought about by an articulating means which is common to all the engine pistons or to their rods and which is responsive to the movement of one of the engine pistons to effect the reversal of the other engine pistons. The engine pistons of the rodless booster pump of the present invention function independently of each other and without the managing influence of a common crank. Their only interdependency resides in the requirement that the total volume of power fluid received in the engine cylinders for driving the engine pistons be the same for any given interval of time. Thus, as one of .the engine pistons decelerates toward the end of its stroke and prior to turn-around, the other engine pistons must at that time accelerate so that an increased volume of power fluid may enter the other engine cylinders to offset the decrease in the volume of power fluid entering the first engine cylinacterized many previous multiplex pumps is eliminated. In order to permit the rodless booster pump of the present invention to function in the manner described, two salient features of the invention are essential. First, it is necessary that means be provided to allow the acceleration of the engine pistons which are not approach ing stroke end to anticipate the deceleration of the engine piston which is undergoing stroke reversal. To this end- I provide dash-pots which. function to decrease the de-' celeration of the engine pistons as they approach the lim its of their stroke, but which do not retard the accelera tion of the pistons as they move away from the ends of their stroke. This cushioning effect of the dash-pots is supplemented by a similar cushioning effect of power fluid transfer valves (subsequently to be described). which are utilized to periodically shift the impress of power fluid to opposite ends of the engine pistons; Thus, the cutback in the volume requirement of power fluid in one engine cylinder at the time of stroke reversal will not be so sudden that the engine pistons in the other engine cylinders will be unable to accelerate fast enough to accommodate an increased volume of power fluid equal to the loss of requirement in the first cylinder.

Second, for the purpose of supplying an unvarying total volume of power fluid to the engine cylinders, a source of a constant volume of power fluid, such as a rotary pump, is utilized, and its discharge is divided with aportion of its output being directed to each of the engine cylinders. A main reversing or transfer valve is associated with each of the engine cylinders and functions to alternately transfer the power fluid to opposite ends of its respective engine cylinder while receiving spent" power fluid from the low pressure end of the engine cylinder and returning it to a power fluid sump or reservoir for recycling. The reversal of the strokes of the enginepistons is thus effected by the throwing of these transfer valves. The slight pressure loss which occurs in the power fluid upon the throwing of the transfer valves tends to supplement the cushioning or decelerating effect which is at that time effected by the dash-pot operative with respect to the engine piston which is associated with the particular transfer valve being thrown. Also, a slight pressureincrease occurs in the portion of the power fi uidwhich is acting upon the other engine pistons whose strokes are not being reversed at that time.

. In order to permit the transfer valves to be thrown eriodically according to a predetermined optimum time sequence, a pilot valve is provided which functions to i it permits the power fluid to accommodate each engine This type of pilot valve managepiston individually according to its ability to travel, and thus compensates for inherent differences in the resistance offered by each of the pistons to the impress of the power fluid upon the working area of the piston. So far as I am aware, all other multiplex pumps supply a driving force to the pistons which is suflicient to drive each of them to the limit of the travel which is allowed by the length of the cylinders in which they move. Since, practicallyspeaking, the pistons will always be different from each other in their resistance to movement in their respective cylinders, some power loss'results from the supplying of a greater volume of power fluid to the lighter pistons (the ones olfering less resistance to the moving force of the power fluid) than is actually required to drive the pistons to the limit of their stroke. In other words, in a situation where dash-pots are utilized to retard the dc celeration of the pistons, an excessive energy loss occurs when the lighter pistons are driven more deeply into their dash-pots in order to prevent the heavy piston from short-stroking.

- The principles of the invention may be applied with equal significance to triplex, quintuplex or multi-cylinder systems, it being only necessary. (a) to make each of the engine pistons a free piston in the sense that no crank- In its broadest aspect, then, the present invention may be said to comprise a plurality of fluid pressure operated 1 engine pistons each having opposed engine piston areas for alternately receiving the impress of a power fluid; pump means connected to the engine piston and responsiveto the movement thereof to pump a fluid; cylinders enclosing each of said pistons; means for decelerating said engine pistons when the engine pistons approach the ends of said cylinders; means for delivering a substantially constant volume of power fluid, and valve means for directing the total volume of power fluid supplied by said delivering means to all of said engine piston cylinders as the individual demands of each piston permit, the latter function being made possible by the utilization, as a part of said valve means, of means for reversing the stroke of the engine pistons periodically according to what period of reversal is' determined to give the greatest output ca-' pacity with the least fluctuation or pulsation in the output volume and pressure over an extended periodof operation. Stated differently, the invention contemplates a multi-cylinder arrangement of directly driven engine-topump piston units wherein the r'eversing valve for each piston unit is controlled by a separate, individual pilot valve, which pilot valves are thrown in a predetermined sequence so as to prorate the total absorbed motive fluid to each unit according to its ability to travel in its cylinder.

In a more specific aspect, the invention contemplates the securement of a pair of pump pistons to each of the engine pistons, which pump pistons are each enclosed in acylinder and are double-acting. This arrangement may be adapted to the pumping of corrosive or gritty fluids, to high capacity needs, or to high output pressure requirements. In every instance, the nonpulsating, steady output characteristic is maintained, and the initial manufacturing cost as well as the maintenance cost of the pump compares very favorably with other pumps which might b characterized by an output which is substantially free of pulsation and surge.

