vickers hydraulic pump pressure adjustment free sample

Things like restrictions and blockages can impede the flow of fluid to your pump. which could contribute to poor fluid flow. Air leak in suction line. Air present in the pump at startup. Insufficient supply of oil in pump. Clogged or dirty fluid filters. Clogged inlet lines or hoses. Blocked reservoir breather vent. Low oil in the reservoir

Now that we’ve ensured that the directional control is not reversed, it’s time to check that the drive motor itself is turning in the right direction. Sometimes incorrect installation leads to mismatched pipe routings between control valves and motors, which can reverse the direction of flow. Check to see that the motor is turning the pump in the right direction and if not - look at your piping.

Check to ensure that your pump drive motor is turning over and is developing the required speed and torque. In some cases, misalignment can cause binding of the drive shaft, which can prevent the motor from turning. If this is the case, correct the misalignment and inspect the motor for damage. If required, overhaul or replace motor.

Check to ensure the pump to motor coupling is undamaged. A sheared pump coupling is an obvious cause of failure, however the location of some pumps within hydraulic systems makes this difficult to check so it may go overlooked

It is possible that the entire flow could be passing over the relief valve, preventing the pressure from developing. Check that the relief valve is adjusted properly for the pump specifications and the application.

Seized bearings, or pump shafts and other internal damage may prevent the pump from operating all together. If everything else checks out, uncouple the pump and motor and check to see that the pump shaft is able to turn. If not, overhaul or replace the pump.

If your pump is having problems developing sufficient power, following this checklist will help you to pinpoint the problem. In some cases you may find a simple solution is the answer. If your pump is exhibiting any other issues such as noise problems, heat problems or flow problems, you may need to do some more investigation to address the root cause of your pump problem. To help, we’ve created a downloadable troubleshooting guide containing more information about each of these issues. So that you can keep your system up and running and avoid unplanned downtime. Download it here.

An object of the present invention is to provide a liquid pressure system utilizing a variable delivery pump, and a pressure compensating device therefor, a liquid pressure motor, a fourway or directional control valve, and an automatic reversing device for said motor. In systems using variable pumps it is a practice to set the pump corresponding to certain load and predetermined speed. Theoretically such a varible pump should maintain a constant speed of the liquid actuated element regardless of a varying load, but practically, there is a change in slippage in said pump which results in a slowing down when an increased load is met or an increase in speed equivalent to the amount of slippage when the load is released. In practice, with systems using a variable pump, fairly good results are often obtained when the work is first set up, but when the oil warms up and thins, and as the tools become dull, there is a noticeable slowing up due to a change in the slippage of the variable pump.

The present invention contemplates a flow control arrangement for a hydraulic system which will provide a constant speed of the hydraulically actuated member regardless of the load conditions and pump slippage. The system, in the main, consists of a combination of a variable delivery pump and a hydraulic speed control member disclosed in my British Patent No. 399,609, dated June 11, 1932, together with an automatic pressure regulator and various control valves.

An illustration of the invention embodied in the hydraulic system having a reversible cylinder motor is set forth in the following description and accompanying drawings.

Referring to Fig. 2 the portion of a housing 4 is an enlarged portion of the liquid supply tank shown diagrammatically at 4a, in which is mounted a variable delivery pump 56and a pressure compensating unit 6. A motor 7 is arranged to drive the pump 5 which is hydraulically connected through a four-way valve 8 to a cylinder e motor 9. A flow control valve 10 and a control valve II are also connected in the hydraulic circuit.

Referring more specifically to the variable delivery pump, the pump part 5 is driven through a drive shaft 12 by the electric motor 7 or any other constant speed power means. The drive shaft 12 runs in ball or roller bearings within a bearing housing 13 and a flange 14, said flange being a part of the drive shaft. The drive shaft carries in a central bore a universal link IS which drives a cylinder block 16 with the same angular velocity as the drive shaft. The cylinder block 16 bears upon a bearing pin I1 and a swivel yoke 18. The swivel yoke 18 is supported by means of a link 19 which is connected to the pressure compensating unit 6. The cylinder block 16 is pivotally mounted on the bearing housing 13 by pins 20. A plurality of pistons 21 are mounted in axial bores or cylinders of the cylinder block 16, said cylinders have axis of symmetry coincident with axis of the cylinder block. By means of piston rods 22 which are provided with universal spherical end bearings, the pistons 21 are in axial positive connection with the drive shaft flange 14.

