CA1334812C - Fluid transfer device - Google Patents

Abstract

A liquid transfer device of the type having a fluid driven linear actuator including a reciprocating output drive member and driven from a double acting cylinder having a pair of fluid ports; a power source providing a fluid flow through an outlet port; and a control device for controlling the rate of fluid flow alternately to and from said fluid ports to thereby control the velocity of travel of the output member. The control device has a valve provided with an inlet port in fluid communication with the outlet port of the power source, a pair of ports in fluid communication one each with the cylinder ports; and an exhaust port. A valve element is disposed in the valve movable between two extreme positions for controlling the rate of flow of fluid flow alternatively from the outlet port of the power source to each of the cylinder ports and from the cylinder ports to the exhaust port, and a control element is provided for cycling the position of the valve element between the extreme positions.

Description

1- 133~12 This invention related to a fluid transfer system, and more particular to a pumping device of a type which may be used as a pump jack in oil fields or for pumping fluids through conduits, such as pipelines.

As is described in Canadian Patent No. 1,032,064, issued May 30th, 1978, entitled "Pump Jack Device" to Minoru Saruwatari, the conventional walking beam type oil lift pump has numerous disadvantages. Because of the large massive parts customarily used in such a pump, the shipping, erection and maintenance costs associated with such a pump are high, and also the start up power demands with which such apparatus are high. High stress variations are developed in the rod string because of the velocity profile developed by the drive system in the walking beam type of pump. The drive system includes an eccentric which is rotated at a constant speed so that the velocity profile produced at the polished rod is sinusoidal. This velocity profile, particularly in pumping heavy oil in deep wells has many inefficient portions. The well characteristics which dictate the speed of the sucker rod at various points along the pumping cycle in the conventional walking beam structures can only be met by increasing or decreasing the input speed at the eccentric thereby effecting the speeds of the entire cycle. Thus, the speed of a sucker rod must travel at much slower speeds than would be possible for other portions of the sucker rod velocity cycle. In order to increase the output of the pump and to minimize the loss of input energy, it is desirable to maintain the speed of a sucker rod at its maximum permissable speed through as many portions of the pumping cycle as possible. The selection of a par-ticular velocity profile during turn around at each end of the pumping stroke is particularly important in obt~; n; ng efficient pumping. In between turn arounds it ~b 133~812 is preferable to achieve quickly the highest possible constant speed to be maintained through the main up and down portions of the strokes.

There exists on the market a number of pump jacks which include other types of drive systems, but in the main, a large percentage of the pump jacks remain to be the walking beam type. The pump jack of the type disclosed in above identified Canadian Patent No.
1,032,064 has been successful in that it has provided a pump jack with significantly increased pumping efficien-cies. There is a need, however, for a hydraulically operated type of pump jack which provides some options with respect to controlling the velocity profile, while maintAi n; ng a reasonable initial cost. There is also a need for a system which does not require operators with a high knowledge of electronic maintenance.

In the pumping actuator or cylinder structure utilized in the hydraulic type pump jack shown in above identified Canadian Patent No. 1,032,064, and also in Canadian Patent No. 1,076,926, issued May 6th, 1980, entitled "Pump Jack Device", to Minoru Saruwatari, the hydraulic cylinder is mounted immediately above the well bore, and the counter balance cylinder and the hydraulic cylinder are mounted in series, i.e. one above the other. When it is desired to use a long pumping stroke, the counter balance and hydraulic m~ch~n;cal structure become very tall. Also, in order to mount the hydraulic cylinder directly on top of the well bore, it is necessary to utilize the piston rod of the power cylinder as the sucker rod in the well bore head.
Problems in alignment of the power cylinder with the well bore head can develope, and this may result in high wear of the seals of the power cylinder.

'A.

~ _ 3 _ 1 33~812 It is an object of the present invention to provide a compact power cylinder which is versatile and which has good operating characteristics requiring a minimum of maintenance.
It is also an object of the present invention to provide a fluid transfer system which includes a relatively simple control system having the ability to produce a velocity profile best suited for the reciprocating output drive of the system.
According to one aspect of the present invention, there is provided a pump actuator and counterbalance device which includes a base member, a cylinder head, a hydraulic cylinder having opposite ends in engagement with the base member and the cylinder head, and a counterbalance inner tube of greater diameter than the hydraulic cylinder being concentrically disposed about the hydraulic cylinder and having opposite ends engaging the base member and the cylinder head. A counterbalance head of generally cylindrical shape and having opposite ends of reduced diameter is mounted for sliding travel longitudinal of the inner tube, and a counterbalance outer tube is mounted on top of the counterbalance head and extends upwardly therefrom, with a counterbalance cap being mounted on an upper end of the counterbalance outer tube. A piston is disposed in the hydraulic cylinder for reciprocation ~~
therein. The cylinder head has a central bore therethrough, -~ ~ 4 ~ 1334812 and a piston rod is connected to the piston and projects upwardly through the central bore and is connected at an upper end to the counterbalance cap.
One form of the invention involves a system wherein the base member has fluid ports and the counterbalance outer tube and cap provide a reciprocating output drive. The system includes a proportional metering valve for controlling the rate of fluid flow alternately to each cylinder port to thereby control the maximum velocity of the output drive in each of its two directions of travel. The valve has a valve body and an axially shiftable spool valve member in the valve body which is movable along a cyclic path of travel to locate alternatively between two extreme positions and wherein the valve member in each of the positions establishes a maximum flow control rate to one of the cylinder ports. The valve member during travel along a first portion of the path from one of the positions to the other decreases the rate of flow of the fluid to one cylinder port and then along a second portion of the path increases the rate of flow to the other cylinder port. A
valve control means is provided for shifting the valve member along the path at a predetermined rate to thereby control the rate of change of velocity of the output drive from maximum velocity in one direction to the maximum velocity in the other direction of travel of the output drive.