An additional object of the present invention is to provide a rodless, multiplex hydraulic pump which is less expensive to manufacture and maintain than other multiplex pumps of the same capacity and pressure characteristics.

A furtherobject of the present invention is to provide a rodless, multiplex hydraulic pump which permits all engine pistons of the pump to travel the maximum distance in their cylinders which they are capable of traveling when subjected to a given volume of power fluid, thus eliminating power losses resulting from interruption of piston stroke by the ends of the cylinders in which they move.

Another object of this invention is to provide a rodless, multiplex hydraulic pump which can absorb an extremely high horsepower input.

Another object of this invention is to provide a rodless,

multiplex hydraulic pump which is adaptable to many specialized pumping requirements.

. Another object of this invention is to provide a rodless, multiplex hydraulic pump, the capacity of which may be. rapidly. and easily altered during the operation of the pump.

Other objects and advantages will be apparent from the following description when considered in conjunction with the accompanying drawings, wherein several embodiments of the invention are illustrated.

In the drawings:

vFIGURE 1 is a schematic view of a rodless, duplex hydraulic pump constructed according to the present invention.

FIGURE 2 is a side view in elevation of the arrangement of the pilot valve in a preferred embodiment of the present invention. A portion of. the pilot valve is shown in section.

FIGURE 3A is an enlarged view in section of one of the main transfer valves. V

- FIGURE 3B is an enlarged view in section of the pilot valveshown in FIGS. 1 and 2.

FIGURES 4A, 4B and 4C are schematic views illustrating modified embodiments of the engine and pump cylinders of the present invention.

Referring now to the drawings in detail, and particularly to FIG; 1, reference characters 10 and 11 designate a pair of engine pistons enclosed in engine piston cylinders Band 13, respectively, 'and each having opposed faces 14 for alternately receiving the impress of power fluid during the operation of the invention. A piston rod 16 extends from each side of each of the pistons 10 and 11 in a direction normal to the faces 14 of each piston and pass through openings 18 at each end of the engine cyl-' inders 12 and 1 3. A reduced diameter pump cylinder 20 is secured to or formed on each end of each engine cylinder l2 and i3 and communicates with the'respective engine cylinder through the respective opening 18. Pump pistons 22 are located inwardly from the ends of the rods 16 so that an end portion 24 of each of the piston rods 16 extends for a substantial distance beyond the pump pistons 22.

According to a construction well understood in the art, the relative dimensions of the engine cylinders 12 and 13, pump cylinders 20 and piston rods 16 are such that the pump pistons 22 will move the length of their respective pump cylinders 20 as the engine pistons 10 and Ill are moving the length of the engine cylinders 12' and 13. Adjacent that end of the pump cylinders 20 which is most remotely located with respect to the engine cylinder 12 or 13 to which the pump cylinders are attached, the diameters of the pump cylinders are reduced to provide dash-pots 26.' The inside diameters of the dash-pots 26 are slightly larger than the diameters of the end portions 24 of the piston rods 16 so that a small clearance exists between the walls of the dash-pot 26 sure of the booster pump increases.

and the piston rods 16 as the latter move into their respective dash-pots.

In order to permit the dash-pots 26 to function in retarding the movement of the pump pistons 22 toward the end of the pump cylinders 20 without impairing movement of the pump pistons away from the ends of the pump cylinders in which the dash-pots are located, a check valve 28 is provided in the outer end of each of the dash-pots and is adapted to prevent the egress of pumped fluid from the respective pump piston by way of its dashpot, while permitting fluid to enter the pump cylinder from an intake conduit 30 by way of the dashpot. The intake conduits 30 directing fluid to each of the several dash-pots 26 merge in a common intake passageway 32 communicating at its intake end (not shown) with the fluid which is to be pumped by means of the invention.

To facilitate the discharge of the pumped fluid from the pump cylinders 20, a discharge conduit 34 communicates with each of the pump cylinders 20 at a point inwardly of the respective dash-pot 26 located in the end of the pump cylinder. Each of the discharge conduits 34 is provided with a check valve 36 adjacent its point of connection to its respective pump cylinder 20 so that loss of suction is prevented during the suction or intake stroke of the pump pistons 22. The discharge conduits 34 leading from the several pump cylinders merge in a common discharge line 37 which leads to the point of application of the pumped fluid.

For the purpose of supplying power fluid to the engine cylinders in the manner essential to the novel operation of the present invention, several major components are necessary and may be designated generically as a source of power fluid 38 capable of delivering a constant volume of power fluid, a pair of main transfer valves 40 and 41 associated with the power cylinders 12 and 13, respectively, and some type of timing mechanism, such as the pilot valve means 42 shown in FIGS. 1, 2 and 3B, functioning to periodically throw the main transfer valves 40 and 41 in a manner hereinafter to be more fully described.