The variable delivery pump is of the well known type wherein the volume of delivery of the pump is dependent on the angle between the axis of the drive shaft 12 and the axis of symmetry of the pistons 21. For example, referring to Fig. 2 the pump would be in neutral position when the cylinder block 16 is in horizontal position, while the maximum volume positioned would be as shown with the swivel yoke 18 bearing against an adjustable stop 23. A flexible conduit 24 connects the inlet of the pump with the tank 4a and a conduit 24a, and a flexible conduit 25 connects the outlet of the pump to a conduit 26 which leads through the four-way valve 8, and conduit 26a which leads to the compensating unit 6.

The compensating unit 6 consists of a small cylinder 27, one side of which is in communication with the main pressure line through the 15 conduit 26a. A smaller cylinder or plunger 28 is slidably fitted co-axial with the bore of the cylinder 27 and a spring 29 normally urges the cylinder 28 to its extreme position shown in Fig. 2.

The flow control valve 10 which consists, in the main, of a balanced orifice valve 30 and an adjustable throttle 31 is adapted to maintain a predetermined differential pressure across the orifice of the throttle 31. A control lever 32 is provided for adjusting the throttle 31. The four5 way valve 8 is a standard type of four-way valve, consisting of a housing 33 and a slidable piston 34 operatively connected to an operating lever 35.

In the operation: The variable delivery pump is driven through the coupling or universal link 15 by the motor 7. When the system is starting, the pressure in the compensating unit 6 is negligible and the spring 29 will serve to displace the pump housing against the stop 23. Oil under pressure will flow from the pump conduit 25 through conduit 26 to the four way valve 8 where it will be directed to the cylinder 9 through the conduit 36 when the valve piston 34 is in the position shown in Fig. 2. This pressure in the piston end of the cylinder will cause the piston 31 to move downwardly and will thereby force the oil in the rod end of the cylinder through a conduit 44, around a check valve 45, and through the control valve II and conduits 46 and 47 to the valve 8 where it will be directed to the tank conduit 24a.

The return of the piston 37 is accomplished by shifting the handle 35 of the four-way valve 8 to move the valve piston 34 downwardly as shown in Fig. 2. In this position the four-way valve V( will direct pressure from the pump to the conduits 47 and 46. Oil under pressure will pass around the plunger 41 and will open the check valve 45 to pass through conduit 44 to the rod end of the cylinder 9. The piston 37 will then 61 be returned to starting position at rapid traverse rate. During this traverse the driving piston is locked between two columns of oil under pressure, thus insuring positive feed rates. The oil at the piston end of the cylinder 9 is forced through the four-way valve back to the tank.

It will be understood that during any given 6 feeding stroke of the piston 37 that the flow controlling unit 10 maintains a fixed rate of exit of liquid from the bottom end of cylinder 9. Accordingly, there is also a constant rate of flow of liquid into the upper end of cylinder 9. It fol- 7 lows then that the compensating mechanism 6 will move the yoke 18 into a position where the pump displacement is equal to the rate of flow into the upper end of cylinder 9. The: yoke 18 is positively maintained in such a position be- 7 cause if it tended to move toward neutral and thus decrease the rate of delivery into cylinder 9, the pressure would fall, thus permitting spring 29 to overcome the force on piston 27 and move the yoke 18 upwardly. If the yoke 18 tended to move upwardly to a position of greater displacement than that corresponding to the rate of flow out of cylinder 9, the pressure would build up in lines 36 and 26, thus permitting piston 27. to overcome the force of spring 29 and bring the yoke back to its intended position.