~ . .

~ _ 5 _ 1 33 ~812 In the accompanying drawings, which show examples of the present invention, Figure 1 is a detailed schematic of a fluid transfer device of the present invention;

Figure 2 is a simplified schematic of one form of the control,device;

Figure 3 is an alternative design of the schematic shown in Figure 2;

Figure 4 is a partial sectional view of one form of proportionally metering valve used in the control device of the present invention;

Figure 4a is a graph showing the relation of the flow of the fluid through the valve to the spool travel, F:~ure 5 is a cross sectional view through another form of a proportional metering valve;

Figure 6 is a more detailed graph than Figure 4a, showing the flow output versus spool position for a ~ ~ 6 1334812 proportional metering valve with spools of different characteristics;

Figure 7 is a schematic of yet a further variation of the control device shown in Figure 2;

Figure 8 is a schematic of yet a further variation of the control device shown in Figure 2;

Figure 9 is a schematic of yet a further variation of the control device shown in Figure 2;

Figure 10 is a schematic of yet a further variation of the control device shown in Figure 2;
Figure 11 is a schematic of yet a further variation of the control device shown in Figure 2;

Figure 12 is a simplified schematic of an alterna-tive embodiment of fluid transfer device shown in Figurel;

Figure 13 is a simplified schematic showing a regenerative hydraulic circuit which is present in one form in the schematic of Figure 12.

Figures 14a and 14b are side views of the top and bottom portions respectively of a double acting cylinder structure according to the present invention, as seen along line 14--14 of Figure 16;

Figures 15a and 15b are upper and lower portions of a cross sectional view of the structures of Figures 14a and 14b but taken along the line 15--15 of Figure 16;

7 1~3~812 Figure 16 is a bottom view of the cylinder struc-ture shown in Figures 14a and 14b and Figures 15a and 15b;

Figure 17 is a cross sectional view similar to Figure 14b and showing only the lower portion of the cylinder structure modified to provide a piston rod projecting from the bottom of the cylinder structure.

lo In Figure 1, there is shown a pump jack structure 30 which is connected by way of a sucker rod to a pump string 31 within a well bore (not shown). The pump jack structure 30 includes a hydraulic actuating cylinder 32 which imparts a reciprocating motion to the pump string 31. The lower portion of the pump jack structure 30 includes a counterbalance 33 which, of course, is of the type for accumulating energy during the down stroke of the pump string and to return the accumulated energy to the system during the up stroke to thereby conserve energy. In the embodiment shown, the hydraulic actuat-ing cylinder 32 is mounted immediately above the counterbalance 33, i.e. the actuating cylinder 32 and counterbalance 33 are in effect in series. The hydraulic actuating cylinder 32 includes a cylinder 35 which has in communication with opposite ends thereof hydraulic lines 36 and 37. Disposed within the hydraulic cylinder 35 is a piston 40 connected through the lower end of the cylinder by way of a piston rod 41 to a piston 42 contained within a cylinder 43 of the counterbalance 33. A piston rod 44 extends through the bottom of the counterbalance cylinder 43 and may provide the polished rod extending through the well head 49 and then being connected to the pumping string 31 within the well bore. The lower end of the counterbalance cylinder 43 is in communication by a fluid line 46 with an accumulator system 45 which may be in the form of a , . ~, ~ - 8 - 1 334812 vessel containing pressurized nitrogen. Thus, the pressurized nitrogen occupies the space in the cylinder 43 below the counter balance cylinder 42. Thus, the chamber in cylinder 43 under the piston 42 together with the container forming the accumulator 45 provide a volume in which the nitrogen is compressed during the downward stroke of the pump.

The fluid transfer system shown in Figure 1 includes a pump 47 which may be a constant output pump provided with relief valve or the pump is preferably a variable flow pump. Although more expensive, a pressure compensated and/or flow compensated pump may be used for more efficient operation then would be provided by a constant flow output pump. The output of the pump is connected to a hydraulic line 48 which is in communication with a hydraulic line 50 extending to a directional control valve 51. The directional control valve 51 has an exhaust port which is connected to a hydraulic line 52 which returns through a cooling system 53 to a reservoir 54. The directional control valve 51 includes a spool (not shown) which is shifted between two extreme positions to direct the flow from the hydraulic supply line 50 either to hydraulic line 36 leading to the upper end of the hydraulic cylinder 35 or the hydraulic line 37 which extends to the bottom of the hydraulic cylinder 35. The position of the spool in the directional control valve 51 is determined by actuation of actuators 55 and 56 disposed at opposite ends of the directional control valve 51 as will be described in more detail below. The actuators are controlled from position sensors 57 and 58 shown as being disposed adjacent to the counter balance cylinder 43. A
transducer device, such as that indicated at 57' in Figure 10, or limit switches may be provided to determine the position of the pumping stroke. Various means may be provided, for example, such as a mechanically operated hydraulic ,:.