In a preferred embodiment of the invention illustrated schematically in FIGS. 1 and 2, the source of power fluid 38 comprises a centrifugal or rotary pump 44 connected to an intake line 46 and to a discharge line 48. When a centrifugal pump is utilized, the pumps inherent slippage will cause a rapidly increasing fall-01f of the pumps volumetric displacement as the discharge pres- It is therefore necessary to provide a suitable metering device 50 (FIG. 1) in the hydraulic circuit of the pump, such as in the power fluid discharge line 48, which functions to constantly monitor the volumetric flow through the line and to control the period of throw of the pilot valves (subsequently to be described) as may be required to maintain the displacement characteristics of the booster pump. The discharge conduit 48 from the pump 44 branches into a pair of conduits 52, each leading to one of the main transfer valves 40 and 41.

Although a number of types of valves may be suitably used as the main transfer valves 40 and 41 of the pres ent invention, I prefer to employ two pairs of three-way piston valves 54 arranged with respect to each other in the manner illustrated in FIG. 1. As shown in FIG. 3A, each of the three-way valves 54 includes a motor cylinder 56 containing a motor piston 58. The motor piston 58 is connected through a valve stem 59 to a two-faced poppet valve head 60. The valve head 60 is located in a valve chamber 62 which is substantially smaller diametrically than the motor cylinder. The valve head 60 during operation of the three-way valves 54 alternately seats upon one or the other of the tapered seats 63 located at each end of the valve chamber 62. The housing portage of each set of three-way valves is arranged so that each three-way valve of each main valve 40 and the action of the shaft 78.

6 41 may receive power fluid from the conduits 52 and may discharge spent power fluid through conduits 64.

A pair of pilot valve passageways 68 communicates. with each pair of the cylinders 56 containing the large diameter motor pistons 58 of the three-way valves, and each pair of the passageways 68 join in a common conduit 70 which leads to the pilot valve means 42. There are thus two conduits 70 leading to the pilot valve means 42-one from each of the main transfer valves 40 and 41. The return flow lines 64 leading from each of the main transfer valves 40 and 41 merge in a common flow return line 46 which carries the spent power fluid discharged from the engine cylinders 12 and 13 back to the pump 44 for recycling. In actual operation, it will be necessary or desirable to provide a suitable power fluid reservoir or storage tank 72, as illustrated in FIG. 2, for receiving power fluid from the return flow line 46 and supplying the power fluid as it is needed to the pump 44.

The arrangements of the two three-way valves 54 in the main transfer valves 40 and 41 are identical. The porting and manifolding of the main transfer valves 40 and 41 should be large enough to prevent the occurrence of large pressure losses at the anticipated speeds of operation. Also, in order for valve effect in the three-Way valves 54 to be identical in both opening and closing, the effective area of the engine piston 58 should be twice the effective area of the valve head 60. Finally, the throw of the three-way valves 54 should be of sufficient length that no obstruction of the valve portage areas is offered.

The pilot valve means 42 which is utilized to eflect the reversal, or throw, of the main transfer valves 40 and 41 is illustrated most clearly in FIGS. 2 and 3B. In the preferred embodiment illustrated in these figures, the pilot valve means 42 comprises a pair of three-way valves 74 and 76 which are set to be thrown at a phase difference of ninety degrees from each other. To accomplish the desired difference in the time of throw of the two valves 74 and 76, each of the pilot valves is offset ninety degrees from the other upon a common shaft 78. The shaft 78 is secured to the periphery of a face plate 80 which is coaxially secured to a shaft 82 driven through a variable speed reducer 84 by the same motor 86 which is utilized to drive the power fluid pump 44. Each of the three-way valves 74 and 76 is connected to receive power fluid from conduit 48 via the conduits 88 and 90, respectively, and to alternately discharge high pressure power fluid to, or receive spent, exhausting power fluid from, the main transfer valves 40 and 41 via the conduits 70. The two three-way valves 74 and 76 share a common discharge port 92 through which the spent power fluid contained in the motor cylinder 56 can be released as the pilot valve means 42 is thrown in the manner hereinafter described. The discharge port 92 communicates with a conduit 94 which joins the return flow line 46, preferably upstream of the reservoir 72, or, optionally, the pump 44, as shown in FIGS. 1 and 2.

The details of construction of the pilot valve means 42 are best illustrated in FIG. 3B. Each of the three-way valves 74 and 76 of the pilot valve means 42 is provided with a high pressure intake port 106 and a common port 108. A valve seat 110 is provided between the port 106 and the port 108 of each of the three-way valves 74 and 76, and a ball valve 112 is seated on a hollow shank 114 so that each valve 112 may seat upon its respective seat 110 when its shank 114 is reciprocated downwardly by It will be noted that the seats 110 are each provided with a bore 116 therethrough which is of larger diameter than the shank 114 which passes therethrough. Thus, when the ball valves 112 are seated on the seats 110 by the movement of the upper ends of the shanks 114 below the upper surface of the seats 110, the hollow interior of the shanks 114 will be in communication with ports 108. Each of the shanks 114 has an opening (not seen) at its inner end 117 which places the lower end of the shank in communication with a crankcase 118 which is common to the two three-way valves 74 and 76. As has previously been explained, spent power fluid is discharged from the common crankcase 118 through a discharge port 92. it will be apparent that the inner end 117 of each shank 114 may be secured to the shaft 78 by a rectangular yoke 119 to convert the rotary motion of the shaft 78 to reciprocating motion of the shanks 114.

