shadi

2009-07-08T10:34:00.000-07:00

Fluid power accessories

Miscellaneous items

Some components used in fluid power systems do not necessarily fall into any of the categories discussed in preceding chapters. These accessory items may be used for powering, modifying, monitoring, or connecting in any type circuit, as the system designer deems appropriate.

Pneumatic accessories

Quick exhaust valves: The speed at which an air cylinder strokes is determined by how fast compressed air enters it and how fast the air already in the cylinder exhausts to atmosphere. System pressure drives air into the cylinder and this does not pose a speed problem in most circuits. Air leaving the cylinder is different because it was at system pressure when the directional valve shifted. Although the air starts exiting quickly, it still holds the piston back. Speeding up a sluggish air-operated cylinder is best accomplished by dealing with its exhaust air. The cross-sectional view and symbol in Figure 18-1 illustrate a quick exhaust valve, which does just that.

Fig. 18-1. Quick-exhaust valve increases air cylinder's stroking speed

The cylinder in Figure 18-1 delivers high impact from low force . . . stamping parts with steel dies and leaving a lasting impression. Cylinder force alone is not capable of making the desired impression -- if any impression at all. Accelerating piston speed over a few inches of travel makes the weight of the tooling act like a hammer swung through the air.

As the cylinder retracts and is held at rest, the shut-off wafer covers the exhaust port and forces air to the cylinder rod end. When the directional valve shifts to extend the cylinder, pressure drops on the left side of the shut-off wafer and trapped pressure in the cylinder forces the wafer to the left. As the shut-off wafer moves left, it closes off flow to the valve and opens a direct path to atmosphere only a short distance from the cylinder port. The rapid exhaust of air reduces backpressure on the cylinder piston, allowing high-pressure inlet air to accelerate and move the piston very quickly.

Any time slowly exhausting air is a problem, look to a quick exhaust valve to remedy the situation.

Mufflers: The air-exhaust mufflers in Figure 18-2 reduce the noise level of air-operated equipment. They are made in several different configurations out of many types of material, but the end result of all of them is the smooth discharge of air.

Fig. 18-2. Typical pneumatic mufflers

The sintered-bronze elements on the left are similar to filters made of the same material. They separate the flowing air into numerous paths to lessen or eliminate the loud crash of air as it leaves an actuator. The sintered-bronze element in the center has a protective metal covering and an adjustable poppet valve to control flow. It works as an inexpensive meter-out flow control when used with a 5-way directional control valve. Because a 5-way valve has two exhaust ports, these speed-control mufflers can regulate speed independently in both directions of travel. The muffler on the right is similar to those used on internal combustion engines. It may be made of plastic or aluminum. It is bulky, but causes less restriction on fast-moving actuators.

Accessory Items for pneumatics and hydraulics

The components described in the rest of this chapter are common to hydraulics or pneumatics. The main difference between them is the materials used to make them. Many pneumatic components can be made of plastic or aluminum to resist corrosion and keep cost down. These materials work well at low pressure. Most hydraulic components see high to very high pressure and need to be much more robust. Cast iron and steel are common materials for hydraulic parts due to their strength and the absence of corrosion. Aluminum is also preferred by some because of its light weight.

Pressure gauges: The gauges shown in Figure 18-3 come in a variety of shapes, sizes, and designs. The most common is the round model that has a moving needle to designate system pressure. The round gauge on the left and the plunger gauge measure psig, not atmospheric pressure. Because atmospheric pressure is in and around an actuator, it doesn't help or hinder performance, so it is not important when determining the amount of work being done.

Fig. 18-3. Four types of pressure gauges

The gauge marked PSIA reads atmospheric pressure instead of zero and can be used to check vacuum as well as pressure. Some of these gauges set on zero and read psi clockwise and vacuum (in inches of mercury) counter-clockwise.

Other designs include battery-operated digital-readout units. These gauges are accurate and very fast reading.

Temperature: Knowledge of the temperature of a fluid or the atmosphere in which it works can be very important. Two styles of temperature gauge are shown in Figure 18-4. When pneumatically operated machines are in atmospheres of 32° F or less, the condensed moisture in them may freeze. When hydraulic circuits operate much above 140° F they can leak or slow down and the fluid in them starts to break down.

Fig. 18-4. Temperature gauges

It is best to keep hydraulic systems between 75° and 130° F. Temperatures above 130° F can vaporize important additives and cause excessive bypass due to reduced fluid viscosity. Fluid temperatures below 75° F can result in sluggish performance.

Flow meters: The cross-sectional view in Figure 18-5 shows a typical inline flow meter that indicates flow in cubic feet per minute (cfm), gallons per minute (gpm), or liters per minute (lpm). This style of meter is made of aluminum or non-magnetic stainless steel to allow the magnet-powered notched steel ring to function.

Fig. 18-5. Cross-sectional view of flow meter

Fluid entering from the left passes through flow holes and against a spring-returned piston fitted with magnets. This piston wraps around a tapered metering cone and has a sharp-edged orifice in contact with it. The only way for fluid to get through is to push the spring-returned piston with magnets to the right. When the piston moves far enough up the tapered metering cone to allow the present rate of fluid to pass, it stops and holds. The magnets in the piston draw the notched steel ring along and the notch reads the flow amount on the clear-plastic cover with flow scales.

This type flow meter is not completely accurate but gives a clear enough indication of flow to meet most troubleshooting needs. Other designs are more accurate but less tolerant of the harsh interaction of a high flow system.

The upper symbol on the right in Figure 18-5 is for a device that only shows whether flow is taking place in the line or not. The middle symbol represents the cross-sectioned device. It indicates both the presence of flow and the flow rate. The lower symbol represents a device that shows the flow rate and keeps a running total of the amount that has passed through it.

Shuttle valves: The circuits in Figure 18-6 illustrate one reason for using shuttle valves. The spring-return cylinder in the upper circuit must be controlled from three locations. This circuit uses pipe tees to interconnect the three normally closed, palm-button-operated, 3-way directional control valves with the cylinder. The only problem is this circuit will not work. When any of the 3-way valves are actuated, input air can flow directly to atmosphere through the other 3-way valves, bypassing the cylinder.

Fig. 18-6. Cylinder circuits with shuttle valves

The lower circuit uses shuttle valves in place of the pipe tees. Air from any of the 3-way directional control valves can only go to the cylinder. The floating ball in the shuttle valve blocks air to the other directional control valves. Exhausting air can go to atmosphere through the valve it entered, go out the opposite valve, or exhaust through both valves. If each 3-way directional control valve has a different pressure at its inlet (as indicated), the cylinder always gets the highest pressure of the valves actuated. The ball in the shuttle valve always moves away from the highest inlet pressure.

Other circuits use shuttle valves to send more than one pilot signal to a directional control valve, read feedback signals from more than one source, or send signals from multiple actuators to a load-sensing pump. Any time multiple inputs are necessary, a shuttle valve will separate them, allow for return flow, and pass the highest input pressure. (Check valves can serve two of these functions but will not allow back flow.)

Rotary unions: Some applications require fluid to flow into or out of rotating parts of a machine. The rotation may be continuous or only part of a turn; the application may have one or many flow paths. Many manufacturers make rotary unions that do this for fluids at pressures as high as 5000 psi, with as many as 20 flow paths. . (Some rotary unions pass electricity as well as fluids if required.) The cross-sectional view in Figure 18-7 is a simplified drawing of a single-path rotary union. The symbol is a circle on a flow line; in this case, the energy triangle indicates hydraulic fluid. Multiple flow paths are shown by multiple lines of whatever type the flow is. (Some rotary unions pass electricity as well as fluids if required.)

Fig. 18-7. Cross-section of rotary union

Quick disconnects: When all or any part of a pneumatic or hydraulic circuit must be removed or changed frequently, a fast way to do so is with quick-disconnect couplings. Quick disconnects usually require a worker to connect and disconnect them manually. However, there are some styles that break away when pulled by mechanical force. Other types only stay connected while held in place by an external force.

The cross-sectional view in Figure 18-8 illustrates the socket-and-plug pair that make up a typical quick disconnect. Sliding the lock-unlock ring to the left allows the detent balls to move out of the way so the plug can be inserted. Inserting the plug all the way into the socket stops leakage as it passes the O-ring seal, opens both check valves, and allows the detent balls to lock in the detent notch to hold the connection together. Sliding the lock-unlock ring to the left again releases the plug as the detent balls lose their backing. The three symbols in the figure show quick disconnects disconnected with dual check valves, connected with dual check valves, and disconnected in a typical air line configuration.

Fig. 18-8. Cross-section of typical quick disconnect coupling

It may be necessary to install oversized quick disconnects because their construction can cause high backpressure. Always check pressure drop in the manufacturer's catalog to assure proper flow capabilities. There are designs that have full flow porting in air and low-pressure hydraulic styles.

Pressure switches: Some fluid power circuits require electrical control signals when pressure reaches specific levels - such as the pressure buildup when a part is clamped or a certain weight is met - or if overpressure may cause damage or is a safety hazard. (Sequence valves -- discussed in Chapter 14 -- can cycle from a pressure buildup, but will not produce a signal to an electrical control circuit when a pressure requirement is satisfied.)

The cross-sectional view and symbols in Fig 18-9 show electrical pressure switches that are set to monitor maximum or minimum pressure and then send a signal to the electric control circuit. (Another electrical output device that reads pressure and sends a signal is a pressure transducer. Pressure transducers are more responsive and have better repeatability, but require additional electronics to read their input.)

Fig. 18-9. Cross-section of pressure switch

The depicted pressure switch includes a plunger that reacts to system pressure by moving. An adjusting screw sets spring pressures against the plunger and allows different settings. When system pressure is high enough to push the plunger upward against spring tension, the plunger closes a limit switch to signal that the set pressure has been reached. When pressure falls, the plunger drops and the limit switch opens again.

Never depend on a pressure switch to indicate actuator position when the actuator positively has to be in a certain position to prevent machine or product damage or to avoid a safety hazard.

Limit switches: Figure 18-10 shows an outline drawing and symbols for a limit switch. While some electrical components are shown to indicate function and location, no wiring appears on fluid power circuit diagrams.

Fig. 18-10. Limit switch

Shock absorbers: The cross-sectional view in Figure 18-11 shows an oil-filled shock absorber. (The figure also includes a proposed symbol.) When cushioned cylinders or other decelerating devices are not satisfactory or desirable, shock absorbers are one viable alternative. Shock absorbers are available in sizes from 3/8 in. or less up to models that can stop a loaded overhead crane traveling at full speed in two feet or less. Some are adjustable, some are self- adjusting. Some use metering orifices (as the figure shows), others use tapered metering cones.

Fig. 18-11. Cross-sectional view of oil-filled shock absorber

Because they may absorb a lot of energy in a short period, most have the ability to transfer fluid from the last stroke to a reservoir for cooling. The same reservoir replenishes the shock absorber with cooled fluid for the next stroke.

The model depicted in Figure 18-11 uses a spring-returned piston with an integral check valve that travels through a bore with several metering orifices in it. As the piston moves through its bore, there are fewer holes for fluid to pass through. Thus, resistance to movement increases throughout the stroke. As the piston strokes, it smoothly decelerates the load at a controlled rate until it stops. Fluid forced out of the bore during the deceleration stroke is sent into an oil chamber that is partially filled with a closed-cell foam accumulator. This accumulator makes it possible for the oil chamber to accept the extra fluid and then force it back to the bore on the return stroke. An oil-fill port allows replenishment of any lost fluid.

Most shock absorber manufacturers offer formulas in their catalogs and/or computer programs to size their products for specific applications.

Hose-break valves: In pneumatic systems, there usually is more air available than is required, so if a hose ruptures or is disconnected suddenly, air will flow profusely and the loose end of the hose can whip about dangerously. A hose-break valve set for a flow greater than working flow will close automatically when flow tries to increase above its capacity. Air hose-break valves never shut off completely so when the line is reconnected, the small bleed bypass fills the repaired section and the hose-break valve opens for use.

Fig. 18-12. Cross-sectional view of hose-break valve

The drawing and symbol in Figure 18-12 represent a typical hose-break valve. Air flows into the right-hand port at a rate up to a certain cfm setting. The distance between the shut-off poppet and its seat determines the maximum flow rate before the shut-off poppet closes and stops flow. Reverse flow is never blocked, but is restricted to cause a pressure drop. When pressure at the right-hand port drops or when pressure at the left-hand port rises, the shut-off poppet opens to pass flow.

Air-oil cylinders, tanks, and intensifiers

Air-oil systems

Compressed air is suitable for many low-power systems, but air’s compressibility makes it difficult to control actuators smoothly and accurately. Some low-power systems need the smooth control, rigidity, or synchronization capabilities normally associated with oil hydraulics. All of these features are available to low-power circuits by using compressed air for power and oil for control. Purchased or specially built air-oil circuits give smooth control when the power requirement is low.

Attached oil-control cylinders

Some manufacturers offer attached oil-filled cylinders to control speed and/or position, Figure 17-1. These units usually work in one direction of travel in a meter-out circuit. They operate such things as drill feeds or other actions that may try to pull the cylinder out. (They also can be used with hydraulic cylinders at higher forces.)

Fig. 17-1. Hydraulically controlled air cylinder – set up for fast advance, controlled feed stroke, and fast retraction

Most manufacturers offer units with valves in the oil line that can stop flow and/or bypass the speed control. The stop control allows an air cylinder to be stopped reasonably accurately with very good repeatability. The bypass control makes it possible to have fast and controlled speeds as the cylinder advances.

The cross-sectional view in Figure 17-1 shows an air cylinder that advances rapidly with airflow controls until its attaching bracket comes in contact with the fast-advance stroke-length adjustment. At this point, air cylinder movement is retarded and controlled by the oil speed-control cylinder as oil flows through a flow control. The air cylinder cannot move any faster than the oil flow allows during this part of the cycle. A spring-loaded oil balance cylinder furnishes oil to make up for the differential loss from rod to cap ends. The air cylinder is controlled by oil flow for the remainder of the cycle.

As the air cylinder retracts and the attaching bracket contacts the rod nut, it pushes the oil speed-control cylinder back to the start position. A flapper-type 1-way check valve on the piston with through holes allows fluid to transfer back to the rod end. Excess cap end fluid is stored in the spring-loaded oil balance cylinder during this part of the cycle.

Some manufacturers offer attached units that are capable of control in both directions of travel. There also are self-contained air powered cylinders with built in oil cylinders and reservoirs. Air produces thrust while oil controls speed and/or mid-stroke stop-and-hold. Some units also have two-speed capabilities. These units look like a standard cylinder with an oversize rod.

Air-oil tank systems

Another common air-oil system uses low-pressure hydraulic cylinders coupled with air-oil tanks, Figure 17-2. These tanks hold more than enough oil to stroke the cylinder one way. An air valve piped to the air-oil tanks introduces compressed air to force oil from the tanks into the cylinder. Add flow controls and shut-off valves to the oil lines to give smooth, accurate cylinder control.

Fig. 17-2. Typical air-oil tank arrangement

When control is only necessary in one direction, the tank on the uncontrolled side can be omitted. This type of circuit requires very good cylinder seals to prevent air or oil transfer.

Air-over-oil tanks do not intensify the oil pressure, regardless of the tank’s diameter or length. The highest possible oil pressure available simply equals the air pressure supplied.

Several cylinder suppliers offer air-oil tanks that consist of a cylinder tube with two cylinder end caps held on the tube with tie-rods. A sight glass can be a length of plastic tubing with air-line fittings attached opposite the air ports. A baffle at the air port keeps oil from being aerated when air blasts in from the valve. A baffle at the oil port keeps any vortex formed from sending air to the cylinder. This baffle also keeps returning fluid from blowing into the air port.

Air-oil tandem cylinders

Tandem cylinders are another approach to using oil for control and air for power. In Figure 17-3, the single-rod cylinder of the tandem runs on air, while the double-rod cylinder is filled with oil. Because volume is equal in both ends of the double-rod cylinder, oil flows from end to end through a flow control and/or shut-off or skip valves for accurate control of speed and stopping.

Fig. 17-3. Typical air-oil tandem-cylinder circuit

Two flow controls in opposite directions provide variable speed in both directions. A bypass flow control around the stop valve would allow for two-speed operation in one direction. (The second speed must be the slower of the two.)

The skip valve option allows a fast approach with deceleration before work contact. The deceleration signal would come from a limit switch or limit valve.

The schematic drawing in Figure 17-4 shows tandem cylinders in a synchronizing circuit. This is a practical way to make two or more air-powered cylinders move in unison. (Using flow controls to do this produces inaccurate results.) When the air valve shifts to extend the cylinders they must move at the same time. This is because the trapped hydraulic oil in the hydraulic cylinders must transfer from the top side of one cylinder to the bottom side of the other one. If one cylinder stops they both must stop at the same time.

Fig. 17-4. Circuit to synchronize air-oil tandem cylinders

Note that the maximum load capability is equal to the capacity of both cylinders’ thrust. With the load placed as shown, the left cylinder transfers energy to the right cylinder through the oil. This gives the right cylinder up to twice as much thrust.

A small make-up tank and check valves replenish any leakage in the plumbing or at the rod seal. If the unit is subject to heating, a small relief valve may be required to keep thermal expansion from over-pressuring the oil-filled chambers. A shut-off valve connecting the transfer lines can re-synchronize the cylinders if the piston seals allow fluid to bypass and the platen gets out of level. Re-synchronization can be handled automatically with a normally closed, 2-way spool valve and limit switches.

(For other air-oil circuits, see the author’s upcoming e-book, "Fluid Power Circuits Explained.")

Some precautions with air-oil circuits

Most air-oil circuits operate at 100 psi or less, so any pressure drop in the circuit can cut force drastically. If oil lines are undersized, cylinder movement will be very slow. Size most air-oil circuit oil lines for a velocity of about 2 to 4 fps. This low speed requires large lines and valves, but is necessary if average travel speed with maximum force is important.

Another common problem with air-oil circuits is that any air trapped in the oil makes the cylinder performance spongy. The air’s compressibility makes accurate mid-stroke stopping and smooth speed control hard to attain. Some arrangement should be provided to bleed any trapped air from the oil chambers. When using an air-oil tank system, it is best to mount the tanks higher than the cylinder they feed. All lines between the cylinder and the tanks should slope up to the tanks. Also, if possible, let the cylinders make full strokes to purge any air. With dual oil-tank systems, incorporate a means for equalizing tank levels into the design.

The cylinder seals must be as leak free and low friction as possible. Any leakage past the seals can cause tank overflow, oil misting, and loss of control.

Intensifiers (or boosters)

In some of the foregoing air-oil circuits, the usual 80- to 100-psi pressure may not be adequate for some operations. This does not mean a hydraulic pump and all the items related to it must be used. Several manufacturers make air-oil intensifiers that convert 80- to 100-psi shop air into 500- to 40,000-psi hydraulic pressure -- in small volumes of fluid.

Single-stroke intensifiers

The simplest intensifier is a single rod-end cylinder with a large piston rod. As explained in Chapter 15, a cylinder with a 2:1 area ratio rod can have pressure as high as twice system pressure in the rod end. This type intensifier is only available in ratios up to 2:1 unless special oversize rods are specified.

Another simple intensifier can be made by coupling the rod of a large-bore cylinder to that of a smaller-bore cylinder with the same stroke, Figure 17-5. Supplying the large bore cylinder with pressurized air or hydraulic fluid forces the hydraulic fluid out of the smaller bore. The upper cross-sectional view is typical of two cylinders assembled in the user’s plant from stock air and/or hydraulic cylinders. The lower cross-sectional view is a purchased assembly that takes less space and eliminates possible mounting and alignment problems. The purchased unit is limited to piston ratios that can have the same size rod in both cylinders.

Fig. 17-5. Two types of differential-cylinder intensifiers

Usually these intensifiers are hydraulic to hydraulic with ratios that are less than 5:1 ratio. Later we’ll see a similar design for air-to-air intensifiers with similar ratios. Never operate these types of intensifiers above the cylinders’ rated pressure. For all intensifier designs, output pressure is directly related to the area ratio between the driving piston and the driven piston (or ram).

The cross-sectional view in Figure 17-6 shows typical construction of two types of 25:1 air-oil intensifiers. They consist of 5-in. bore air cylinders with 1-in. rods displacing oil from high-pressure oil chambers. The upper cross-sectional view is a dual-head intensifier that requires some sort of blocking valve to isolate its inlet oil from its outlet oil. This is usually done with a pilot-operated check valve so flow can return when the actuator reverses.

Fig. 17-6. Ram-type single-stroke intensifiers

The lower cross-sectional view is a triple-head intensifier that has an integral high-pressure seal to isolate inlet oil from high-pressure oil after the rod moves approximately 2 in. There is no need for external isolation because oil can flow freely either way anytime the ram is retracted.

A single-stroke intensifier must be sized to supply enough oil to make the working cylinder perform its work before the air piston bottoms out. It is good practice to size the intensifier for 10 to 15% more fluid than required. Avoid long fluid conductors if possible because the oil’s compressibility and stretching hose can use up the small-volume safety output quickly.

The circuit in Figure 17-7 shows a typical high-pressure air-oil circuit using the components described so far. This could be a press operation that requires a 10-in. total stroke. The stroke concludes with a 0.25-in. high-pressure stroke that generates 25 tons of force. Based on a maximum pressure of 2000 psi, a cylinder with a 6-in. bore is needed to produce the required force. The piston area of a 6-in.-bore cylinder is 28.274 in.2, so the 0.25-in. work stroke produces a volume of 7.07 in.3 of high-pressure oil. Using a standard 5-in. intensifier with a 1-in. ram, this requires

Fig. 17-7. Typical high-pressure air-oil circuit

(7.07) (110%) / (0.7854) = 9.9-in. stroke plus 2 in. more for passing the high-pressure seal for a total stroke of 12 in. The volume of the 6-in. bore X 10-in. stroke high-pressure hydraulic cylinder is (28.275 in.2) X (10 in.) or 283 in.3, so the air-oil tanks should have 6-in. bores and be 12-in. long.

The cycle starts when the solenoid on the 4-way directional control valve is energized to send air to the left-hand air-oil tank. Simultaneously, the valve exhausts air from the right-hand air-oil tank. Oil at air pressure is pushed through the triple-head intensifier to the high-pressure hydraulic cylinder. The cylinder advances rapidly at low force until it contacts the work.

At work contact, pressure builds in the left-hand air-oil tank and in the pilot line to the 4-way sequence valve. With supply-air pressure at 80 psi and the sequence valve set for 65 to 75 psi, the valve shifts and cycles the intensifier. As the intensifier extends, after it travels approximately 2 in., it passes through the high-pressure seal to block low-pressure oil and force high-pressure oil into the cylinder. Pressure in the work cylinder can now go as high as 2000 psi to produce the required 25 tons of force.

When the solenoid on the 4-way directional control valve de-energizes, air exhausts from the left-hand air-oil tank and from the 4-way sequence-valve pilot. The sequence valve shifts to its original position and the triple-head intensifier retracts. Air also is directed to the right-hand air-oil tank to pressurize it for the retract stroke of the high-pressure hydraulic cylinder. After the intensifier retracts past the high-pressure seal, the work cylinder can retract quickly to end the cycle. Note: only 80 psi acting on the area of the work cylinder develops retraction force. While as much as 25 tons of force was generated during the short extension stroke, only 1869 lb are generated during retraction.

The intensifier could be cycled by other means -- such as a limit switch or a pressure switch and solenoid valve combination. It could even be operated manually.

Also note: any of the above units could be cycled with hydraulic oil as the driving force. Usually such hydraulic-to-hydraulic intensifiers are only between 2:1 and 5:1 because the input pressure can be much higher than typical compressed air.

Reciprocating intensifiers

For higher volumes of intensified fluid, several manufacturers make reciprocating units. The cross-sectional view and circuit in Figure 17-8 show a typical single-ram intensifier that uses compressed air for power and pumps oil in the high-pressure side. These units often are supplied in a ready-to-run condition as pictured. They may cycle as soon as air is supplied or they may require an external signal to start. Most reciprocating units supply less that 3-gpm maximum at low pressure and slow to a stop at maximum pressure.

Fig. 17-8. Reciprocating air-to-hydraulic intensifier

To produce higher pressures, some units incorporate more than one air cylinder in series to raise the intensification ratio. These units also come with pressure chambers and rams on both ends to provide a greater volume of high-pressure oil.

Some manufacturers build reciprocating hydraulic-to-hydraulic intensifiers with ratios as high as 20:1 to generate pressures up to 12,000 psi. These units supply small volumes of high-pressure oil from low- to high-pressure input fluids.

Special air-oil units

Several companies manufacture special self-contained air/hydraulic cylinders with integral tanks and intensifiers that produce low-pressure advance, high-pressure work, and low-pressure retract strokes. Externally, they appear to be over-length air cylinders, but they can have output forces as high as 150 tons.

Air-to-air intensifiers

When an application requires a small volume of high-pressure air, try an air-to-air intensifier instead of a high-pressure compressor. The cross-sectional view and circuit in Figure 17-9 shows the makeup of a 2:1 intensifier that can almost double output pressure. Inlet air is delivered to the driving cylinder by a double pilot-operated valve and to the intensifying cylinder through check valves. As the two pistons stroke to the right, the full area of the left piston and the annulus area of the right piston are pushing the right piston’s full area at almost double force. Thus, air exiting the right piston is at about twice input pressure. The discharged air flows through a check valve and on to the high-pressure circuit.

Fig. 17-9. Air-to-air intensifier with 2:1 ratio

When the pistons complete their strokes, the one on the right contacts a small integral limit valve that sends a signal to the double pilot-operated valve and shifts it to reverse the pistons’ strokes. The same areas and forces push in this direction but they work against a smaller intensifying area. The intensifier will continue cycling until pressure at the pressure-air outlet port reaches full pressure. At that point, the pistons stall and hold pressure until the downstream pressure drops.

These intensifiers will stroke considerably more slowly at about 80% of their maximum pressure so it is best if the output air pressure is at least 20% above what is required. A regulator at the working machine can control the actual working pressure so less air is wasted.

Intensification ratios and output volumes are functions of piston ratios, bore sizes, and stroke lengths. Outputs up to 250 psi are standard with most manufacturers. Some offer higher pressures. Very high-pressure units use hydraulic cylinders to drive gas cylinders to reach pressures as high as 45,000 psi.

(For more air-oil and intensifier circuit designs, see the author’s upcoming e-book, "Fluid Power Circuits Explained.")

قوانين وجداول

2009-07-08T10:19:00.000-07:00

Ch 22: Fluid Power Formula

Click here to download the PDF Version of this chapter.

رسومات هيدروليكيه

2009-07-08T10:16:00.000-07:00

Ch 21: Sample Circuits

For each of the circuit diagrams, identify the numbered components and describe the circuit's operation.
A link to the answers follows at the bottom of the page.

المكدسات الهيدروليكيه4

2009-07-08T10:13:00.000-07:00

Hydro-pneumatic accumulators

Hydraulic accumulators

Accumulators make it possible to store useable volumes of almost non-compressible hydraulic fluid under pressure. The symbols and simplified cutaway views in Figure 16-1 show several types of accumulators used in industrial applications. They are not complete representations but they illustrate general working principles.

Fig. 16-1. Cross-sectional views and symbols for hydraulic accumulators

A 5-gal container completely full of hydraulic oil at 2000 psi will only discharge a few cubic inches of fluid before the pressure drops to 0 psi. If the same container were filled half with oil and half with nitrogen gas, it could discharge more than 1 1/2 gallons of fluid while pressure only dropped 1000 psi. This is the great advantage of hydro-pneumatic accumulators.

Accumulator types

No separator: Some original accumulators were high-pressure containers with a sight glass to show fluid level. They were filled approximately half with oil and half with nitrogen gas -- with no separation barrier between them. Before stopping the pump, a shut off valve at the accumulator discharge port was closed to prevent fluid and gas from escaping. This type of accumulator is not used on new circuits today, but there still are many in service.

Gas-charged bladder: Many accumulators now use a rubber bladder to separate the gas and liquid. A poppet valve in the discharge port keeps the bladder from extruding when the pump is off. The original design was the bottom-repair style, shown on the left in Figure 16-1. It is still offered by most manufacturers. The top-repair style on the right is now available and makes bladder replacement simple and fast.

Gas-charged piston: The gas-charged piston accumulator has a free-floating piston with seals to separate the liquid and gas. It operates and performs similarly to the bladder type, but has some advantages in certain applications. A gas-charged piston accumulator can cost twice as much as an equal-sized bladder type.

Spring-loaded piston: A spring-loaded piston accumulator is identical to a gas-charged unit, except that a spring forces the piston against the liquid. Its main advantage is that there is no gas to leak. A main disadvantage is that this design is not good for high pressure and large volume.

Weight loaded: All gas-charged accumulators lose pressure as fluid discharges. This is because the nitrogen gas was compressed by incoming fluid from the pump and the gas must expand to push fluid out. The weight-loaded accumulator in Figure 16-1 does not lose pressure until the ram bottoms out. Thus 100% of the fluid is useful at full system pressure. The major drawback to weight-loaded accumulators is their physical size. They take up a lot of space and are very heavy if much volume is required. They work well in central hydraulic systems because there usually is room for them in the power unit area. However, central hydraulic systems are falling out of favor, so only a few facilities use weight-loaded accumulators. (Rolling mills are one application where space to place large items is not a problem.) Note that there is often a long dwell time to fill these monsters.

Diaphragm accumulators: There are also diaphragm accumulators with resilient or metal diaphragms. They are used where the stored volume is small.

Why are accumulators used?

To supplement pump flow: The most common use for accumulators is to supplement pump flow. Some circuits require high-volume flow for a short time and then use little or no fluid for an extended period. Generally speaking, when half or more of the machine cycle is not using pump flow, the application is a likely candidate for an accumulator circuit.

The circuit in Figure 16-2 uses several accumulators to supplement pump flow because the dwell time is 45 seconds out of the 57.5-second cycle time. This circuit’s 22-gpm fixed-volume pump operates on pressure during most of the cycle to fill the cylinder and the accumulators. Without the accumulators, this circuit would require a 100-gpm pump driven by a 125-hp motor. The first cost of the smaller pump and motor plus the accumulators is very close to that of the larger pump and motor. However, energy savings over the life of the machine make the pictured circuit much more economical.

Fig. 16-2. Accumulator circuit that supplements pump flow

One drawback of using accumulators to supplement pump flow is that the circuit must operate at a pressure higher than needed to perform the work. In the circuit in Figure 16-2, a minimum of 2000 psi is necessary to perform the work. This means the accumulators must be filled to a higher pressure so they can supply extra fluid without dropping below the minimum pressure. This circuit uses 3000-psi maximum pressure to store enough fluid to cycle the cylinder in the allotted time and still have ample force to do the work. The flow control in the circuit is necessary to keep the cylinder from cycling too rapidly. An accumulator discharges fluid at any velocity the lines can handle at whatever the pressure drop is when a flow path is opened.

The circuit in Figure 16-2 uses a fixed-volume pump and an accumulator unloading-and-dump valve. The valve forces pump flow to the accumulators when pressure drops approximately 15% below its maximum set pressure. At set pressure, the unloading valve opens and all pump flow bypasses to tank at 25- to 50-psi pressure drop. While the pump is bypassing, a check valve keeps the accumulators from unloading to tank. The dump valve (which is a high-ratio, pilot-to-close check valve) is held closed by pump idle pressure until the pump shuts down.