switch or a mech~n;cally operated electrical switch. As shown in Figure l, the sensors 57 and 58 are electrical, receiving a current from a battery 60 by way of an electrical lead 61. When lower sensor 57 is activated, electrical lead 62 provides current to a solenoid actuator 63 located at one end of a directional control valve 64 in a control section 59 of the hydraulic circuit of the present invention. When the upper sensor is sensed as the piston 42 of the counter balance reaches the upper end of the stroke, electrical lead 65 then carries a current to the solenoid actuator 66 at the opposite end of the directional control valve 64.
The directional control valve 64 has an input port connected by a supply line 67 to hydraulic output line 48 of the pump 47. Depending on which solenoid is activated, i.e. solenoid actuator 63 or solenoid actuator 66, output line 68 or output line 69 of the directional control valve 64 is activated. The output flow from line 68 is directed through a flow control valve 70 and the flow of output line 69 is directed through a flow control valve 71. From flow control valve 70 fluid is conducted to actuator 56 by way of hydraulic line 72 and the fluid flow from flow control valve 71 is conducted to the hydraulic actuator 55 of the directional control valve 51 by hydraulic line 73.
Thus, it can be seen that the spool valve member of the directional control valve is shifted from one direction to the other as hydraulic sensors 57 and 58 are alterna-tively activated. As flow control valves 70 and 71 are adjustable to control the rate of flow therethrough, the rate at which the pressurized fluid is supplied to the actuators 55 and 56 is controlled, and therefore the rate of speed of travel of the spool from one of its position to the other is controlled. By providing two separate flow control valves, the speed in one direction can be adjusted to be different from the speed in the other direction for the valve spool.

133~812 Figure 2 shows the directional control valve and its relation to the hydraulic actuating cylinder 32 and the source of the pressurized fluid in the form of pump 47 separate from the remaining circuit. It can be readily seen that if actuator 55 is energized to shift the spool member of the directional control valve 51 to the left, the fluid output of the pump 47 is directed to line 36 while line 37 is connected to exhaust. Thus, the piston 40 of the hydraulic actuating cylinder descends. The actual rate of descent is not necessarily controlled solely by the fluid flowing from the pump 47 into the line 36. The rate at which the fluid is allowed to return to the reservoir from line 37 through the directional control valve 51 may in fact control the rate of descent. A number of factors, such as the setting of the counter balance, the weight of the sucker rod, etc. determine whether in fact the piston 40 is actually being driven downwardly by the pressurized fluid entering the cylinder through the line 36.
A simplified schematic of the present invention is shown in Figure 3 wherein the directional control valve is shown as a proportional metering valve 51a. The circuitry is otherwise the same as in Figure 2.
However, a proportional meter valve is capable of allowing the circuitry to provide a more sophisticated velocity profile for the output of the hydraulic actuating cylinder 32. Figure 4 shows, a partial cross sectional view of a proportional metering valve for use 30 as valve 51a in Figure 3, and it includes a housing 75 including a spool 76 which has a plurality of spaced lands, which permit metered flow from one spaced annular groove about the bore to another as will be described in more detail in conjunction with Figure 5.
In the proportional metering valve of Figure 4, the housing 75 has fluid chambers at the opposite ends of .

, ~ . . .

~ - 11 - 1334812 the spool and provide the actuators 55 and 56 of the directional control valve. Ports 77 and 78 communicated with the chambers and are connected to hydraulic lines 73 and 72, respectively, in Figure 1. The proportional metering valve 5la of Figure 4 is provided with threaded members 80 and 81 which extend into opposite ends of the chambers of the actuators 55 and 56 so as to be engaged by the end of the spool 76 as it is shifted in either direction. Thus, if pressurized fluid is admitted through port 78 via hydraulic line 72 (Figure 1) to shift the valve to the position shown in Figure 4, its travel to the right is limited by engagement with the threaded member 80. Thus, the manually adjusted member 80 determines the final position of the spool valve 76 in its final position at the right hand end of its cycle. Similarly the adjustable member 81 determines the spools final position when shifted to the left as fluid enters port 77 and the fluid is exhausted through port 78.
As will be described in more detail in conjunction with Figure 5, the flow through the valve from its inlet to the output port is controlled by an interaction between spaced lands which allow flow from one circum-ferential groove in the bore containing the spool toanother. Similarly the flow which would be entering a port and going to the exhaust port of the directional control valve would be controlled by the interaction between the lands and the grooves. Figure 4a shows the flow relationship to the spool of travel which is the result of the rate of flow past the spaced lands. For example, as the spool travels from the left towards the right in Figure 4a, the flow to one outlet port of the proportional metering valve, such as that connected to line 37 in Figure 1, would follow the line of the graph , . .
.

~ - 12 - 1334812 down to the zero value and then the flow to the other outlet port such as that connected to line 36 in Figure 1 would follow the line of the graph up to the maximum flow on the right hand side of the draft. At the extreme travel in either direction, the flow to the lines 36 and 37 would alternatively be at the m~;mum shown at the top of each side of the graph. If the member 80 is threaded further in to stop the travel of the spool to the right, for example, then the maximum flow output for the output port controlled when the spool is at that end would be less than the permitted m~; mum because the maximum spool travel would be prevented, and the flow would be at a lower value as indicated in the line of the graph, such as X.
The cross sectional view of the proportional control valve shown in Figure 5 provides more details with respect to the manner in which the flow output versus spool position is established. In this Figure the proportional metering valve has a body or housing 75 on which are mounted actuators 55 and 56 which control the flow of fluid via passages 82 and 83 to opposite ends of the spool 76b. The spool is centered by a spring mechanism 84 attached to one end of the spool 76b. The housing has an inlet port 85 which communicates with a central annular groove. Located on either side of the central annular groove 85' and spaced therefrom are two inner grooves 86' and 87' which communicate with outlet ports 86 and 87. The spool has spaced lands 88 and 89 which have at their opposed ends V shaped notches 91 and 92. Thus, as the spool is shifted in one direction, the flow from inlet 85 to outlet 86 is controlled by flow through the groove 91.
When the valve is shifted in the opposite direction, the flow to the outlet 87 is controlled by groove 92. At the same time, control of fluid flow from an inlet to the exhaust port of the proportional metering valve is being controlled in the same manner. Thus, depending on B

~ - 13 - 1334812 the shape of the notches, for example, a variation in the flow output versus spool position can be achieved.