Operation As has been previously explained, the pump 44 delivers a constant volume of power fluid to the conduit 48. To achieve constant volume operation using a centrifugal pump, it is necessary to provide a metering device 5% (FIG. 1) which will sense fluctuations of the volume of power fluid flowing in conduit 48 and control the pilot valve means 42 to compensate for variations in the pump requirements. A constant displacement pump, such as a rotary pump, can, of course, be used without being affected by changes in head and thus there is no need for such a volume metering device.

The power fluid supplied by conduit 48 is divided and flows through conduit 52 to each of the main transfer valves 40 and 41. Simultaneously, a small portion of the power fluid in conduit 48 is directed through conduits 88 and 90 to the three-way valves 74 and 76 of the pilot valve means 42. In the position of the pilot valve means 42 and main valves 40 and 41 shown in FIG. 1, both of the three-way valves 74 and 76 of the pilot valve means 42 are open to high pressure power fluid due to the position of the eccentric shaft 7%, although the valve 74 is just on the verge of closing following its open period corresponding to 180 degrees of rotation of the eccentric shaft 78. Threeway valve '76 has been open to high pressure power fluid from conduit dtl during 90 degrees of the rotation of the eccentric shaft '78 and will remain open for an additional 90 degrees of the shafts rotation. In other words, each three-way valve 74 and 76 of the pilot valve means 42 is open during one half the period of one revolution of the eccentric shaft 78 and is closed for one half of such period, with the throws of the two valves 74 and 76 being 90 degrees out of phase with each other or differing in time by one fourth the time required for the shaft 73 to make one revolution.

With the throw of the pilot valve 42 occurring in this manner, it will be apparent that power fluid will be directed via line 88 through three-way valve 74 and conduit 70 to the main transfer valve 4% during one half theperiod of revolution of the eccentric shaft 78, and, therefore, of shaft 82, which'is connected through the variable speed reducer 84 to the motor 36. The same is true of power fluid that is directed from conduit 90,

through three-way valve 76 and conduit 7% to the main transfer valve 41. It will further be apparent that power fluid so directed to the main transfer valves 4% and 41 from the pilot valve 42 will act on the large diameter motor pistons 58 in the main transfer valves 48 and 41 to throw the valves to the positions shown in FIG. 1. The concurrent supply of power fluid from the pilot valve means 42 to both main transfer valves 4% and 41, will take place during 90 degrees or one fourth of the rotation of the eccentric shaft 78, after which one of the valves 40 or 41 will be thrown to the opposite posi tion due to the closure of its respective three-way valve 74 or 76, in pilot valve means 42, thus denying access of high pressure power fluid to the large diameter motor piston 58 of the main transfer valve so thrown. The concurrent closure of both of the three-way valves 74 and 76 in pilot valve means 42 will likewise occur during 90 degrees of the rotation of the eccentric shaft 73 so that during this phase of the shaft rotation, both the main transfer valves 40 and 431 will be thrown to the opposite position from their positions illustrated in FIG. 1.

The pattern of actuation of the main valves 40 and 3 41 in'response to the pilotvalve means 42 may thus be summarized as follows:

(a) Both of the main valves 49 and 41 are concurrently positioned as shown in FIG. 1 for the period of time which it takes eccentric shaft 78 to make one-fourth revolution;

(12) One of the main transfer valves 4th or 41 will then be thrown, due to the closure of its respective three-way valve 74 or 76, and its pistons 58 and 60 will occupy the opposite position from that shown in FIG. 1. This relationship of the two main transfer valves 44B and 41 will continue for the period of time required for the eccentric shaft 78 to complete another quarter revolution;

(0) The other main transfer valve 49 or 41 will then be thrown due to the closure of its respective three-way valve 74 or 76, so that both main transfer valves will occupy the opposite position from that shown in PEG. 1. Both main transfer valves will remain in this position during the period of time required for the eccentric shaft to rotate degrees;

(d) During the last 90 degrees of rotation of the eccentric shaft 73, one of the three-way valves '74 and 76 will be reopened, and its respective main transfer valve 49 or 41. will be returned to the position shown in HS. 1.

It will be apparent, of course, that when the three-way valves 74 and 76 of pilot valve means 42 are closed, the pistons of the three-way valves 54. of the respective main valves 40 and 41 are reciprocated under the influence of power fluid from the conduits 52, and that exhausted or low pressure power fluid in the valve motor cylinders is returned to the pilot valve means 42 by way of the respective conduit 74). This exhausted power fluid returns ultimately to the pump 44 by way of the hollow shank 114, crankcase 118, port 92 and the conduit 94.

For each of the phases of the main transfer valve operating cycle that has been described'above, a corresponding condition characterizes themovement of the engine pistons 16 and 11. Thus, with the main transfer valves 40 and 41 occupying the positions shown in FIG. 1 and corresponding to phase (a) of their operating cycle as described above, both engine pistons 10 and 11 travel in the same direction (as illustrated in FIG. 1). If both the main transfer valves 46 and 41 were thrown to the opposite position as described in phase (0) above, the engine pistons ill and 11 would both travel in the opposite direction, that is, toward the bottom of the page as they are viewed in FIG. 1. When only one of the main transfer valves 40 and 41 is reversed from the position shown in FIG. 1, the engine pistons 10 and 11 travel in opposite directions from each other.