To maintain pressure: Another common application for accumulators is to maintain pressure in a circuit while the pump is unloaded. This is especially useful when using fixed-volume pumps on long holding cycles. The laminating-press circuit in Figure 16-3 clamps material and holds it at force for one to five minutes. If the pump were flowing across the relief valve at high pressure for this length of time, a lot of heat would be generated, wasting energy. With a pressure-compensated pump, energy loss would be less, but the system might still overheat in a short time.

Fig. 16-3. Using an accumulator to maintain pressure and/or make up for leakage

Adding an accumulator, flow control, and pressure switch to the fixed-volume pump circuit allows the pump to unload when pressure is at or above the pressure switch’s minimum setting. If leakage at the valve or cylinder seals allows pressure to drop about 5%, the pressure switch shifts the directional control valve to pressurize the cylinder cap end and build pressure back to maximum. The only time the pump is loaded is when fluid is required. This circuit will laminate parts continuously and does not need a heat exchanger. The flow control should be set at a reduced rate so the accumulator does not dump too rapidly when the directional control valve shifts to retract the platen. Flow to make up for leakage is minor and does not need a high rate.

The accumulator dump valve in Figure 16-3 is a high-ratio pilot-to-close check valve that is held closed by the low pressure when the pump is unloaded. It opens to discharge any stored energy when the pump shuts down.

To absorb shock: Fast-moving hydraulic circuits can produce pressure spikes that cause shock when flow is stopped abruptly. Accumulators can be installed in such shock-prone circuits to reduce damaging pressure and flow spikes to an acceptable rate -- or eliminate them completely. (Accumulators can handle other pressure-spike concerns with some valve additions for special instances.)

Figure 16-4 depicts an accumulator installed to eliminate the pressure spike caused by sudden flow blockage. The nitrogen charge in this installation should be 5 to 10% above the working pressure. This keeps the accumulator out of the circuit except during pressure spike situations. A bladder-type accumulator works best here because of its fast response to pressure changes. (Use caution when applying accumulators to shock situations. It is possible to actually increase shock instead of reducing or eliminating it.)

Fig. 16-4. Using an accumulator to eliminate shock caused by a sudden flow stoppage

As an emergency power supply: Some hydraulically operated machines may always need to stop in the open position to keep from damaging product or equipment. When a power failure shuts the hydraulic pump off and the machine happens to be some position other than open, there needs to be some way to get it open. An engine-driven standby pump could fill the bill and in some instances might be the best remedy. Another option is to use accumulators that are charged before the first cycle and held that way until the machine shuts down. The stored energy is ready to cycle the machine to the open position in case of a power failure.

The circuit in Figure 16-5 operates a slide gate on a waste material bin that opens hydraulically to fill a transfer truck. The circuit is located in a remote area that is prone to power failure, so it was designed to automatically close the gate in case power went off.

Fig. 16-5. Using an accumulator as an emergency power supply

The schematic diagram shows the cylinder at rest with the pump running. When the unit starts, solenoids C and C2 on the normally open 2-way directional valves are energized. They stay energized while the pump is on. The first pump flow goes through the check valve and fills the accumulator with enough fluid to extend the cylinder from any open position. When electrical power is available, the gate can be opened and closed to dump waste material into the waiting truck. If a truck is filling and a power failure occurs, the pump stops and all solenoids de-energize. At this point the accumulator is ported to the cylinder cap end and fluid in the cylinder rod end has a free path to tank.

Notice the manual drain connected to the line between the check valve and the accumulator. This drain must be opened before working on the circuit. A placard on the machine warns maintenance personnel of the potential danger if the accumulator is not drained. Emergency power supplies are the only accumulator circuit that cannot be drained automatically in most cases.

Accumulator precautions

Always arrange some method to drain the accumulator at shut down. (At the end of this section, several ways to drain an accumulator automatically are shown. Plus, there is always the old standby, a manual drain.) Never work on a circuit with an accumulator until you are sure it is depressurized.

Make sure accumulator flow is restricted to a reasonable rate during operation and shut down to avoid damage to the machine or piping. Accumulators will discharge fluid at any rate the exit flow path will allow. Such high flow does not last long, but the damage it causes is done quickly.

Always isolate the pump from the accumulator with a check valve so fluid cannot back flow into the pump. Without a check valve, accumulator back flow can drive the pump backward -- and overspeed it to destruction in some instances.

Check the accumulator’s pre-charge pressure at installation and at least once a day for the first week of operation. If there is no noticeable loss of pressure during this time, do the next check a week later. If all is well then, do a routine check every three to six months thereafter. Whenever the accumulator pre-charge drops below nominal pressure, the volume of available fluid is reduced and finally the cycle slows.

One way to check accumulator pre-charge is to turn off the pump, allow the accumulator to empty all oil back to tank, and then connect the items in a charge kit, Figure 16-6. First remove the gas-valve cap and install the charge kit gauge, hose, and tee-handle assembly on the gas valve. Next, turn the tee handle in to open the valve and read gauge pressure. However, every time this operation is performed there is the chance the valve will not reseat and gas will start to leak.

Fig. 16-6. Charging an accumulator or checking its pre-charge pressure with a charge kit

To avoid potential gas leakage, Figure 16-7 illustrates two noninvasive methods to check pre-charge. Both are fast, simple, and can be done almost anytime without a lengthy interruption of production. Either of these ways gives a fast reasonably close check without invading any plumbing. They are not 100% accurate, but will be within ±5% of the gauge reading -- with almost anyone doing them. The method on the left is the least accurate -- especially when using a glycerin-filled gauge.

The Pump Just Starting method on the left shows a jump in pressure after the pump starts then a steady climb to set pressure. This first jump is the pre-charge pressure and the steady climb is during compression of the gas in the bladder or behind the piston. The length of time between the first pressure jump and reaching system pressure depends on the volume of the accumulator and the pump output.

Fig. 16-7. Two non-invasive procedures for checking accumulator pre-charge pressure

The Pump Shutoff From Full Pressure method is easiest and most accurate, especially if the accumulator dump valve is manually operated. Fluid can be bled off slowly with a manual dump so the gauge reaches pre-charge pressure slowly.

With this method the system must be at pressure and the accumulator charged at least above pre-charge pressure. At system shut down either an automatic or manual drain is opened and pressure starts to fall. Because the gauge is reading oil pressure and the only reason there is pressure is because of trapped gas above it, pressure will fall to a point then suddenly drop to zero. Read the pressure as the gauge suddenly drops to zero to determine gas pre-charge.

This method is the most accurate but is not precise like a gauge reading, so use it for a cursory check as often as necessary to see if the gas charge is holding.

Accumulator pre-charge pressure

Normally, gas-charged accumulators are pre-charged to approximately 85% of the system’s minimum working pressure. This assures that the bladder or piston does not discharge all the fluid during every cycle. If all fluid is evacuated at high rates, bladders can get caught in the poppet valves and pistons can be deformed when metal hits metal.

In certain applications, this 85% figure may be low because minimum system pressure is low. In such a case, use a piston-type accumulator because the piston can move up the bore almost any distance without damage. A bladder accumulator should not be used when pre-charge pressure is less than half the maximum pressure. This avoids compressing the bladder so tightly that rubbing action on itself wears holes in it.

Applying accumulators

Many applications can use any type accumulator with equally satisfactory results. However, there are some cases where one particular style is more responsive or offers a longer service life. As mentioned in the previous section, the amount of pre-charge pressure is one reason for selecting a bladder or piston accumulator.

Weight-loaded accumulators respond to pressure buildup slowly so they do not work well as shock absorbers. Weight-loaded accumulators will reduce but not stop pressure spikes. Piston accumulators are not as fast as bladder types at responding to fast increases to pressure. So in these situations, the best choice is a bladder-type accumulator.

Some accumulator circuits are installed to dampen high-pressure spikes at the outlet of piston pumps. A piston accumulator in this application cannot respond quickly enough to do the job. Also, the short stroking distance of the piston and seals can cause excessive wear to the bore and seals. A bladder accumulator works best in this type circuit.

Sizing accumulators

Most accumulator suppliers offer information in their literature about sizing accumulators for any of the above circuits. Many offer computer programs that only require the input of system requirements. The program then figures accumulator size and outputs a part number. One company offers a formula and software for use on the Internet.

Accumulator dump valves

In all the foregoing accumulator applications (except the one for emergency power supply), the accumulator fluid was drained automatically at shut down. This is very important because accumulators store energy that can be a safety hazard and can cause damage to the machine. Here are examples of different types of accumulator dump valves and circuits.

Figure 16-8 shows one frequently used circuit. A normally open, solenoid-operated, 2-way directional control valve is teed into the pump line between the isolation check valve and the accumulator. The solenoid is wired so that it is energized when the pump starts and de-energized when the pump stops. An orifice in front of the 2-way valve controls flow when the accumulator is discharging to prevent damage to the valve. This arrangement works equally well with fixed-displacement or pressure-compensated pumps.

Fig. 16-8. Circuit that uses a solenoid-operated valve to dump an accumulator

A note of caution: Some solenoid valves, even though they are designed for continuous duty, get very hot when energized for long periods. Such overheating can cause varnish deposits to form and lock the valve’s internal parts in the closed condition after the pump shuts down. This means the trapped energy does not get discharged and the accumulator can cause harm to anyone working on the circuit.

The dump circuit in Figure 16-9 is only for pressure-compensated pumps. A packaged set of valves isolates the accumulator while the pump is running and automatically dump it at shut down. The package consists of an isolation check valve, a pilot-to-close check valve, and a flow-control orifice.

Fig. 16-9. Hydraulically operated circuit that isolates and dumps an accumulator supplied by a pressure-compensated pump

At pump startup, flow goes to the circuit and the accumulator. Pressure from the pump outlet shifts the pilot-to-close check valve, blocking flow to tank. When the accumulator is full, the pump compensates to no flow and the circuit waits for a new cycle. When pressure drops, the pump comes back on stroke and makes up for flow going to the circuit. At pump shut down, pilot pressure to the pilot-to-close check valve drops and the valve shifts to open. Now, stored energy in the accumulator is ported to tank through the orifice. This circuit is very reliable because it depends on system or pump pressure to close and/or open valves.

A fixed-volume pump must be ported to tank at very low pressure when its flow is not doing work. A common circuit for unloading a fixed-volume pump and dumping an accumulator is shown in Figure 16-10. An internally piloted unloading relief valve with integral check valve forces all pump flow to the circuit and the accumulator until the system reaches the set pressure. As the control ball starts to relieve, system pressure pushes against the unloading piston and forces it off its seat. This takes all pressure off the top of the relief valve poppet. The pump unloads to tank at 25 to 100 psi until system pressure drops approximately 15%. After that drop, spring force pushes the unload piston back and pump flow goes to the circuit again.

Fig. 16-10. Hydraulically operated circuit that isolates, unloads, and dumps an accumulator supplied by a fixed-displacement pump

The accumulator dump valve blocks fluid from going to tank while the pump is running and opens to discharge stored energy when the pump shuts down. The accumulator dump valve is a high ratio (up to 200:1) pilot-to-close check valve that is held shut by the pump's unloaded or work pressure. With a 200:1 area ratio between the poppet and the pilot piston, 25-psi pressure at the pilot port will stop as much as 5000 psi at the poppet shut off. This keeps fluid in the accumulator circuit until the pump is shut down. Then, all stored pressurized fluid flows to tank quickly and safely. (One supplier offers the unloading relief valve and the accumulator dump valve in a single body. This combination simplifies piping while offering the same effect.)

Other accumulator applications

Accumulators are also used for systems where thermal expansion could cause excessive pressure. Cylinders with blocked ports in a high ambient heat area can go to high pressure if there is no place for expanding fluid to go.

Another use for accumulators is as a barrier between two different fluids. The pump that uses hydraulic fluid keeps pressure on a circuit that uses water or another incompatible medium.

One supplier offers low-pressure accumulators as breathing devices for sealed reservoirs. This keeps airborne contaminants out of the hydraulic oil as the fluid level rises and falls.

For more circuits and other information on accumulators, see the author’s upcoming e-book Fluid Power Circuits Explained.

الاسس الهيدروليكيه

2009-07-08T09:51:00.000-07:00

Any media (liquid or gas) that flows naturally or can be forced to flow could be used to transmit energy in a fluid power system. The earliest fluid used was water hence the name hydraulics was applied to systems using liquids. In modern terminology, hydraulics implies a circuit using mineral oil. Figure 1-1 shows a basic power unit for a hydraulic system. (Note that water is making something of a comeback in the late '90s; and some fluid power systems today even operate on seawater.) The other common fluid in fluid power circuits is compressed air. As indicated in Figure 1-2, atmospheric air -- compressed 7 to 10 times -- is readily available and flows easily through pipes, tubes, or hoses to transmit energy to do work. Other gasses, such as nitrogen or argon, could be used but they are expensive to produce and process.

Fig. 1-1: Basic hydraulic power unit.

Of the three main methods of transmitting energy mechanical, electrical, and fluid fluid power is least understood by industry in general. In most plants there are few persons with direct responsibility for fluid power circuit design or maintenance. Often, general mechanics maintain fluid power circuits that originally were designed by a fluid-power-distributor salesperson. In most facilities, the responsibility for fluid power systems is part of the mechanical engineers' job description. The problem is that mechanical engineers normally receive little if any fluid power training at college, so they are ill equipped to carry out this duty. With a modest amount of fluid power training and more than enough work to handle, the engineer often depends on a fluid power distributor's expertise. To get an order, the distributor salesperson is happy to design the circuit and often assists in installation and startup. This arrangement works reasonably well, but as other technologies advance, fluid power is being turned down on many machine functions. There is always a tendency to use the equipment most understood by those involved.

Fig. 1-2: Basic pneumatic power arrangement.

Fluid power cylinders and motors are compact and have high energy potential. They fit in small spaces and do not clutter the machine. These devices can be stalled for extended time periods, are instantly reversible, have infinitely variable speed, and often replace mechanical linkages at a much lower cost. With good circuit design, the power source, valves, and actuators will run with little maintenance for extended times. The main disadvantages are lack of understanding of the equipment and poor circuit design, which can result in overheating and leaks. Overheating occurs when the machine uses less energy than the power unit provides. (Overheating usually is easy to design out of a circuit.) Controlling leaks is a matter of using straight-thread O-ring fittings to make tubing connections or hose and SAE flange fittings with larger pipe sizes. Designing the circuit for minimal shock and cool operation also reduces leaks.

A general rule to use in choosing between hydraulics or pneumatics for cylinders is: if the specified force requires an air cylinder bore of 4 or 5 in. or larger, choose hydraulics. Most pneumatic circuits are under 3 hp because the efficiency of air compression is low. A system that requires 10 hp for hydraulics would use approximately 30 to 50 air-compressor horsepower. Air circuits are less expensive to build because a separate prime mover is not required, but operating costs are much higher and can quickly offset low component expenses. Situations where a 20-in. bore air cylinder could be economical would be if it cycled only a few times a day or was used to hold tension and never cycled. Both air and hydraulic circuits are capable of operating in hazardous areas when used with air logic controls or explosion-proof electric controls. With certain precautions, cylinders and motors of both types can operate in high-humidity atmospheres . . . or even under water.

When using fluid power around food or medical supplies, it is best to pipe the air exhausts outside the clean area and to use a vegetable-based fluid for hydraulic circuits.

Some applications need the rigidity of liquids so it might seem necessary to use hydraulics in these cases even with low power needs. For these systems, use a combination of air for the power source and oil as the working fluid to cut cost and still have lunge-free control with options for accurate stopping and holding as well. Air-oil tank systems, tandem cylinder systems, cylinders with integral controls, and intensifiers are a few of the available components.

The reason fluids can transmit energy when contained is best stated by a man from the 17th century named Blaise Pascal. Pascal's Law is one of the basic laws of fluid power. This law says: Pressure in a confined body of fluid acts equally in all directions and at right angles to the containing surfaces. Another way of saying this is: If I poke a hole in a pressurized container or line, I will get PSO. PSO stands for pressure squirting out and puncturing a pressurized liquid line will get you wet. Figure 1-3 shows how this law works in a cylinder application. Oil from a pump flows into a cylinder that is lifting a load. The resistance of the load causes pressure to build inside the cylinder until the load starts moving. While the load is in motion, pressure in the entire circuit stays nearly constant. The pressurized oil is trying to get out of the pump, pipe, and cylinder, but these mechanisms are strong enough to contain the fluid. When pressure against the piston area becomes high enough to overcome the load resistance, the oil forces the load to move upward. Understanding Pascal's Law makes it easy to see how all hydraulic and pneumatic circuits function.

Fig. 1-3: How Pascals Law affects a cylinder

Notice two important things in this example. First, the pump did not make pressure; it only produced flow. Pumps never make pressure. They only give flow. Resistance to pump flow causes pressure. This is one of the basic principles of fluid power that is of prime importance to troubleshooting hydraulic circuits. Suppose a machine with the pump running shows almost 0 psi on its pressure gauge. Does this mean the pump is bad? Without a flow meter at the pump outlet, mechanics might change the pump, because many of them think pumps make pressure. The problem with this circuit could simply be an open valve that allows all pump flow to go directly to tank. Because the pump outlet flow sees no resistance, a pressure gauge shows little or no pressure. With a flow meter installed, it would be obvious that the pump was all right and other causes such as an open path to tank must be found and corrected.

Fig. 1-4: Comparison of mechanical and hydraulic leverage

Another area that shows the effect of Pascal's law is a comparison of hydraulic and mechanical leverage. Figure 1-4 shows how both of these systems work. In either case, a large force is offset by a much smaller force due to the difference in lever-arm length or piston area.

Notice that hydraulic leverage is not restricted to a certain distance, height, or physical location like mechanical leverage is. This is a decided advantage for many mechanisms because most designs using fluid power take less space and are not restricted by position considerations. A cylinder, rotary actuator, or fluid motor with almost limitless force or torque can directly push or rotate the machine member. These actions only require flow lines to and from the actuator and feedback devices to indicate position. The main advantage of linkage actuation is precision positioning and the ability to control without feedback.

At first look, it may appear that mechanical or hydraulic leverage is capable of saving energy. For example: 40,000 lb is held in place by 10,000 lb in Figure 1-4. However, notice that the ratio of the lever arms and the piston areas is 4:1. This means by adding extra force say to the 10,000-lb side, it lowers and the 40,000-lb side rises. When the 10,000-lb weight moves down a distance of 10 in., the 40,000-lb weight only moves up 2.5 in.

Work is the measure of a force traversing through a distance. (Work = Force X Distance.). Work usually is expressed in foot-pounds and, as the formula states, it is the product of force in pounds times distance in feet. When a cylinder lifts a 20,000-lb load a distance of 10 ft, the cylinder performs 200,000 ft-lb of work. This action could happen in three seconds, three minutes, or three hours without changing the amount of work.

When work is done in a certain time, it is called power. {Power = (Force X Distance) / Time.} A common measure of power is horsepower - a term taken from early days when most persons could relate to a horse's strength. This allowed the average person to evaluate to new means of power, such as the steam engine. Power is the rate of doing work. One horsepower is defined as the weight in pounds (force) a horse could lift one foot (distance) in one second (time). For the average horse this turned out to be 550 lbs. one foot in one second. Changing the time to 60 seconds (one minute), it is normally stated as 33,000 ft-lb per minute.

No consideration for compressibility is necessary in most hydraulic circuits because oil can only be compressed a very small amount. Normally, liquids are considered to be incompressible, but almost all hydraulic systems have some air trapped in them. The air bubbles are so small even persons with good eyesight cannot see them, but these bubbles allow for compressibility of approximately 0.5% per 1000 psi. Applications where this small amount of compressibility does have an adverse effect include: single-stroke air-oil intensifiers; systems that operate at very high cycle rates; servo systems that maintain close-tolerance positioning or pressures; and circuits that contain large volumes of fluid. In this book, when presenting circuits where compressibility is a factor, it will be pointed out along with ways to reduce or allow for it.

Another situation that makes it appear there is more compressibility than stated previously is if pipes, hoses, and cylinder tubes expand when pressurized. This requires more fluid volume to build pressure and perform the desired work. In addition, when cylinders push against a load, the machine members resisting this force may stretch, again making it necessary for more fluid to enter the cylinder before the cycle can finish.

As anyone knows, gasses are very compressible. Some applications use this feature. In most fluid power circuits, compressibility is not advantageous; in many, it is a disadvantage. This means it is best to eliminate any trapped air in a hydraulic circuit to allow faster cycle times and to make the system more rigid.

Boyle's Law

Boyle's Law for gasses states: It is the principle that, for relatively low pressures, the absolute pressure of an ideal gas kept at constant temperature varies inversely with the volume of the gas. In down-home language this means if a ten cubic foot volume of atmospheric air is squeezed into a one cubic foot container, pressure increases ten times. (10 X 14.7 psia = 147 psia.) Notice that pressure is stated as psia.

Fig. 1-5: Measurement of gauge and absolute pressure

Normally, pressure gauges read in psi (with no additional letter). Commonly called gauge pressure, psi disregards the earth's atmospheric pressure of 14.7 psia, because it has no effect either negative or positive on a fluid power circuit. The a on the end of psia stands for absolute, and would be shown on a gauge with a pointer that never goes to zero unless it is measuring vacuum. Another type of gauge that shows both negative and positive pressures would have a pointer with an inches-of-mercury (in. Hg) scale below zero and a psig scale above zero. Both of these gauges could read pressure or vacuum. (They are always found in a refrigeration repairperson's tool kit. Refrigeration units have both vacuum and pressure in different sections of the system at the same time.) Figure 1-5 pictures a typical psig gauge and one type of psia gauge.

In the example above, when ten cubic feet of air was squeezed into a one cubic-foot space, both pressures were given in psia. To see what gauge pressure (psig) would be, subtract one atmosphere from the 147-psia reading. (147 psia 14.7 psia = 132.3 psig.) To calculate the amount of compression of air in a system, always use absolute pressure, or psia, not psig. For example: the cylinder in Figure 1-6 contains eight cubic feet of air at 70 psig. To what will pressure increase when an external force pushes the piston back until the space behind the piston is two cubic foot? It is obvious the pressure will rise four times. At first it might look easy to take 70 psig X 4 = 280 psig, but this answer is wrong. For the correct answer, gauge pressure must be changed to absolute pressure. In this case by adding one atmosphere to the 70-psig reading. (70 psig + 14.7 psia = 84.7 psia.) Now multiply the 84.7-psia pressure by 4 to see what the absolute pressure is when the cylinder stops at one cubic foot volume. (84.7 X 4 = 338.8 psia.) Finally, to return to gauge pressure, subtract one atmosphere from the absolute pressure. (338.8 psia 14.7 psia = 324.1 psig.) Notice that the correct pressure is 44.1 psig higher than when gauge pressure is the multiplier.

Fig. 1-6: Pressure change as air is compressed

Temperature was not considered in both preceding cases, but notice that the law says kept at constant temperature. Compressing a gas always increases its temperature because the heat in the larger volume is now packed into a smaller space. The next law says that increasing temperature increases pressure if the gas cannot expand. This means the pressures given are measured after the gas temperature returns to what it was originally.

Gauges today read in psi and bar. Bar is a metric or SI unit for pressure and is equal to approximately the barometer reading or one atmosphere. One atmosphere is actually 14.696 psi but the SI unit for bar is 14.5 psi.

Charles' Law

Heating a gas or liquid causes it to expand. Continuing to heat a liquid will result in it changing to the gaseous state and perhaps spontaneous combustion. If the gas or liquid cannot expand because it is confined, pressure in the contained area increases. This is stated in Charles' Law as: The volume of a fixed mass of gas varies directly with absolute temperature, provided the pressure remains constant. Because fluid power systems have some areas in which fluid is trapped, it is possible that heating this confined fluid could result in part damage or an explosion. If a circuit must operate in a hot atmosphere, provide over pressure protection such as a relief valve or a heat- or pressure-sensitive rupture device. Never heat or weld on any fluid power components without proper preparation of the unit.

Static head pressure

The weight of a fluid in a container exerts pressure on the containing vessel's sides and bottom. This is called static head pressure. It is caused by earth's gravitational pull. A good example of head pressure is a community water system. Figure 1-7 shows a water tower with a topmost water level of 80 feet. A cubic inch of water weighs 0.0361 pounds. Therefore a one square-inch column of water will exert a force of 0.0361 psi for every inch of elevation. This works out to .433 psi per foot of elevation. For the water tower in Figure 1-7, the pressure at the base would be: 80 ft X 0.433 psi/ft = 34.6 psi. This pressure is always available, even when no pumps are running. Of course, if the water level drops, static head pressure also will drop.

Fig. 1-7: Pressure measurement for water tower

The specific gravity of hydraulic oil is approximately 0.9, so multiplying water's 0.433 psi per foot by 0.9 shows oil exerts 0.39 psi per foot of elevation. Usually this fraction is rounded to 0.4 for simplicity. If the water tower were filled to 80 ft with oil, it would exert a pressure of 32 psi at ground level. Other fluids would develop a higher or lower static pressure according to their specific gravities.

This pressure is only realized at ground level at the tower. Outlets at other levels would be higher or lower according to their distance below the fluid surface.

Tanks seen on most water towers simply store volume. Pressure does not drop rapidly or require frequent pump starts to maintain the fluid level. The size or shape of the tank does not affect pressure at the base. Pressure at the base of a straight 80-ft pipe would be the same, but useful volume before pressure drop would change drastically. Always remember: it is not the physical size of a body of fluid that determines pressure but how deep it is.

Head pressure can have an adverse effect on a hydraulic system because many pumps are installed above the fluid level. This means the pump must first create enough vacuum to raise the fluid and then create even higher vacuum to accelerate and move it. Therefore there is a limit to how far a pump can be located above the oil level. Most pumps specify a maximum suction pressure of 3 psi. At 4- to 5-psi suction pressure, pumps start to cavitate . . . causing internal damage. At 6- to 7-psi vacuum, cavitation damage is severe and noise levels increase noticeably. (The effects of cavitation are covered fully in Chapter 8, Fluid power pumps and accessory items.) Axial- or in-line-piston pumps are especially vulnerable to high inlet vacuum damage and should be set up below the fluid level to produce a positive head pressure.

Many modern hydraulic systems place the pump next to the reservoir so the fluid level is always above the pump inlet. With this type of installation the pump always has oil at startup and has a positive head pressure at its inlet. A better arrangement puts the tank above the pump to take advantage of even greater head pressure. Everything possible should be done to keep pressure drop low in the pump inlet line because the highest possible pressure drop allowable is one atmosphere (14.7 psi at sea level).

The earth's atmosphere the air we breathe exerts a force of 14.7 psi at sea level on an average day. This pressure covers the whole earth's surface, but at elevations higher than sea level, it is reduced by approximately 0.5 psi per 1000 feet. This pressure of earth's atmosphere is the source of the power of vacuum. The highest possible vacuum reading at any location is the weight of the air above it at that time. A reading of maximum vacuum available is given during the local weather forecast as the barometer reading. Divide the barometer reading by two to get the approximate atmospheric pressure in psi. This force could be directly measured if it were possible to isolate a one square-inch column of air one atmosphere tall at a sea level location. Because this is not possible, the method used to measure vacuum is demonstrated in Figure 1-8.

Fig. 1-8: Vacuum measurement with mercury

Submerge a clear tube with one closed end in a container of mercury and allow it to fill completely. (The tube must be more than 30-in. long for this example to work when mercury is the liquid.) After the mercury displaces all the air in the tube, carefully raise the tube's closed end, keeping the open end submerged so the mercury can't run out and be replaced by air. When the tube is positioned vertically, the liquid mercury level will lower to give the atmospheric pressure reading in inches of mercury (29.92-in. H_g at sea level). The mercury level will fluctuate from this point as high and low-pressure weather systems move past. If the tube had been 100-in. tall, the mercury level would still have dropped to whatever the atmospheric pressure was at its location. The reason the mercury does not all flow out is that atmospheric pressure holds it in.

This barometer could have been built using another liquid but the tube would have to be longer because most other liquids have a much lower specific gravity than mercury's 13.546. Water, with a specific gravity of 1.0, would require a closed-end tube at least 33.8 ft long, while oil, with a specific gravity of approximately 0.9, would have to be even longer.

Vacuum pumps can be similar in design to air compressors. There are reciprocating-piston, diaphragm, rotary-screw, and lobed-rotor designs. (See air compressor types in Chapter 8, Fluid power pumps and accessory items.) Imagine hooking the inlet of an air compressor to a receiver tank and leaving the outlet open to atmosphere. As the pump runs, it evacuates air from the receiver and causes a negative pressure in it.

Fig. 1-9: Cross-sectional view of venturi vacuum generator

Vacuum pumps are an added expense and normally are only found in facilities that use a constant supply of negative pressure to operate machines or make products.

Vacuum generators that use plant compressed air as a power source are also available. These components have no moving parts but use plant air flowing through a venturi to produce a small supply of negative pressure. Figure 1-9 shows a simplified cutaway view of a venturi-type vacuum generator. The device consists of body A with compressed-air inlet B that passes air flow through venturi nozzle C. The air exhausts at a higher velocity to atmosphere through orifice D. As air at increasing velocity flows past opening E near the venturi nozzle, it creates a negative pressure and draws in atmospheric air through port F. Port F can connect to any external device that needs a vacuum source. A vacuum gauge at port F shows negative pressure when compressed air is supplied to port B.

Vacuum generators are inexpensive, but can be costly to operate. For every 4 cfm of air supply required to power them, they use approximately one compressor horsepower. For this reason, venturi-type vacuum generators usually are installed with a control valve to turn them on only when needed.

Vacuum is limited to one atmosphere maximum at any location, and standard vacuum pumps only reach about 85% (approximately 12 psi) of this on average. As a result, vacuum is not powerful enough to do much work unless it acts on a large area.

Fig. 1-10: Simplified representation of lifting with vacuum

Many industrial vacuum applications have to do with handling parts. Large-area suction cups can lift a large heavy part with ease, as illustrated in Figure 1-10. When the lift rises, negative pressure (vacuum) inside the suction cups causes atmospheric pressure on the opposite side of the part to push it up.

Fig. 1-11: Simplified representation of work holding with vacuum

Industries such as glass and wood manufacturing use vacuum to hold work pieces during machining or other operations, as shown in Figure 1-11. The pieces are held firmly in place as the negative pressure under them causes atmospheric pressure to push against them. A resilient seal laid in a groove in the fixture keeps atmospheric air from entering the cavity beneath the part. This groove can be cut to match the contour of the part. In machining operations, the seals can isolate interior cutouts, allowing them to be removed while firmly holding the rest of the piece.

Fig. 1-12: Simplified representation of plastic-sheet forming with vacuum

Heated plastic sheet can be vacuum-formed to make some products at a much lower cost than other types of plastic forming, as suggested in Figure 1-12. Forming heated plastic sheet in a cavity or over a shape is quick and positive. When atmospheric pressure tries to fill the negative-pressure area under the softened sheet, the sheet is forced into the desired shape. Large parts such as pickup-truck bed liners are formed by this method.

Two types of fluid power circuits

Most fluid power circuits use compressed air or hydraulic fluid as their operating media. While these systems are the same in many aspects, they can have very different characteristics in certain ways.