The force of the spring 29 and area of piston 27, of course, are so chosen as to maintain in line 36 a pressure greater than that required to overcome the maximum resistance encountered at tool 39 plus the friction of movement of piston 37. Under conditions where the tool resistance is less than maximum, the additional resistance to movement of piston 37 is produced by building up pressure on the liquid in the bottom of cylinder 9 and in pipe 49. It is inherent that the pressure above piston 37 must be equal to the sum of the pressure below the same plus the tool resistance. Thus as the tool resistance decreases, the pressure in line 49 must increase so as to maintain this sum constant, Since the action of the flow control device 10 is independent of the pressure in line 49, it is obvious that the piston 37 moves at a fixed rate.

It is also obvious that since the stroke of the pump is automatically controlled to pump only the quantity of fluid required for driving the motor that no power is wasted in the device except the small quantity which escapes to the suction side of the system by the slippage in the pump itself. Of course, should a variation in slippage occur at the pump 5, the compensating mechanism 6 will automatically increase the stroke setting of yoke 18 slightly and sufficiently to make Sup for the additional slippage. The rate of flow from pipe 36 into cylinder 9, however, remains constant at any rate of slip in pump 18 because liquid cannot flow into cylinder 9 faster than it can flow out of the same. If the tool 39 meets Swith any unusual resistance or if the operator fails to reverse the four-way valve when the piston 37 reaches its lowermost position, the pressure in the system will rise and the volume of the pump 5 will decrease. If the pressure reach0 es a maximum, the pump will deliver no liquid at all thus insuring no rupture of the system.

What I claim is: 1. In combination in a hydraulic system, a hydraulic motor having a piston and cylinder, a variable delivery pump for preloading the inlet side of the motor during the power stroke, said pump having a movable part for varying the delivery capacity, a spring pressed piston arranged to urge the movable part 0 to on-stroke position, means connecting the pressure on the outlet side of the pump between the pump and the motor to said spring pressed piston to act in opposition to the spring whereby the volume output of said pump is pro5 portional to the requirements of the system, and a speed control device including a chamber for receiving the liquid flow from the outlet side of the motor, a variable discharge orifice member, a chamber on the intake side of said orifice mem0 ber, a valve between said two chambers, and a spring pressed piston operatively connected to said valve for controlling the pressure in said second named chamber to control the pressure differential and the flow across said orifice mem5 ber, one side of said piston being in direct communication with said second named chamber and the other side of said piston, which is the side upon which the spring acts, being in direct communication with the discharge outlet on the discharge side of the variable orifice member, said speed control device and said means cooperating to effect constant speed movement of said motor piston irrespective of pump slippage and resistance met thereby.

2. In a hydraulic power transmission system the combination of a variable displacement pump, a fluid motor, conduits forming a circuit extending from the pump outlet to the motor and from the motor to the pump inlet, means forming a predetermined restriction in said con- 16 duits, valve means responsive solely to the pressure differential across said restriction and connected to maintain such differential substantially constant, and means responsive to slight variations in pump outlet pressure for controlling the pump displacement comprising a spring pressed piston arranged to put said pump on-stroke, and means connecting the pressure on the outlet side of the pump between the pump and the motor to said spring pressed piston to act in opposition to the spring, said means coacting together to maintain the motor speed substantially constant independently of variations in motor load and of variations in pump slippage.

This invention relates to power transmissions and particularly to hydraulic circuits for actuators such as are found on earth moving equipment including excavators and cranes.

The invention more particularly relates to hydraulic systems for automatic braking of preselected braking pressures of swing devices found, for example, in excavators and cranes. Swing drives usually comprise a hydrostatic drive having a hydraulic pump and motor, and associated gearing and controls that direct the horizontal rotation of booms found on excavator and cranes.

Swing drive arrangements have utilized the control of fluid velocity or flow to the motor through a directional control valve. With velocity or flow control, the operator selects the direction and flow of fluid at system pressure.