As will be apparent from FIG. 1, the end portion 24 of one of the piston rods 16 of engine piston 10 is just entering its dash-pot 26, and the direction of engine piston travel will soon be reversed due to closure of three-way valve 74 of pilot valve means 42. On the other hand, engine piston 11 still has about one half its maximum possible stroke to travel before the end portion 24 of its piston rod 16 enters its respective dashpot 26. The closure of the three-way valve 74 will cause the direction of travel of engine piston 10 to be reversed, but will not change the direction of travel of engine piston 11. From what has previously been said, it will be recalled that the throw of three-Way valves '74 and '76 of pilot valve means 42 is accomplished after fixed time intervals (and/or power fluid volume intervals) which depend only upon the rotation of the eccentric shaft 73 whose rotation depends, in turn, upon the speed of motor 36 and the reduction ratio of variable speed reducer 84. It will therefore be apparent that the stroke reversal of each of the engine pistons 19 and 11 is accomplished after a fixed time interval and without regard to the position of the engine piston in its respective cylinder. It is thus possible to adjust the engine piston stroking characteristics; both as to speed of travel and distance, by varying the speed of rotation of eccentric shaft 78 through adjustment of the variable speed reducer 84.

When the operation of the pump is commenced, the variable speed reducer 84 will be set for a relatively small reduction in speed from the motor 86 to eccentric shaft 78. This will cause rapid stroke reversal of the engine pistons and 11 so that they will travel only a portion of the total length of their respective engine cylinders before being reversed. The Weights of the two engine pistons 10 and 11 will differ slightly, as will the frictional resistance of each to movement in'its respective cylinder and other factors affecting the resistance which each offers to the impress of the power fluid tending to move the pistons. In other words, one of'the engine pistons 10 and 11 may be termed lighter and the other heavier, if theseterms are used to describe the relative ease with which each may move in its cylinder. The lighter piston will, of course, be caused to travel farther during the time interval between reversals than will the heavier piston-that is, its stroke will be longer. Thus, at a given setting of the variable speed reducer 84, the lighter piston may travel eighty percent of its possible total travel, Whereas the heavier piston may travel only forty-five percent of its possible travel. It is therefore apparent that due to the influence of the dash-pots 26 in smoothing out the stroke reversals, each piston will base its stroke at, or into, at least one of the dash-pots 26.

After starting the pump with a low speed reduction ratio as described above, the frequency of stroke reversal is gradually decreased by gradually increasing the speed reduction between motor 86 and shaft 82. The volumetric input of power fluid from the pump 44 is maintained constant. A decrease in the frequency of stroke reversal causes the length of the engine piston strokes to be increased until the stroke of the light piston is terminating at each of its ends with the end portions 24 of its shafts 16 in the dash-pots 26 at the ends of the pump cylinders 20. On the other hand, only one of the end portions 24 of the piston rods 16 attached to the heavy engine piston will enter one of the dash-pots 26 unless the frequency or stroke reversal is decreased excessively. In the'latter event, all of the dash-pots 26 may be entered by the respective piston rod end portions 24a result which is undesirable in practically every instance, since power is wasted in driving the light piston rods 16 forceably against the bottoms of the dash-pots 26, or at least very deeply into the dash-pots.

With the stroke reversal frequency established so that each engine piston is traveling its optimum stroke length in accordance with both its and the other pistons indi= vidual ability to travel, and not as limited by the length of its engine cylinder, the dash-pots will function to control the movement of the engine pistons so that the ends 24 of the piston rods 16 will not bang against the ends of the pump cylinders 29, and so that the movement of one engine piston will be decelerated slowly enough at stroke reversal to allow the other engine piston to accelerate a corresponding amount. Having, in the manner described, established the speed reduction ratio between motor 86 and shaft 82 which gives optimum stroking action by the engine pistons 10 and 11 for a given volumetric input of the pump 44, the speed of the motor 86 may be increased or decreased to provide a larger or smaller volumetric input of power fluid from pump 44. This intentional variation in volume of power fluid pumped through conduit 48 from the pump 44 will be I reflected in a proportionate increase or decrease in the total amount of pumped fluid which is pumped by the pump pistons 22 so that the capacity of the rodless multiplex pump of the invention may be easily increased or decreased. Moreover, at all capacities of the pump, the

orifice, is, as a practical matter, free of fluctuations in 10 volume and pressure. In other words, a nonpulsating discharge is achieved. 1

In situations where the pressure demand upon the booster pump will vary between a known maximum and minimum, the pilot valve cadence or period of throw may. be set to give an engine piston stroke which will answer an average pressure demand between such maximum and minimum values. That is, less energy will be wasted in dash-pot entry than is Wasted when the pressure demand falls off to allow longer strokes, and more energy is wasted in dash-pot entry than is wasted when the pressure demand builds up to cause shorter stroking. In this way, a modest margin of safety in the extent of engine piston travel is established at an average pressure demand so that the booster pump can tolerate the slight over-travel of the pistons which will be produced by a low pressure surge in output demand.