For example: remote outdoor applications may use dry nitrogen gas in place of compressed air to eliminate freezing problems. Readily available nitrogen gas is not hazardous to the atmosphere or humans. Because nitrogen is usually supplied in gas cylinders at high pressure, it has a very low dew point at normal system pressure. The gas may be different but the system's operating characteristics are the same.

Hydraulic systems may use a variety of fluids -- ranging from water (with or without additives) to high-temperature fire-resistant types. Again the fluid is different but the operating characteristics change little.

Pneumatic systems

Most pneumatic circuits run at low power -- usually around 2 to 3 horsepower. Two main advantages of air-operated circuits are their low initial cost and design simplicity. Because air systems operate at relatively low pressure, the components can be made of relatively inexpensive material -- often by mass production processes such as plastic injection molding, or zinc or aluminum die-casting. Either process cuts secondary machining operations and cost.

First cost of an air circuit may be less than a hydraulic circuit but operating cost can be five to ten times higher. Compressing atmospheric air to a nominal working pressure requires a lot of horsepower. Air motors are one of the most costly components to operate. It takes approximately one horsepower to compress 4 cfm of atmospheric air to 100 psi. A 1-hp air motor can take up to 60 cfm to operate, so the 1-hp air motor requires (60/4) or 15 compressor horsepower when it runs. Fortunately, an air motor does not have to run continuously but can be cycled as often as needed.

Air-driven machines are usually quieter than their hydraulic counterparts. This is mainly because the power source (the air compressor) is installed remotely from the machine in an enclosure that helps contain its noise.

Because air is compressible, an air-driven actuator cannot hold a load rigidly in place like a hydraulic actuator does. An air-driven device can use a combination of air for power and oil as the driving medium to overcome this problem, but the combination adds cost to the circuit. (Chapter 15 has information on air-oil circuits.)

Air-operated systems are always cleaner than hydraulic systems because atmospheric air is the force transmitter. Leaks in an air circuit do not cause housekeeping problems, but they are very expensive. It takes approximately 5 compressor horsepower to supply air to a standard hand-held blow-off nozzle and maintain 100 psi. Several data books have charts showing cfm loss through different size orifices at varying pressures. Such charts give an idea of the energy losses due to leaks or bypassing.

Hydraulic systems

A hydraulic system circulates the same fluid repeatedly from a fixed reservoir that is part of the prime mover. The fluid is an almost non-compressible liquid, so the actuators it drives can be controlled to very accurate positions, speeds, or forces. Most hydraulic systems use mineral oil for the operating media but other fluids such as water, ethylene glycol, or synthetic types are not uncommon. Hydraulic systems usually have a dedicated power unit for each machine. Rubber-molding plants depart from this scheme. They usually have a central power unit with pipes running to and from the presses out in the plant. Because these presses require no flow during their long closing times, a single large pump can operate several of them. These hydraulic systems operate more like a compressed-air installation because the power source is in one location.

A few other manufacturers are setting up central power units when the plant has numerous machines that use hydraulics. Some advantages of this arrangement are: greatly reduced noise levels at the machine, the availability of backup pumps to take over if a working pump fails, less total horsepower and flow, and increased uptime of all machines.

Another advantage hydraulic-powered machines have over pneumatic ones is that they operate at higher pressure -- typically 1500 to 2500 psi. Higher pressures generate high force from smaller actuators, which means less clutter at the work area.

The main disadvantage of hydraulics is increased first cost because a power unit is part of the machine. If the machine life is longer than two years, the higher initial cost is often offset by lower operating cost due to the much higher efficiency of hydraulics. Another problem area often cited for hydraulics is housekeeping. Leaks caused by poor plumbing practices and lack of pipe supports can be profuse. This can be exaggerated by overheated low-viscosity fluid that results from poor circuit design. With proper plumbing procedures, correct materials, and preventive maintenance, hydraulic leaks can be virtually eliminated.

Another disadvantage could be that hydraulic systems are usually more complex and require maintenance personnel with higher skills. Many companies do not have fluid power engineers or maintenance personnel to handle hydraulic problems.

5-1. Schematic drawing of a hydraulic circuit, and physical drawing of the components in the circuit.

Typical pneumatic circuit

Figure 5-1 includes a pictorial representation and a schematic drawing of a typical pneumatic circuit. It also has a pictorial and schematic representation of a typical compressor installation to drive the circuit (and other pneumatic machines). Seldom, if ever, is the compressor part of a pneumatic schematic. Power for a typical pneumatic circuit comes from a central compressor facility with plumbing to carry pressurized air through the plant. Pneumatic drops are similar to electrical outlets and are available at many locations.

Why schematic drawings?

Schematic drawings make it possible to show circuit functions when using components from different manufacturers. A 4-way valve or other component from one supplier may bear little physical resemblance to one from other suppliers. Using actual cutaway views of valves to show how a machine operates would be fine for one circuit using a single supplier's valves. However, another machine with different parts would have a completely different-looking drawing. A person trying to work on these different machines would have to know each brand's ins and outs . . . and how they affect operations. This means designing and troubleshooting every circuit would require special and different knowledge. Using schematic symbols requires learning only one set of information for any component.

Schematic symbols also give more information than a picture of the part. It may almost impossible to tell if a 4-way valve is 3-position by looking at a pictorial representation. On the other hand, its symbol makes all features immediately clear. Another advantage is that by using ISO symbols the drawing can be read by persons from different countries. Any notes or the material list may be unreadable because of language differences, but anyone trained in symbology can follow and understand circuit function.

Parts of a typical pneumatic system

The schematic in Figure 5-1 starts at the filter, regulator, and lubricator (FRL) combination that is connected to the plant-air supply. FRL units are important because they assure a clean, lubricated supply of air at a constant pressure. It's important to keep these units supplied, drained, and set correctly to keep the circuit operating smoothly and efficiently.

The filter is first in line to remove contamination and condensed water. It should be drained regularly or fitted with an automatic drain. The regulator should be set at the lowest pressure that will produce good parts at the cycle rate specified. The lubricator should be adjusted to allow oil to enter the air stream at a reasonable rate. In poorly maintained plants, the filter may be completely full of contaminants, the regulator is screwed all the way in, and the lubricator is completely empty.

Air-logic controls

Air-operated miniature valves called air-logic controls control the circuit in Figure 5-1. Air-logic controls run on shop air and are actuated by air palm buttons and limit valves to start and continue a cycle.

This circuit has an OSHA safe anti tie-down dual palm button start control. The two palm buttons must be operated at almost the same time or the cylinder will not extend. Tying down one palm button renders the circuit inoperative until it is released. The rest of the logic circuit causes the drills to extend and keeps the clamp cylinder down until they have all retracted and stopped. This circuit also has an anti-repeat feature, which means the cycle only operates once, even if the operator continues to hold the palm buttons down. Safety features such as these are easy to implement.

Directional-control valves

A 5-way, double-pilot-operated directional control valve operates the cylinder. This valve extends and retracts the cylinder according to signals from the air logic controls in the cabinet. Movement also requires inputs from the palm buttons to make sure the operator is safely clear of the cylinder before it operates. This directional control valve has speed-control mufflers in its exhaust port to control cylinder speed in both directions. These devices also reduce noise from exhausting air.

A limit valve at the extend stroke of the cylinder makes sure it has reached the part before the drills start. A limit valve monitors position but it cannot tell if the cylinder has reached full clamping force. In most applications when the cylinder is close enough to make the limit valve, it will be at or near clamping force before the next operation gets to the work. In some applications it might be necessary to add a pressure sequence valve to make sure the cylinder reaches a certain pressure before the cycle continues.

Air drills

Rotary output devices such as air motors with built-in cycling valves and rotary actuators that make only a fraction of a turn are available to perform many functions. Because compressed air is the driving force, these devices are explosion-proof and can operate in dirty or wet atmospheres without the problems posed by electrical equipment. Carefully applied air-operated devices can be an improvement in many situations.

These and other air-operated components are explained and applied in the following chapters.

5-2. Schematic drawing of an air circuit with air-logic controls, and physical drawing of the components in the circuit.

Typical hydraulic circuit

Figure 5-2 provides a pictorial representation and a schematic drawing of a typical hydraulic circuit. Notice that the hydraulic power unit is dedicated to this machine. Unlike pneumatic circuits, most hydraulic systems have a power unit that only operates one machine. (As mentioned before, some new installations are using a central hydraulic power source with piping throughout the plant to carry pressurized and return fluid.)

Why a schematic drawing?

Schematic drawings make it possible to show circuit functions when using components from different manufacturers. A 4-way valve or other part from a different supplier may bear little resemblance to one from other suppliers. Using actual cutaways of a valve to show how a machine operates would be fine for one circuit using one supplier's valves. Nevertheless, another machine with different parts would have a completely different looking drawing. A person trying to work on these different machines would have to know each brand and how they affect operations. This means designing and trouble shooting every circuit would require special different knowledge. Using schematic symbols requires learning only one set of information for any component.

Schematic symbols also give more information than a picture of the part. It may be hard to impossible to tell if a 4-way valve is 3-position by looking at a pictorial representation while its symbol makes all features immediately clear. Another feature is by using ISO symbols the drawing can be read by persons from different languages. Any notes or the material list may be in a language foreign to you but following and understanding circuit function should not be a problem.

Parts of a typical hydraulic schematic

A good starting point for any hydraulic schematic is at the power unit. The power unit consists of the reservoir, pump or pumps, electric motor, coupling and coupling guard, and entry and exit piping, with flow meters and return filter. It also might include relief valves, unloading valves, pressure filters, off-line filtration circuits, and control valves. The power unit must be able to cycle all functions in the allotted time at a pressure high enough to do the work intended. A well-designed circuit will run efficiently with little to no wasted energy that generates heat. It will run many years with minimum maintenance if its filters are well maintained and it is not overheated.

When items such as pressure gauges and flow meters are installed, it is easy to troubleshoot any system malfunction quickly and accurately. Flow meters always show pump flow (or lack thereof) and eliminate premature pump replacement. They can indicate impending pump failure well in advance of system failure. Also quick-disconnect plug-in type ports at strategic locations make it easy to check pressure at any point.

Directional control valves

The circuit in Figure 5-2 has only one directional control valve to extend and retract the main cylinder. Pressure-control valves make the hydraulic motor and rotary actuator operate in sequence after the cylinder extends and builds a preset pressure. (This is not the best way to control actuators, but it is shown here to demonstrate the use of different valves.)

An isolation check valve between the pumps keeps the high-pressure pump from going to tank when the low-volume pump unloads. A pilot-operated check valve in the line to the cap end of the main cylinder traps fluid in the cylinder while the motor and rotary actuator operate.

Pressure-control valves

A pressure-relief valve at the pumps automatically protects the system from overpressure. An unloading valve dumps the high-volume pump to tank after reaching a preset pressure. A kick-down sequence pressure-control valve forces all oil to the cylinder until it reaches a preset pressure. After reaching this pressure, the valve opens and sends all pump flow to the hydraulic motor first. A sequence valve upstream from the rotary actuator keeps it from moving until the hydraulic motor stalls against its load. A pressure-reducing valve ahead of the hydraulic motor allows the operator to set maximum torque by adjusting pressure to the motor inlet. (All of these controls are covered in the text of this manual.)

Another pressure-control valve -- called a counterbalance valve -- located in the rod end line of the main cylinder keeps it from running away when the directional control valve shifts. The counterbalance valve is adjusted to a pressure that keeps the cylinder from extending, even when weight on its rod could cause this to happen.

Accumulators

Because hydraulic oil is almost non-compressible, a gas-charged accumulator allows for storage of a volume of fluid to perform work. The expandable gas in the accumulator pushes the oil out when external pressure tries to drop. The accumulator in this circuit makes up for leakage in the cylinder cap-end circuit while pump flow runs the hydraulic motor and rotary actuator. Use care when specifying and using accumulators because they can be a safety issue.

These and other hydraulic components are explained and applied in the following chapters.

Parallel and series circuits

There are parallel and series type circuits in fluid power systems. Pneumatic and hydraulic circuits may be parallel type, while only hydraulic circuits are series type. However, in industrial applications, more than 95% of hydraulic circuits are the parallel type. All pneumatic circuits are parallel design because air is compressible it is not practical to use it in series circuits.

In parallel circuits, fluid can be directed to all actuators simultaneously. Hydraulic parallel circuits usually consist of one pump feeding multiple directional valves that operate actuators one at a time or several in unison.

Figure 5-3 shows a typical pneumatic parallel system schematic. All actuators in this circuit can operate at the same time and are capable of full force and speed if they have ample supply. The filter, regulator, and lubricator combination must be sized to handle maximum flow of all actuators in motion at the same time, When the air supply is insufficient, the cylinder with the least resistance will move first.

5-3. Schematic drawing of three cylinders in a typical pneumatic parallel circuit.

Figure 5-4 shows a typical hydraulic parallel system schematic. Any actuator in this circuit can move at any time and is capable of full force and speed when the pump produces sufficient flow. Parallel circuits that have actuators that move at the same time must include flow controls to keep all flow from going to the path of least resistance.

5-4. Schematic drawing of three cylinders in a typical hydraulic parallel circuit.

Flow controls are usually required to keep single cylinder movement from over speeding. The circuit in Figure 5-4 shows a meter-in flow control at each directional control valve's inlet to control speed in both directions. Placing flow controls at the cylinder ports would allow separate speeds for extension and retraction.

Figure 5-5 illustrates cylinders or hydraulic motors in typical series circuits. These synchronizing circuits are the most common use for actuators in series. The schematic drawing at left shows how to control two or more cylinders so they move simultaneously at the same rate. Oil is fed to the cylinder on the left and it starts to extend. Oil trapped in its opposite end transfers to the right cylinder, causing it to extend at the same time and rate. Oil from the right cylinder goes to tank. The platen moves and stays level regardless of load placement. Notice that this circuit uses double-rod end cylinders so the volumes in both ends are the same. (Other variations of this circuit are shown in the chapter on cylinders, which also explains synchronizing circuits in detail.)

5-5. Schematic drawings of two synchronizing hydraulic circuits.

The hydraulic motor circuit on the right in Figure 5-5 shows a simple way to run two or more motors at the same speed. Fluid to the first motor flows into the inlet of the second motor to turn it at the same time and speed. Except for internal leakage in the motors, they will run at exactly the same rpm. As many as ten motors can operate in series -- based on their loads and speeds.

Hydraulics vs. pneumatics

Pressurized fluids act in a certain manner in most situations. However, there are instances where a gas-type fluid does not perform as its liquid counterpart does. As mentioned earlier in this chapter, a pneumatic actuator is incapable of holding a position against increasing external forces because the air can be compressed more. Other situations such as flow-control circuits, return-line backpressure, energy-transfer considerations, and more are covered and explained in the text.

Conventions used in this manual

All schematic symbols and drawings are in accordance with the International Standards Organization (ISO) format. These symbols and representative parts are laid out in Chapter 4 either in whole or in part. Some symbols are made up of several standard parts and are not shown in their entirety in Chapter 4.

When a symbol is not shown it is good practice to use the symbol shown in the suppliers catalog. If no symbol is given there then use standard symbol parts to make a representation of the new item.

As in all cases of drawings using schematic symbols, the circuit designer may use his or her experience or opinion to interpret some parts. This usually does not make the schematic harder to read, just different. If a part representation is not clear, refer to the material list and check the supplier's catalog for an explanation of the valve's function.

Color coding

To better understand how a part or circuit works, consider using color coding for the lines and components. Color coding is instituted by the instructor, designer, or engineer and is according to his or her interpretation, so it might not be consistent in each case. Most training manuals and manufacturers use the following color code.

Red: Working fluid flow lines, usually from the pump to a device. This line is always solid. It can represent plastic tubing as small as 5/32-in. OD for air or any size pipe or tubing for hydraulics.
Blue: Return lines from valves and other devices for hydraulic circuits. This line always is solid, and can represent any size pipe or tubing.
Yellow: Metered or flow-controlled fluid that is at a reduced speed in relation to the same line without a restriction. This line could be solid or a series of long dashes if pilot flow must be metered.
Orange: A reduced-pressure line, such as a pilot-pressure line or one carrying accumulator precharge gas. This line could be a solid after a reducing valve or a long-dashed line for pilot flow.
Green: Pump inlet lines (suction lines) or drain lines. These lines would be solid for the pump inlet and a series of short dashes for drains. Two types of lines with the same color are not confusing -- even when in close proximity to each other.
Purple or indigo: These colors usually indicate working fluid that has been pressure-intensified by area differences or load-induced conditions. These pressures are usually greater than the setting of the main relief valve or reducing valve that feeds the circuit.
Lines without color are considered non-working or to have no flow at present.

This color-coding technique is used in the transparencies used with this manual

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2009-07-08T09:36:00.000-07:00

Fluid power pumps

A fluid power system’s prime mover is a pump or compressor that converts electricity or some form of heat energy into hydraulic or pneumatic energy. These devices can be rotary or reciprocating, single or multiple stage, and fixed or variable volume. They may move a variety of fluids and come in many different designs. Some pump designs offer unique features that make them especially suitable for a particular application.

Figure 8-1 shows several types of compressors in simplified cutaway form. These cutaways represent many standard designs used in industrial applications. They are not complete representations but simply show general working principles.

Fig. 8-1. Several designs of rotary air compressors

Reciprocating-piston air compressors

The single-piston/single-stage, dual-piston/single-stage, and dual-piston/dual-stage compressors illustrated in Figure 8-1 are typical designs for piston-type air pumps. Compressors of these designs may be rated as low as horsepower or as high as 1000 or more horsepower. The smaller sizes are air cooled while larger ones are water cooled.

Single-stage compressors normally operate at 125 psi or less and produce approximately 4 scfm (standard cubic feet per minute) of flow at 100 psi. (One scfm is 1 ft³ of gas at 68°F, 14.69 psia, and a relative humidity of 36%.

Diaphragm air compressors keep lubricating fluids out of the air or gas they are compressing. This arrangement often makes the air suitable for breathing and it can be used in applications where contamination from compressor oil cannot be tolerated. The cutaway view in Figure 8-1 shows an oil-driven diaphragm compressor that is capable of very high pressure. As the oil piston extends, it forces oil against the diaphragm to compress the gas. On the retract stroke, pressure inside the diaphragm plus vacuum returns the bladder to pick up more atmospheric air.

Piston-type reciprocating compressors below a 15- to 25-hp range usually start and stop at preset low and high pressure settings. Larger reciprocating compressors typically continue to run after pressure reaches the preset maximum, but they then stop compressing by holding their inlet valves open. This arrangement is called unloading. It saves wear on the electric motor because the motor only has to start one time.

Rotary compressors

Rotary compressors employ lobed rotors, vanes, screws, or impellers to draw in ambient air and compress it. Figure 8-1 also shows these devices. While these types of air pump are more compact and produce less vibration, they have lower efficiency than other types. All these designs (except the multi-stage centrifugal compressor) are limited to a maximum of 150 to 200 horsepower.

Rotary compressors run continuously and are capable of no flow to full flow at any time. An inlet-restricting valve closes or opens in response to pressure changes. Many rotary compressor installations do not require a receiver tank, due to their ability to change flow in relation to demand.

Pneumatic pump efficiency

Using atmospheric air as a means to transmit energy is very inefficient. A 1-hp air motor requires between 7 and 15 compressor hp while it runs. A hydraulic motor that produces the same output would only need 1½ to 2 hp input.

Air cylinders are more efficient than air motors, but still require three to four times more prime mover energy than their hydraulic counterparts. The general rule of thumb is: use hydraulic cylinders when an air-cylinder circuit would require a 4- or 5-in. or larger bores to produce the necessary force. This is especially important when the cylinders must operate at high cycle rates. Up-front cost of the hydraulic system is more, but operating cost savings soon pay for the added expense.

On the other hand, a 20-in. bore air cylinder used to maintain tension on a conveyor belt (with minimal cycling) would be a very efficient system.

Complete air compressor installation

Figure 8-2 combines the schematic diagram and picture representation of a typical air compressor installation. (The compressor could be a reciprocating or rotary type.) The aftercooler may not be required on installations under 50 hp, and it could be air-cooled instead of water-cooled. An air dryer is necessary in certain applications, but is often left out due to added cost. As noted earlier, a receiver tank might be eliminated with a rotary compressor is there never is a demand for short bursts of high-volume air. Water traps with drains are required on all systems because a compressor takes in a lot of water with the ambient air. (Even with an air dryer there is always the time when the dryer needs service but the system cannot be shut down. A trap will help during these times.) Other components, such as isolation or bypass valves for the aftercooler and air dryer, often are part of the circuit.

Fig. 8-2. Pictorial (at left) and schematic representations of typical air compressor installation

Hydraulic pumps

Most hydraulic pumps are positive-displacement devices. Pumps with positive sealing parts -- whether rotary or reciprocating -- move fluid every time they operate. This means that if the pump is turning, it produces flow. (Conversely; blocking flow stops the pump’s rotation mechanically.) Positive-displacement pumps have higher efficiencies than their non-positive-displacement counterparts, such as impeller or centrifugal designs. Figure 8-3 illustrates some non-positive-displacement designs that could be used to run hydraulic circuits. Because these pumps only run at 50 to 75% efficiency, they are not used in high-pressure circuits. They are frequently found in systems with high-water-content fluids (HWCF), such as 95% water and 5% soluble oil, because these pumps require little or no lubrication. Also, these systems usually operate at or below 400 psi.

Fig. 8-3. Two types of non-positive-displacement pumps

Some positive-displacement pumps are paired with centrifugal pumps to pressurize their inlets to keep them from cavitating. Or, when a positive-displacement pump is run at higher rpm than specified, the inlet may not be large enough to let in enough fluid at atmospheric pressure. In this case a non-positive-displacement pump can force fluid into the undersized inlet and eliminate cavitation.

A non-positive-displacement pump does not require a relief valve in many installations. There is enough slippage in most designs to allow for stopping flow while not over pressuring the circuit. However, if the pump operates at no flow for more than two or three minutes, simple bypass circuit to move fluid for cooling purposes should be added. The bypass circuit could be a small relief valve, a manual petcock, or a normally closed solenoid valve operated by a timer or pressure switch.

The propeller design is the least efficient of these pumps because there is a direct path from inlet to outlet through the blades. The minimum rpm of this type pump is high due to this open path. The centrifugal-impeller design operates at much closer tolerance so it slips less fluid while operating.

Fixed-displacement pumps

Fixed-displacement pumps are found most commonly in circuits with a single actuator. This allows the pump to be unloaded at little or no pressure when not performing work. A multiple-actuator circuit, where only one device moves at a time, can also be practical for fixed-displacement pumps if the actuators use about the same volume of fluid. This means total pump flow is either doing work at load pressure or is being sent to tank at very low pressure.

Avoid using meter-in or meter-out flow controls with fixed-volume pumps because a flow restriction increases pressure and the increase sends fluid to tank at the relief valve setting. This produces excess heat and all the problems associated with it. One way to use fixed-displacement pumps with multiple-actuator circuits is to include an accumulator with an unloading and dump valve. With this circuit, the pump is only on pressure when fluid is required. The accumulator accepts excess pump flow and provides working flow when the pump is unloaded. Figure 8-12 shows a fixed-volume pump with an accumulator.

Fixed-displacement pumps are usually less expensive and more contamination tolerant than pressure-compensated pump. Note: this does not mean they should be run with dirty fluid or that cheaper is really less expensive. It only means they fill the bill in many applications where cost is a factor.

Gear-on-gear fixed-displacement pumps

One of the oldest hydraulic pumps is the gear-on-gear design shown in Figure 8-4. As the driven gear turns, the idler gear turns in the opposite direction. At first, air trapped between the teeth and housing is moved to the outlet and forced out by the meshing teeth in the center. This starting action creates a negative pressure (vacuum) at the inlet. Atmospheric pressure then pushes oil into the pump. Now hydraulic fluid flows around the teeth and out to the circuit. Because the sealing action -- between the gear teeth and the housing, and where the teeth mesh -- has minimum clearance, when fluid is blocked, the gears stop turning.

Fig. 8-4. Gear-on-gear positive-displacement pump

A standard gear pump is unbalanced because there is high pressure on one side and low pressure or vacuum on the other side of the gears. This causes high bearing loads and shortened service life at pressures above 1500 psi. Some newer designs reduce this unbalance by clearing the housing (or clearance area) and only having a short sealing area. This greatly reduces bearing forces so that pressures up to 4000 psi continuous are commonplace today. However, even with this new design there is no compensation for gear or housing wear.

Gear-on-gear pumps can have more than one pumping section within a common housing. This allows for different flows or pressures to some circuits for speed and force changes.

Internal-gear fixed-displacement pumps

Figure 8-5 shows a cutaway view and the symbol for an internal-gear pump. The standard design is unbalanced and has no way to compensate for tooth or housing wear. Most pumps of this type are limited to 1000 psi or less. They are often used as transfer or supercharging pumps at low pressure due to their less efficient design. (There is a German-designed internal-gear pump that has a wear-compensating feature and a special bearing arrangement that allows it to operate continuously at up to 5000 psi and with more than 95% overall efficiency throughout its life.) Standard gear pumps start out at 85 to 90% efficiency when new. As the gears and housing wear, their efficiency deteriorates until they no longer supply enough fluid to maintain cycle time.

Fig. 8-5. Internal-gear positive-displacement pump

Gerotor fixed-displacement pumps

The newest design of a gear pump is called a gerotor (combining the words generated and rotor). A cutaway view and symbol is shown in Figure 8-6. This pump design is not common in the marketplace. At present there are only one or two manufacturers that offer this type. On the other hand, as a fluid motor it is one of the most common designs and is offered by more than 15 different companies.

Fig. 8-6. Gerotor-type positive-displacement pump

A gerotor pump uses a driven gear of, say, seven teeth inside an internal-tooth gear with eight teeth. The driven gear rotates inside the internal tooth gear and they both turn in the same direction. Because of the machined shapes, the driven gear always makes contact with the internal tooth gear at different points as they rotate. As the example in Figure 8-6 shows, this allows cavities to open and close as the gears turn.

In the example, as the driven gear turns clockwise, the internal tooth gear turns the same direction, but at one tooth per revolution slower speed. This action causes cavities to form on the left hand that start reducing pressure in this area. This reduced pressure (vacuum) allows higher atmospheric pressure to push fluid into the pump and fill the forming cavities. Kidney-shaped cavities in this sector, on both sides of the teeth, accept fluid to fill them for 180° around the inlet side. As the gears continue to turn, the cavities formed on the left side start closing on the right hand side. This forces fluid through the kidney-shaped openings and to the outlet port.

Like other gear pumps, gerotor pumps are unbalanced and have no way to compensate when clearances become worn. Although a new gerotor pump starts out at 85 to 90% efficiency, it deteriorates as it runs and constantly loses volume.

Gerotor pumps also can have more than one pumping section in a common housing, again allowing for different flows or pressures to some circuits for speed and force changes.

Another point on gear pumps: their output flow cannot be varied -- except by changing them physically or running them at a different speed. The next two types of pumps are capable of changing volume while running the same speed. These pumps can also reduce flow on a pressure build-up signal and almost eliminate the need for a relief valve.

Multi-screw fixed-displacement pumps

The pump in Figure 8-7 is similar to a gear pump but uses helical gears or screws to move the fluid. The driven screw is in close fit mesh with the idler screws and all gears have minimum clearance in the housing. As the driven screw turns, the idler screws also turn and the cavities between the screws move toward the outlet. This action forms a vacuum at the inlet. Atmospheric pressure then pushes fluid into the cavities and the fluid moves to the outlet. This pump has very smooth flow -- without the pulses produced by the other positive-displacement pumps in this manual. Flow from the outlet is smooth and continuous. However, screw pumps are not highly efficient. There is a lot of bypass in the original design and as the screws and housing wear, bypass increases. This design pump often is used to supercharge other pumps, as a filter pump, or a transfer pump at low pressure.

Fig. 8-7. Multiple-screw pump

Vane-type fixed-displacement pumps

The most common pump for industrial applications is the vane design shown in Figure 8-8. The left-hand cutaway view illustrates the original unbalanced design. Today, most vane pumps are of the balanced design shown on the right. Balanced vane pumps operate at higher pressures and have long bearing life. All vane designs compensate for wear, so their efficiency stays in the 90 to 95% range throughout their service lives. Vane pumps are efficient, quiet, and inexpensive. They have great longevity when supplied with clean fluid.

Fig. 8.8 Two designs of vane pumps

As a prime mover turns the rotor, centrifugal force slings the vanes outward. (Most manufacturers recommend a minimum speed of 600 rpm to make the vanes extend.) Now, as the vanes follow the off-center cam ring, a chamber is formed between the cam ring and the rotor. This chamber gets larger as the vanes extend, creating a negative pressure (vacuum) at the inlet port. Atmospheric pressure then forces fluid into these enlarging voids and fluid starts to move. As a vane passes the highest point on the cam ring, it is forced back into its slot and the chambers between the vanes decrease. As a chamber size decreases, fluid is forced out through the kidney-shaped openings to the outlet. Even though vane tips wear, they still touch the cam ring, so efficiency is not affected for a long time.

The other leakage and wear point is at the sides of the gears or rotors of these pumps. Most modern vane pumps have pressure-loaded floating plates that are hydraulically forced against the turning members. Hydraulic pressure tries to push the plates away from the gears or rotors in a certain area, but a slightly larger area on the opposite side of the plates pushes back under the same pressure. This keeps the side areas sealed without applying excess force against the turning members. (Some inexpensive low-pressure pumps may not have floating side plates but depend instead on manufacturing tolerances to control leakage.)

Note that the unbalanced vane pump in Figure 8-8 has pressure on one side of the rotor and vacuum on the other side. This pump has to have large bearings or operate at lower pressures. The balanced-design pump pictured on the right has pressure on opposite sides of the rotor. As a result, the bearing load is the same at 0 psi, 2000 psi, or any pressure at which the pump runs. The balanced design also produces twice the flow for the same overall package size.

Vane pumps are available with two or three pumps in one housing to give more flow or different rates of flow to satisfy the needs of some circuit designs. These pumps have a common inlet and separate outlets as required.

Typical circuits for fixed-volume pumps

Figure 8-9 shows a circuit using a fixed-volume pump in a simple, single-cylinder circuit. A tandem-center directional control valve routes all pump flow to tank at low pressure when the cylinder is idle. When the cylinder cycles, pressure never goes higher than necessary to do the work at hand, so energy waste is minimal. With an efficient pump, this circuit operates all day without a heat exchanger and fluid temperature never increases more than 10° or 15°F above ambient.

Fig. 8-9. Schematic diagram of open-center circuit with fixed-volume pump supplying single cylinder

Figure 8-10 shows a multiple-cylinder circuit supplied by a fixed-volume pump. Here, the tandem-center valves are connected in series, so all pump flow can go to tank when the actuators are idle. This circuit works best when the actuators do not move simultaneously. When two or more actuators move at the same time, the pressure to make the cylinders move is additive and may exceed the relief valve setting. Also, downstream actuators only get fluid from the actuators upstream from them. As a result, stroke lengths may be limited.

Fig. 8-10. Schematic diagram of open-center circuit with fixed-volume pump supplying multiple cylinders

Use caution when selecting directional valves for this circuit. Pay particular attention to pressure-drop charts because pressure drop is additive for each valve. This circuit could start up with a 200-psi drop at idle. With more valves in series, pressure drop at idle and running can cause sluggish operation and generate heat. Also, choose valves that are able to operate at tank line pressure. Every upstream valve sees pressure at pump and tank ports while a downstream actuator is working.