Typically, flow control of the swing drive provide free swing or coasting of the boom on cranes. That is, in the absence of a command signal in the hydraulic system, the boom or the boom and load will coast to a stop due to frictional forces without excessive oscillation of the boom cable or the load.

Excavators are usually arranged with flow control to provide blocked center braking of the boom. That is, the boom or the boom and load will immediately decelerate to a stop in the absence of a command signal. In such use, return flow from the motor is relieved at the motor work port by a relief valve at a predetermined pressure setting. The blocked center braking allows rapid alignment of the boom and load and also provides for maintaining the boom stationary with the excavator operating on an inclined surface.

It is also desirable, under certain conditions of operation, to brake the swing drive at a preselected reduced pressure; i.e. a pressure setting below the relief valve pressure setting.

In view of the foregoing, it is an object of this invention to provide a hydraulic circuit arrangement for automatic braking at preselected pressures of swing drives wherein an operator may selectively choose, by means of a simple adjustment, a free swing braking arrangement, or reduced pressure braking anywhere between the free swing and blocked center braking arrangements.

In accordance with the invention, the velocity control braking arrangement disclosed herein comprises a hydraulic control valve system, such as disclosed in U.S. Pat. No. 4,201,052 having a common assignee with the present application, including a pilot controller, a pump, and a hydraulic actuator. The actuator includes a movable element and a pair of openings adapted to function alternately as inlets or outlets for moving the element in opposite directions. The pilot controller supplies fluid to the system at pilot pressure and the pump supplies fluid at pump pressure to the motor. The control system includes a line adapted for connection to each of the openings. A meter-in valve means controls fluid flow from the pump to the motor and is selectively operable by pilot pressure from the pilot controller. A meter-out valve is associated with each of the lines for controlling fluid flow from the motor. The meter-out valves are each selectively pilot operated by pilot pressure from the pilot controller. In accordance with the invention, the supply fluid being supplied to the actuator is applied, at a predetermined pressure, to the meter-out valve means controlling flow from the actuator in opposition to the pilot pressure which tends to open the meter-out valve means.

Referring to FIG. 1, the hydraulic system embodying the invention comprises an actuator 20, herein shown as a rotary hydraulic actuator, having an output shaft 21 that is moved in opposite directions by hydraulic fluid supplied from a variable displacement pump system 22 which has load sensing control in accordance with conventional construction. The hydraulic system further includes a manually operated controller 23 that directs a pilot pressure to a valve system 24 for controlling the direction of movement of the actuator, as presently described. Fluid from the pump 22 is directed to the line 25 and line 26 to a meter-in valve spool 27 that functions to direct and control the flow of hydraulic fluid to one or the other end of the actuator 20. The meter-in valve spool 27 is pilot pressure controlled by controller 23 through lines 28, 29 and lines 30, 31 to the opposed ends thereof, as presently described. Depending upon the direction of movement of the valve, hydraulic fluid passes through lines 32, 33 to one or the other end of the actuator 20.

The hydraulic system further includes a meter-out valve 34, 35 associated with each end of the actuator in lines 32, 33 for controlling the flow of fluid from the end of the actuator to which hydraulic fluid is not flowing from the pump to a tank passage 36, as presently described.

The hydraulic system further includes spring loaded poppet or drop check valves 37, 38 in the lines 32, 33 and spring-loaded anti-cavitation valves 39, 40 which are adapted to open the lines 32, 33 to the tank passage 36. In addition, spring-loaded poppet valves 41, 42, are associated with each meter-out valve 34, 35 acting as pilot operated relief valves. A bleed line 47 having an orifice 49 extends from passage 36 to meter-out valves 34, 35.

The system also includes a back pressure valve 44 associated with the return or tank line. Back pressure valve 44 functions to minimize cavitation when an overrunning or a lowering load tends to drive the actuator down. A charge pump relief valve 45 is provided to take excess flow about the inlet requirements of the pump 22 and apply it to the back pressure valve 44 to augment the fluid available to the actuator.