Although the present invention has been described by referring to a drawing illustrating a preferred embodiment of the invention in which only two engine pistons are utilized, the principles of the invention are not confined to such an arrangement, and a rodless triplex or quintuplex pump may be constructed in accordance with the principles of the invention. It is only necessary to provide a main transfer valve for each of the engine pistons which is utilized, and to modify the pilot valve means in a manner which will permit it to throw the main transfer valves in proper sequence after fixed time intervals. However, due to the nonpulsating output characteristics of the rodless duplex pump described above, and its high capacity potential, the duplex embodiment will, in most adaptations, be equivalent to a crank-actuated triplexpump of the prior art. The need for a triplex or quintuplex rodless hydraulic pump constructed according to the present invention will probably not often arise, except perhaps in instances Where extreme discharge pressures of sizable volumes are demanded.

The duplex, double-acting embodiment of the rodless pump of this invention is quite adaptable to a variety of pumping requirements, and several adaptations of the pump are suggested by the modifications of the engine piston and associated pump pistions illustrated in FIGS; 4A, 4B and 4C.

In FIG. 4A, each of the pump pistons 22 is single-acting. Pumped fluid is introduced to the pump cylinders 20 from the intake conduits via check valves 122 and dash-pots 26. The pump cylinders 20, contrary to the arrangement illustrated in FIG. 1, are segregated from the engine cylinder and suitable packing 124 is provided around the piston rods '16. Each of the pump pistons 22 has a passageway 126 extending therethrough to permit pumped fluid to pass from one 'side of the piston to the other upon actuation of a check valve 128 located in each pump piston. Fluid is discharged from the pump cylinders 20 by way of conduits 130. t

The arrangement depicted in FIG. 4B makes each of the piston pumps 22 double-acting with pumped fluid being introduced viathe conduits 132 and discharged via the conduits 134. Suitable check valves are provided in the intake and discharge conduits 132 and 134, respectively.

The pump" illustrated in FIG. 4C utilizes a closed system containing a captive hydraulic fluid; Thus, a single conduit 138 connects the outer'end of each of the pump cylinders 20. Pumped fluid is introduced to each pump cylinder 20 from an intake conduit 140 and is discharged from the cylinders through conduits 142.

I As a basis for comparing the characteristics attributable to the different pump end designs, let it be assumed that in each instance, 4A, 4B and 4C, the inside diameter of the engine cylinder is five inches, the outside diameter of the piston rod is one and one-fourth inches, the inside diameter of the pump cylinder is-two inches, and the stroke length is forty-two inches. With these assumed uniform dimensions, the characteristics of the three pump types 1 i illustrated in FIGS. 4A, 4B and 40 may be tabulated as follows:

Pump 4A 4B 4C Pressure, Horsepower 115 115 115 Engine Horsepower Required to Produce 1,000 p.s.i. Output Pressure at One Stroke Per Minute 1.332 2.137 .812

From the tabulated data, it will be seen that in making each of the two pump pistons double-acting, as shown in FIG. 43, a very substantial increase in volumetric capacity is achieved. On the other hand, much higher output pressure is achieved with the single-acting pump piston ar- 'rangement shown in FIG. 40 than with either the FIG. 4A or FIG. 43 types. The FIG. 4A pump end arrangement is, in effect, a compromise between the FIG. 4B and 4C types insofar as volume and pressure outputs are concerned, but is moreeconornical to construct than the latter pumps.

From the foregoing description, it will be perceived that the present invention presents a novel rodless, hydraulic reciprocating pump which is capable of adaptation to a variety of pumping needs and which in every application delivers a substantially constant, nonpulsating output. The pump is capable of absorbing extremely high input horsepower, and the power wasted in moving the engine pistons in their cylinders is reduced to aminimumby virtue of the ability of the mechanism to supply power fluid to the engine pistons only as their ability to travel in their pistons may require.

Changes in the details of construction of the pump of the present invention will occur to those skilled in the art, and, insofar as such changes employ only the use of equivalent elements and structures and do not depart from the principles of the invention, they are deemed to fall within the scope of the invention as defined by the following claims.

I claim: 7

1. A hydraulic pressurebooster comprising a plurality of fluid pressure operated engine pistons each having opposed engine piston areas for alternately receiving the impress of power fluid during operation of the booster; pumping means connected to each of said engine pistons and responsive to the movement thereof to pump a fluid; cylinders enclosing each of said engine pistons; decelerating means operative to decelerate the movement of said engine pistons when said engine pistons'approach the ends of said cylinders, and allowing free, unimpaired movement of said pistons away from the ends of said cylinders; means for delivering a substantially constant volume of power fluid; main valve means interconnected between the ends of said cylinders and said power fluid delivering means, said main valve means being shiftable between a first position to transfer power fluid to one of the ends of said cylinders, and a second position to transfer power fluid to the opposite ends of said cylinders; and timing means connected to said main valve means for shifting said main valve means periodically, independently of the movement of the engine pistons and after predetermined intervals of time whereby said engine pistons may be reciprocated in their respective cylinders with each stroke acting over a predetermined time interval which is independent of the time required for said pistons to travel any specific distance.