Figure 8-11 shows a multiple-cylinder circuit that uses a normally open solenoid-operated relief valve to unload the pump when the actuators are idle. Anytime an actuator cycles, a solenoid on its directional control valve and the solenoid on the normally open solenoid-operated relief must be energized at the same time. This circuit often requires flow controls -- and may need a heat exchanger to get rid of wasted energy.

Fig. 8-11. Schematic diagram of closed-center circuit with relief valve and fixed-volume pump supplying multiple cylinders

The circuit in Figure 8-12 has a fixed-volume pump with an accumulator to store energy and allow the pump to unload when no fluid is required to do work. It is similar to a pressure-compensated pump circuit because there is only pump flow at pressure when the circuit calls for it. The pump-unloading-and-accumulator-dump valve sends pump flow to the circuit until pressure reaches its set level. After reaching set pressure, the valve opens fully and dumps all pump flow to tank at minimum pressure. When circuit pressure drops about 10 to 15%, this valve closes and again directs pump flow to the circuit. (A normally open solenoid-operated relief valve controlled by a pressure switch could be used in place of the pump-unloading-and-accumulator-dump valve.)

Fig. 8-12. Schematic diagram of closed-center circuit with pump-unloading and accumulator-dump valve, and fixed-volume pump supplying multiple cylinders

Pressure-compensated, variable-volume vane pumps

Figure 8-13 shows cutaway views and symbols for a pressure-compensated vane pump. Vane pumps are one type of fixed-volume pump that can be made to function as variable volume and/or pressure compensated. The pumping action is the same as the fixed-volume, unbalanced vane pump previously discussed. The difference is that the cam ring is not fixed but can move in relation to the rotor. An adjustable force spring holds the cam ring in its offset position until enough pressure builds inside it to push against the spring and drive it toward center. As the cam ring moves closer to center, output flow decreases until it finally stops. The cam ring never makes it all the way to center because some flow is always needed to make up for internal bypass.

Fig. 8-13. Cross-sectional views of vane pump at full flow and at no flow

Internal leakage in fixed-volume pumps passes into the case and back into the inlet flow. Internal leakage in variable-volume pumps also passes into the case but has no passageway to return to the inlet line. All internal leakage must be drained from the case directly to tank through a full-flow drain line. This case-drain line should exit from the highest point on the pump so the case stays full of fluid at all times. Always fill the case of a newly installed pump to make sure it has lubrication at startup. Also, make sure the case-drain line terminates below fluid level in the tank so it cannot suck air.

Some pressure-compensated pumps have a maximum-volume adjusting screw to prevent the cam ring from going to full stroke. This feature makes it possible to adjust the maximum flow when pressure is below the compensator setting. The feature could be used to limit maximum horsepower when only a small portion of a higher flow pump is required. (In most circuits this feature has no use because flow is usually controlled by flow controls or actuator size.)

Two symbols can indicate pressure-compensated pumps schematically. The complete symbol on the left shows all the functions, while the simplified symbol on the right omits the case drain and shows the compensating arrow inside the pump circle. Because most schematic drawings now are done on CAD systems that automatically produce the complete symbol, the simplified symbol seldom appears today.

Pressure-compensated pumps normally do not need a relief valve to protect the system from over pressure. However, many circuits with pressure-compensated pumps use a relief valve just in case the pump hangs on flow. When a relief valve, for whatever reason, is used with a pressure-compensated pump, it is imperative that it be set 100 to 150 psi higher than the pump compensator. If the relief valve is set lower than the compensator, the circuit will operate as a fixed-volume setup and quickly overheat the fluid. If the relief valve is set at the same pressure as the compensator, it is possible that the relief valve will start to dump at the same time the compensator starts to reduce flow. Then the pressure drop lets the relief valve shut and the compensator asks for more flow. These oscillations can continue until the pump fatigues and fails.

Setting the relief valve and compensator is a four-step operation.

Set the relief valve at maximum pressure.
Set the pump compensator at a pressure that is 200 to 300 psi higher than final system pressure.
Set the relief valve 100 to 150 higher than the final compensator setting.
Set the pump compensator at system pressure.

The other reason often stated for using a relief valve in a pressure-compensated pump circuit is because of pressure spikes. When a pressure-compensated pump has to instantaneously shift from full flow to no flow, fluid leaving the pump while it is shifting to center has no place to go. Because pressure is resistance to flow and resistance is a maximum at this point, pressure can climb very high. These full-flow to no-flow spikes can easily go as high as five to seven times the pump compensator setting (depending on the pump volume). Adding a relief valve to this scenario can reduce the spikes because a relief valve will respond much faster than a pressure-compensated pump. However, a pilot-operated relief valve still has some response time and will often spike two to three times its setting before opening fully.

A better way to protect the pump and circuit is to install a small accumulator at the pump outlet and pre-charge it to approximately 80% of set pressure. Now when the pump must react rapidly, the accumulator provides a place for excess fluid to go. An accumulator also helps actuator response time at cycle start because there is a ready supply of fluid even though the pump is at no flow.

Piston-type, fixed-displacement pumps

There are two types of piston pumps in use today. The oldest design is the radial-piston type. Radial-piston pumps come in two different configurations. The one shown in Figure 8-14 is sometime called a check valve or eccentric pump. The design in Figure 8-15 is what usually comes to mind when radial pumps are mentioned.

Fig. 8-14. Cross-sectional view of radial-piston pump (check valve or eccentric type)

The cutaway in Figure 8-14 shows how the pistons move fluid when the eccentric turns and strokes them forward, while springs return them. Check valves at the piston ends allow flow from the inlet chamber and exit flow to the outlet port.

Many of these type pumps are capable of very high pressures -- up to and exceeding 10,000 psi. At the same time they usually flow low volume -- below 6 gpm. They are highly efficient pumps, with unidirectional flow. In fact cw or ccw shaft rotation produces the same flow rate and direction. (An eccentric pump can be made pressure compensated and/or variable volume by restricting inlet flow or pressurizing the area under the pistons to keep the springs from fully extending them.)

Fig. 8-15. Two cross-sectional views of variable-displacement radial-piston pump

Variable-displacement radial-piston pumps

Figure 8-15 shows a cutaway view of a basic radial-piston pump that can function as fixed volume, variable volume, pressure compensated, and bidirectional flow, or a combination of these functions. The pump in Figure 8-15 is variable volume only. As a fixed-volume pump it would have the reaction ring offset as shown in the right hand cutaway view, with no method of changing that condition. (This is one configuration that will probably never be used with this design pump.)

As the cylinder block and pistons rotate, centrifugal force pushes the pistons against the reaction ring. When the pump is in the on-flow condition (as in the right-hand cutaway view), the pistons are moving out of their bores in the lower half of the picture and forming a vacuum. Fluid is forced into the inlet and fills these voids. As the pistons pass left center, they stop extending and begin to be pushed back into their bores. During the top half of their travel, the pistons force the trapped fluid through the outlet to the circuit. Moving the reaction ring’s centerline closer to the cylinder block’s centerline reduces flow

Pressure-compensated, radial-piston pumps

The radial-piston pump in Figure 8-16 is pressure compensated. This pump produces flow when the outlet pressure falls below the level set by the pressure-adjusting screw. When pressure in the pilot line increases enough to compress the compensator spool’s spring, pilot flow is connected to the compensator piston, and its drain to the case is blocked. Pilot flow to the compensator piston forces the reaction ring to move against the return spring and reduce outlet flow. The reaction ring never reaches center because the circuit, pilot control, and internal leakage must be overcome to hold pressure.

Fig. 8-16. Cross-sectional view of pressure-compensated radial-piston pump, with symbols

Two symbols can be used to show pressure-compensated pumps schematically. The complete symbol at the lower right of Figure 8-16 shows all the functions, while the simplified symbol above it omits the case drain and places the compensating arrow inside the pump circle. Again, because most schematic drawings are done on CAD systems now, the simplified symbol is seldom used.

A radial-piston pump can also produce bi-directional flow. It can take in or force out fluid from either port while turning the same direction. This design pump is used in closed-loop circuits where all outlet flow goes to an actuator and return flow from the actuator goes back to the pump inlet. A common circuit of this type is a hydrostatic drive. Fluid from a bi-directional pump goes to a bi-directional motor to give infinitely variable output speed and force in either direction of rotation without requiring a directional control valve.

Bi-directional, radial-piston pumps

The pump in Figure 8-17 has a small opposing piston that pushes continuously against a larger control piston on the opposite side of the reaction ring. The control piston can be pressurized or exhausted by a 3-way servovalve, thus infinitely varying the reaction ring position to either side of center. Input signals to the servovalve can come from manual, mechanical, or electronic controllers. A common circuit produces four manually variable flows and directions, using four single-solenoid directional control valves.

Fig. 8-17. Radial-piston pump used in bi-directional flow circuit

A charge pump, driven off the main pump shaft, supplies pilot oil to maintain pressure on the opposing piston. It also supplies oil to the mechanical-feedback servovalve that pressurizes or exhausts the control piston. The charge and pilot circuits usually run at 250 to 400 psi. Notice that the “A” and “B” ports are only connected to the actuator -- not to tank -- when using a hydraulic motor or double rod-end cylinder. (The pump must have added tank ports to operate a single rod-end cylinder circuit.)

Figure 8-18 shows a cutaway view and schematic drawing of a bi-directional pump driving a single rod-end cylinder. Because there is less volume in the rod end of a single rod-end cylinder, flow to and from that end is less in relation to the cap end. This poses a problem when using a closed-loop circuit.

Fig. 8-18. Cross-sectional view and schematic diagram of closed-loop circuit with bi-directional pump supplying single rod-end cylinder

The pump cutaway and schematic show how adding suction check valves, a shuttle valve, and a bypass relief valve allow the pump to bypass excess flow from the cap end and take in added flow for the rod end. This is a common circuit for this type pump. With this circuit, cylinder speed is infinitely variable and direction change requires no directional control valve. Direction change is very smooth because flow must go to zero in one direction before it can reverse. Because of this, the actuator rapidly and smoothly decelerates to a stop condition. When flow reverses, it increases steadily to full flow in the opposite direction without system shock.

Wobble-plate piston pump

The wobble-plate piston pump design shown in Figure 8-19 is one type of inline or axial-piston pump. As the wobble plate turns, the spring-loaded pistons reciprocate -- drawing in fluid as they spring return and discharging it as they are forced to extend. Direction of rotation is not important for this pump because flow is the same when it turns either way.

Fig. 8-19. Cross-sectional view of wobble-plate pump

Many pumps of this design operate at very high pressure and can flow high volume as well. Another feature is the ability to isolate the outlet of one or more pistons to give more than one flow volume to a circuit. This allows a single pump to function like other double or triple pumps in hi-lo circuits or to operate different actuators at various flows and pressures. This design pump can also be made variable volume and/or pressure compensated. Some designs use a restricted inlet to accomplish both functions because the spring-loaded pistons will not fill as far if their inlet is restricted.

Inline or axial-piston, fixed-volume pumps

Figure 8-20 shows a more common design for piston pumps. This design is seldom used as a fixed-volume pump because it can be made pressure compensated -- which many circuits require. This design can be fixed-volume, variable volume, pressure compensated and bi-directional flow, the same as the radial-piston design. The main reasons for its popularity are its compact design and its lower price. A radial-piston pump of the same flow will normally cost four to six times as much as the inline design and be three to four times larger physically.

Fig. 8-20. Cross-section view and symbol for fixed-volume inline or axial-piston pump

An inline piston pump like the one in Figure 8-20 is similar in design to the wobble-plate pump. The main difference is in the way the pistons move and stroke. An inline pump uses a fixed-angle swashplate instead of a wobble plate. The pistons are not spring loaded but are held against the swashplate by piston shoes and a shoe plate. The pistons are pulled out of and pushed into their bores mechanically.

The pistons are fitted in the cylinder block, which is splined to the drive shaft, and they turn along with the shoes and the shoe plate. As the pistons slide down the swashplate, they are pulled out of their bores and create a vacuum at the inlet port. Atmospheric pressure forces fluid to fill the piston bores until the pistons reach the bottom of the swashplate angle. Fluid enters through the kidney-shaped openings half way around one revolution. As the cylinder block continues to turn, the pistons are forced back into their bores and fluid discharges through the outlet. A kidney-shaped opening on the other half of the valve plate allows fluid to flow until the pistons are fully returned. Inline pumps always have an odd number of pistons, so one never can be directly across from another at the transition from being pulled out to being pushed in.

Inline piston pumps require a case drain to send bypass and/or control oil back to tank. The drain line should be unrestricted at all times and should terminate below the fluid level in the tank. If the drain line terminates above fluid level, the pump housing can be vacuumed dry, causing damage to the pump.

It is good practice to install a flow meter in the drain line. The flow meter indicates when to change out the pump before it loses efficiency or is worn beyond repair. A flow meter with an integral limit switch can be set to give a warning when case drain flow goes above a specified volume. Usually a pump should be changed when case flow is greater than 7 to 10% of maximum rated flow.

Inline piston pump efficiency runs in the 95 to 98% range. They, are very versatile, have many control options, and would work well on any type circuit. They are more expensive than gear and vane pumps so they lose out when price is the deciding factor.

Variable-volume inline or axial piston pumps

Most inline pumps have some way to change the angle of the swashplate. This makes the pump capable of variable volume, pressure compensation, and bi-directional flow. Figure 8-21 shows a variable-volume setup with a manual control. Low-flow pumps (those under 20 gpm) can use manual controls. Higher-flow pumps need hydraulically powered pistons to move against the higher forces in the pump.

Fig. 8-21. Cross-section view and symbol for variable-volume inline or axial-piston pump

The basic operation of this pump is the same as a fixed-volume inline piston pump. The difference here is the angle of the swashplate can be changed manually to allow longer or shorter piston strokes for more or less volume while the pump turns at the same speed. This feature can conserve energy when an actuator needs variable speeds. It replaces a flow control that limits flow and either sends excess fluid across a relief valve or forces a pressure-compensated pump to go to high pressure and reduced flow. Other controls include manual servo, manual handwheel, and electronic servo, to name a few.

If this pump is in an open-loop circuit, make sure the control cannot go past center -- or no flow -- condition. If the lever is moved left of perpendicular, flow reverses and the pump tries to take fluid from the circuit and send it to tank. Very soon the pump will run dry and be damaged due to lack of lubrication. (Later in this text, a bi-directional pump circuit is shown with all the necessary additions to make the pump work properly in a bi-directional mode.)

Notice that the symbol in Figure 8-21 duplicates the standard pump symbol with a sloping arrow added to it. This indicates a pump with variable or adjustable flow.

Pressure-compensated inline or axial piston pumps

The pressure-compensated pump shown in Figure 8-22 can change outlet flow when pressure tries to go above a predetermined setting. This design pump only has outlet flow when there is a pressure drop due to circuit demand. (Most manufacturers offer an option to limit maximum flow when pressure drops to add versatility to the circuit.) The maximum-volume adjusting screw keeps the volume-destroking piston from retracting all the way even when pressure drops.

Fig. 8-22. Cross-section view and symbol for pressure-compensated inline or axial-piston pump

Pump operation is the same as previously explained for fixed-volume inline or axial piston pumps. The difference is that this design has a moveable swashplate that is held on stroke by the on-stroke spring. These pumps always produce full flow when pressure is below the compensator setting.

When this pump’s outlet flow meets resistance, pressure builds in the pump-pressure communicating port. This pressure pushes against the spring-loaded compensator spool. In its normal position, this spool allows fluid behind the volume-destroking piston to go to tank through the case drain. When pressure is high enough to force the compensator spool against its spring, the spool allows fluid to flow into the chamber behind the volume-destroking piston while it blocks flow to tank. Enough fluid enters the chamber behind the volume-destroking piston to push the swashplate against its spring and start destroking the pump. The swashplate moves to a position to stroke the pistons just enough to makeup for the bypass and control fluid used by itself and any fluid used in the circuit. This could be any amount of flow -- even zero. Because of this, the pump never sends fluid to tank across a high-pressure relief valve, so heat generation is minimal. Pump stroke varies anytime fluid is required -- from maximum to minimum depending on circuit use.

Two symbols can be used to show pressure-compensated pumps schematically. The complete symbol on the left shows all the functions, while the simplified symbol on the right omits the case drain and puts the compensating arrow inside the pump circle. Because most schematic drawings are done on CAD systems now, the simplified symbol is seldom used.

The inline pump design is subject two common problems:

Operating the pumps at high vacuum inlet can quickly deteriorate the swaged connection between the piston and shoe (see Figure 8-23). When this joint is subjected to extra pulling and then pushing 12 to 1800 times per minute, it wears and comes apart quickly. When it does come apart, it wrecks the swashplate surface and the rest of the piston shoes. Most manufacturers recommend 1 psi or less vacuum at the inlet, and indicate longer life if the pump is supercharged by another pump at 5- to 30-psi inlet pressure.
The shoe has hollowed out areas on its face that receive oil through an orifice in the piston as it forces fluid out. This bypass oil lubricates the shoe and causes it to float a few micrometers off the swashplate. This happens because the shoe’s area is greater than the area of the opposing piston. Because there is no metal-to-metal contact between these parts, the pump has long life expectancy. If contamination stops flow of pressurized fluid, the shoes will contact the swashplate at high force while there is minimum lubrication. The pump fails shortly thereafter. When possible, feed the pump with at least tank head pressure by mounting it alongside or under the tank. Also, keep fluid cleanliness level at least that specified by the pump manufacturer.

Fig. 8-23. Potential problem areas within inline-piston pumps

As noted in Figure 8-23, always fill the case of a new or repaired piston pump with fluid before startup. The pump needs lubrication and will have very little until bypass fills the case. Also, a filled case will seal clearances and make it easier for the pump to prime.

Fixed-volume bent-axis pumps

Another type piston pump is the bent-axis design shown in Figure 8-24. Like the radial piston pumps previously discussed, this is a very expensive pump and it is physically large when all its optional features are installed. Therefore, this design pump is not a common sight in industry.

Fig. 8-24. Cross-section view and symbol for fixed-volume bent-axis hydraulic pump

Its main advantage over an inline piston pump is that it holds up much better when the inlet sees high vacuum. The piston connections are stronger and are not prone to separating from the drive.

This type pump is manufactured in fixed-volume, variable-volume, and pressure-compensated models, as well as with bi-directional flow and combinations these functions. It has an efficiency range from 95 to 98% and gives long service life when supplied with clean fluid. Most manufacturers make this pump in low- to high-volume sizes. Most are capable of 4000 psi and more.

The cutaway view of a fixed-volume bent-axis pump in Figure 8-24 shows that as the drive shaft turns, the cylinder block also turns at the same rate through the universal drive link. Because the cylinder block is at an angle to the drive shaft, the pistons reciprocate in their bores. The pistons draw in fluid during one half of each revolution and discharge fluid during the other half. Kidney-shaped openings in the valve plate direct the fluid in and out of the piston bores. Because the housing is a single piece, the angle and volume is fixed for a given rpm.

Variable-volume, pressure-compensated bent-axis pumps

The cutaway view in Figure 8-25 shows a bent-axis pump that is capable of variable volume as well as pressure compensation. This means the pump output can be varied by a manual control or it can automatically change as pressure increases to a predetermined setting. (This cutaway represents only one way such a pump might be built.) The operates in the same way as the pump in Figure 8-24, but the angle of the cylinder block can vary to reduce flow on a pressure demand. Also, the maximum-volume screw can limit the maximum angle of the cylinder block to establish maximum flow. This is an option on many manufacturers’ designs.

Fig. 8-25. Cross-section view and symbol for variable-volume pressure-compensated bent-axis pump

As pressure at the outlet builds to the setting of the pressure adjustment, the compensator spool is pushed back. The spool forces the compensator piston to push the cylinder block to a lesser angle. When pressure reaches the preset level, the cylinder block stays in any position required to maintain the flow needed at the preset pressure.

Two symbols can be used to show pressure-compensated pumps schematically. The complete symbol on the left shows all the functions, while the simplified symbol on the right omits the case drain and shows the compensating arrow inside the pump circle. Because most schematic drawings are done on CAD systems now, the simplified symbol is seldom used.

Figure 8-26 shows the symbol and a cutaway view of a bi-directional, bent-axis pump for closed-loop circuits. This pump operates in the same manner as the previously described bent-axis pumps, but is capable of drawing in and discharging fluid from either port while turning the same direction. This design needs an external pilot supply because it has no integral pilot pump.

Fig. 8-26. Cross-section view and symbol for bi-directional bent-axis pump

The cutaway shows optional features such as: maximum-volume screws in both flow directions and a proportional-control valve for infinitely variable flow from either port. Manual, mechanical, and solenoid controls also are available. A control piston that is offset by a resisting piston with a smaller diameter moves the cylinder block. The 3-way servovalve ports fluid to or exhausts fluid from the larger piston to position the cylinder block. This pump design is not readily available currently, but there still are many of them operating in the field.

Load-sensing function

All pumps that can be pressure compensated can also be made load sensing. Load sensing is a control technique that keeps the pump compensator from holding full pressure until an actuator stalls. Normally a pressure-compensated pump circuit operates at full compensator pressure setting unless an actuator is using all the pump flow. While an actuator is using all pump flow, pressure is whatever it takes to move the load. This is an ideal setup because all energy -- except for component inefficiencies -- is being used to do work. There is no wasted energy except for inefficiencies and very little heat is generated. A load-sensing circuit uses a feedback signal from the actuator that keeps pump pressure at 100 to 300 psi above the load. Some load-sensing pumps have a fixed differential while others are adjustable. When no actuator is moving, system pressure is at the load-sensing setting of 100 to 300 psi instead of the compensator setting. Energy savings is the main advantage of a load-sensing function, but it also makes a non-compensated flow control perform like it is pressure compensated.

Fig. 8-27. Schematic diagram of pressure-compensated closed-center load-sensing circuit

The schematic drawing in Figure 8-27 is a typical load-sensing circuit with two actuators. Notice that the sensing lines from the actuator flow lines to the pump compensator. A load-sensing pump must be able to read any load it is powering so that ample pressure can be maintained. Also notice that the load-sensing lines go through check valves to isolate the flow lines from each other.

All flow controls in a load-sensing circuit must be meter-in type so pressure at the actuator is always high enough to move the load. In the case of the vertical cylinder, a counterbalance valve keeps it from running away while extending. Notice that the load-sensing line from the rod end of the vertical cylinder is connected between the counterbalance valve and the directional control valve so it does not see a load when the circuit is at rest.

Because the pump is pressure compensated, the directional control valve’s pump port is blocked in center position. This circuit uses a bar manifold with modular meter-in flow controls and a modular “B” port counterbalance valve sandwiched under float-center directional control valves for piping convenience and leak prevention. The lines connected to the “A” and “B” ports below the meter-in flow controls go to isolation check valves, then on to the load-sensing connection.

With the circuit at rest as shown, the load-sensing connection sees little or no pressure because the actuator ports are connected to tank. At this point, circuit pressure is equal to the load-sensing bias spring, regardless of the setting of the adjustable compensator spring. At this low pressure, the circuit consumes very little horsepower and generates little heat. The pump’s internal parts are subject to low stress, which makes them last longer and maintain high efficiency

When a cylinder cycles, the load-sensing connection sees whatever pressure it takes to move it. Pump outlet pressure rises to load pressure plus load-sensing bias-spring force. When both cylinders operate simultaneously, the load-sensing connection receives pressure from the highest load through the isolation check valves. Pump pressure is always that needed to move the highest load plus a value added by the load-sensing bias spring. (Some load-sensing bias springs are adjustable within a narrow range.)

When load-sensing valves have low or no bypass flow, use shuttle valves in place of the isolation check valves. Shuttle valves will not trap backflow when a directional control valve shifts to center position. Check the chosen pump to see if this feature is standard, must be specified, or is not available.

Load-sensing, fixed-volume pumps

The oldest load-sensing circuits for fixed-volume pumps are like those diagrammed in Figures 8-9 and 8-10. The pumps in these circuits never operate at a higher pressure than work resistance and never send fluid across the relief valve unless there is a malfunction in the hydraulics or control circuit.

Figure 8-28 diagrams a simple load-sensing pump circuit using standard valves. A pilot-operated relief valve with a 70-psi spring dumps pump flow to tank at 70 psi when the vent port is at 0 psi. A shuttle valve receives pressure feedback from the actuator and signals the pilot-operated relief valve’s vent port with the actual working pressure. As the actuator moves at a reduced speed, pump pressure stays 70 psi above actual load pressure, so excess flow that goes to tank wastes less energy.

Fig. 8-28. Schematic diagram of fixed-volume pump in closed-center load-sensing circuit

Several manufacturers offer fixed-volume pumps with integral load-sensing valves. Hookup is simple for these pumps, and in some designs, bias pressure can be adjusted.

This setup is not as efficient as a pressure-compensated pump with load sensing, but it always provides an advantage in fixed-volume pump circuits. The results are best when maximum system pressure is high and actuator’s extension and retraction speeds are low.

Horsepower- and/or torque-limiting pumps

Horsepower or torque limiting is another control technique that only works on pumps that are capable of variable volume. Its main application is in the mobile-equipment field, where most hydraulic circuits are powered by gas or diesel engines. These engines usually must move the machine as well. To maximize actuator speed and force while minimizing horsepower drain, all actuators can be fast at low loads but still able to move heavy loads without pulling excess horsepower. Each manufacturer that supplies this setup may have a different way of doing it.

Fig. 8-29. Schematic diagram of horsepower- or torque-limiting pump control

Figure 8-29 shows the schematic diagram of a circuit with a horsepower-limiting pump. A pressure-compensator adjustment still controls maximum output pressure but the preset limiting valve can reduce flow as pressure increases. Reducing flow as pressure increases keeps horsepower or torque from exceeding a preset limit. The horsepower/torque limiter is preset for a given pressure and flow. This system could be useful in an accumulator circuit to allow higher flow as pressure decreases while limiting horsepower draw as pressure climbs.

Typical circuit for pressure-compensated pumps

Most pressure-compensated pumps use a closed-center circuit such as the one in Figure 8-30. These circuits could have load sensing or other controls. They usually include multiple actuators. Closed-center circuits typically operate at maximum system pressure and output flow matches the circuit requirement. Flow controls keep actuators at operating speed because maximum flow may make them move too rapidly. Flow controls also make it possible for more than one actuator to move simultaneously without affecting their stroke times. Note that flow controls also increase heat generation because the moving or work force may not require full system pressure. Also, some actuators may require pressure-reducing valves to lower the maximum force so that it doesn’t cause damage.

Fig. 8-30. Schematic diagram of typical pressure-compensated pump circuit

Normally, pressure-compensated pumps do not need relief valves to protect their systems from overpressure. However, many circuits with pressure-compensated pumps include a relief valve just in case the pump hangs on flow. When a relief valve, for whatever reason, is used on a pressure-compensated pump, it is imperative that the relief valve is set 100 to 150 psi higher than the pump compensator. If the relief valve is set lower than the compensator, the circuit will operate as a fixed-volume setup and quickly overheat the fluid. If the relief valve is set at the same pressure as the compensator, the relief valve can start to dump as the compensator starts to reduce flow. Then pressure drop lets the relief valve shut and the compensator ask for more flow. This oscillating action can continue until the pump fatigues and fails.

Setting the relief valve and compensator is a four-step operation:

Set the relief valve at maximum pressure.
Set the pump compensator at a pressure 200 to 300 psi higher than the final relief valve pressure.
Set the relief valve 100 to 150 psi higher than the final compensator setting.
Set the pump compensator at system pressure.

Another reason often stated for using a relief valve in a pressure-compensated pump circuit is because of pressure spikes. When a pressure-compensated pump has to instantaneously shift from full flow to no flow, fluid leaving the pump while it is shifting to center has no place to go. Because pressure is resistance to flow and resistance is maximum at this point, pressure can climb very high. These full-flow-to-no-flow spikes can easily go up to five to seven times the pump compensator setting, depending on the pump volume.

Adding a relief valve to this scenario can reduce the spikes because a relief valve will respond much faster than a pressure-compensated pump. However, a pilot-operated relief valve still has some response time and will often spike two to three times its setting before opening fully.

A better way to protect the pump and circuit is to install a small accumulator at the pump outlet, pre-charged to approximately 80% of set pressure. Now, when the pump must react quickly, the excess fluid can go into the accumulator with very little pressure spike. An accumulator also helps actuator response time at cycle start because there is a ready supply of fluid even though the pump is at no flow.

Another consideration is pump priming when a pressure-compensated pump is mounted above fluid level. When a system first starts (and sometimes when it has not been operated recently), the inlet line holds no oil above tank level. Atmospheric air in this line above fluid level must be evacuated before atmospheric pressure can push fluid in. Because most pressure-compensated pumps operate against a closed-center circuit, there is no place for this trapped air to go. Hydraulic pumps may easily move 100 gpm of fluid at 3000 psi, but they are very poor air movers. At startup, the pump never primes and could be damaged from lack of lubrication -- especially if the case has not been filled. Usually the outlet line is opened at a union or some other fitting and the pump primes as soon as the trapped air can leave.

A better approach is to install the air-bleed valve shown in Figure 8-31. This valve is not required in most cases when the pump is along side or below the tank because filling the tank should also fill the inlet line. It is also seldom required with a fixed-volume pump in an open-center circuit because the pump outlet has a direct path to tank. However, when priming is a problem and there are no inlet line leaks or restrictions, then the air-bleeds valve may be required. The circuit in Figure 8-30 shows the correct location and piping for this valve.

Fig. 8-31. Cross-section view and symbol for typical air-bleed valve

The cutaway view in Figure 8-31 shows the internal configuration of a typical air-bleed valve. The poppet in this valve normally is held open by a light spring, so trapped air can flow easily through its flow orifices to tank. When the pump primes and oil tries to flow through these orifices, pressure builds and the poppet closes. The poppet stays closed as long as the pump is running.

Always pipe the air bleed valve as close to the pump outlet as possible. Any oil in the line must be pushed out before trapped air can be exhausted, so the closer the better. Always terminate the air bleed valve’s outlet below fluid level. If it terminates above fluid level, air can pass through the valve and let oil in the pump return to tank.

Closed-loop circuits

The circuit in Figure 8-32 is a typical hydrostatic-transmission setup. It uses a variable volume, bi-directional pump to drive a hydraulic motor at infinitely variable speed. Hydrostatic drives are normally used to drive vehicles but can be used in industrial applications where smooth acceleration, deceleration, and reversing are required. These circuits usually incorporate an inline or axial-piston pump coupled to a variety of hydraulic motors. As a closed-loop circuit, all pump flow goes to the motor and all motor flow returns to the pump. With 100% efficient parts, the circuit could run with the same oil its whole life. In the real world however, the hydraulic motor and bi-directional pump have internal bypass so a fixed-volume charge pump is placed in the circuit to make up for leaks. The charge pump can also supply fluid to control circuits and accessory devices.

Fig. 8-32. Schematic diagram of typical hydrostatic drive circuit

The charge pump inlet draws fluid from a reservoir through a low-micron filter and sends it to the inlets of the charge check valves. When the hydraulic motor is not turning, any fluid not used by the closed loop goes through the charge-pump relief valve, then back to tank through the pump case and a heat exchanger. When the hydraulic motor is turning, all charge flow goes to the low-pressure side of the loop through one of the charge check valves, the hot-oil bypass valve, and the hot-oil bypass relief valve at a lower pressure. This action makes sure the closed loop receives cooled, filtered oil that can carry away heat and contamination. It also sends cool, clean oil through the motor and pump case to flush contamination and dissipate heat.