Looking at the graph of Figure 6, if a spool design is selected to provide the middle line shown as K, for example, and if the spool travel is adjusted so that in the travel in the direction of the arrow A shown at the bottom of the graph, the flow to one outlet valve is decreased from 15 gallons per minute as shown at the extreme left hand vertical line, following the middle line to zero output flow at the bottom of the graph.
The flow to the other outlet port increases along the line on the right hand side of the graph until it reaches the vertical line C which represents an output flow of approximately 12 gallons per minute at the point X. When the valve is shifted in the opposite direction, then the flow from the last mentioned output valve is decreased from 12 gallons per minute to zero and then the flow to the first mentioned outlet valve is in-creased back to the rate of 15 G.P.M.

It may be readily appreciated, therefore, that thedownward maximum speed of the piston 40 may be governed by the end setting which would in turn control the m~imum flow rate through the valve to the outlet port controlling the flow through hydraulic line 36. The upward stroke could be set at a different value as il-lustrated in the graph represented in Figure 6 so as to allow, for example, for a higher flow through the outlet port connected to conduit 37. Again, it should be kept in mind that the flow through the line 36 may not in fact govern the downward speed of the piston 40 as it may be the flow rate permitted for the fluid returning through line 37 to exhaust which governs the flow rate.
However, the principle is the same.

The characteristics of the turn around at the opposite ends of the travel of the piston 40 is also B

.., ~ - 14 - 1334812 controlled by the rate that the spool travels along the lines represented by the arrows A and B in Figure 6.
This in turn is controlled by the manner in which actuators 55 and 56 operate. In the valve of Figure 5, if connected into the circuit of the diagram shown in Figure 1, the action of the actuators 55 and 56 would be controlled by flow control valves 70 and 71. If the flow, for example, through valve 70 is permitted to be at a higher rate, then the actuator 56 would move more quickly to open the flow into passage way 82 more quickly so that a spool valve would be moved to the left faster than it is moved to the right as viewed in Figure 5. This, of course, would result in a quicker turn around time, for example, when the piston 40 reaches the top, comes to a stop and then starts to descend, than when the piston reaches the bottom and comes to a stop and starts to ascend.

It may be seen, therefore, that with a relatively simple circuitry, including a proportional metering valve, a very customized velocity profile can be selected for the output of the polished rod of the pump jack structure.

Referring to Figure 7, this is a simplified hydraulic schematic in which the actuators 55 and 56 may be in the form of hydraulic actuators or solenoid actuators but which might receive the signal for shifting the spool of the valve directly from limit switches actuated by the movement of the piston assembly of the pump jack actuator. The proportional metering valve 51c, however, in this arrangement is shown with the adjustable stops for the spool, as previously described as threaded members 80 and 81. A simple arrangement of the type depicted by Figure 7 would allow for a variation between the upward and downward strokes ..~
., - 15 - 133~
because it would, as previously described, establish the limit of travel of a spool in either direction to thereby in effect move the stopping point of the spool inwardly from the maximum shown at the outer edges of the graph of Figure 6.

In the arrangement shown in Figure 8, the propor-tional metering valve 51d is provided with actuators 55d and 56d which are the variable solenoid type. Such solenoids could obtain there energization directly through limit switches controlled by the piston 40. By adjusting the variable solenoids, however, their rate of activation can be changed so that the speed of travel of the spool in the proportional metering valve is con-trolled. Thus, with such an arrangement, the turnaround characteristics of the piston 40 may be adjusted by changing the rate at which the spool of the propor-tional metering valve moves from one end to the other.
This in effect makes it possible to select the speed at which the spool travels in the direction indicated by the arrow A in the graph of Figure 6 and/or in the direction of the arrow indicated by B in the graph of Figure 6.

In the schematic of Figure 9, the proportional metering valve 51e is provided with hydraulic actuators, i.e. hydraulic pilots to control the flow of fluid to opposite ends of the main spool of the metering valve.
The hydraulic actuators 55e and 56e, however, are shown as being controlled by variable solenoids 93 and 94, respectively, which may be activated directly through limit switches which are controlled by the position of the piston 44. The variable solenoids control the rate at which the hydraulic actuators are shifted, and therefore, it is possible with this arrangement to independently control the turn around times for the output rod of the pump jack structure.

~ - 16 - 1~3~

Figure 10 schematic of a system similar to Figure 1, but simplified to depict a proportional metering valve 51f which is provided with adjustable end limit members for the spool, such as thread members 80 and 81. Thus, in this arrangement the velocity along the main stroke may be independently controlled in opposite directions, and the turn around time can be independent-ly controlled as well as discussed in connection with proportional metering valve 51a shown in Figure 4.

In the embodiment shown in Figure 11, the same functions can be obtained as in the embodiment of Figure 10 wherein the adjustable limit members were provided.
However, there are provided valve members 94 and 95 which are the type to provide variable pressure reduc-ing. The use of such valves make it possible to control the pressure of the flow to hydraulic actuators 55 and 56 so as to control the maximum spool displacement in either direction. The variable pressure reduction valves are used, therefore in place of the mechanical adjusting devices 80 and 81 of Figures 4 and 10.