2. A hydraulic pressure booster comprising a plurality of fluid pressure operated engine pistons each having opposed engine piston areas for alternately receiving the impress of power fluid during operation of the booster; pumping means connected to each of said engine pistons and responsive to the movement thereof to pump a fluid;

cylinders enclosing each of said engine pistons; decelerating means operative to decelerate the movement of said engine pistons when said engine pistons approach the ends ofsaid cylinders, and allowing free, unimpaired movement of said pistons away from the ends of said cylinders; means for delivering a substantially constant volume of power fluid; main valve means interconnected between the ends of said cylinders and said power fluid delivering means, said main valve means being shiftable between a first position to transfer power fluid to one end of said cylinders, and a second position to transfer power fluid to the opposite ends of said cylinders; and timing means responsive to the delivery of a predetermined amount of power fluid by said power fiuid supplying means to throw said pilot valves whereby the strokes of said engine pistons may be reversed after an optimum total volume of power fluid has been delivered to one end of each of said engine cylinders.

3. A hydraulic pressure booster comprising a plurality of fluid pressure operated engine pistons each having opposed engine piston areas for alternately receiving the impress of power fluid during operation of the booster; pumping means connected to each of said engine pistons and responsive to the movement thereof to pump a fluid; cylinders enclosing each of said engine pistons; dash-pots cooperating with each of said pumping means to decelerate the movement of said engine pistons toward the ends of their respective cylinders; valve means in each of said dash-pots removing the vacuum in said dash-pots when said engine pistons move away from the ends of said cylinders to permit free, unimpaired movement of said engine pistons away from the ends of said cylinders; means for delivering a substantially constant volume of power fluid; main valve means interconnected between the ends of said cylindersand said power fluid delivering means, said main valve means being shiftable between a first position to transfer power fluid to one of the ends of said cylinders, and a second position to transfer power fluid to the opposite ends of said cylinders; and timing means connected to said main valve means for shifting said main valve means periodically, independently of the movement of the engine pistons and after predetermined intervals of time.

4. A hydraulic pressure booster as claimed in claim 1 wherein said decelerating means comprises a plurality of dash-pots to decelerate the movement of said pistons toward the ends of said cylinders; and valve means in each of said dash-pots removing the vacuum in said dash-pots when said pistons move away from the ends of said cylinders.

5. A hydraulic pressure booster as claimed in claim 1 and further characterized to include adjusting means for varying the period of said timing means while said booster is operating.

6. A hydraulic pressure booster as claimed in claim 1 wherein said power fluid delivering means comprises a rotary pump and said timing means comprises a pilot valve having its period of throw synchronized with the period required for said rotary pump to deliver a predetermined volume of power fluid.

7. A hydraulic pressure booster as claimed in claim 1 wherein said power fluid delivering means comprises a centrifugal pump; and regulating means connected to the discharge of said centrifugal pump for varying the period of said timing means to maintain the volume of power fluid supplied by said centrifugal pump constant despite fluctuations in the head of said power fluid downstream from said regulating means.

8. A hydraulic pressure booster as claimed in claim 1 wherein said pumping means comprises a pair of pump pistons connected to opposite sides of said engine pistons and a cylinder enclosing each of said pump pistons.

9. A hydraulic pressure booster as claimed in claim 8 wherein said cylinders enclosing said pump pistons are hydraulically segregated from said engine piston cylinders, and further characterized to include an intake conduit and a discharge conduit communicating with each end of each of said pump piston cylinders; and valve means in said conduits adapted to permit pumped fluid to enter said pump piston cylinders from said intake conduits and to be discharged from said pump piston cylinders into said discharge conduits upon reciprocation of said engine pistons.

10. A hydraulic pressure booster as claimed in claim 8 wherein said cylinders enclosing said pump pistons are hydraulically segregated from said engine piston cylinders, and further characterized to include a fluid passageway interconnecting the remote ends of the two pump piston cylinders adjacent each engine cylinder; a captive liquid enclosed in said remote ends of said pump piston cylinders and said conduit; an intake conduit and a discharge conduit communicating with the end of each of said pump piston cylinders opposite said remote end and adjacent said engine piston cylinders; and valve means in said conduits adapted to permit pumped fluid to enter said pump piston cylinders from said intake conduits and to be discharged from said pump piston cylinders into said discharge conduits upon reciprocation of said engine pistons.

11. A hydraulic pressure booster as claimed in claim 1 wherein there are two of said engine pistons, and said reversing valve means comprises a first pair of three-way valves each having two of its ports communicating with two of the ports of the other three-way valve in said first pair; said first pair of three-Way valves being responsive to said timing means to direct a portion of the power fluid delivered by said power fluid delivering means to one end of one of said engine piston cylinders, and, alternately, to the other end of said one engine piston cylinder; and a second pair of three-way valves each having two of its ports communicating with two of the ports of the other three-way valve in said second pair; said second pair of three-way valves being responsive to said timing means to direct the remaining portion of the power fluid delivered by power fluid delivering means to one end of the other of said engine piston cylinders, and, alternately, to the other end of said other engine piston cylinder.