Small hydrostatic pumps can be controlled manually, hydraulically, or electro hydraulically. Larger systems cannot be controlled manually, due to the high force required to move the swashplate.

The system-relief valves protect the hydraulic motor and bi-directional pump from excess pressure when the motor is powered. When the pump center’s motor outlet flow is blocked, the motor may be driven by external forces and cannot stop immediately. At this time, the system-relief valves allow fluid from the motor -- now acting like a pump -- to bypass at high pressure to the opposite motor port. This allows the motor to stop smoothly even when an operator tries to stop it abruptly. (Other options to protect the circuit from bypassing through the system-relief valves during deceleration are available from most suppliers.)

Bi-rotational pumps

Unirotational pumps can only move fluid when rotating in one direction. These pumps usually have a larger inlet port in relation to the outlet port size. They are limited as to inlet-outlet function because internal bypass is always ported to the housing on the inlet side. This means all internal bypass goes to the case and then back to the pump inlet. Because all pumps have a shaft sticking out of the housing, there must be a seal to stop fluid leak when the pump is at rest and vacuum leak when it is running. A unirotational pump could move fluid when turning either way but the shaft seal would blow above 25- to 50-psi outlet pressure in reverse flow.

Bi-rotational pumps can move fluid while turning in either direction of rotation if they are piped correctly. Both ports on these pumps are usually sized as inlets. Figure 8-33 shows a cutaway view of a bi-rotational pump with internal check valves that allow bypass to go to the inlet side of the pump. Bi-rotational pumps are mainly used on mobile equipment where the prime mover cannot easily change direction of rotation. This means right- and left-hand rotation pumps would have to be kept to satisfy different pieces of equipment.

Fig. 8-33. Cross-sectional view and symbol for bi-directional pump

In industrial applications where a 3-phase motor’s direction of rotation can be easily changed, pump rotation direction is not important. There is only one instance where a left-hand rotation pump must be specified. This is the case where a double-shafted electric motor drives a pump at both shafts. One of the pumps in this application must be setup for left-hand rotation.

Pump horsepower

Two formulas often used to figure hydraulic pump horsepower are:
hp = (psi)(gpm)/1714 -- (to calculate pure horsepower), and
hp = (psi)(gpm)/1714(actual pump efficiency).
Normally efficiency is assumed to be 85% because most new industrial pumps are at or above this figure.

The first formula is for a known pump volume -- figured from its cubic inches/revolution times the number of revolutions per minute. Say this displacement at 1200 rpm came to 12 gpm, but a flow meter at the pump outlet only shows 10.6 gpm at 1000 psi. The pump is still moving 12 gpm as far as its horsepower requirement is concerned, but the speed of the driven device will be that produced by 10.6 gpm. The 12-gpm pump was picked because the actual flow required was at least 10 gpm.

The second formula is applied when pump efficiency is known and horsepower is being figured for the actual 10-gpm requirement. Now pump efficiency must be considered because theoretical flow is greater than 10 gpm, and the electric motor must be able to pump the extra fluid even though it does not get to the actuator.

These two formulas can be simplified to:
hp = 0.000583 (gpm)(psi) -- (for pure horsepower), and
hp = 0.0007 (gpm)(psi) -- (for an 85% efficiency pump).
A common rule of thumb is: 1 gpm at 1500 psi = 1 hp.

Most suppliers’ catalogs show the horsepower required to drive a given pump at different pressures. These figures are usually conservative so designers can use them with confidence. Also, most electric motors can operate continuously at 110% of nameplate rating (and up to 140% for short bursts). Remember too that the only time a fixed-volume pump will be at full flow and full pressure is when the device it is driving has stalled. A pressure-compensated pump draws the highest horsepower just before it starts reducing flow slightly below its pressure setting. That event usually is not of long duration.

Many formula-data books have horsepower charts that make picking an electric motor simple. These charts are usually based on the 85% efficiency formula.

Cavitation

Next to contamination, cavitation causes more pump damage than anything else. Cavitation occurs when a pump needs 10.8 gpm at its inlet, but only gets 10.5 gpm. The missing 3 gpm winds up as voids or vacuum bubbles that implode when they go from suction to pressure. The implosions are rapid and damaging to adjacent surfaces when outlet pressure is high. They can take a pump out of service in hours. When outlet pressure is low -- under 200 psi -- there is still some noise and damage but it is minimal.

Some mobile equipment shuts the inlet to their pumps when the equipment travels, only allowing 1or 2 % of pump flow. This small volume goes through an open-center circuit at less that 15 psi, so implosions are not a problem. Another advantage is fuel savings. Because pump flow is so low, horsepower drain is much less.

Cavitation comes from several situations that are easy to rectify:

Long suction lines with many turns.
Undersize suction lines.
The pump mounted too far above the fluid.
Fluid viscosity too high (either wrong viscosity or low temperature).
A collapsed suction hose.
Turning the pump faster than the manufacturer recommends.
A clogged inlet strainer
A blocked air breather (especially in circuits with oversize rods or single-acting cylinders).

Any of the above could be eliminated immediately with a supercharging pump. This is a separate pump operating at low pressure (usually under 30 psi), that forces fluid to the system pump inlet.

Most of these conditions also can be eliminated by good design practices:

Locate the pump close to the tank -- preferably alongside or under it.
Never use a suction line smaller than the pump inlet port.
Use the fluid recommended by the pump supplier, and install tank heaters if the system will be exposed to temperatures below 65°F.
Never use pressure hose for suction lines. The lining of a pressure hose is not firmly attached to its body and can collapse under vacuum. Use hose specifically designed for suction service.
When a pump must turn faster than recommended, install a supercharging pump or elevate the tank to provide head pressure. Make sure a vacuum gauge at the pump inlet never goes above 1.5 to 3 psi.
Use a good filtering system – rated at least 10 µ -- so the suction strainer cannot block flow. Consider the suction strainer as insurance against startup contamination large enough to wreck a pump instantaneously.

Another situation that occurs in suction lines is air leaks. Air leaks are not cavitation but make the same noises and damage as vacuum cavitation. The only way to tell the difference in these situations is to look at the oil in the tank. If the oil is foamy from aeration, there is an air leak in the circuit. If the fluid is clear or almost clear of bubbles, there is a vacuum cavitation problem.

Air leak problems can come from poor piping practices. It is best to never use a standard pipe union in the inlet line. It is practically impossible to seal a standard union against an air inlet leak. Plumb the inlet line with as few fittings as possible and make sure any joints are sealed. If a plumbing connection is suspect, apply some of the system fluid to each joint to see if the noise stops. This type air leak problem usually shows up at system start. It seldom happens to a running circuit.

On systems that have been running for some time, a good place to look for air leaks is at the pump shaft seal. Fixed-volume pumps have their drive shaft sticking out of the housing and inside the housing is suction vacuum. When a shaft seal wears or is damaged from heat, it may let atmospheric air in before it lets oil leak out. The oil application test works here also, but can be messy because of shaft rotation speed.

The suction line is the most important line on the hydraulic circuit. Fluid can be pushed through pressure lines but a suction line only has one atmosphere (approximately 14.7 psi at sea level) with which to work. Most pumps are slightly damaged above 3 psi. At 4 psi and higher, cavitation noise is evident and pump damage escalates. Higher vacuum accelerates the damage.

What causes cavitation damage?

Erosion is the result of cavitation implosions as fluid passes from the inlet side of a pump to the outlet side. Figure 8-34 shows how the change from vacuum to pressure makes the vacuum or air voids collapse or implode. At low pressure, these voids merely close up and no damage is done. At high pressure, the fluid does not stop when the void is full but continues at high velocity through the void and impinges metal surfaces to the point of getting into the metal’s pores. Pressure drops as the next pulse of fluid approaches and high-pressure fluid in the metal pores rushes back out. During this part of the cycle, some very small particles of metal are dislodged and a cavity starts to form. Because this high- to low-pressure cycle can happen more than 200 times per second on a 12-vane pump at 1200 rpm, it is easy to understand how a pump can be physically damaged so quickly. These implosions are in the area where metal against metal is the only sealing action between vacuum and system pressure, so once the metal erodes, pump efficiency decreases because fluid bypasses through the damaged area.

Fig. 8-34. Representation of erosion from implosions impinging on metal caused by cavitation or air leaks

Pump-motor alignment

Most hydraulic pumps have light bearings while the electric motors by which they are driven have heavy-duty bearings. This makes it extremely important that the alignment of the pump and motor shaft be near perfect. Angular or offset misalignment always results in pump bearing failure, followed by internal failure soon after startup. Shaft couplings can take care of minor inconsistencies in shaft alignment, but they wear out very soon when not properly applied.

Figure 8-35 illustrates examples of misalignment. When the pump and motor are mounted separately, they must be aligned as nearly perfect as possible. Straight edges, dial indicators, and lasers give accuracy ranging from low tech to high tech, but they are only part of the answer. The pump and motor must sit on a rigid base and must be held down with ample force so they do not slip around during operation. The best alignment job possible can be rendered useless by inadequate mounting hardware.

Fig. 8-35. Correct way to belt-drive a pump

A simple way to overcome alignment problems is to use the pump-motor adapter shown in Figure 8-35. The pump-motor adapter is attached to a “C” face electric motor that has a flat-machined face and pilot protrusion. This face and pilot are perpendicular to the shaft and concentric to very close tolerances. A matching pilot and face are machined on the pump. The pump-motor adapter has matching machined faces and pilot bores. It is purchased for a particular motor and pump, so it is the right length for the shafts specified and matches the motor and pump mounting flanges. When this assembly is bolted together all parts align perfectly.

The shaft coupling then is slipped together and its setscrews tightened through the access port provided. The motor can be mounted on almost any surface without a chance of misalignment, and the pump can be changed without alignment problems anytime or place. Always use a coupling guard with an open coupling arrangement. Install the access-port cover before operating the pump when using the pump-motor adapter setup.

Figure 8-35 also shows the correct way to drive a pump with a belt. Light bearings on the pump cannot stand the side loads from belts so the pump fails very soon. Use pillow-block bearings to take the side load and couple the pump to the bearing guided shaft. This arrangement gives long service in applications where belts must be used.

Testing a pump

Figure 8-36 shows a typical setup for testing a pump that is suspect, has been out of service, or has been rebuilt. The flow meter could be an added device if the unit does not have one. It could be part of a test stand setup but is a necessary item when checking pump efficiency. The loading valve could be a ball valve as shown in the figure or another type valve as long as it can take the maximum pressure it will see. The relief valve must be in place and set for maximum rated pressure or operating pressure as needed. A pressure gauge is required to indicate system pressure. The filter should be part of a standard hydraulic power unit, but would usually be an off-line setup on a test stand.

Fig. 8-36. Test set-up for repaired pumps

To test a pump, lower the relief valve pressure setting to minimum. Then start the electric motor and check for flow on the flow meter. The meter should read at or very near catalog rating with all flow going directly to tank.

If the meter shows the pump producing ample flow, start closing the loading valve and watch the pressure gauge as it climbs. The reading should be low because the relief valve is set low. When the loading valve is closed completely, reset the relief valve to test pressure and observe the flow meter. Flow will drop somewhat, depending on the type of pump being tested. Most manufacturers publish rated flow at pressure in their literature. If the flow meter reads at or near cataloged rated flow, the pump is ready to put in service. If not, the pump should be checked or rebuilt to bring it up to specification.

Other pumps

Chapter 18 covers air- and hydraulic-driven intensifiers or boosters, which technically are pumps. These units usually are associated with air-oil systems. That is why their descriptions are in Chapter 18.

Air-to-hydraulic intensifiers are 100% efficient in the hydraulic end and are pressure compensated. They usually produce low volume so they are not normally used as a system’s prime mover. Their main advantage is they can hold pressure for long periods without generating heat or consuming energy. (Check out Chapter 17 to learn more about this unique pumping system.)

الصمامات الهيدروليكيه

2009-07-08T09:18:00.000-07:00

SLIP-IN CARTRIDGE VALVES

The term cartridge valves commonly refers to screw-in types of pressure, directional, and flow control valves. Screw-in type cartridge valves are mostly low-flow valves -- 20 gpm or less, although some manufacturers’ valves can handle more than 100 gpm. Screw-in cartridges are very compact, develop low-pressure drop, have little leakage, and produce inexpensive circuits that are reliable and easy to maintain. Screw-in cartridges are most often part of a drilled manifold but also are available in individual bodies. The function and performance of screw-in cartridge valves are the same as in-line or subplate-mounted valves.

Slip-in cartridge valves are different because -- except for pressure controls -- they are simply 2-way, bi-directional, pilot-to-close check valves. Most circuits using slip-in cartridge valves flow at least 60 gpm and can go as high as 3000 gpm. Slip-in cartridges are compact, develop low-pressure drop, and operate at pressures to 5000 psi. Slip-in cartridges can function as pressure, flow, and directional control valves.

Figure 4-1. 1:1 Poppet-type cartridge valve

Figure 4-1 shows a cutaway view and symbol of a 1:1 area ratio, poppet-type cartridge valve. Pressure relief, sequence, unloading, and counterbalance functions normally use a 1:1 area ratio poppet. The area ratio is the relation of the pilot area to the A port area. The 1:1 area valve stays closed when pilot pressure is equal to or greater than the A port pressure.

Figure 4-2 shows a cutaway view and symbol for a 1:1.1 area ratio valve. Here the pilot area is 1.1 times the A port area. Use this 1:1.1 ratio for special directional controls where system pressure at the pilot area must hold against excess pressure at the B port. Some pressure control applications also use this area ratio. Flow is possible from A to B, or B to A with low or no pilot pressure.

Figure 4-2. 1:1.1 Poppet-type cartridge valve

Figure 4-3 shows a cutaway and symbol for a 1:2 area ratio cartridge valve. Most directional-valve functions use this area ratio. Here, pilot area is twice the A or B port area. The 1:2 ratio valve allows flow from A to B or B to A with the same pressure drop. When the pilot area sees the same pressure as the A and/or B, all flow stops.

Figure 4-3. 1:2 Poppet-type cartridge valve

Slip-in cartridge pressure-relief valves

The schematic symbol and cutaway in Figure 4-4 are for a slip-in cartridge relief valve. The symbol for a cartridge is more pictorial than for spool valves, though the pressure-adjusting section uses a conventional ISO symbol.

Figure 4-4. Slip-in cartridge relief valve

Pressure relief cartridges can only flow from port A to port B. Port A is always connected to the pump while port B is always connected to tank. The spring that holds the poppet in place allows it to open at about 30 psi. This internal spring seats the poppet regardless of valve mounting position.

A slip-in cartridge valve has a cover that contains porting relative to the function the valve will perform and an adjustable spring-loaded poppet (the adjustable relief). This cover also holds the slip-in cartridge in place. The slip-in cartridge has a bushing with seals to prevent leakage to the outside or across the ports. This bushing fits in a machined cavity and contains the poppet that moves to allow fluid to pass. The poppet on a relief valve has a ratio of 1:1, which means the areas at the working fluid side, at the A port, and at the pilot side are equal.

Drilled pilot passages allow fluid to flow through control orifices to the pilot area of the poppet and to the adjustable relief in the cover. As system pressure increases, the poppet sees the same pressure on both sides and stays closed . . . held by the 30-psi spring. When system pressure reaches the relief setting, the adjustable relief opens a small amount, allowing pilot flow to tank. When pilot flow to tank is greater than control orifice flow from the A port, pressure on top of the poppet lowers. Then the poppet unseats to pass excess pump flow to tank.

Figure 4-5 shows the same cartridge relief valve with a single-solenoid directional valve -- or venting valve -- mounted on the cover. This solenoid-operated relief holds maximum pressure with the solenoid energized and unloads the pump to tank at approximately 30 psi when the solenoid is de-energized. Reversing the solenoid coil and spring keeps the pump loaded until the venting valve is energized.

Figure 4-5. Slip-in cartridge relief valve (solenoid-operated, normally vented)

Figure 4-6 shows the symbol for a dual-pressure relief valve with pump unloading. Pressures are set at the two manually adjustable relief covers and the solenoids select which relief to use. When both solenoids are de-energized, the pump unloads.

Figure 4-6. Slip-in cartridge relief valve (solenoid-operated, normally vented) for two different pressures

The symbol in Figure 4-7 is for an infinitely variable cartridge relief valve. A proportional solenoid valve is mounted on the cover of this 1:1 cartridge. The proportional solenoid valve controls vent flow, which in turn controls pressure. An electronic signal sets infinitely variable pressure to protect the system in varying conditions. The manually adjusted relief cover under the proportional solenoid sets maximum system pressure regardless of electrical input.

Figure 4-7. Slip-in cartridge relief valve. (Relief valve is proportional-solenoid operated.)

Figure 4-8 shows the symbol for a relief valve with a low-pressure unloading port. Set the relief cover for maximum pressure as before. Then, when it reaches maximum pressure, the relief cartridge opens to unload the pump at approximately 30 psi. Venting pressure comes from piping the unloading port downstream of a check valve that holds fluid in the accumulator. Until there is about a 15% pressure drop in the accumulator holding circuit, the pump will stay unloaded. When pressure drops about 15%, the relief cartridge closes until system pressure reaches maximum setting again.

Slip-in cartridge pressure-reducing valves

Figure 4-8. Slip-in cartridge -- pressure relief and unloading (infinitely variable pressure)

The schematic symbol and cutaway in Figure 4-9 are for a cartridge pressure-reducing valve. The ISO symbols for the cartridge and the pressure-reducing section are conventional. Pressure-reducing cartridges only flow from port B to port A. Port B always sees inlet or system pressure, while port A is the reduced-pressure outlet. If reverse flow is necessary, add a bypass check valve to allow return flow around the reducing valve. The spring directly holding the spool in place keeps it open regardless of valve mounting position when pressure is below the adjustable relief setting.

Figure 4-9. Slip-in cartridge reducing valve

The slip-in cartridge reducing valve has a cover that contains porting relative to the function to be performed. An adjustable spring-loaded poppet (the adjustable relief) in the cover sets outlet pressure. This cover also holds the slip-in cartridge in place. The cartridge has a bushing with seals to prevent leakage to the outside or across the ports. This bushing fits in a machined cavity and contains the spool that closes as pressure increases. The spool on a reducing valve has a ratio of 1:1 -- which means that the A port area and pilot area are equal.

A drilled pilot passage allows fluid to flow through a pressure-compensated control orifice to the adjustable relief in the cover, as well as to the top of the spool. As pressure builds, the spool stays open because of the spring and the equal pressures on equal areas, thus letting flow continue through the valve. When the A port reaches the reduced pressure setting, the adjustable relief opens and pilot fluid flows to tank through the drain. When pilot flow is greater than control orifice flow, lower pressure on top of the spool allows it to rise, blocking flow from the B port to the A port. Pressure at the A port will not exceed that set on the adjustable relief unless a load-induced pressure tries to force flow back through the closed spool. There will be pilot flow out the drain port whenever the reducing valve is at reduced pressure. Blocking or closing the drain port causes the spool to fully open and allow outlet pressure to reach system pressure.

A pressure-reducing valve will not allow reverse flow after it has reached its set pressure. For example, if the reduced pressure is 500 psi at a cylinder and some outside force starts pushing against the cylinder, there is no place for most of the fluid to go. About 50 to 100 in.3/min of excess fluid passes through the pilot circuit and out the drain port while the valve is reducing. Fluid in excess of drain flow becomes trapped and pressure builds, possibly to dangerous levels. If there is a chance of outside forces that can increase outlet pressure, add a relief valve bypass at the outlet. A bypass relief valve relieves trapped fluid before excessive pressure can damage the valve or machine.

Figure 4-10 shows a cartridge reducing valve with dual-pressure capabilities. A solenoid-operated selector valve and a second adjustable relief mounted on the cover give the option of two pressures. Always use the first adjustable relief above the spool for maximum pressure setting. A single-solenoid directional valve (as shown) allows default to maximum pressure. Using a 2-position detented directional valve maintains the last pressure selected.

Figure 4-10. Slip-in cartridge reducing valve

Figure 4-11 shows a proportional solenoid valve mounted on the adjustable relief. Such a valve allows selection of infinitely variable pressures via an electrical command. Allowing pilot flow to bypass the adjustable relief gives a reduced pressure of anything lower than the adjustable relief setting. An electronic signal to the proportional solenoid varies pilot flow that controls pressure on top of the spool.

Figure 4-11. Slip-in cartridge reducing valve (proportional operated, Infinitely variable pressure)

Slip-in cartridge directional control valves
Slip-in cartridge check valves

The simplest directional control valve is a check valve. Figure 4-12 shows the symbol and cutaway for a cartridge check valve. A check valve has a cover with a control orifice to control pilot fluid. The control orifice dampens the poppet movement. It is available in several diameters. The cover also holds the cartridge in place and seals it with an O-ring. The cartridge has a bushing with seals to prevent leakage to the outside or across the ports. A machined cavity holds the bushing that contains the poppet that will open when fluid flows in the right direction. The poppet on a check valve has a 1:2 ratio, which means the area at the two working ports (A or B) is one half of the pilot area. A 1:2 ratio poppet allows flow in either direction as long as pilot pressure is off or slightly less than half the working pressure.

Figure 4-12. Slip-in cartridge check valve

There are several spring forces available -- from as low as 5 psi to more than 70 psi. The lowest spring pressure possible is best for normal check valve operation.

In Figure 4-12, a drilled pilot passage senses the pressure at the A port. Flow from the B port to the A port passes with a slight pressure drop caused by the volume of flow plus the spring force. When flow tries to reverse (from the A port to the B port) as pressure on the A port half area increases, it goes through the pilot passage to the main pilot area. Because the A port area is only half the pilot area, the poppet stays closed and blocks reverse flow.

Figure 4-13 shows the same valve with the pilot passage drilled to the B port. With this valve, flow is free to go from A to B, but not from B to A.

Figure 4-13. Slip-in cartridge check valve

The symbol and cutaway in Figure 4-14 are for a cartridge pilot-operated check valve. The cartridge is the same as a standard check valve, but with a different cover. The cutaway shows the works of the cartridge pilot-operated check valve cover. On the left of the cover, a pilot piston pushes a simple ball check from the left seat to the right seat. The ball check stays to the left -- its normal position -- held by a light spring.

Figure 4-14. Slip-in cartridge pilot-operated check valve

If oil tries to pass from the B port to the A port, the same pressure that is trying to open the check on the half area also is applied to the pilot area, keeping the poppet closed.

Flow from the B port to the A port requires a pilot pressure equal to at least 30% of pressure at the B port to shift the pilot piston. When there is sufficient pressure on the pilot piston, it will move the ball check off the left seat, opening a path to the drain. At the same time, closed flow at the right seat blocks flow from the B port. With little or no pressure at the pilot area, the 1:2 poppet opens, allowing flow from the B port to the A port. If pilot pressure drops while oil is reverse flowing, the poppet shuts due to pressure on the pilot area like any check valve. Various sizes of dampening control orifices control shifting speed of the poppet to help reduce system shock.

Slip-in cartridge directional control valves

The symbol and cutaway in Figure 4-15 are for a simple 2-way cartridge valve. Most cartridge directional valves have a 1:2 pilot ratio, although a 1:1.1 ratio works better in certain circuits. In either case, pilot pressure equal to the working pressure at port A and/or port B closes the poppet.

Figure 4-15. Slip-in cartridge directional valve with plain cover

From the cutaway view in Figure 4-15, it is plain to see that fluid pressure at port A will push the poppet off its seat and allow flow to port B. Although it is less obvious, fluid pressure at port B will also open the poppet and allow flow to port A. To stop flow in either direction, apply pilot pressure to the pilot area opposite port A. If any pilot pressure generates a closing force equal to the opening force at the A and/or the B ports, the spring bias closes the poppet.

Although slip-in cartridge directional valves appear to be normally closed, they open easily without pilot pressure. A vertically mounted cylinder controlled by slip-in cartridge valves can free-fall when the pump stops and pilot pressure drops. This problem is easy to fix, as will be shown in some later circuits.

The cutaway view in Figure 4-15 shows a plain cover with a pilot passage and a control orifice. (Pilot pressure in this type of valve would come from another solenoid valve or control valve in the circuit.) Control orifices come in a variety of sizes to provide smooth, non-shock movement of the poppet. To control shock even more, add a skirt with V notches to the poppet. Figure 4-23 shows the symbol and cutaway for a cartridge poppet with a V-notched skirt for flow control or dampening function. Different manufacturers have other ways to achieve this dampening effect.

Figure 4-16. Slip-in cartridge directional valve with single-solenoid operator

Figure 4-16 shows the symbol for a 1:2 slip-in cartridge with an interface for a solenoid-operated directional valve on the cover. This solenoid valve directly operates the cartridge beneath it. It also can pilot other cartridge valves through drilled passages in the manifold. The single solenoid pilot valve can keep the poppet normally closed or normally open. Figure 4-16 shows a normally closed configuration.

Figure 4-17 shows a double-solenoid, detented pilot operator. The cartridge poppet stays in its last position even with both solenoids de-energized. With this type of solenoid operator there is no need to maintain current on the solenoid after the valve shifts.

Figure 4-17. Slip-in cartridge directional valve with double-solenoid. detented operator

Figure 4-18 shows a double-solenoid, spring-centered pilot operator. The center condition of the pilot operator allows the cartridges it pilots to open when both solenoids are de-energized. This could allow a cylinder to relax in case of power failure or when activating the emergency stop.

Figure 4-18. Slip-in cartridge directional valve with double solenoid operator N.O.

Conversely, Figure 4-19 has a double-solenoid pilot operator that closes all cartridges when both solenoids are de-energized. The actuator would stop suddenly and be locked in place. This type of pilot operator could cause system shock without some means of decelerating the cylinder.

Figure 4-19. Slip-in cartridge directional valve with double solenoid operator N.C.

Slip-in cartridge valves with 1:2 area ratios appear to be normally closed because of the spring inside the poppet. However, one half the pilot area is connected to port A or port B, and pressure at these ports can open the poppet. The only way to keep a slip-in cartridge valve closed is to keep pilot pressure on the pilot area at all times.

Anytime the pump is running, there should be enough pilot pressure to keep a poppet closed. However, when the pump stops or pilot pressure drops for any reason, the poppet may open, allowing an actuator to move. This might cause a safety hazard or machine damage.

The symbol and cutaway in Figure 4-20 are for a cover with an integral shuttle valve. A shuttle valve will take signals from two sources and send the higher pressure signal to the pilot area. At the same time, the shuttle valve will not let either signal pass through to the other signal passage.

Figure 4-20. Slip-in cartridge directional valve with shuttle-valve cover

The cutaway of the shuttle operator shows that a pilot signal to pilot passage 1 or pilot passage 2 goes to the pilot area, but not out the line with little or no pilot signal. This happens because the shuttle poppet closes the inactive or low-pressure opening and only allows pilot oil from the active or higher-pressure side to flow to the pilot area. Because the area of the shuttle ball is equal to that of both pilot passage ports, the strongest signal always goes to the pilot area. In most applications, this is an important feature.

The vertically mounted, rod-down cylinder shown in Figure 4-21 is holding a heavy weight. This is an example of an over-running load. With standard externally piloted slip-in cartridges, the weight will fall when the pump stops or anytime pilot pressure drops below approximately 275 psi. This is because the 23,000-lb weight, acting on the 40.06 square inches of rod end area, produces a static pressure of 574 psi (23,000/40.06 = 574 psi). This 574 psi would act against half the area of the poppets to push them open. It takes approximately 275 psi on the pilot area plus the spring force to hold the poppets shut. For safety’s sake, change this circuit to one with shuttle-valve covers.

Figure 4-21. Plain-cover, solenoid-operated and plain-cover, remotely piloted slip-in cartridge directional valves -- dropping a load when the pump is off

The circuit in Figure 4-22 is the same as above except for a shuttle valve in the cover. One pilot supply is from the pump, while the second pilot supply is from the cylinder’s rod end. While the pump is on and the system is at pressure, pilot supply is from the pump. In case of low or no system pressure, pilot oil comes from the cylinder’s rod end. With the cylinder’s rod end as the pilot source, pressure that is trying to open the poppets on the half area of port A also acts on the pilot areas. Because the pilot areas are twice the A port area, the poppets stay closed. The shuttle valve cover assures there is always pilot pressure on the pilot area when the cylinder is not fully extended.

Figure 4-22. Shuttle-cover, solenoid-operated and plain-cover, remotely piloted slip-in cartridge directional valves -- holding a load when the pump is off

Slip-in cartridge directional valves (continued)

The symbol and cutaway for a slip-in cartridge valve in Figure 4-22 include a stroke-adjusting screw that limits poppet travel. Restricting flow by limiting poppet movement controls the actuator’s maximum speed. The filled triangle in the poppet symbol shows the skirted or modified poppet that allows smooth flow change as it shifts.

The cutaway in Figure 4-23 shows one design of a slip-in cartridge with a stroke limiter. The cartridge function is identical to any 1:2-ratio poppet except for the limited movement. Restricting the poppet movement makes the cartridge function as a flow control as well as a directional valve.

Figure 4-23. Slip-in cartridge directional valve with plain cover, stroke limiter, and dampening function

Figure 4-24 shows the symbol for an adjustable-stroke cartridge valve with a directional control valve cover. This particular valve only comes in a single-solenoid configuration as shown. Also, it cannot pilot other cartridge valves in the manifold. (Figures 4-27 and 4-28 show an adjustable stroke-cartridge in a circuit. These examples also show a problem that can occur when using a stroke limiter as a flow control in a meter-out circuit.)

Figure 4-24. Slip-in cartridge directional valve with single-solenoid operator (N.C.)

Figure 4-25 shows the symbol and cutaway for an internal-poppet, orifice-type cartridge valve. The internal poppet orifice supplies pilot oil from the A port only. Standard orifices that meet most needs are available. The internal pilot supply cartridge valve provides a check-valve function without drilling pilot passages in the manifold. As a check valve, it always allows free flow from the B port to the A port and blocks flow from the A port to the B port.

Figure 4-25. Slip-in cartridge directional valve with orifice port in poppet for pilot supply

The 2-way cartridge shut-off valve in Figure 4-26 is for high flow systems. This 2-way shut-off might allow pump flow to a circuit as shown in the schematic. Also use a 2-way shut-off to let fluid flow from a large cylinder to tank for rapid advance. Using a normally open solenoid valve in place of the normally closed one shown allows flow through the cartridge valve until the solenoid is energized.

Figure 4-26. Slip-in cartridge directional valve with NC orifice port in poppet for pilot supply

One advantage of the internal-pilot-supply-type cartridge valve is that it is not necessary to keep the pump running to have pilot pressure. This can eliminate a shuttle valve when an over-running load tries to move the cylinder.

The internally piloted slip-in cartridge always controls flow from the A port to the B port. Fluid is free to flow from the B port to the A port because pilot supply comes only from the A port.

When using a stroke-adjusted poppet to meter-out flow from a cylinder with an oversize piston rod, look out for the problem that appears in Figure 4-27. This circuit pictures a horizontal cylinder with a 2:1 rod that needs a meter-out flow control. This is good circuit design for spool-type valves, but when using an adjustable-stroke slip-in cartridge valve, it can cause trouble. This circuit can actually increase the cylinder speed when making an adjustment to slow it.