As will be described in conjunction with the embodiment of the system shown in Figure 12, the control of the proportional metering valve 51g may be carried out by way of a fluid circuit utilizing compressed air.
There are various other means available for controlling the manner and extent to which the spool of the propor-tional metering valve is shifted, for example, thetransducer 57' (Figure 10) of a type commercially available, which would continually sense the position of the piston assembly of the pump jack structure could be provided which could continually feed information to a programmable electronics device for providing an output signal to control the actuators of the proportional metering valve. Alternatively, a servo mechanism could ` ~ - 17 - 133~8~2 be utilized for providing a current for controlling the equivalent of the actuators described above. While the above description referred to variable solenoids for controlling the rate of activation of the activators, a varying current could be provided which would determine the amount of shift in addition to the rate of shift of the output of the actuators, thus, making it possible to accomplish the exact profile of the velocity output of the pump jack structure.
Referring now to Figure 12, the shown control means of this embodiment includes an air operated section in lieu of the hydraulic control section 59 of the embodi-ment shown in Figure 1. The fluid transfer device of Figure 12 includes an air pump 103 which provides air to a pressure tank 104 from which pressurized air is provided via an air line 105 not only to the control system but also to the counter balance portion of the double acting cylinder structure 200 described further below. As shown in the upper left hand corner of Figure 12, the cylinder structure of this embodiment is of a type which combines the hydraulic cylinder 201 with an outer counter balance cylinder 202 described below. Two double acting cylinder structures 200 are used in parallel with the polish rod 44a connected to a cross beam 100 attached to the top of the twinned cylinder structures. The air passing through line 105 is provided with a lubricant mist by the device shown at 105a. As will be described in more detail below, the mist ladened air enters the accumulator chamber and as there is a continual feed of air to the counter balance chamber, the mist collects in the chamber in order to ensure a proper level at all times, and excessive lubricant which might collect there is dispelled automatically.

~ - 18 - ~33~12 There is provided an accumulator tank 108 which is in communication with the counter balance chambers of the double acting cylinder structure 200 by way of line 107. The accumulator tank 108 is maintained charged by a connection line 109 to the air supply. Manual control valves 110 are provided in the air lines 107 connecting the counter balance chambers to the accumulator tank 108 in order that during maintenance the position of the polish rod may be controlled at a location to the well bore head. An air supply line 111 is connected to limit switches 112 and 113 activated by the upward and downward movement of the outer liner of the accumulator of one of the double acting cylinder structures 200.
Thus, the output of pressurized air to air lines 114 and 115 is controlled, the lines 114 and 115 being connected to the actuators at the end of directional control valve 64 which controls the flow of hydraulic fluid from line 116 which is connected to the pump 47. The output ports of the directional control valve are connected to pressure reducer valves 94 and 95 functioning in the manner described in connection with Figure 11. The flow of fluid to the pressure reducer valves 94 and 95 passes through flow control valves 70 and 71 which provide the same function as those shown in Figure 1. The flow from flow control valves 70 and 71 and the pressure reducer valves 94 and 95 is conducted to the actuators 55 and 56 at the end of the proportional metering valve 51g so that the shifting of the spool is controlled individual-ly in opposite directions both with respect to the rate of shifting, by way of flow control valve 70 and 71 and the final position of the spool by way of pressure reducer valves 94 and 95.

The hydraulic lines 36 and 37 extend from the proportional metering valve 51g to inlet passages 226 and 236 of the hydraulic power cylinder portion of the !j~, ~ , .

133~812 double acting cylinder structure 200, which are described in more detail below. Lines 36 and 37 may be part of a regenerative hydraulic circuit as will be described in more detail below. The selection of the piston rod and the piston of the double acting cylinder, such as that shown in Figures 14 to 17 may be such that the area of the piston subjected to the pressurized fluid below the piston is twice the area to which the pressurized fluid is subjected above the piston. The result of the this effective area ratio for the piston is that during the up stroke of the piston twice as much fluid flows through inlet 236 (Figure 14b) of the double acting cylinder structure 200 to raise the piston as flows through inlet passage 226 (Figure 14b). In order to be able to utilize a smaller pumping system as well as smaller hydraulic lines, the regenerative hydraulic circuit 120 is utilized. A more detailed schematic of this circuit is shown in isolation in Figure 13. This circuit includes a pressure relief valve 121 which allows the flow of fluid being expelled through line 36 to flow through line 122 to the line 37. Normally flow is prevented from flowing back to the proportional metering valve 51f because of the presence of a valve 123 in the line 36. The valve 123 is a type which is spring loaded to a closed position as shown in Figure 13 but is moveable to a position if an actuator 124 thereof senses a pressure above a predetermined amount in line 37. When the flow is to the upper end of the cylinder through line 36, the flow by passes valve 123 by way of line 126 through a pressure relief valve 127.
Accordingly, it can be seen that with the two to one ratio indicated above, the flow from the outlet port of proportional metering valve 51f connected to the line 37 is the same as to the line 36, the difference being required in the lower end of the cylinder being made up by the hydraulic fluid passing directly from line 36 to line 37.

I 3~812 Referring now to Figures 14a to 16, there is shown a double acting cylinder structure 200 which is in the form of a combined hydraulic cylinder 201 centrally disposed within an outer counterbalance cylinder 202, thus, it is readily apparent that the hydraulic power cylinder and counterbalance cylinder are in effect in parallel rather than in series such as the one shown in the schematic diagram of Figure 1. The cylinder structure 200 would normally be used, therefore, with an identical twin cylinder disposed parallel to it and spaced from it. A cross beam may be mounted across the upper ends of the twin cylinders, with a polish rod connected to the cross beam midway therebetween.
Alternatively, a cross beam could be connected between counter balance outer tubes 273 of a pair of cylinder structures 200 at a location closer to the lower ends thereof which allows for better cylinder stiffness. In other words, the longitudinal axes of the twin power cylinders 200 would be parallel to but spaced on opposite sides of the longitudinal axes of the well bore.