Meter-in valve comprises a bore in which a spool 27 is positioned and the absence of pilot pressure maintained in a neutral position by springs. The spool normally blocks the flow from the pressure passage 26 to the passages 32, 33. When pilot pressure is applied to either passage 30 or 31, the meter-in spool is moved in the direction of the pressure until a force balance exists among the pilot pressure, the spring load and the flow forces. The direction of movement determines which of the passages 32, 33 is provided with fluid under pressure from passage 26.

When pilot pressure is applied to either line 28 or 29, leading to meter-out valves 34 or 35, the valve is actuated to throttle flow from the associated end of actuator 20 to tank passage 36.

It can thus be seen that the same pilot pressure which functions to determine the direction of opening of the meter-in valve also functions to determine and control the opening of the appropriate meter-out valve so that the fluid in the actuator can return to the tank line.

Provision is made for sensing the maximum load pressure in one of a multiple of valve systems 24 controlling a plurality of actuators and applying that higher pressure to the load sensitive variable displacement pump 22. Each valve system 24 includes a line between lines 32, 33 having a shuttle valve 50 therein that receives load pressure from one of the adjacent passages 32, 33. Shuttle valve 50 senses which of the pressures is greater and shifts to apply the higher pressure to pump 22. Thus, each valve system in succession incorporates another shuttle valve 51 which compare the load pressure therein with the load pressure of an adjacent valve system and transmit the higher pressure to the adjacent valve system in succession and finally apply the highest load pressure to pump 22.

The details of the preferred construction of the elements of the hydraulic circuit are more specifically described in the aforementioned U.S. Pat. No. 4,201,052 which is incorporated herein by reference.

In accordance with the invention, when the meter-in valve spool 27 is operated to provide supply pressure to one of the openings of the actuator, the supply pressure is also applied to prevent venting of the spring loaded poppet valves 41, 42 which serve as relief valves for meter-out valves 34, 35. As shown in FIG. 1, an adjustable relief valve 52 is connected by line 53 through lines 54, 55 having check valves 56, 57 therein to the poppet valves 41, 42 that control the meter-out valves 34, 35.

When an operator commands an output pressure or flow by introducing pilot pressure at C1 from a remotely located hydraulic remote control, for example, to shift the meter-in spool of meter-out valve 27 to the right, FIG. 1, and fluid would flow from "P" to acutator Port "B". The pilot pressure would also cause meter-out valve 34 to open permitting flow out of the actuator. The load would be accelerated up to a speed determined by the level of pilot pressure. When the operator desires to stop the load, he removes the pilot pressure at "C1" by centering the hydraulic remote control. The flow being supplied to the chamber between the meter-in valve spool 27 and the load check valve 38 for cylinder Port "B" will cease and the chamber will be allowed to drain through pilot line C2. The spring chamber of the pilot relief valve 52 will be at low pressure. The relief valve 52 will establish a back pressure acting on the balance piston of poppet valves 41 or 42, and will allow the pilot piston to open, thereby allowing the meter-out element 34 or 35 to function as a relief valve for application of load pressure at Port "A" or "B".

When a high inertia load has been accelerated up to full speed by flow supplied from Port "B", and the command at C1 ceases, the load will tend to keep running and cause flow into Port "A". The balance piston of poppet valves 41 will be allowed to open at a pressure determined by the pilot relief valve 52 which drains into the chamber between the meter-in valve 27 and "B" port load check valve 37.

The level of braking pressure can be preselected by adjusting the spring force of the relief valves. The range can be from very low pressure, or free coast, up to the maximum relief valve setting. When a load is being driven and pressure is present in the chamber between the meter-in spool and the load check, the additional relief valve will not function.

Although the invention is especially applicable to a hydraulic circuit utilizing pilot operated meter-in and meter-out valves; it may also be utilized with manually operated, mechanically operated or electrically operated valves. Also, the system can be applied to loads other than swing drives such as vehicle propulsion drives and winch drives.