Figure 4-27. Slip-in cartridge valves with adjustable-stroke poppet at CV1 (extending with poppet on CV1 full open)

The circuit in Figure 4-27 shows the valves shifted to extend the cylinder. Flow from the pump is passing through CV3 to the cylinder’s cap end. Oil from the cylinder’s rod end is flowing to tank freely through CV1 because the stroke adjuster is fully open. Pressure gauge PG1 shows a system pressure of 700 psi. Gauge PG2 in the cylinder’s cap line reads 700 psi, and PG3 at the cylinder’s rod line reads 0 psi. The 700-psi reading is from the load’s resistance (the cylinder is moving with no flow restriction). Pilot pressure is always the same as system pressure. Flow to the cylinder’s cap end is 50 gpm and flow from the rod end to tank is 25 gpm.

With the stroke limiter screwed in to restrict tank flow to 12.5 gpm, the conditions shown in Figure 4-28 will prevail. In a normal meter-out circuit with a flow control and a spool-type directional valve, the cylinder speed slows and system pressure increases.

Figure 4-28. Slip-in cartridge valves with adjustable stroke poppet at CV1 (extending with poppet on CV1 set at 12.5 gpm)

With a cartridge-valve circuit, however, restricting flow from the cylinder’s rod end increases system pressure. Gauges PG1 and PG2 register approximately 1600 psi -- or a little more than twice the non-restricted flow pressure. This is because the load now is being moved by pressure on half the piston area in a regeneration circuit. Gauge PG3, at the cylinder’s rod end, climbs to approximately 1800 psi due to area-ratio intensification.

This intensified pressure acts on the half A port area at CV2, while half the B port area sees system pressure. Pilot pressure on the full pilot area of CV2 is 1600 psi, plus a spring force of, say, 75 psi. If the full pilot area is one square inch, the poppet has a closing force of 1675 lb. The 800-lb opening force on the poppet is generated by 1600 psi on half the area. The opening force on the other half area of the poppet is 900 pounds (1800 psi X 1/2 sq. in.), making the total opening force 1700 lb. With 1700-lb opening force and 1675-lb closing force, the poppet opens to allow rod-end oil to regenerate to the cap end. Instead of the cylinder slowing to half speed, it moves 150% faster due to regeneration. The more the flow from CV1 to tank decreases, the faster the cylinder extends. Note that restricting flow at CV3 as a meter-in flow-control circuit would allow infinite control of cylinder speed. Another option would be to use a shuttle cover and take pilot pressure from the pump or the cylinder’s rod end. As pressure intensified at the rod end, pilot pressure to CV2 would increase also.

Slip-in cartridge directional valves compared to spool-type 4-way directional valves

Figure 4-29 pictures a circuit with a 300-gpm pump powering a large-bore, 2:1 rod-diameter cylinder. This is the type of circuit that uses an important feature of slip-in cartridge valves. Flow from the rod end is only 150 gpm as the cylinder extends, but while retracting, flow from the cap end is 600 gpm. A conventional 4-way valve to operate the cylinder in Figure 4-32 must be capable of 600-gpm flow. A 4-way valve with this capacity is large and expensive. Its delivery may involve a long lead time.

Figure 4-29. Slip-in cartridge valves for high-flow circuits (at rest, pump running)

Four slip-in cartridge valves can duplicate the function of the 600 gpm 4-way valve. This may sound expensive and inefficient, but with a circuit such as the one in Figure 4-29, it actually is more efficient, less expensive, and saves space.

Figure 4-30. Slip-in cartridge valves for high-flow circuits (cylinder extending)

The cylinder is extending in Figure 4-30, with 300 gpm going to the cap end through CV3. Simultaneously, the rod end of the cylinder is discharging 150 gpm to tank through CV1. With this difference in flow, it costs less and saves space to use cartridges of different sizes. Size the cartridges for nominal pressure drop at their maximum flow.

Figure 4-31. Slip-in cartridge valves for high-flow circuits (cylinder retracting)

The cylinder is retracting in Figure 4-31. Flow to the rod end is 300 gpm while flow from the cap end is 600 gpm. For this higher cap-end flow, use a larger cartridge to minimize backpressure. This circuit will have three different-sized cartridges to carry the flow required during each phase of the cycle. Size CV1 for 150 gpm, CV4 for 600 gpm, and CV2 and CV3 for 300 gpm. When using a regeneration circuit, size CV3 for 600-gpm flow also.

The small amount of space taken by the cartridges, plus the lower cost and better availability of the parts make this system superior to one with a spool-type 4-way valve for high flows.

Another advantage of slip-in-type cartridge valves is their short response time. Cartridge poppets do not have land overlaps like spool valves have. Without land overlap there is flow when pilot pressure drops. Also, when the poppet opens, it only moves far enough to allow system flow to pass. When applying pilot pressure again, the poppet closes quickly without the extra travel often seen in spool valves. A spool valve, without stroke limiters, shifts full stroke. This full shifting may be far enough to pass several times the flow required. Then, when the spool starts returning to center, there is extra spool travel just to get back to controlling flow. This does not sound like much but faster response of cartridge valves can shorten cycle time and increase production.

Figure 4-32. Conventional 4-way directional valve for high-flow circuit (at rest, pump running)

The circuit in Figures 4-29 to 4-32 uses only one pilot control valve. This limits the versatility of the slip-in cartridges. Multiple pilot control valves, shown in the following circuits, make the use of cartridge valves even more attractive.

Slip-in cartridge directional valves on running-away loads

Figures 4-33 to 4-36 show a vertically mounted (rod down) cylinder holding a heavy platen and tooling. This cylinder will run away if oil discharges to tank uncontrolled.

Figure 4-33. Slip-in cartridge valve circuit for running-away load (at rest, pump running)

A pressure-control cover (that makes CV6 a counterbalance valve) prevents rapid flow to tank. Cartridges CV1 through CV4 control cylinder flow and direction, and solenoid-operated directional valves shift their positions. Cartridge check valve CV5 bypasses normally closed counterbalance valve CV6 to retract the cylinder. With all solenoids deenergized, the pump unloads to tank through cartridge valves CV3 and CV4. Counterbalance valve CV6 keeps the load from falling at this time.

Figure 4-34. Slip-in cartridge valve circuit for running-away load (cylinder extending, regeneration)

Figure 4-34 shows valve positions and likely pressures as the cylinder is regenerating forward. Energizing solenoids A1 and B2 closes CV4 and opens CV3, porting pump flow to the cylinder. This action also opens CV2 to allow cylinder’s rod end flow to go through CV3, combine with pump flow, and regenerate the cylinder forward. Counterbalance valve CV6 keeps the forward motion of the cylinder from going faster than the pump and regeneration volume as indicated by the 750 psi seen on gauge PG3. The cylinder is extending rapidly at low or no force.

Figure 4-35. Slip-in cartridge valve circuit for running-away load (cylinder extending at full tonnage)

As the cylinder extends, it makes a limit switch to deenergize solenoid B2. When solenoid B2 drops out, Figure 4-35, pilot pressure closes CV2, while CV1 opens to tank. The cylinder slows to about half the regeneration speed, but is now able to generate full force. Counterbalance CV6 still keeps the cylinder from free-falling as it approaches the work. When the cylinder starts to form a part, pressure increases to whatever it takes to do the work. The cylinder continues extending until it finishes the work stroke.

Figure 4-36. Slip-in cartridge valve circuit for running-away load (cylinder retracting)

To retract the cylinder, the valve conditions shown in Figure 4-36 prevail. Energizing solenoids B1 and B2 allows CV2 and CV4 to open, and closes CV1 and CV3. Pump flow now goes to the cylinder’s rod end through cartridge check valve CV5. That bypasses normally closed counterbalance valve CV6. Oil from the cylinder’s cap end goes to tank through cartridge valve CV4.

Anytime all solenoids are deenergized, the cylinder stops and holds position. Counterbalance valve CV6 holds the cylinder in place as long as its pressure setting is greater than the load-induced pressure in the cylinder’s rod end.

When sizing the counterbalance valve, be sure to consider the cylinder’s static pressure. Slip-in cartridge valves have high flow capacity at nominal pressure drops. When available pressure drop is high, flow can increase to a point that the counterbalance valve’s response is too slow to stop the cylinder quickly.

Slip-in cartridge directional valves with prefill valves

Figures 4-37 to 4-40 shows cartridge valves controlling a 50-in. cylinder with a 48.75-in. piston rod.

In Figure 4-37 the circuit is at rest with all solenoid valves deenergized. The cylinder maintains its position because load-induced pressure on the CV1 and CV2 pilot areas holds them closed. Pilot pressure reaches the pilot valves through CK1, while CK2 blocks flow to tank. When system pressure is higher than pressure in the cylinder’s rod end, CK2 lets this higher pressure into the pilot circuit. The pump unloads to tank through CV3 and CV4.

Figure 4-37. Slip-in cartridge valve for vertically mounted cylinder with prefill valve (at rest, pump running)

The size of CV1 is important, because controlling flow through it sets the cylinder free-fall speed. Size the valve for the pressure drop generated by load-induced pressure. A stroke limiter in CV1 actually sets maximum cylinder extension speed.

Energizing solenoid A1 shifts CV4 closed to block tank flow and leaves CV3 open to send pump flow to the cylinder’s cap end. Figure 4-38 shows the cylinder in a controlled free fall. Energizing solenoid B2 lets CV1 open, while holding CV2 closed. Oil from the cylinder’s rod end now has a path to tank through CV1.

A prefill valve lets oil from the tank into the cylinder’s cap end. The cylinder will advance as fast as the stroke limiter on CV1 allows. Free-fall speed can be in excess of 15 in./sec.

Figure 4-38. Slip-in cartridge valve for vertically mounted cylinder with prefill valve (cylinder extending, controlled fall)

As the cylinder extends in free fall, it contacts a limit switch that de-energizes solenoid B2, Figure 4-39. CV2 remains closed and CV1 tries to close. As CV1 is closing, backpressure on the cylinder’s rod end will build to 900 psi and the pressure control will keep CV1 from fully closing. Because CV1 is restricting flow at 900 psi, the cylinder decelerates and tries to stop. While the cylinder is slowing, decreased vacuum in the cap end lets the pre-fill valve close. After the prefill closes, pump flow forces the cylinder to keep moving and rod-end pressure keeps the pressure control on CV1 open. Deceleration is smooth and rapid. The cylinder continues extending toward the work at the slower pump rate.

Figure 4-39. Slip-in cartridge valve for vertically mounted cylinder with prefill valve (cylinder approaching work, decelerating)

Figure 4-40 shows the cylinder at work. Solenoid A1 is still energized and solenoid B2 has been energized again. (Solenoid B2 could be reenergized by a limit switch or by a pressure switch when the cylinder contacts the work.) Energizing solenoid B2 lets CV1 open fully, taking away the 900-psi backpressure that decreases tonnage. The cylinder extends at the force required to do the work (up to relief pressure setting).

Figure 4-40. Slip-in cartridge valve for vertically mounted cylinder with prefill valve (cylinder pressing)

To retract the cylinder, energize solenoids A2 and B1. Solenoid B2 closes CV3, pilots the pre-fill valve open, and opens CV4 to tank. Solenoid A2 closes CV1 and opens CV2, sending pump flow to the cylinder’s rod end. The cylinder retracts rapidly with most of the cap-end flow going to tank through the pre-fill valve. In case of power failure or emergency stop, the cylinder stays where it is, or if it is moving, it decelerates, stops, and holds its position.

Pressure-control valves

Several types of pressure-control valves are found in fluid power circuits. Some keep the whole system from excess pressure while others only protect a portion of the system. Others allow flow to an isolated circuit after reaching a preset pressure. Some bypass fluid at low or no pressure when activated.

This chapter only covers relief valves and unloading valves because they are closely associated with hydraulic pumps. The other pressure-control valves are part of the control circuit and will be dealt with after directional control valves.

Why relief valves?

All fixed-volume pump circuits require a relief valve to protect the system from excess pressure. Fixed-volume pumps must move fluid when they turn. When a pump is unloading through an open-center circuit or actuators are in motion, fluid movement is not a problem. It is when the actuators stall with the directional valve still shifted that a relief valve is essential.

Pressure compensated pump circuits could run successfully without relief valves because they only move fluid when pressure drops below their compensator setting. (Most designers still use a relief valve in these circuits for reasons explained later.)

In either case, a relief valve is similar to a fuse in an electrical system. When circuit amperage stays below the fuse amperage, all is well. When circuit amperage tries to exceed fuse amperage, the fuse blows and disables the circuit. Both devices protect the system from excess pressure by keeping it below a preset level.

The difference is that when an electrical fuse blows it must be reset or replaced by maintenance personnel before the machine can cycle again. This requirement alerts the electricians to a possible problem and usually causes them to look for the reason before restarting the machine. Without the protection of a fuse, the electrical circuit would finally overheat and start a fire.

In a hydraulic circuit, a relief valve opens and bypasses fluid when pressure exceeds its setting. The valve then closes again when pressure falls. This means a relief valve can bypass fluid anytime . . . or all the time . . . without intervention by maintenance. (It also means the system can run hot even with a heat exchanger installed.)

Many fixed-volume pump circuits depend on this bypassing capability during the cycle, and some even bypass fluid during idle time. A well-designed circuit never bypasses fluid unless there is a malfunction, such as a limit switch not closing or an operator overriding the controls. This eliminates most overheating problems and saves energy.

Relief valve operation

There are two different designs of relief valves in use: direct acting and pilot operated. Both types have advantages and work better in certain applications.
Some terms relating to relief valves and their function are:

Overshoot: The actual pressure reading when a relief valve first opens to bypass fluid. (It can be up to twice the actual pressure setting.)
Hysteresis: The difference in pressure between when a relief valve starts letting some flow pass (cracking pressure) and when full flow is passing.
Stability: The fluctuation of pressure as a relief valve is bypassing at set pressure.
Reseat pressure: The pressure at which a relief valve closes after it has been bypassing.
Pressure override: The difference in the pressure reading from the time a relief valve first opens (cracking pressure) until it is passing all pump flow to tank.

Direct-acting relief valves

Figure 9-1 shows a cutaway view and the symbol for a direct-acting relief valve. The valve has a poppet that is pressed against its seat by an adjustable spring. An adjusting knob can be change the force on the spring to raise or lower maximum pressure. The poppet remains seated while pump flow goes to the circuit and pressure is lower than the relief valve setting. If pressure tries to go above spring setting, the poppet is forced off the seat just enough to pass excess pump flow to tank.

Fig. 9-1. Cutaway drawing and symbol for direct-acting relief valve.

The symbol shows a single box with a flow arrow offset from the inlet P and outlet T flow lines. The dashed pilot line from the inlet line to the bottom of the box indicates inlet pressure can push against the flow arrow. On the opposite side of the box is a spring with a sloping arrow through it to show an opposing force on the flow arrow. When pressure at port P builds enough to overcome spring pressure, it forces the flow arrow up until there is a path from P to T. Although there is no pilot passage in the actual valve, the function is implied and thus is part of the symbol.

The main advantage of direct-acting relief valves over pilot operated relief valves is that they respond very rapidly to pressure buildup. Any relief valve does not know there is a problem until pressure is very near or at its setting. Then it must open to relieve excess flow as quickly as possible to keep pressure overshoot low. Because there is only one moving part in a direct-acting relief valve, it can open rapidly, thus minimizing pressure spikes. Figure 9-2 shows typical performance graphs from direct-acting and pilot-operated relief valves. Notice the difference in response time and pressure spikes as the valves open to send excess flow to tank.

Fig. 9-2. Typical performance plots for direct-acting and pilot-operated relief valves

The main disadvantage of direct-acting relief valves is that they open partially at about 150 psi below set pressure. Because the poppet is in direct contact with the spring that sets maximum pressure, when the poppet opens it forces the spring back and increases pressure. The amount depends on the spring’s length and stiffness. The plot in Figure 9-3 shows the flow/pressure relationship of a typical direct-acting relief valve. With a direct-acting relief valve setting of 1500 psi at 10 gpm, it is very possible that some fluid will start to pass when pressure is as low as 1350 to 1400 psi. Continued pressure increase allows more flow until all pump flow goes to tank at 1500 psi. If work is still being performed at 1450 psi, it will be at a reduced speed because some flow is going to tank. When this valve is set at 1500-psi cracking pressure, no flow will bypass until pressure reaches that level, but final pressure would be as high as 1650 psi. (Pilot-operated relief valves . . . discussed next . . . do not start to open until pressure is within 25 to 50 psi of their settings.)

Fig. 9-3. Plot of flow-pressure relationship of a typical direct acting relief valve.

Direct-acting relief valves often are quite noisy due to the high velocity of the fluid bypass and the instability inherent in their design.

Direct-acting relief valves are not normally used on industrial hydraulic systems, except for those with flows under 3 gpm, and as pilot control devices. Most industrial designs use long springs that gain little force per compression increment to keep pressure override low.

When a direct-acting relief valve is specified as preset, non-adjustable, always specify whether the valve is to be set for cracking pressure or full flow. If full flow is desired, a flow must be specified also.

Pilot-operated relief valves

Figure 9-4 shows cutaway views of two common types of pilot-operated relief valves. There are many variations of these designs but the function and symbol are the same. The pilot section on each valve is a low-flow direct-acting relief valve that sets maximum system pressure. Because the valve is small and passes very little flow, it has less than 50-psi pressure override as it operates.

Fig. 9-4. Cutaway view and symbol for two common types of pilot-operated relief valves.

The control orifice in the balanced piston or poppet usually has a diameter around 0.040 in. This size gives good relief-flow stability and is not prone to becoming blocked with contamination. If the orifice is plugged, the balanced piston or poppet will open at approximately 20 psi and dump all pump flow to tank.

A flow path from the outlet of the control orifice . . . on top of the balanced piston or poppet . . . leads up to the pilot section, which contains a spring-loaded poppet. Adjusting the tension on the spring-loaded poppet sets the pressure in the circuit. Fluid used by the pilot section returns to tank through the tank port. The balanced-piston type has a hole through it that lets control fluid flow to tank. The vent port in the pilot section is normally plugged. (Removing the plug allows this valve to perform other functions.)

Many inline-mounted valves have two inlet ports as a piping convenience. Pump flow comes in one inlet and exits through the opposite one. This eliminates the need for a tee in the pump line plumbing.

How a pilot-operated relief valve works

Pump flow enters the inlet port and flows to the circuit and through the control orifice to the top side of the balanced piston or poppet. It also travels up to the pilot section’s spring-loaded poppet, where it is blocked. When pressure is too low to unseat the spring-loaded poppet, pressure is the same on either side of the balanced piston or poppet. Because hydraulic forces are equal on both sides of the balanced piston or poppet, the light spring holds them in their normally closed position. This condition continues until pressure reaches approximately 25 to 50 psi below the pressure set at the relief valve pressure-adjusting knob.

For example, if pressure was set at 1000 psi, at around 950 psi the spring-loaded poppet in the pilot section will crack open and allow a small amount of fluid to pass to tank. At this point the amount of fluid passing the spring-loaded poppet can easily flow through the control orifice so pump flow to tank is blocked. As pressure continues to increase, it finally forces the spring-loaded poppet in the pilot section to open far enough so that flow through it is greater than flow through the control orifice. When flow through the spring-loaded poppet is more than flow through the control orifice, pressure on top of the balanced piston or poppet decreases. When the pressure imbalance is great enough, the balanced piston or poppet moves toward the decreased pressure and opens a flow path to tank. Flow to tank is just enough to bypass any excess fluid the system is not using. As a relief function, this valve never opens more than enough to bypass excess flow.

When system pressure decreases, the spring-loaded poppet in the pilot section reseats. Fluid trapped on top of the balanced piston or poppet forces it to close and block pump flow to tank.

A pilot-operated relief valve allows all pump flow to go to the actuators almost to its final setting. This means the valve can operate at a lower maximum pressure and it will not slow actuator speed when forces increase.

Remote pilot operation

Another capability of pilot-operated relief valves is that they can be operated remotely. Figure 9-5 shows the vent port connected to a direct-acting relief valve at a remote location for easy pressure adjustment. Because a relief valve is normally mounted at or very near the pump outlet, it can be difficult to reach. When it is necessary to change pressures on a regular basis, the setup in Figure 9-5 works well. The vent port of the pilot-operated relief valve is connected to a direct-acting relief valve at a distance of 15 ft maximum. The pilot-operated relief valve is set for maximum pressure and the remote adjustment can set at any pressure lower than this maximum.

Fig. 9-5. Pilot-operated relief valve connected for remote control.

Using a 4-way directional control valve and three remote adjustments could allow electrical selection of three different pressures. Using more directional controls and more remote adjustments could give multiple pressure selections electrically.

Solenoid-operated relief valves

Figure 9-6 shows how a directional control valve attached to the pilot section and piped to the vent port and tank can bypass or block flow from the control orifice. Bypassing the control-orifice fluid allows pump flow to unload to tank at about 20 psi. Blocking control-orifice flow forces fluid to the circuit at pressures up to relief valve setting. This is one way to keep a fixed-volume pump from overheating the fluid when it is not performing work. (See Chapter 8, Figure 8-11 for a circuit that uses a normally open solenoid-operated relief valve to unload a fixed-volume pump in a multiple cylinder circuit.)

Fig. 9-6. Normally open solenoid-operated relief valve.

Solenoid-operated relief valves can be purchased in normally open mode (as shown), normally closed mode, and double-solenoid dual- or tri-pressure setups. (See Chapter 4 for symbols.) A solenoid-operated relief valve also can be used as a 2-way normally open or normally closed directional valve in high-flow circuits.

Proportional-solenoid relief valves

The relief valves in Figure 9-7 are electronically adjusted by using a proportional solenoid instead of an adjusting knob. A proportional solenoid produces increased force with increased voltage. These solenoids usually operate at 0 to.10 V on DC current. They can produce infinitely variable force. The direct-acting type is for low (below 3 gpm) flow. It also can serve in the pilot section of high-flow pilot-operated valves. Operation of a proportional relief valve is the same as for manually controlled valves. The difference is how the force on the control poppet is generated.

Fig. 9-7. Relief valves operated by proportional solenoid.

Unloading valves

Unloading valves are pressure-control devices that are used to dump excess fluid to tank at little or no pressure. A common application is in hi-lo pump circuits where two pumps move an actuator at high speed and low pressure, the circuit then shifts to a single pump providing high pressure to perform work.

Another application is sending excess flow from the cap end of an oversize-rod cylinder to tank as the cylinder retracts. This makes it possible to use a smaller, less-expensive directional control valve, while keeping pressure drop low.

Direct-acting unloading valves

The cutaway view in Figure 9-8 shows the construction of a direct-acting unloading valve. The valve consists of a spool held in the closed position by a spring. The spool blocks flow from the inlet to the tank port under normal conditions. When high-pressure fluid from the pump enters at the external-pilot port, it exerts force against the pilot piston. (The small-diameter pilot piston allows the use of a long, low-force spring.) When system pressure increases to the spring setting, fluid bypasses to tank (as a relief valve would function). When pressure goes above the spring setting, the spool opens fully to dump excess fluid to tank at little or no pressure. (The example circuit in Figure 9-10 illustrates this function.)

Fig. 9-8. Direct-acting unloading valve

Pilot-operated unloading valve

The cutaway view in Figure 9-9 shows a pilot-operated unloading valve. A pilot-operated unloading valve has less pressure override than its direct-acting counterpart, so it will not dump part of the flow prematurely. It also will go from no flow to maximum flow quickly, thus using all the flow from the high-volume pump flow for a longer period, and rapidly dropping horsepower draw from the high-volume pump.

Fig. 9-9. Pilot-operated unloading valve.

(This valve design is also used as an unloading relief valve in accumulator circuits. Chapter 16 on Accumulators will have a circuit using this valve.)

A pilot-operated unloading relief valve is the same as a pilot-operated relief valve with the addition of an unloading spool. Without the unloading spool, this valve would function just like any pilot-operated relief valve. Pressure buildup in the pilot section would open some flow to tank and unbalance the poppet, allowing it to open and relieve excess pump flow.

In a pilot-operated unloading valve, the unloading spool receives a signal through the remote-pilot port when pressure in the working circuit goes above its setting. At the same time, pressure on the spring-loaded ball in the pilot section starts to open it. Pressure drop on the front side of the unloading spool lowers back force and pilot pressure from the high-pressure circuit forces the spring-loaded ball completely off its seat. Now there is more flow going to tank than the control orifice can keep up with. The main poppet opens at approximately 20 psi. Now, all high-volume pump flow can go to tank at little or no pressure drop and all horsepower can go to the low volume pump to do the work. When pressure falls approximately 15% below the pressure set in the pilot section, the spring-loaded ball closes and pushes the unloading spool back for the next cycle.

An unloading valve requires no electric signals. This eliminates the need for extra persons when troubleshooting. These valves are very reliable and seldom require maintenance, adjustment or replacement.

Hi-lo pump circuit

Often a cylinder needs very little force to stroke to and from the work -- and only a short high-force stroke to perform the work. When this is the case, the hi-lo circuit in Figure 9-10 works well and costs less.

Fig. 9-10. Typical hi-lo circuit using two pumps.

For example: if a single-pump circuit needs 60 gpm to make the required cycle time and 3000 psi to perform the operation, the circuit would require a 110-hp electric motor to drive it. (60 X 3000 X 0.000583 = 104 hp)

The circuit in Figure 9-10 is a typical hi-lo pump circuit that consumes less horsepower while maintaining fast cycle times. It uses a 25-hp motor and supporting equipment for less expense up front, as well as during its useful life. The motor drives a 50-gpm low-pressure pump and a 15-gpm high-pressure pump -- for a total of 65 gpm. The extra flow is required to maintain cycle time because the work stroke is slower. The tank, valves, and line sizes are still rated for 65-gpm flow and 3000 psi, but the electric motor and controls are much smaller.

As shown in Figure 9-10, the hi-lo circuit also has a relief valve, an unloading valve, and a check valve. The relief valve protects the low-volume/high-pressure pump from pressure above 3000 psi. The unloading valve is set at 500 psi to divert flow from the high-volume/low-pressure pump to tank when system pressure climbs above this setting. A check valve after the high-volume/low-pressure pump isolates system pressure from the unloading valve circuit while performing work at maximum pressure.

A 4-way, 3-position, solenoid pilot-operated, spring-centered, all-ports-open directional control valve sends all pump flow to tank while the system is idle. This power unit and valve arrangement send a double-acting cylinder through a fast-approach, high-force work stroke and fast return – driven by a 25-hp electric motor. The unloading valve cutaway view shows the pipe connections to this in-line mounted valve.

Energizing solenoid A1 on the directional valve sends flow from both pumps to the cap end of the double-acting cylinder. The cylinder advances rapidly at low pressure until it contacts work. At this point, contact pressure builds quickly and when it passes 500 psi, the unloading valve is forced open. Now, all high-volume pump flow is diverted to tank at very low pressure (and horsepower). Up to this point, the highest horsepower draw would be: (65 gpm)(500 psi)(0.000583) = 19 hp.

With the high-volume pump unloaded, there is plenty of horsepower to raise the high-pressure pump to the 3000-psi pressure required to do the work. The work requires (15 gpm)(3000 psi)(0.000583) = 26 hp. This is well within the capability of the 25-hp motor specified.

A hi-lo circuit makes it possible to replace a high-horsepower motor and its control components with a much smaller less-expensive setup.

Other applications for relief valves

Relief valves are used in circuits to protect components from excess pressure due to heat or external forces where pressure buildup in a blocked flow circuit could damage an actuator or be a safety hazard.

In hydraulic motor circuits, relief valves can eliminate shock when the motor must be decelerated quickly. In this function, fluid is ported from the high-pressure outlet port of the motor to the low-pressure inlet port, while holding ample backpressure to stop the motor without damage.

Fig. 9-11. Symbols for modular relief valves. (Note that these symbols do not show X and Y ports for solenoid pilot-operated valves.)

Most relief valve functions are available as modular or sandwich valves that mount between the directional control valve and sub-plate. Figure 9-11 shows most of the common configurations presently offered by fluid power suppliers. These modules are usually available in all valve sizes up to D08 (3/4 in.) ports.

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2009-07-08T09:13:00.000-07:00

Accumulators used for fast response and over-pressure control of pressure-compensated pumps

Because most pressure-compensated pump circuits have closed-center or two-position directional valves (such as the one shown in Figure 1-16), they stay at full-pressure, no-flow until a valve shifts. After any directional valve shifts to start an actuator’s movement, pressure in the circuit starts to drop. When the pump sees a pressure drop, its internal mechanism starts shifting as fast as possible to start fluid flowing. Pump shifting times vary, but no matter how fast they shift, the actuator’s initial response will be slowed down.

Fig-1-16

With an accumulator installed, as shown in Figure 1-17, the pump is still at no-flow when the circuit is at rest. However, there is a ready supply of oil at pressure available. As a cylinder starts to cycle, as seen in Figure 1-18, fluid flows directly to the actuator from the accumulator and pressure starts to drop. This pressure drop causes the pump to go on stroke, but now pressure drop is minimal. The cylinder takes off quickly and smoothly, and the pump has time to respond to the flow need.

Fig-1-17

On the other end of the cycle, if the pump is at full flow and all valves center or all the actuators hit the end of stroke, the flow requirement suddenly drops to zero. The pressure-compensated pump is still flowing at the maximum rate and pressure starts to climb. The pump will continue at full flow until pressure reaches 80-98% of the compensator setting. There has been zero flow needed for some time, but the pump does not know this until pressure is near maximum. When pressure reaches compensator setting, the pump starts to shift to no flow. All pump flow during shifting time has no place to go, so this excess flow generates a pressure spike of five to ten times the compensator setting. This pressure spike can cause premature failure of the pump, plumbing, and actuators.

Fig-1-18

A common fix for this pressure spike is to add a relief valve near the pump outlet, set 150 to 200 psi higher than the pump compensator (as shown in Figure 1-16). This relief valve should reduce the pressure spike, but it does not lower it as much as it appears. A relief valve remains closed until pressure reaches 90 to 98% of its setting. Once the relief reaches maximum pressure, it starts to open, but by the time it actually relieves, the pressure may be 11/2 to 3 times its set pressure. This reduced spike is better, but still is not as good as what an accumulator could provide.

Other problems can occur with relief valves. For example, if the relief valve setting is at or near the pump-compensator setting, the pump can start oscillating on-off flow. As the pump nears its pressure-compensator setting and starts to compensate, the relief valve starts to relieve. A flow path is created when the relief valve begins to open, so downstream pressure drops, causing the pump to go back on stroke. The drop in pressure allows the relief valve to close, so downstream pressure builds up again. This oscillation cycle repeats rapidly, causing damage to the pump and possible line failure due to shock. In another example, if the relief valve setting is lower than the pump compensator, all pump flow goes to tank at relief pressure, generating excess heat. To avoid these problems, use the correct procedure when setting pressures on a relief valve used to reduce pressure spikes.

An accumulator absorbs excess pump flow with minimal pressure override or shock. While fluid from the pump compensates from full flow to no flow, as seen in Figure 1-19, it has a direct path to the accumulator. Because the accumulator has a compressible gas in it, it takes in the small amount of excess flow produced while the compensator is reacting. Pressure increase from this additional fluid is imperceptible.