Also, as will be described in more detail below, cylinder structure 200 may be easily modified so as to be mountable, as a single unit, directly over the well bore with the piston of the cylinder structure connected to the polished rod of the well bore.

The cylinder structure 200 has a main base portion 203, (Figure 14B) having a central bore 204 passing perpendicular therethrough. The bore 204 is provided with a larger upper counter bore 205 and a lower larger counter bore 206. As shown in Figures 14b and 15b, a plug member 207, which is provided with an upwardly projecting annular portion 208 is secured to a bottom surface 210 of the base 203 by way of bolts 211 passing ~ ~ - 21 - 1 3 ~4 812 through openings in an annular flange 212 and being threaded into threaded bores (not shown) in the bottom of the base 203. An O ring 213 is provided in the outer annular groove of the bore 206, so that when the plug 207 is bolted into place, it closes the lower end of bore 204. A lower end of a main hydraulic cylinder 215 is received in the upper counter bore 205 and is sealed relative to the bore 205 by way of a seal ring 216 received in a peripheral groove in the counter bore 205. The hydraulic cylinder 215 is slightly longer than the total stroke required for the output of the cylinder structure 200. The upper end of the hydraulic cylinder 215 is received in a bore 217 in a cylinder head 220.
The cylinder head 220 is bolted to the base by way of elongated bolts, the upper end of one of which is denoted as 221 in Figure 14a. There is provided between the upper end of the hydraulic cylinder 215 and the end of the bore 217 a spacer 222, and a seal ring 223 is located between the end of the spacer and the end surface of the hydraulic cylinder 215. An additional seal 224 is provided between the outside of the end portion of the hydraulic cylinder 215 and the cylinder head 220, this seal ring being located in an annular groove in the bore 217. The base 203 is provided with an inlet passage 226 which is connected to hydraulic line or conduit 36 of the hydraulic system described above. The passage 226, which is horizontally disposed and has a vertical component 226a communicates with a counter bore 227 in the upper surface of the base 206 (see Figure 14b). The cylinder head 220 is provided with a passage 230 which communicates with an upward extension of the bore 217 of the cylinder head. Thus, the inner end of the passage 230 is in communication with the interior of the hydraulic cylinder 215.
Communicating with passage 230 is a bore 231 which extends upwardly from the lower surface of the cylinder ` - 22 - 133 4812 head. A hydraulic passage way 232 is provided between passage 226 and 230 by way of a rigid tubular member 233 which is received at its lower end in counter bore 227 and at its upper end in bore 231. The ends of the tubular member, 233, are sealed relative to the base 203 and the cylinder head 220 by way of seals 234 and 235 received in the annular grooves in counter bore 227 and bore 231. It may be seen, therefore, that pressurized hydraulic fluid from line 36 connected to inlet or passage 226 communicate with the upper interior end of the hydraulic cylinder 215 by way of passage 232 communicating at its bottom end with passage 226 and at its upper end with passage 230, which in turn, communi-cates with the interior of the upper end of the hydraulic cylinder 220.

The base 203 is provided with a second inlet passage 236 for connection at its outer end to the hydraulic line 37 described above. The inner end of the inlet passage 236 which is diametrically opposed to the inlet passage 226, communicates with the central bore 204 of the base 203. A~cordingly, the inlet passage 236 is in communication with the lower end of the hydraulic cylinder 215 so that pressurized fluid from hydraulic line 36 is free to pass through inlet passage 236 into the portion of the hydraulic cylinder projecting into the base 204.

The upper surface of the base portion 203 is provided with a large bore 237 and the lower surface of the cylinder head 220 is provided with a large bore 239 as well, the bores 237 and 239 being of substantially the same diameter. A large tubular member forms an inner tube 240 of the counter balance chamber, the opposite ends of the inner tube being received in large bores 237 of the base 203 and the large bore 239 of the a ` ~ - 23 - 133 481~

cylinder head 220. The outer periphery of the lower end of the inner tube is sealed relative to the base by way of an annular seal 241. The hydraulic cylinder 215 and the counter balance inner tube 240 are concentrically disposed and define an annular space 242 therebetween.
The tubular member 233 is located in the annular chamber 242.

As may be best seen in Figure 15B, the base 203 is also provided with an inlet passage 244 which may be disposed 90 relative to the inlet passages 226 and 236. The inlet passage 244 is adapted to be connected to air line 46 discussed above. Inlet passage 244 which extends horizontally into the base communicates at its inner end with a vertical passage 245 connected to a port or opening 246 which is in communication with the interior of the inner tube 240. As may be seen in Figure 15A, the cylinder head 220 is provided with a port 247 so as to communicate the area above the cylinder head 220 with the annular chamber 242.
Therefore, the area above the cylinder head and the annular chamber 242 are placed in communication with the accumulator system 45 by way of passages 244 and 245 and fluid line 46.
The base is further provided with a makeup air inlet passage 250, which is connected to above described air line 105. The inner end in the passage 250 is connected to a threaded vertical passage 251. As can be seen from Figure 14A, the lower surface of the cylinder head 220 is provided with a threaded vertical passage 252 which communicates with a horizontal outlet nozzle 253. An air hose 254 having threaded couplers 255 at opposite ends thereof are threaded into the threaded 35 vertical passages 251 and 252 of the base 203 and the cylinder head 220 thereby placing the inlet passage 250 ~0 ` - 24 - 133~81~

in communication with the outlet nozzle 253 so that the air which is laden with hydraulic mist expels into the area above the cylinder head 220.