12. A hydraulic pressure booster comprising a pair of fluid pressure operated engine pistons each having opposed engine piston areas for alternately receiving the impress of power fluid during operation of the booster; an engine cylinder enclosing each of said engine pistons; a piston rod extending from each of'the opposed engine piston areas; a pump piston carried by each of said piston rods and offset axially inward from the free ends of said rods; a pump cylinder enclosing each of said pump pistons; a dash-pot at the end of each of said pump cylinders most remotely located with respect to the nearest adjacent engine cylinder, said dash-pots each being of larger diameter than the free end of the adjacent piston rod and smaller diameter than the adjacent pump piston; an intake conduit communicating with each of said dash-pots at the bottom thereof; a check valve associated with each of said dash-pots operative to permit fluid to enter said pump cylinders from said intake conduits via said dash-pots while preventing the discharge of fluid from said pump cylinders into said intake conduits; a discharge conduit communicating with each of said pump cylinders; a check valve associated with each of said discharge conduits for preventing the ingress of fluid to said pump cylinders from said discharge conduits during theintake stroke of said pump pistons; means for supplying a substantially constant volume of power fluid to said engine cylinders; fluid actuated transfer valves associated with each of said engine cylinders for alternately directing power fluid from said power fluid delivering means to the opposite ends of each of said engine cylinders; a pilot valve hydraulically connected to each of said fluid actuated transfer valves and to said power fluid supplying means and adapted to throw said transfer valves when said pilot valves are thrown; and timing means responsive to the delivery of a predetermined amount of power fluid by said power fluid supplying means to throw said pilot valves whereby the strokes of said engine pistons may be reversed after an optimum total volume of power fluid has been delivered to one end of each of said engine cylinders.

References Cited in the file of this patent. UNITED STATES PATENTS 2,145,854 Bijur Feb. 7, 1939 2,420,896 Meyers May 20, 1947 2,486,079 Tucker Oct. 25, 1949 2,508,298 Saari May 16, 1950 2,819,835 Newhall Jan. 14, 1958

Patent Citations
Cited PatentFiling datePublication dateApplicantTitle
US2145854 *Jul 18, 1930Feb 7, 1939Auto Research CorpLubricating installation
US2420896 *Oct 4, 1945May 20, 1947Vulcan Soot Blower CorpReciprocating expansible chamber motor with pilot controlled distributing valve
US2486079 *May 18, 1945Oct 25, 1949Hpm Dev CorpHydraulic booster
US2508298 *Apr 16, 1948May 16, 1950Saari Oliver JFluid pressure intensifying device
US2819835 *Nov 26, 1954Jan 14, 1958Harwood Engineering CoSystem for delivering a continuous and steady flow of a compressible fluid at high pressure
Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US3256827 *Dec 21, 1964Jun 21, 1966James E SmithHydraulic power converter
US3319654 *Jan 20, 1964May 16, 1967Giovanni FaldiAir distributor for fluid pumping plants operated by compressed air
US3386384 *Jun 27, 1966Jun 4, 1968Cicero C BrownMultiple power consuming devices
US4011723 *Jun 28, 1974Mar 15, 1977Ross James JFluid power system
US4241581 *Dec 5, 1978Dec 30, 1980The Boeing CompanySynchronizer for hydraulic actuators
US4350082 *Mar 5, 1979Sep 21, 1982Dresser Industries, Inc.Pump piston having ring lubrication means
US4439114 *Mar 19, 1981Mar 27, 1984Kimmell Garman OPumping system
US4527959 *May 10, 1983Jul 9, 1985Whiteman Manufacturing CompanyDrilling fluid pump providing a uniform, controlled pressure and flow rate
US4666374 *Dec 14, 1984May 19, 1987Cooper Industries, Inc.Methods and apparatus for producing uniform discharge and suction flow rates
US5520520 *Mar 28, 1995May 28, 1996Nakamoto; TomijikoPneumatically operated double acting pump for viscous food stuffs
US7713033 *Nov 10, 2004May 11, 2010Halliburton Energy Services, Inc.Double-acting, duplex pump controlled by two, two position spool valves
US7967578Jan 27, 2005Jun 28, 2011Richard Frederick McNicholHydraulic gravity ram pump
US8454325Jan 30, 2008Jun 4, 2013Richard F. McNicholCoaxial pumping apparatus with internal power fluid column
US8535017Jun 27, 2011Sep 17, 2013Richard Frederick McNicholHydraulic gravity ramp pump
US8932030Aug 12, 2013Jan 13, 2015Mcnichol, Richard FrederickHydraulic gravity ram pump
DE2546600A1 *Oct 17, 1975Jul 29, 1976Tyrone HydraulicsSteuereinrichtung fuer eine hydraulikanlage
WO2005073555A1 *Jan 27, 2005Aug 11, 2005Richard Frederick McnicholHydraulic ram pump
Classifications
U.S. Classification417/347, 92/10, 417/397, 417/390, 91/40, 91/39, 91/36
International ClassificationF01L25/00, F04B9/117, F01L25/06, F04B9/113, F04B9/00
Cooperative ClassificationF04B9/113, F01L25/06, F04B9/1172
European ClassificationF04B9/113, F01L25/06, F04B9/117A