Fig-1-19

To size an accumulator for fast response of the circuit, plan to have somewhere between 1 and 5 sec of actuator flow before pressure drops below the minimum it takes to move it. A rule of thumb is to have 1 gal of accumulator for every 10 gpm of pump flow.

Using an accumulator as an emergency power supply

A conventional hydraulic system will not operate unless the pump is running. Some machines must be able to cycle to a safe condition after a power or pump failure. Use an accumulator to store enough energy to move the actuators to a safe condition after the pump quits. The operator or setup person can manually cycle the machine into a safe condition by using the stored energy.

The hopper gate cylinder shown in Figure 1-20 must close in case of a power failure. If the gate stays open, the entire hopper could overflow the truck under it, then dump on the ground. This circuit uses a pressure-compensated pump that maintains pressure with minimal heating during normal operation. An accumulator F stores the first pump flow, check valve D stops accumulator back flow, and normally open directional valves C isolate the accumulator from the cylinder and tank during normal operation.

Fig-1-20

The gate cylinder needs at least 1500 psi, so the pump compensator is set for 2000 psi. This ensures that the accumulator has enough fluid to extend the cylinder when necessary. Because the solenoids on valves C are energized by the pump start command, the accumulator is completely isolated from the cylinder and tank as long as the pump runs. When solenoid B of the 4-way directional valve shifts (as seen in Figure 1-21), the gate opens as fast as the pump moves it.

Fig-1-21

When solenoid A shifts the 4-way directional valve, as seen in Figure 1-22, the gate closes as fast as the pump moves it. When the power is on, the cylinder extends or retracts partially or all the way at the operator’s command.

Fig-1-22

If the gate cylinder is partially or completely open and power fails, the circuit automatically goes to the condition shown in Figure 1-23. In this condition the pump stops, the 4-way directional valve centers, and the normally open 2-way shutoff valves C open.

Fig-1-23

When power fails, the accumulator has a direct path to the cap end of the cylinder while rod-end oil flows to tank. The cylinder will extend and close the gate using the stored energy in the accumulator. Place warning signs at the gate indicating this equipment can operate at any time without operator intervention.

When using an accumulator for emergency power supply it is difficult to automatically drain it during normal operation. Automatically draining the accumulator would defeat its purpose as an emergency power supply. Add a manual drain valve E, with warning signs to tell maintenance persons to manually drain the accumulator before working on the gate circuit.

Size emergency-power accumulators to hold enough oil to move all actuators to the home position before pressure drops to dangerous levels. Most manufacturers provide formulas in their catalogs and offer several offer excellent computer programs to size accumulators for emergency-power supplies.

Using accumulators for leakage makeup

Some hydraulic circuits, such as in laminating presses, need to hold at pressure for long periods. A pressure-compensated pump could maintain pressure, but energy loss from pump leakage generates heat. Another way to hold pressure for long periods is with a fixed-volume pump and an accumulator. Figure 1-24 shows a press cylinder that must stay extended under pressure for several minutes.

Fig-1-24

Tee small accumulator D into the cylinder cap-end line through flow control C. Flow control C allows the accumulator to fill quickly but discharge slowly when directional valve A centers or shifts to retract the cylinder. Flow control C should pass enough flow to let the accumulator discharge quickly without system shock when directional valve A shifts to retract the cylinder. Any oil left in the accumulator when the directional valve centers will make the cylinder extend a small amount. Tee dump valve B into the cylinder cap-end line to automatically discharge the accumulator when the pump stops. Tee pressure switch E into the cap-end cylinder line to set pump load and unload pressure. Pressure switch E sets high and low pressures to control maximum and minimum tonnage.

Fig-1-25

When the pump starts, Figure 1-25, backpressure check valve F gives 75 psi pressure, closing accumulator dump valve B and supplying pilot oil for solenoid pilot-operated directional valve A. When directional valve A shifts, the cylinder starts to extend, Figure 1-26, at whatever pressure it takes to overcome the counterbalance valve. The signal to the extend coil of directional valveA goes through the normally closed contacts on pressure switch E. Because gas pre-charge pressure in the accumulator is approximately 85% of working pressure, no fluid will enter it yet.

Fig-1-26

When the cylinder contacts the work, Figure 1-27, pressure increases and oil fills the accumulator. Upon reaching the maximum working pressure set by pressure switch E, the normally closed contacts open, de-energizing the solenoid on directional valve A. Directional valve A spring centers, the pump unloads, and oil stored in the accumulator maintains pressure while making up for cylinder and valve leakage.

Fig-1-27

Bypass at the cylinder seals and/or valve causes pressure to drop slowly to the low-pressure setting of pressure switchE. This low-pressure setting is normally adjustable but must be high enough to keep the parts firmly together. Upon reaching the low-pressure setting, pressure switch E shifts, allowing the normally closed contacts to shift directional valve A to refill the accumulator. Upon reaching maximum working pressure, directional valve A again spring centers to unload the pump, while the accumulator holds its pressing force and makes up for leaks.

A pilot-operated check valve in the cap-end cylinder line between the directional valve and the pressure switch would have less leakage than the blocked port of the spool valve. With a pilot-operated check valve and resilient seals in the cylinder, it is possible to maintain pressure for 2 to 5 min or more. Use an all-ports-open directional valve with the pilot-operated check valve. This accumulator circuit maintains pressure in the cylinder while unloading the pump. It also conserves energy while using an inexpensive fixed-volume pump.

Using accumulators as shock absorbers

Accumulators can reduce damage from shock in some circuits if correctly applied. In other applications, an accumulator may add shock by releasing stored energy too quickly.

The top half of Figure 1-28 illustrates one way shock is produced. Flow velocity in a hydraulic circuit may be 25 to 30 fps and not cause any problems. However, if oil flow stops abruptly, as seen in Figure 1-28’s middle example, damaging shock can rip out tubing, blow seals, and split pump housings with ease. A column of moving fluid has a lot of energy that can get out of control.

Fig-1-28

The third example in Figure 1-28 shows a small accumulator teed into the line at the shut off that stops flow suddenly. An accumulator spreads the shock energy over a short period of time and eliminates the potential for damage.

To absorb flow shock, the accumulator is usually pre-charged at about 70 to 80% of system pressure. At this pre-charge pressure, only a small amount of fluid enters the accumulator subsequent to a shock situation. There is also little fluid transfer to take away from or add to the normal pump flow.

When it is necessary to stop a heavy load, such as shown in Figure 1-29, try using an accumulator and a hydraulic cylinder. The accumulator’s pre-charge pressure holds the cylinder extended, thus making it ready for the advancing mass. When the load contacts the cylinder, it mechanically forces it to retract. As the cylinder retracts, fluid flows into the accumulator and gas pressure increases. As pressure increases, the higher resistance slows the mass more. After the load decelerates, the cylinder might try to push the part away. Add valves between the accumulator and the cylinder to control the shock absorber after it finishes decelerating the load.

Fig-1-29

Some large, slow-turning piston pumps send a shock wave into the circuit every time a piston discharges oil. On the left of Figure 1-30, the piston pump does not have an accumulator at the discharge port. Pressure at the gauge will fluctuate from less than system pressure to well above it without an accumulator.

Fig-1-30

On the right side of Figure 1-30, adding a small accumulator reduces discharge flow and shock damage. A portion of the sudden discharge flow from an advancing piston goes into the accumulator and discharges smoothly while waiting for the next stroke. The pre-charge pressure for this type of accumulator circuit is about 60 to 75% of maximum system pressure.

Accumulator manufacturers have formulas in their brochures to calculate any situation mentioned here. Some suppliers have computer programs that do all the math after asking for circuit parameters.

Pump supplementing circuit with full pressure when work is contacted

In some cases, a pump-supplementing accumulator circuit can speed up cylinder extension and/or retraction without having to go above working pressure. Normally in a pump-supplementing circuit, the relief valve is set as high as possible above the working pressure to store ample fluid. As the cycle progresses, oil from the accumulator and pump move the actuator quickly, but circuit pressure drops steadily. If pressure drops below the actuator’s need, the pump must refill the accumulator before the cycle finishes. To overcome this problem, a larger pump and/or more accumulators are necessary.

The next circuit shows an accumulator arrangement that provides high volume to move the cylinder rapidly with the relief valve set at working pressure. The accumulator and pump supply volume to fill the large bore cylinder as it extends. The cylinder then moves to working pressure while a check valve isolates the accumulator.

Like all accumulator circuits, there must be time for refilling between cycles, as shown in Figure 1-31. Pre-charge the accumulator to a pressure slightly higher than it takes to retract the cylinder. The cylinder will then retract when directional valve A and normally open, solenoid-operated relief valve H shift. (Also see Figure 1-34.) The large piston rod reduces the return volume, although retract pressure will be higher. When the cylinder fully retracts, pressure climbs and the accumulator starts to fill through check valve E and the bypass check valve around flow control C. Piston-type accumulators are best for this circuit because they can have a low pre-charge pressure and a high final pressure without internal damage. The accumulator can discharge a large volume of oil because the pressure in it is not important when the cylinder needs full tonnage.

Fig-1-31

When pressure in the circuit reaches 2000 psi, pressure switch G de-energizes the solenoid on normally open, solenoid-operated relief valve H, unloading the pump to tank.

When directional valve A and normally open, solenoid-operated relief valve H shift, Figure 1-32, pump flow and accumulator flow provide a large volume of oil to quickly stroke the cylinder to the work. Because accumulators can discharge at a very high rate, use flow control C to set the desired advance speed. Pressure in the circuit will fall as the cylinder extends and will be well below working pressure before the cylinder meets the work.

Fig-1-32

When the cylinder contacts the work, Figure 1-33, check valve F keeps pump flow from going to the accumulator. The pump will continue filling the cylinder and pressure will build to whatever it takes to do the work. Check valve F blocks flow to the accumulator to isolate it during the high-pressure work stroke.

Fig-1-33

When directional valve A shifts to the retract position, Figure 1-34, pump flow goes to the cylinder rod end. The accumulator pre-charge is high enough to force all pump flow to the cylinder, causing it to quickly retract.

Fig-1-34

Figure 1-31 shows the cylinder reaching the top of the stroke. The accumulator now accepts all pump flow through check valve E until pressure switch G unloads the pump.

Two other pump-supplementing circuits with full pressure when work is contacted

Figures 1-35 and 1-36 depict two more ways to use an accumulator for volume and still have immediate high pressure for doing work. Either circuit works equally well with the two pump types shown.

Fig-1-35

These circuits would normally require a piston-type accumulator. Notice the pre-charge is less than one-third of maximum pressure. The large pressure difference would squeeze the bladder in a bladder-type accumulator so much that holes caused by chafing would allow the nitrogen gas to leak. The minimum pressure in the circuit could be even lower than shown here. If the actuators can move at 300 psi, then use 150 to 200 psi pre-charge.

Fig-1-36

The circuit in Figure 1-35 uses a pressure-compensated pump and a normally open, poppet-type, 2-way directional valve. All flow goes directly to the accumulator, filling it to maximum pressure with the pump operating. When the cylinders start to cycle, flow from the pump and accumulator move them rapidly. When the cylinders contact the work, pressure is well below the required amount. To get full force, energize solenoid C1. This stops pump flow to the accumulator and raises the cylinders to full pressure. De-energize solenoid C1 when the cylinders finish their work to allow the accumulator to refill.

Energizing solenoid C1 when the actuators are moving is possible with a correctly designed poppet valve. Notice the blocked position of the valve has a check valve symbol, meaning it only stops flow to the accumulator. This type of poppet valve provides accumulator volume to the actuators when pressure is low. However, maximum pressure is immediately available when the cylinders meet resistance. De-energize solenoid C1 at the end of the cycle to refill the accumulator. Some poppet-type directional valves have a very high pressure drop when flowing through the closed check valve. Use a brand designed for low pressure drop in this circuit.

The circuit in Figure 1-36 has a fixed-volume pump with a normally open,, solenoid-operated relief valve and pressure switch to unload the pump at maximum pressure. Minimum system pressure for this circuit is 1500 psi. Therefore, it is important to set the sequence valve in front of the accumulator to this pressure. Set the pressure switch to unload the pump at 1700 psi. Then set the normally open, solenoid-operated relief valve at approximately 1900 psi. Because no oil can go to the accumulator if the system pressure is below 1500 psi, the actuators will always have maximum force anytime they meet resistance. When the cylinders are moving to and from the work, pump and accumulator flow can combine to give rapid movement at reduced pressure. Flow from the accumulator can always go to the cylinders through the bypass check valve. Fluid only goes to the accumulator when pump flow is greater than the system requires. This circuit fills the accumulator anytime the cylinders stop or anytime required volume is less than pump output.

There will be some heating of the oil while the accumulator is filling until system pressure reaches 1500 psi or above. One advantage is that no control circuitry is necessary, even while the accumulator fills anytime actuator volume is less than pump flow.

Non-invasive way to check accumulator pre-charge

It is important to check accumulator pre-charge pressure at regular intervals. Check a new installation each shift for a few days to see if there is a gas-pressure loss. It the gas charge is holding, check pre-charge pressure weekly for the next month. If all is well at the end of a month, then monthly checks should be more than satisfactory.

The normal way to check pre-charge pressure is: (1). Shut down the system. (2). Attach a gauge and charging kit to the accumulator. (3). Open the gas valve and check the pressure reading.

However, this procedure is time consuming, allows some gas to discharge, and may damage the charging valve, which can result in a continuous leak. Outlined below is a simple, non-invasive way to check accumulator pre-charge pressure to see if gas is leaking.

Fig-1-37

Figure 1-37 shows a partial accumulator circuit. This figure shows an operating hydraulic system, just as the pump stops. At this point, the accumulator relief/unload/dump valve is open, draining pressurized oil stored in the accumulator. As fluid in the accumulator discharges, pressure at gauge PG1 starts dropping. By controlling the flow with a fixed orifice or a flow control, pressure deteriorates slowly when there is oil in the accumulator.

Fig-1-38

When all fluid is out of the accumulator, Figure 1-38, pressure at gauge PG1 will suddenly drop to zero. Carefully note gauge pressure when it suddenly drops. The pressure seen at the sudden drop is the present pre-charge pressure of the accumulator. This reading is only as accurate as the gauge and the person reading it. It is not a perfect reading, but will be close enough to see if a full-fledged check is needed.

Fig-1-39

If there is more than one accumulator on the machine, as in Figures 1-39 and 1-40, this test will show the lowest pre-charge pressure. When a low pre-charge pressure shows up, check each accumulator individually until finding those at a lower pressure than required.

Fig-1-40

Another way to check pre-charge pressure is to note the gauge’s pressure reading when turning on the pump. With an accumulator in the circuit, the first pressure reading should be pre-charge pressure. It is difficult to obtain an accurate reading this way with glycerin-filled or orifice-dampened gauges in the circuit. The gauge should also be at or close to the accumulator to keep line losses from adding to the reading.

Hydraulic type accumulator dump valves

When using an accumulator, there must be a way to discharge stored oil before safely working on the circuit. Even when using the accumulator for emergency power supply, install a manual drain valve for safe operation.

A manual drain valve with a gauge near it is the best way to ensure a safe operation. Mark the manual drain valve and place warning signs at all hydraulic component locations indicating there is an accumulator in the circuit and to open the manual drain before performing maintenance.

A common way to discharge stored energy is to use a normally open, solenoid-operated, 2-way directional valve teed into the pressure line with its outlet hooked to tank. Wire the solenoid on the 2-way valve to close when the pump is running. Any time the pump stops, the 2-way solenoid valve de-energizes and discharges stored oil to tank.

A solenoid-operated valve works well in most cases but can cause problems. First, if the valve fails to close or only partially closes, oil dumps across it, generating heat and making it operate sluggishly or not at all. Second, if the valve fails to open when the pump stops, the circuit is unsafe. This is a safety hazard for an inexperienced person who might not detect the problem. Third, additional wiring creates additional costs.

If the circuit uses a fixed-volume pump as shown in Figures 1-41 through 1-44, use an accumulator relief/unload/dump valve for most applications. This valve has an integral adjustable 2-way unloading valve A to unload the pump when reaching set pressure. Also, there is a pilot valve to close shut-off B that stays closed while the pump is running and opens any time the pump stops. Isolation check valve (C) keeps accumulator oil from back flowing to the pump when it stops.

Fig-1-41

In Figure 1-41 the pump has just started, so pressure jumps to accumulator pre-charge pressure and all flow goes to the accumulator through check valve C. Pilot-operated 2-way shut-off B pilots closed when the pump is running. The pilot-operated, adjustable-spring shut-off Astays closed until set pressure is reached.

Pressure continues to climb until the accumulator is full, as seen in Figure 1-42. When pressure reaches that set on 2-way adjustable-spring valve A, it opens, unloading the pump to tank at low pressure. Even while unloading there is enough pressure to keep pilot-operated 2-way shut-off B closed.

Fig-1-42

When pressure in the circuit drops approximately 15%, Figure 1-43, unload-valve A closes, again forcing oil to the circuit and accumulator. The pump will load and fill the system any time pressure drops about 15%. This pump load pressure is non-adjustable so it will not work for all circuits.

Some manufacturers offer an accumulator relief/unload/dump valve with an adjustable differential setting. Setting these valves’ load-unload pressure by more or less than the 15% differential is possible.

Fig-1-43

When the pump shuts off, as in Figure 1-44, pilot pressure to 2-way valve B drops, allowing it to open. Now all stored fluid from the accumulator has a path directly to tank. The accumulator will quickly discharge, making it safe to work on the circuit.

Fig-1-44

CAUTION! ALWAYS CHECK AN ACCUMULATOR CIRCUIT FOR PRESSURE BEFORE WORKING ON IT. NEVER ASSUME THE AUTOMATIC UNLOADING SYSTEM WORKED!

Hydraulic-type accumulator dump valves (continued)

When using an accumulator with a pressure compensated pump, the packaged dump valve shown works well. (See Figures 1-45 through 1-48.)

A pressure-compensated pump maintains pressure while flow changes to meet the needs of the circuit. When the first actuator in the system starts to move, there is no flow for it until pressure drops. As pressure drops, a pressure-compensated pump will go on stroke quickly but there will be a slight pause before flow actually starts. The addition of the small accumulator shown in Figure 1-45 nearly eliminates the startup pause. This enhances system response while reducing cycle time and pressure fluctuations.

Fig-1-45

On the other end of the cycle, if the pump is at full flow and all the valves center or all the actuators hit end-of-stroke, flow requirement suddenly goes to zero. The pressure compensated pump is still flowing at maximum and pressure starts to climb. The pump will continue at full flow until pressure reaches 80 to 98% of the compensator setting. There has been zero flow needed for some time, but the pump does not know this until pressure is near maximum. When pressure reaches the compensator setting, the pump starts to shift to no-flow. All pump flow during the shifting time has no place to go, so this excess flow makes a pressure spike of five to ten times the compensator setting. This pressure spike can cause premature failure of the pump, plumbing, and actuators. An accumulator as shown will take in this small volume of oil to minimize the spike.

As with any accumulator installation, safety is important. When shutting a circuit down for maintenance, always drain the accumulators. A manual drain valve works, but the automatic drain shown on the facing page is better. When the pump starts -- and as long as it is running -- a pilot valve closes check valve B to block the drain port. Check valve A isolates the pump from accumulator back flow when it stops or fails. There is no electrical wiring needed, so the accumulator dump valve is invisible to the control circuitry.

The pump is just starting in Figure 1-45, so pressure immediately climbs to accumulator pre-charge pressure. Flow continues until the accumulator is full and system pressure is at its maximum. Pilot-to-close check valve B blocks the drain path to tank when the pump starts. The drain path stays closed as long as the pump is running.

Fig-1-46

Figure 1-46 shows flow while the circuit is working. Accumulator and/or pump flow will go to the actuators to quickly start them and move them through their cycle. During the working part of the cycle, the accumulator smooths out flow fluctuations, while reducing pressure drops and spikes.

With the system at rest as shown in Figure 1-47, pump flow is zero and the accumulator is full and ready for another cycle.

Fig-1-47

Figure 1-48 shows how the circuit responds when the pump stops. Check valve A closes to stop back flow and pump motoring. Pressure to pilot-to-close check B drops out, allowing it to open. All accumulator volume now has a path to tank through an orifice that keeps flow at a reasonable rate. In a very short time the accumulator’s stored energy dissipates, making it safe to work on the system.

Fig-1-48

CAUTION! ALWAYS CHECK AN ACCUMULATOR CIRCUIT FOR PRESSURE BEFORE WORKING ON IT. NEVER ASSUME THE AUTOMATIC UNLOADING SYSTEM WORKED!

Linear pressure-type accumulators

The following circuits use accumulator types with little or no pressure drop as they discharge fluid.

Gas- or spring-loaded accumulators lose pressure as fluid discharges and the gas or spring expands. In a typical circuit using this type of accumulator, the maximum system pressure must be higher than working pressure to allow for this pressure drop. Some circuits cannot operate at these elevated pressures or may need high pressure for the entire stroke. Therefore, they can’t use gas or spring loaded accumulators.

The circuit in Figure 1-49 shows a weight-loaded accumulator, a fixed-volume pump, and a normally open, solenoid-operated relief valve that can replace either circuit shown in Figures 1-10 and 1-11. Notice the maximum pressure and working pressure are at 2000 psi. This is possible because the weight-loaded accumulator does not lose pressure as fluid discharges. Until the accumulator piston reaches bottom, system pressure stays constant.

Fig-1-49

With a weight-loaded accumulator, the amount of weight on a given piston area sets maximum pressure. To raise or lower maximum pressure, weight must be added or taken off. Set the relief valve on this type circuit 100 to 150 psi higher than system pressure so it does not bypass during normal operation.

The main disadvantage of a weight-loaded accumulator is its physical size. An accumulator for the circuit shown in Figure 1-49 would require a 10-in. ram with a 60-in. stroke for the cylinder to have full force for its entire cycle. This size accumulator needs almost 160,000 lb of weight on the ram to get the required volume and pressure stated. A block of concrete approximately 1080 ft3 in size or about 10 X 10 X 11 ft would be necessary to meet this need. Such high mass eliminates the use of this type accumulator for mobile equipment and also rules out many industrial applications. Using a smaller accumulator ram with a longer stroke reduces weight, but you must make sure column strength is adequate when reducing ram diameter.

The air-cylinder-loaded accumulator shown in Figure 1-50 works the same as a weight-loaded accumulator. There is a slight pressure drop as fluid starts to flow due to piston and ram seal friction but this is usually not enough to cause problems.

Fig-1-50

Physical size can also be a problem with air-cylinder-loaded accumulators, especially when using low air pressure. Most plant systems operate at 100 to 125 psi so the unit required to handle the cylinder in Figure 1-50 might be a 40-in. bore air cylinder driving a 9-in. ram with a 75-in. stroke. Using air pressure at 250 psi could reduce the accumulator to a 30-in. air cylinder driving a 10_-in. ram for a 55-in. stroke. In either case, these accumulators are still too large for mobile equipment and for many industrial applications.

Air-cylinder-loaded accumulators work best and are more economical to operate using a surge tank for the air cylinder. Surge tanks provide fast flow for discharging high oil volumes with minimal pressure drop. They also make it possible to use a small air compressor because it only has to make up for leaks after the system gets up to pressure. Size the surge tank to allow for a 3- to 8-psi drop when the accumulator discharges during a normal cycle.

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2009-07-08T09:06:00.000-07:00

Hydraulic accumulators make it possible to store useable volumes of non-compressible fluid under pressure. A 5-gal container completely full of oil at 2000 psi will only discharge a few cubic inches of fluid before pressure drops to 0 psi. The same container filled with half oil and half nitrogen gas would discharge over 11/2 gal of fluid before pressure dropped to 1000 psi.

Figures 1-1 through 1-4 show symbols used for different types of accumulators. Figures 1-5 through 1-8 are simplified cutaways showing construction of different types of accumulators.

Fig-1-1 to 1-8

All accumulators except Figure 1-4 will have a pressure decrease as fluid discharges. A weight-loaded accumulator maintains pressure until all oil is used.

When using an accumulator, it is necessary to install a manual or automatic function to de-pressurize all fluid before working on the circuit. Several manufacturers make automatic discharge valves that work well. These automatic discharge valves are explained at the end of this section.

Most hydraulic accumulators are used in one of four applications: 1. Supplement pump flow in circuits with medium to long delays between cycles. 2. Hold pressure in a cylinder while the pump is unloading or stopped. 3. Have a ready supply of pressurized fluid in case of power failure. 4. Reduce shock in high velocity flow lines or at the outlet of pulsating piston pumps.

The following circuit images show some circuits using accumulators for the operations mentioned in 1 to 4 above. Other accumulator circuits and information follow.

Using accumulators to supplement pump flow

Some hydraulic circuits require a large volume of oil for a short time; for example to move a large cylinder rapidly to clamp a part. After clamping, the circuit needs little or no additional fluid for period of time while curing takes place. When a circuit has extended dwell time, an accumulator can be used to downsize the pump, motor, tank, and relief valve. The cost of accumulators usually offsets savings on these smaller components, but downsizing saves on operating costs.

The conventional pump, directional valve, and cylinder pictured in Figure 1-9 show horsepower and flow requirements needed for a 12.5 sec cycle time. The advance cycle requires full power, while returning the cylinder needs minimal force. Reduction of the pump and motor size is not possible if the cylinder cycles rapidly. However, if there was a 45 sec wait between cycles, the pump and motor could be almost 70% smaller with an accumulator circuit.

Fig-1-9

This reduced flow and horsepower are possible when using accumulators and the circuit shown in Figure 1-10. The extra expense of the accumulators offsets the reduced price for the power unit, but operating cost is less for the life of the machine. The directional valve and piping from the accumulators to the cylinder still has to handle the 125 gpm flow.

Fig-1-10

Using a gas charged accumulator in a pump supplementing circuit will increase maximum system pressure. The extend portion of the cycle needs at least 2000 psi working pressure, which requires filling the accumulators with fluid above 2000 psi so they can discharge oil and not drop below minimum pressure. The maximum system pressure should be as high as can be tolerated. The higher the maximum allowable system pressure, the smaller the accumulators. The drawback of high pressure is that the circuit is at this pressure when the cycle starts. If this higher pressure can cause damage or other problems, it should be lowered to a safe level.

Accumulator circuits normally have flow controls because there is a volume of oil at elevated pressure that can discharge almost instantaneously. Placing a flow control at the accumulator outlet allows free flow from pump to accumulator and adjustable flow to system.

The circuit in Figure 1-10 has a minimum pressure of 2000 psi and a maximum pressure of 3000 psi. This pressure is the limit of most hydraulic components. A 22-gpm pump driven by a 40-hp motor now meets the force and cycle time specified. All pump flow continuously goes to the circuit instead of being unloaded most of the time as in conventional circuits.

As the cylinder cycles, the accumulators supply fluid at a rate set by the flow control. Pump flow adds to accumulator flow to set the required cycle time. Cylinder cycling could be made faster than specified by increasing outlet flow from the accumulator.

The fixed-volume pump in Figure 1-10 unloads through a special accumulator relief/unload/dump valve, which sends all pump flow to the accumulators and cylinder until the system reaches set pressure. After reaching set pressure, the valve opens and unloads the pump to tank at approximately 50 psi. The pump will continue to unload until the system pressure drops about 15%. This pressure drop might be from leakage or at the start of a new cycle. Any time pressure drops, the pump will load and stay loaded until pressure tries to go above 3000 psi. With this valve, stored oil in the accumulators automatically discharges to tank when the pump stops, which makes the circuit safe to work on shortly after locking and tagging off the pump.

Fig-1-11

Figure 1-11 shows a variation of the accumulator circuit in Figure 1-10. Here a 1-gpm fixed-volume pump and a 5-gpm pressure-compensated pump supply oil until the accumulators fill. A pressure switch, set at about 2900 psi, unloads the fixed-volume pump through a solenoid-operated relief valve. After the fixed-volume pump unloads, the pressure-compensated pump finishes filling the accumulators and holds maximum pressure without fluctuations and with minimal heating.

The accumulator dump valve in this circuit will stay closed as long as the pumps are running. When the pumps stop, this valve quickly and automatically discharges the accumulators to tank.

Full-time pressure with fixed-volume pumps

Some circuits need pressure at all times to hold position or maintain force. The circuit in Figure 1-12 holds pressure on the cylinders when they stop, but excessive heat generation makes it a poor choice. Flow controls keep pressure in the circuit while a cylinder is moving.

Fig-1-12

Some designers use the circuit shown in Figure 1-13 to simultaneously reduce energy loss and maintain holding pressure. This double-pump circuit provides high flow (to move the cylinders rapidly) and low flow (for pressure holding). While the system is at holding pressure, the high-flow pump goes to tank through an unloading valve. Only the low-flow pump goes across the relief valve. Although energy loss is drastically reduced, it is still excessive.

Fig-1-13

The circuits shown in Figures 1-14 and 1-15 use a small accumulator to hold pressure on the actuators while unloading the pump at minimum pressure. This makes it possible to use a less expensive fixed-volume pump instead of a pressure-compensated pump, with little or no energy loss or heat generation.

Fig-1-14

Fig-1-15

The pump in Figure 1-14 unloads through an accumulator relief/unload/dump valve. This valve sends all pump flow to the accumulator and cylinders until the system reaches set pressure. After it reaches set pressure, the valve opens and unloads the pump to tank at approximately 50 psi. The pump will continue to unload until the system pressure drops about 15%. This pressure drop might be from leakage or it could be at the start of a new cycle. The pump loads again and fills the circuit until pressure tries to go above 2000 psi. While the pump unloads, the accumulator makes up for any leakage so pressure at the cylinders only drops about 15% maximum. The length of time the pump unloads depends on the size of the accumulator and the amount of system leakage. With the accumulator relief/unload/dump valve, stored oil in the accumulator discharges to tank when the pump stops. This makes the circuit safe to work on shortly after locking and tagging out the pump.

Notice the variation of the above pressure holding circuit in Figure 1-15. Here the pump unloads through a normally open, solenoid-operated relief valve controlled by a pressure switch. The accumulator and actuators fill from the pump until system pressure reaches 2000 psi. At 2000 psi, the pump unloads through a solenoid operated relief valve at approximately 50 psi. The main advantage of the circuit in Figure 1-15 is that pressure drop is adjustable by more or less than the fixed 15% allowed by the unloading valve in Figure 1-14.

To have a safe accumulator circuit, it is necessary to have a means to discharge stored energy at shutdown. The circuit in Figure 1-15 uses a high-ratio pilot-to-close check valve. The pilot ratio is about 200:1, which means 25 psi in the pilot line can hold as much as 5000 psi in the circuit. Most unloading circuits have at least 25 psi while unloaded, so this valve works well. When the pump shuts off, pressure drops to zero, the pilot-to-close check valve opens, and stored energy dumps to tank.

Another way to automatically discharge the accumulator at shutdown is with a normally open, solenoid-operated, 2-way directional valve. This directional valve connects to the accumulator pressure line and on to tank. Starting the pump motor also energizes the solenoid on the normally open 2-way valve, causing it to close. As long as the pump runs, this valve blocks the flow path to tank. When the pump stops, the solenoid is deenergized, and the valve shifts to port stored energy to tank.

CAUTION! ALWAYS CHECK AN ACCUMULATOR CIRCUIT FOR PRESSURE BEFORE WORKING ON IT. NEVER ASSUME THE AUTOMATIC UNLOADING SYSTEM WORKED!