Slidably mounted on the exterior of the inner tube 240 of the counter balance chamber is a cylindrical shaped counter balance head 257. The counter balance head 257 has an enlarged central portion and end portions of smaller diameter provided with inwardly projecting circumferentially spaced, longitudinally extending ribs 258 which support annular wear rings 260 engaging the outer surface of the outer cylindrical surface of the inner tube 240. At the lower most portion of the counter balance head 257, there are provided annular grooves receiving annular seals 261.
Immediately above the seals 261 the counter balance head is provided with radially ex~ ;ng threaded bores 262 which receive plugs 263 which may be removed to permit oil maintained in the annular space 264 between the counter balance head 257 and the inner tube 240 to be drained and thereby flush any wear particles accumulated at the bottom of the counter balance head. Near the upper end of the annular chamber 264, the counter balance head 257 is provided with a threaded bore 265 which receives a site plug 266.

Near the upper end of the counter balance head 257, there is provided a threaded bore 267 which receives an automatic drain device 268. The drain device 268 may be of a commercially available type which periodically expels any liquid accumulated therein together with a blast of air from the pressurized air contained within the interior of the counter balance head 257.

It may be seen, therefore, that by maintaining the oil level within the annular chamber 264 at appropriate ~ ~ - 25 - 1 3~ ~ 8 ~ 2 height, which can be monitored by way of site plug 266 the wearing surfaces provided by wearings 260 and the outer surface of the inner tube 240 are continually exposed to the lubrication. As further indicated above, periodic partial draining of the chamber 264 by opening the plugs 263 guarantees removal of any wear particles from the main wear area between the reciprocating counter balance chamber and the inner tube 240. The level of the oil within the chamber may be observed through site plug 266 but, in any event, because of the continued admission of lubricant to the chamber via the discharge of lubricant particles with the air issuing from outlet nozzle tube 53 there is a continued feel of lubricant to the chamber. The rise of the level of lubricant within the chamber 264 is governed, however, by excess oil being expelled through the automatic drain device when it reaches the threaded bore 267.

The counter balance cylinder 202 further includes a counter balance cap 270 which has a downwardly project-ing cylindrical portion 271 having a peripheral groove thereabout. The upper end of the counter balance head 257 terminates in an upwardly extending cylindrical portion 272 having the same outside diameter as the cylindrical portion 271 of the counter balance cap 270.
The cylindrical portion 272 is also provided with a peripheral groove. A counter balance outer tube 273 which is substantially of the length of the working stroke of the hydraulic cylinder piston is affixed between the cap 270 and the head 257 by way of bolts 274 which extend through aligned openings in an outer flange 275 of the cap 270 and an upper flange 276 of the counter balance head 257. Seals 277 and 278 are provided in the peripheral grooves of the cylindrical portion 271 of the counter balance cap and the cylindrical portion 271 of the counter balance head. Thus, as ~.~

` ~ - 26 - 1 ~3~812 the piston of the hydraulic cylinder travels to the lower end of its stroke, the air accumulated within the counter balance chamber, including the volume between the cylinder head 220 and the counter balance cap 270 is compressed, and this compressed air is in communication with air line 46 through port 257, inner chamber 242 and passages 245 and 244.

Telescopically ext~n~ing into the hydraulic cylinder 215 from its upper end is a piston rod 280 which has ext~n~;ng into its lower end a threaded bore 281 (Figure 14B). A piston member 282 having an axial bore 282a in its upper surface extends over the lower end of the piston rod and is secured thereto by way of a bolt 285 ext~n~;ng through a central bore 286 of the piston and being threaded into the threaded bore 281.
The piston has a elongated annular grooves adjacent opposite ends thereof which receive piston rings 284.
The piston further has a groove around the centre portion thereof receiving a wear ring 283. The cylinder head 220 is provided with an axial bore 287 there-through. The bore 287 has annular grooves therein which receive seals 288. An elongated annular groove at the lower end of the bore receiving a wear ring 289 in engagement with the piston rod 280. The upper end of the piston rod has a reduced threaded portion 291 ext~n~;ng through a central bore 292 of the counter balance cap 270. A nut 293 threaded onto the reduced threaded portion 291 secures the outer or upper end of the piston rod 280 to the cap.

As previously explained, the piston rod may be selected with a cross sectional area which is one-half of the cross sectional area of the lower end of the piston 282 so that the cross sectional area of the annular chamber 295 provided between the piston rod 280 ~r!

~ ~ - 27 - 133~812 and the hydraulic cylinder 215 is one-half the cross sectional area of the hydraulic cylinder 215 which equals the effect of lower surface of the piston 282.

It can be seen that as fluid is admitted through the inlet passage 236, the piston is driven upwardly and the piston rod therefore extends, moving with it the portion of the counter balance structure including the counter balance cap 270, the counter balance outer tube 273 and the counter balance head 257 which slides along the outer surface of the inner counter balance tube 240.

As may be seen from Figure 17, it is possible to replace the lower cylinder plug 207 with a gland casting 300 which is bolted to the bottom of the base in the same manner as the plug 207. The gland casting, however, has a central axial bore 301 extending there-through, the internal bore being provided with a first long annular groove and a pair of outer narrower annular grooves. The long groove contains a wear ring 302 and the outer grooves contain seals 303. A piston rod 304 extends through the bore 301 and has a reduced threaded portion 305 which extends through the central bore 286 of the piston 282 and is threaded into the threaded bore 281 of the piston rod 280 so as to secure the piston rod 304 to the piston rod 380. Thus, the cylinder structure can be converted to one of a through rod type. The structure illustrated in Figure 17, for example, could be used to mount a cylinder structure immediately over the well head, the piston rod 304 thus providing the polished rod of the well head and being connected at its lower end to the sucker rod string.

4~

Claims (12)

1. A pump actuator and counterbalance device compris-ing:
a base member, a cylinder head, a hydraulic cylinder having opposite ends in sealing engagement with said base member and said cylinder head, a counterbalance inner tube of greater diameter than said hydraulic cylinder being concentrically disposed about said hydraulic cylinder and having opposite end engaging said base member and said cylinder head, a counterbalance head member of generally cylindrical shape and having opposite ends of reduced diameter mounted for sliding travel longitudinal of said counterbalanced inner tube, a counterbalance outer tube mounted on top of said counterbalance head and extending upwardly therefrom, a counterbalance cap mounted on an upper end of said counterbalance outer tube, a piston disposed in said hydraulic cylinder, for reciprocation therein, said cylinder head having a central bore there-through, and a piston rod connected to said piston and project-ing upwardly through said central bore and being connected at an upper end to said counterbalance cap.

2. A pumping apparatus including a pair of said pump actuators and counterbalance devices as defined in claim 1, said devices being mounted in parallel, side by side and further including a horizontal cross beam connected therebetween for movement with said counterbalance heads of the pair of devices, and an output drive member connected to said cross beam.

3. A pump actuator and counterbalance device as defined in claim 1, wherein said counterbalance head includes a mid-portion having an inner surface of greater diameter than said outer diameter of said counterbalance inner tube for thereby defining an annular lubricant chamber therebetween.

4. A pump actuator as defined in claim 3, and including an inlet in communication with said lubricant chamber for admission of air carrying a lubricant mist.

5. A pump actuator as defined in claim 4 and including an automatic drain device communicating with said lubricant chamber at a level above a lower portion of said lubricant chamber.

6. A pump actuator as defined in claim 1, wherein said base member includes a pair of hydraulic ports, first passage means extending from one of said ports and communicating with a low end of said hydraulic cylinder, second passage means communicating with an outlet port in said base member within said counterbalance inner tube, said cylinder head having passage means in communication with an inlet port within said counterbalance inner tube and with an upper end of said hydraulic cylinder, and hydraulic conduit means within said counterbalance inner tube and placing said outlet port in said base member in communication with said inlet port of said cylinder head.

7. A pump actuator as defined in claim 1, wherein said cylinder head is provided with an opening placing a chamber provided between said counterbalance inner tube and said hydraulic cylinder in communication with a counterbalance chamber within said counterbalance outer tube.

8. A pump actuator as defined in claim 1, wherein said base member has a passage formed therein and connected to an accumulator connection port, said passage communicating with a port in said base member opening into the chamber between said counterbalance inner tube and said hydraulic cylinder.

9. A pumping system comprising:
a fluid driven linear actuator;
a power source providing a fluid flow through an outlet port;
said fluid driven linear actuator being a pump actuator and counterbalance device including a base member, a cylinder head, a hydraulic cylinder having opposite ends in sealing engagement with said base member and said cylinder head, a counterbalance inner tube of greater diameter than said hydraulic cylinder being concentrically disposed about said hydraulic cylinder and having opposite end engaging said base member and said cylinder head, a counterbalance head member of generally cylindrical shape and having opposite ends of reduced diameter mounted for sliding travel longitudinal of said counterbalanced inner tube, a counterbalance outer tube mounted on top of said counterbalance head and extending upwardly therefrom, a counterbalance cap mounted on an upper end of said counterbalance outer tube, a piston disposed in said hydraulic cylinder, for reciprocation therein, said cylinder head having a central bore there-through, and a piston rod connected to said piston and project-ing upwardly through said central bore and being connected at an upper end to said counterbalance cap, said counterbalance outer tube and cap providing a reciprocating output drive, said base member having ports, and means placing said ports in communication with opposite ends of said hydraulic cylinder;
a proportional metering valve for controlling the rate of fluid flow alternatively to each said port to thereby control the maximum velocity of said output drive in each of its two directions of travel, said valve having a valve body and an axially shiftable spool valve member in said valve body and being movable along a cyclic path of travel to locate alternatively between two extreme positions, said valve member in each of said positions establishing a maximum flow control rate to one of said cylinder ports, said valve member during travel along a first portion of said path from one of said positions to the other decreasing the rate of flow of the fluid to one cylinder port and then along a second portion of said path increasing the rate of flow to the other cylinder port; and a valve control means for shifting said valve member along said path at a predetermined rate to thereby control the rate of change of velocity of the output drive from maximum velocity in one direction to the maximum velocity in the opposite direction of travel of said output drive,

10. the pumping system of claim 9 wherein said valve control means includes actuator means for alternately shifting said spool member and sensors for energizing said actuator means in response to predetermined sensed positions of said output drive, and said proportional metering valve has adjustable means for selectively and independently determining the limit of travel of the spool member in either direction.

11. the pumping system of claim 9 wherein said valve control means includes hydraulic actuator means for alternatively shifting said spool member, sensors for energizing said actuator means in response to predetermined sensed positions of said output drive, and said valve control means includes a directional valve for controlling the flow of fluid to said hydraulic actuators through separate hydraulic lines, each of said hydraulic lines including a separate variable pressure reduction valve for controlling maximum spool displacement in either direction.

12. the pumping system of claim 9 wherein said valve control means includes hydraulic actuator means for alternately shifting said spool member and sensors for energizing said actuator means in response to predetermined sensed positions of said output drive, and said valve control means includes a directional valve for controlling the flow of fluid to said hydraulic actuators through separate hydraulic lines, each of said hydraulic lines including a separate variable pressure reduction valve for controlling maximum spool displacement in either direction.