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Frequently, pumps can be downsized if the circuit uses an accumulator. The accumulator volume adds to that of both pumps to speed downward travel of the press ram. When the ram meets sufficient resistance, the pressure switch shifts the solenoid valve. This directs fluid from the large pump to recharge the accumulator, while the small pump continues to supply high-pressure fluid to the ram. When the manually operated, 4-way valve is shifted for the return stroke, pressure is relieved, the solenoid valve is de-energized, and both pumps and the accumulator deliver fluid for rapid return.
	Traverse and clamp In this arrangement, fluid from the large accumulator combines with pump output to extend the cylinder rapidly. Fully extending it trips the limit switch to actuate solenoid (c). The small accumulator maintains high clamping pressure on the cylinder for a timed period, during which the pump recharges the large accumulator. Any fluid lost by the small accumulator will also be replaced during this time.
	Standby power source In this diesel engine starting circuit, maximum power is required for a short period, with long time between operations. Power for starting is stored in the accumulators. During operation, the main pump charges the accumulators to the pressure setting of the unloading valve. The pump is unloaded for the remainder of running time. For starting, the manual valve is opened, connecting the combined output from the accumulators to drive the fluid motor. The hand pump recharges the accumulators in case of leakage.
	Surge reduction Operating the 4-way, closed-center valve in this circuit can cause the formation of shock pressures several times the value of the maximum pressure setting on the relief valve. Because the relief valve cannot act fast enough to drain off fluid, the high pressures can be dangerous to personnel and equipment. The accumulator in this circuit absorbs the surge pressures generated when the valve is placed in the neutral position
	Clamping Holding pressure, leakage compensation, and power savings are obtained by using the accumulator in this vise circuit. While the vise jaws are in the clamp position, pressure is held by the accumulator, and pump output is unloaded at low pressure. The accumulator compensates for any leakage past the piston seals in the cylinder. When clamping pressure drops below the setting of the unloading valve, the valve closes, and the pump recharges the accumulator.
	Increased speed Using a pilot-operated check valve allows adding fluid from the accumulator to pump output at the proper time within a cycle. Operating the manual valve directs fluid to retract the cylinder, exerting a pulling force. When pressure increases, the check valve opens, connecting the accumulator to the cylinder for fast action. Releasing the manual valve allows the pump to recharge the accumulator to the pressure setting of the unloading valve.
	Safety device These mill rolls are loaded by hydraulic pressure. Using an accumulator allows running the pump unloaded most of the time, which saves power. The accumulator also protects the rolls from damage if a large piece of foreign matter enters the mill by absorbing fluid displaced when the roll rises. This fluid returns to the circuit when the foreign matter has passed through.

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Pressure- or fatigue-testing machines often require high pressure for long periods of time. Other circuits might need a small volume of high-pressure fluid for a short period while most of the cycle only needs low pressure. Other machines can use air cylinders to manipulate a part but need very high pressure to perform one operation. Some manufacturers make high-pressure rotary pumps — rated up to approximately 10,000 psi — but these pumps are expensive and may heat the fluid. Another choice for low-volume/high-pressure circuits is an intensifier.

When a circuit calls for a small volume of high-pressure oil or air, consider using an intensifier — sometimes called a booster. Most cylinder manufacturers build air- or hydraulic-powered intensifiers. Or you can use off-the-shelf cylinder parts to assemble your own booster. Also, intensification is a natural function of single-rod cylinders and motor-type flow dividers.

Fig. 13-1. Air-oil intensifier symbol.

Figure 13-1 pictures the symbol for an air-oil intensifier. While the symbol shows two pistons with different diameters, the actual intensifier consists of a piston pushing a rod. The large-area air piston pushes a small-area hydraulic ram against trapped oil. The difference between the two areas gives high-pressure capability at the small ram. This capability is indicated by the area ratio. If the air piston has a 5-in. diameter and the oil piston has a 1-in. diameter, the area ratio is 25:1. With this area ratio, 80 psi acting on the air piston produces 2000 psi at the hydraulic piston.

Fig. 13-2. Reciprocating air-oil intensifier.

Stroke length dictates the maximum volume of high-pressure fluid from an intensifier configured as in Figure 13-1. The booster in Figure 13-2 produces the same pressure but an unlimited volume. A reciprocating intensifier takes fluid from a reservoir and forces it into the circuit. In effect, the reciprocating intensifier is a single-piston pressure-compensated pump. The area ratio and air pressure determine the maximum hydraulic pressure. This pump is close to 100% efficient, so oil heating is not a problem. Intensifiers do not need relief valves because they stall at maximum pressure.

Fig. 13-3. Oversize-rod intensifier.

The oversize-rod cylinder shown in Figure 13-3 also is an intensifier. Any single-rod cylinder intensifies pressure with the rod end port blocked. The larger the rod diameter, the greater the intensification. For low intensification — say 1.5 to 2 times system pressure — a single-rod cylinder is inexpensive and readily available.

Fig. 13-4. Motor-type flow-divider/intensifier.

Figure 13-4 depicts the symbol for a motor-type flow divider used as an intensifier. This type intensifier produces a continuous flow of higher-pressure oil at a reduced flow rate. The reduced flow rate is the same ratio as the pressure increase. (A 2:1 intensifier reduces the flow by 50%.) A motor-type flow divider intensifier is less efficient than a piston-type intensifier and is not recommended for applications with long holding periods.

Fig. 13-5. Air-to-air intensifier.

Figure 13-5 shows the symbol for an air-to-air intensifier. These intensifiers produce small volumes of higher-pressure air from the plant air supply. Ratios up to 4:1 are common. Hydraulically driven designs with higher ratios are available from some manufacturers.

Intensifier circuit using standard cylinders

The schematic diagram in Figures 13-6 through 13-9 suggests how to use standard cylinders as an air-hydraulic intensifier. This is a quick way to get high ratio intensification for a rush job. A 6-in. bore air cylinder driving a 1.5-in. bore hydraulic cylinder gives an intensification ratio of 16:1. With 80-psi input air, hydraulic output pressure is approximately 1300 psi.

Fig. 13-6. Air-oil intensifier circuit using standard cylinders. System on and ready.

Mount the cylinders to a beam or machine member and pipe them as shown in the Figures 13-6. This circuit allows a hydraulic cylinder to operate at low pressure during extension and retraction, with a short high-pressure work stroke to clamp, punch, or do other work. The circuit includes shop-made intensifier A, air-oil tank B, air-pilot-operated hydraulic check valve C, solenoid-operated 5-way air valve D, sequence operated 5-way air valve E, and work cylinder F. With solenoid S1 deenergized, the cylinder and intensifier stay fully retracted, ready for a work stroke.

Fig. 13-7. Air-oil intensifier circuit using standard cylinders. Work cylinder advancing at low pressure.

Energizing solenoid S1 on valve D, as in Figure 13-7, directs air to air-oil tank B and exhausts the rod end of cylinder F. Oil from the air-oil tank free-flows through check valve C to extend the cylinder rapidly. Pressure in the line to the cylinder’s cap end remains low as the cylinder moves toward the work, so sequence valve E stays in its normal position. The cylinder extends until it contacts the work.

Fig. 13-8. Air-oil intensifier circuit using standard cylinders. Work cylinder holding at high pressure.

After the cylinder contacts the work, pressure in its cap-end port increases. Figure 13-8 shows the circuit condition after this pressure buildup shifts sequence valve E. When sequence valve Eshifts, air goes to the cap end of the 6-in. cylinder on intensifier Aand exhausts from its rod end. Cylinder Aextends to stroke the 1-1/2-in. hydraulic cylinder. This forces high-pressure oil to the cap end of work cylinder F. Check valve C is held closed by its spring to block high-pressure oil from going to air-oil tank B. Pressure in the cap end of cylinder F rises to approximately 1300 psi — and is available to power any high-force operation.

The intensifier’s hydraulic cylinder must provide enough oil to move the work cylinder through its high-pressure stroke. A 3.25-in. bore work cylinder with a high-pressure work stroke of 0.75 in. requires a minimum 6.22 in.3 intensifier volume. Calculate volume by multiplying the area of the working cylinder by the length of the high-pressure work stroke. To figure the minimum intensifier stroke, divide the volume required for the work cylinder by the area of the intensifier. In this example, the minimum intensifier stroke is 3.5 in. To make sure there is always enough high-pressure oil to do the job, add 1.0 to 1.5 in. to the intensifier stroke to allow for oil compressibility, hose stretch, and possible future needs. Choose an intensifier stroke of at least 5 in. for this application.

Fig. 13-9. Air-oil intensifier circuit using standard cylinders. Work cylinder retracting at low pressure.

Deenergizing solenoid S1 on valve D, Figure 13-9, directs air to the rod end of cylinder F and to the pilot port of air-pilot-operated check valve C. Check valve C opens, providing oil from the cap end of cylinder F with a free path to tank. Pilot pressure to sequence valve E drops when valve D shifts. When sequence valve E returns to its normal position, intensifier A retracts and fills the intensifier cylinder with oil for the next cycle.

Notice that as cylinder F retracts, only 80-psi air pressure drives it. There is ample hydraulic pressure to extend the cylinder for the high-force work stroke, but only air pressure to retract it. If a higher retracting force is needed (to disengage tooling or for other reasons), external help or other circuit changes may be necessary.

Adjust hydraulic pressure to the cylinder with a regulator in the air line connected to sequence valve E. With a regulator to adjust the air pressure, changing hydraulic force is simple.

Hydraulic cylinder F should have resilient seals that keep oil from leaking to the air side or air to the oil side. Some circuits use two air-oil tanks on cylinder F to prevent aeration of the oil. (Chapter 3 has information about sizing and hooking up air-oil tanks.)

Three-head intensifier circuit with tandem cylinder

Several manufacturers produce 3-head intensifiers that eliminate external pilot-operated check valves. The first head on a 3-head intensifier has an air seal on its rod facing the air side and a hydraulic seal facing the oil side. The second head has an oil port into the rod chamber and a resilient seal facing the third head. The third head has a welded-on oil chamber that the piston rod enters. When the piston rod advances, it displaces oil from this chamber to create high pressure. The ratio of the air-piston area to the rod area intensifies the pressure by up to 40:1, or even higher. A standard 5-in. bore air cylinder with a 1-in. diameter piston rod produces 25:1 intensification. (This is a standard size for several manufacturers.) Three-head intensifiers supply a small volume of oil for short high-pressure work strokes. Calculate the high-pressure oil volume by multiplying the rod area by the stroke length after the rod passes the seal between the second and third head.

Fig. 13-10. Tandem-cylinder air-oil intensifier circuit with typical 3-head intensifier. System is on and ready.

Figure 13-10 shows how 3-head intensifier C pressurizes air-oil tandem cylinder A. This circuit provides rapid low-force advance and retract strokes, with a short high-force work stroke when the cylinder meets resistance. Solenoid-operated directional valve B extends the air part of the tandem cylinder. Sequence valve D operates intensifier C. Sealed expansion tank Ereceives oil from the tandem cylinder while it extends at high pressure. For an expansion tank, mount an air filter with a clear bowl upside down, and remove the filter element. The transparent bowl makes it easy to check oil levels. This circuit eliminates air-oil tanks to make the system more compact. Figure 13-10 shows the circuit at rest.

Fig. 13-11. Tandem-cylinder air-oil intensifier circuit with typical 3-head intensifier. Tandem cylinder is advancing rapidly.

Shifting solenoid S1 on valve B, as in Figure 13-11, makes the air-oil tandem cylinder advance rapidly to the work. Oil in the double rod-end cylinder transfers from front to back through the center head of intensifier C. (Keep these transfer lines short with oil velocity below 4 fps to minimize pressure drop.) As the cylinder advances, pressure at the cap port stays low. Adjust the spring on sequence valve D to cycle the intensifier after the tandem cylinder contacts the work.

Fig. 13-12. Tandem-cylinder air-oil intensifier circuit with typical 3-head intensifier. High pressure in tandem cylinder.

At this contact, sequence valve D shifts to start intensifier C stroking forward, as in Figure 13-12. When the intensifier rod passes through the seal between heads 2 and 3, pressure intensification begins on the back of the double rod-end cylinder. As the tandem cylinder extends, trapped oil from the front chamber goes into expansion tank E. (Pressure increases slightly in the tank because the air trapped above the oil is compressed.) Use a tank with three to four times the volume displaced by the cylinder during the high-pressure work stroke. As the intensifier continues to stroke, increased pressure performs the work.

It is important that the intensifier contains enough oil to move the tandem cylinder through its high-pressure stroke. If the double-rod cylinder has a 3.25-in. bore with a 1.375-in. rod, and the high-pressure stroke is 0.375 in., then a minimum of 2.55 in.3 of oil is needed. Add considerations for oil compressibility plus line and cylinder tube expansion to the cylinder high-pressure stroke volume. Remember: line expansion is greater when using flexible hose. Determine the volume of oil in the high-pressure portion of the piping and cylinder, and then increase this volume by 0.5% per thousand psi of pressure. Often it requires 0.5 to 1.5 in.3 of oil at 2000 psi to make up for oil compressibility. Calculate oil compressibility and add it to the stroke volume so the intensifier does not bottom out before the oil reaches the desired high pressure. On most 3-head intensifiers, add 2.0 in. to the stroke required for volume to make up for oil that bypasses the rod before it enters the high-pressure seal between head 2 and head 3.

Fig. 13-13. Tandem-cylinder air-oil intensifier circuit with typical 3-head intensifier. Tandem cylinder is retracting rapidly.

Figure 13-13 shows the intensifier and cylinder retracting. Deenergizing solenoid S1 on valve B lets sequence valve D spring-return to its normal condition. The intensifier starts retracting at high speed, but the tandem cylinder moves slowly. When the intensifier passes the high-pressure seal between heads 2 and 3, the tandem cylinder quickly returns to its home position.

Reciprocating intensifier for increased volume

A single-stroke intensifier produces a limited volume of high-pressure fluid. Pressure stops building when a single-stroke intensifier reaches the end of its stroke. If cylinder seals or piping leak, a single-stroke intensifier may build pressure, but then quickly lose it. When a circuit needs unlimited high-pressure volume at low flow, use a reciprocating intensifier.

Figures 13-14 through 17 show a reciprocating intensifier powering a cylinder that must hold clamping pressure for days. Reciprocating intensifier A, air-oil tank B, pilot-operated check D, solenoid valve E, and sequence valve F advance cylinder C to the work rapidly. This arrangement can hold as much as 3200 psi for long periods without wasting energy or generating heat.

Fig. 13-14. Air-oil intensifier circuit with purchased reciprocating intensifier. System is on and ready.

Several companies assemble reciprocating intensifiers with a directional valve, limit valves, and check valves in a unit. Special-order units may come with air-oil tanks, special valves, or accumulators — all pre-piped for operation. When supplied with pressurized air, the unit in Figure 13-14 cycles and pumps oil until it reaches a maximum pressure. Other units operate from a pilot signal whenever the machine requires intensified pressure. For even higher pressures, dual or triple air pistons give higher ratios. Double-acting intensifiers increase oil volume while using less air. Most manufacturers offer single-acting intensifiers as standard and double-acting intensifiers as an option. When a machine needs a low to medium volume of high-pressure oil and has long holding times, use a reciprocating intensifier.

Fig. 13-15. Air-oil intensifier circuit with purchased reciprocating intensifier. Intensifier is filling with oil.

The circuit changes as shown in Figure 13-15 after the cylinder contacts the work. The intensifier starts cycling because pressure buildup shifts sequence valve F to open. Pilot-operated check valve D closes, blocking pressure fluid from going to tank. Pressure in cylinder C is already at 800 psi. As the intensifier retracts, suction draws oil in through the right-hand check valve to fill the oil chamber. Its spring and the pressure already in the work cylinder hold the left-hand check valve closed. A reciprocating intensifier is a low-volume, single-piston, pressure-compensated pump that continues to move fluid until it reaches maximum pressure. Because output from the intensifier is intermittent, cylinder movement is jerky, as is the rate of pressure increase.

Fig. 13-16. Air-oil intensifier circuit with purchased reciprocating intensifier. Intensifier is reversing direction.

Figure 13-16 shows the intensifier changing from filling mode to pumping mode. The reciprocating air piston depresses the upper cam valve, reducing pressure on the right end of the double-bleed valve and causing it to shift. Both check valves close at this time, trapping oil in the cylinder. The intensifier now starts its pumping stroke.

Fig. 13-17. Air-oil intensifier circuit with purchased reciprocating intensifier. Intensifier is filling work cylinder with high-pressure oil.

The intensifier is extending and discharging fluid through the left check valve to the actuator in Figure 13-17. Fluid fills the actuator and pressure increases. The intensifier continues to reciprocate until it reaches maximum pressure. At maximum pressure, the intensifier stalls but continues to make up for internal or external leakage.

With the addition of an accumulator, a reciprocating intensifier could supply cylinders or motors that operate intermittently. The accumulator stores oil during machine idle time, and then discharges it at high flow without pulses for short periods. Use flow controls to slow the rapid uncontrolled movements likely to occur when using an accumulator.

Oversize-rod cylinder as an intensifier

There are times when the operating pressure of a hydraulic system is too low to produce enough force on a cylinder. The pump’s rated pressure may be inadequate or the electric motor has too little horsepower for the higher pressure. Also, other actuators in the system may not be able stand higher pressure. One answer to this problem is a hydraulic cylinder piped as an intensifier.

When a single-rod cylinder extends, pressure in the rod end intensifies if there is any resistance to flow out of it. Resistance could be from a flow control, counterbalance valve, or simply a restriction. The amount of intensification depends on the area differential of the cap end to the rod end of the cylinder. A typical 4.0-in. bore cylinder with a 2.5-in. oversize rod is sold as a 2:1 ratio. All standard interchangeable cylinders use standard bore and rod sizes that are close to but not greater than a 2:1 ratio. The 4.0-in. bore, 2.5-in. rod combination actually has 1.64:1 area differential. With the rod-end port blocked, a 1.64:1-ratio cylinder produces 1640 psi at the rod end if the cap-end pressure is 1000 psi. This intensified fluid might cause problems in a typical circuit, but could supply a small volume of higher-pressure oil for a short, high-force work stroke from a cylinder.

Fig. 13-18. 2:1 rod cylinder serving as an intensifier. At rest with pump running.

The volume of oil entering and leaving the intensifier cylinder has the same ratio as the intensification. In the 1.64:1 example above, with a cylinder cap-end flow of 10 gpm, pressure intensified flow from the rod end is 6.1 gpm. The larger the cylinder rod, the higher the intensified pressure — and the lower the flow.

Figure 13-18 shows a schematic diagram of an oversize-rod cylinder used as an intensifier. Intensifier cylinder A has 5.0-in. bore with a 3.5-in. diameter rod. The area of the cap side is 19.64 in.2 and the rod annulus area is 10.01 in.2, giving a ratio of 1.96:1. Every 100 psi in the cap end produces 196 psi in the rod end. Also, 10 gpm entering the cap end pushes 5.1 gpm from the rod end. Stroke length of intensifier cylinder A must give enough volume to move work cylinder B through its high-pressure work stroke. If cylinder B has a 10.0-in. bore and a 0.5-in. stroke, the required volume is approximately 40 in. 3. Dividing the 40-in.3 work-stroke volume by a 10-in.3 intensifier volume indicates that a minimum stroke of 4 in. is needed from cylinder A. To allow for oil compressibility and leakage, specify an intensifier stroke of 6 to 8 in.

Fig. 13-19. 2:1 rod cylinder serving as an intensifier. Work cylinder is extending rapidly.

The cycle is automatic because sequence valves D and E control extension and retraction of the intensifier. Cycle time is slightly slower than the original low-force circuit.

Figure 13-19 shows solenoid A1 on directional valve C energized. Fluid flows directly to work cylinder B through the free-flow check on sequence valve E. Work cylinder B advances rapidly toward the work at low pressure.

Fig. 13-20. 2:1 rod cylinder serving as an intensifier. Work cylinder is extending under high pressure.

At work contact, pressure builds to the setting of sequence valve D, Figure 13-20. Intensifier cylinder A extends and pressurizes oil in the cap end of work cylinder B to approximately twice system pressure. Before the intensifier bottoms out, it must give enough volume to complete cylinder B’s work stroke. For long holding cycles, calculate valve and cylinder leakage, then add extra intensifier stroke so pressure holds.

Fig. 13-21. 2:1 rod cylinder serving as an intensifier. Both cylinders retracting.

To retract work cylinder B, energize solenoid B1 to direct oil to its rod end, as in Figure 13-21. As cylinder B retracts, sequence valve E forces oil from its cap end to retract intensifier cylinder A. This saves pump fluid and retracts the intensifier within normal cycle time. When intensifier cylinder B retracts fully, external pilot-operated sequence valve E opens and the remainder of the oil in the work cylinder cap end goes to tank. The only added cycle time is while the intensifier boosts pressure in the work cylinder.

Motor-type flow divider as an intensifier

A motor-type flow divider intensifies pressure at one outlet when the other outlet is at a lower or no pressure. In the case of a 2-outlet motor flow divider with equal displacements, when inlet pressure is 1000 psi, one outlet can be at 2000 psi while the other outlet is 0 psi. While pressure doubles, flow from the intensified outlet is one half that at the inlet. The energy from the zero outlet motor transfers to the other motor to produce intensified pressure.

With more than one section going to tank, say from a 4-outlet divider with three outlets to tank, intensification would be four times. While the intensified fluid is four times inlet pressure, volume is only one-fourth inlet flow.

Using motor dividers with unequal sections is another way to get high intensification. If the motor in one section discharges 3 gpm to tank and the other section sends 1 gpm, intensification is still 4:1.

Fig. 13-22. Motor-type flow divider used as an intensifier. At rest with pump running.

Figures 13-22 through 25 show how to use this feature of motor-type flow dividers in a circuit. This circuit has equal flow divider C and 3-way directional valve D in the cylinder cap end line. In the at-rest condition, both outlets of the flow divider connect to the cap-end port.

Fig. 13-23. Motor-type flow divider used as an intensifier. Cylinder extending at full speed.

In Figure 13-23, the cylinder is extending at full speed and low thrust. Shifting solenoid A1 of the directional valve ports oil through one side of the divider and 3-way valve to the rod-end port. Fluid from the other side of the divider goes directly to the cylinder rod-end port. Size the pump and valves to provide enough flow for the speed required in the fast-forward portion of the cycle. Normally, motor horsepower is low for a cylinder moving a light load.

Fig. 13-24. Motor-type flow divider used as an intensifier. Cylinder extending at full power.

When the cylinder makes a limit switch, as in Figure 13-24, it energizes solenoid C1 on the 3-way valve. When the valve shifts, oil from one section of the motor flow divider goes to tank. Pressure doubles, while cylinder speed drops to half what it was before energizing solenoid C1.

This circuit works best on actuators that do not stall. Using this setup for a fast advance and clamping operation might result in excess heat from internal leakage in the flow divider during the clamping part of the cycle.

Fig. 13-25. Motor-type flow divider used as an intensifier. Cylinder retracting.

Energizing solenoid B1, Figure 13-25, makes the cylinder retract. Oil from the cap-end port flows through both sections of the flow divider, then back to tank through the directional valve.

When using a motor-type flow divider as an intensifier, make sure it is capable of operating at the elevated pressure. Pressure rating of an inexpensive gear motor flow dividers may be only 2000 psi intermittent and 1500 psi continuous. Some gerotor flow dividers go as high as 4500 psi intermittent and 3000 psi continuous.

Special air-oil intensifier cylinder

Some manufacturers build self-contained, air-driven, high-force hydraulic cylinders. These units look like a very long stroke air cylinder. Typically, they have 2 to 10 in. total strokes with 1.0- to 1.5-in. high-force strokes. They often replace a hydraulic unit on a machine that needs high tonnage for one operation on an otherwise air-powered circuit. Because these special intensifiers are self-contained, they only require an air supply and a signal to start them. They have sealed reservoirs so they operate in any position. They normally have an indicator to monitor oil volume for preventive maintenance. According to bore size and stroke length, cycle rates go as high as 150 per minute. The bigger the bore and longer the stroke, the fewer the cycles per minute.

Fig. 13-26. Special air-oil intensifier cylinder. System on and ready.

As with other air-oil devices, return power is only cylinder net rod-end area multiplied by air pressure. The unit may have 50 tons to punch a hole but only 0.5 ton to retract the punch. For high retraction force use springs or urethane strippers, or add short-stroke return cylinders.

Figure 13-26 has a cutaway view of the intensifier cylinder at rest. (This view only shows function, not necessarily an actual assembly.) Air piston and rod C with attached hydraulic ram D move rapidly at low force to the work and return the tooling at the end of the cycle. Ram D is the area that intensified oil pushes on to get the short, high-force work stroke. Spring-loaded, floating piston A forms the top of a variable-volume, sealed oil tank. Spring-return air piston B drives its piston rod into trapped oil to intensify pressure for the work stroke. Directional control valve E cycles the advance and return strokes of cylinder C, and supplies air to pilot sequence valve F, starting the high-pressure work stroke.

Fig. 13-27. Special air-oil intensifier cylinder. Fast advance at low force.

When directional valve Eshifts, as in Figure 13-27, air piston-and-rod C extends the tooling to the work rapidly. As the piston-and-rod extend, ram D advances and fills with oil from the variable-volume tank. Vacuum forms in the chamber behind ram D, and the spring behind the floating piston forces oil into the void. Piston-and-rod C continues to advance and oil transfers until the work is met. This low-force advance stroke moves quickly (and uses air flow controls when necessary). Seals on ram D separate oil and air.

Fig. 13-28. Special air-oil intensifier cylinder. Starting high-pressure cycle.

Figure 13-28 shows the intensifier after contacting the work. When air piston-and-rodC stop against the work, pressure build-up behind the piston shifts sequence valve F. When sequence valve F shifts, shop air extends spring-return air piston B. The first movement of the spring-loaded air piston advances the rod to the flow port connecting the tank to the chamber behind hydraulic ram D. As the rod enters this flow port, it passes through a resilient seal, stopping flow to tank and sealing the chamber behind ram D. This action automatically isolates the low-pressure chamber — eliminating the need for a pilot-operated check valve.

Fig. 13-29. Special air-oil intensifier cylinder extending at low speed with high force.

As the spring-return air piston continues to extend, as in Figure 13-29, the rod displaces oil in the chamber behind hydraulic ram D. In this case, the area of spring-return piston B is 15 times the area of the rod entering the sealed chamber. The air piston and rod continue to displace oil and move hydraulic ram D until the pressure behind the ram becomes 15 times greater than the air pressure on the piston. The stroke of spring-return piston B and the diameter of its rod set the maximum high-pressure work stroke. The higher the intensification ratio and the shorter the stroke, the less the high-pressure stroke capability.

Deenergizing directional valve E allows the spring loaded air piston and the work cylinder to return home. The work cylinder returns slowly while spring return air piston B retracts past the high-pressure seal.

Air-to-air intensifiers

Instead of buying a high-pressure compressor when only a small volume of compressed air is needed, consider using an air-to-air intensifier. Air-to-air intensifiers are small self-contained units that operate automatically as long as they have a supply of compressed air. Figure 13-30 shows a generic schematic of a simple air-to-air intensifier made from stock cylinders and valving. The arrangement has two cylinders connected at their rod ends and mounted on a beam, with limit switches or limit valves, a directional control valve, and four check valves. As long as compressed air is supplied to the intensifier, it takes in atmospheric air, compresses it, and sends it to a receiver and/or the system.

Fig. 13-30. Typical piping arrangement for air-to-air intensifier (shown running with air on).

As the cylinders in Figure 13-30 stroke to the left, the intensifier takes in atmospheric air at the cap end of the 3.0-in. bore cylinder. Compressed air discharges from the rod end to a high-pressure receiver. After the cylinders fully stroke to the left, a limit valve pilot-shifts the directional control valve to stroke the cylinders to the right. When the cylinders stroke to the right, the opposite check valves take in and discharge air. Reciprocation continues until outlet pressure from the 3.0-in. cylinder reaches approximately twice the inlet pressure at the 4.0-in. bore cylinder.

The piping arrangement in Figure 13-30 produces less intensified air per compressor horsepower than the circuits in Figures 13-31 and 13-32. Taking in and compressing atmospheric air to a higher pressure gives a minimal high-pressure volume for each stroke. When compressing a gas, reducing volume by one half doubles absolute pressure. If the 3-in. bore cylinder has a 6-in. stroke and intake pressure is 14 psia, then as the cylinder moves through 3 in. of stroke, pressure climbs to 28 psia. As the 3-in. bore cylinder continues to stroke, pressure goes to 56 psia 1.5 in. from the end and to 112 psia 0.75 in. from the end. The cylinder finally starts discharging 160-psia air about 0.625 in. from the end of its stroke. Volume entering the high-pressure receiver is minimal for each stroke, and continues to decrease as the pressure level increases.

Fig. 13-31. Air-saving piping arrangement for air-to-air intensifier (shown running with air on).

Using shop-air pressure in the intensifying cylinder, Figures 13-31 and -32, greatly reduces this high-pressure/low-flow problem. First, the high-pressure receiver starts with 80 psig and the air in the intensifying cylinder starts at 94 psia. This circuit discharges intensified air for more than half its stroke, making it a smaller, more-efficient package.

Notice also, the approximately 2:1 intensification from a 4-in. cylinder driving a 4-in. cylinder. This is possible because two areas, pressurized by shop air, push against one area of the intensifier cylinder. The actual intensification of the unit in Figure 13-31 is 2.06:1 when stroking to the left, and 1.93:1 when stroking to the right.

Fig. 13-32. Air-saving piping arrangement for air-to-air intensifier (shown running with air on).

For higher pressure use a smaller-bore intensifier cylinder or a larger-bore driving cylinder. Figure 13-32 depicts a 4-in. bore driving cylinder and a 2-in. bore intensifier cylinder. This combination increases inlet air pressure about five times. The actual intensification ratios are 6.33:1 as the cylinders stroke left, and 4.74:1 as the cylinders stroke right.

When specifying cylinders to build an air-to-air intensifier, be careful not exceed their pressure rating. Pre-lubed cylinders are best for this type of operation because they keep excess lubricator oil out of the high-pressure circuit.

As an air-to-air intensifier pumps air to a maximum pressure, the volume decreases as the pressure increases. It is best to operate the intensifier to produce a pressure 15 to 20% higher than the system needs, with a regulator to set the maximum pressure at the work.

الدوائر المحرّكة السائلة

2009-07-08T08:50:00.000-07:00

Fluid Motor Circuits

One way to change fluid energy into useful work is through air or hydraulic motors. These fluid motors produce rotary power that can drive conveyors, operate long transfers, power fan blades, run a winch, drill and tap a hole, and handle many other applications.

Compared to electric motors, quick reversal of rotation or stalling does not damage a fluid motor. Changing motor speed (within the limits of its specifications) does not adversely affect torque. Repeated starting and stopping air or hydraulic motors does not cause damage or overheating. Also, hydraulic motors commonly operate at low speed without a gear reducer. These features — along with compact size — make fluid motors the best choice in many applications.

To size a fluid motor, calculate torque instead of horsepower. However, remember that fluid motors have approximately 50% less starting torque than electric motors of the same horsepower. Make allowance for this reduced torque if the motor must start under load.

Vane-design air motors are common in low-torque/high-speed applications. For some high-torque applications, vane motors with integral gear reducers work well. Piston-type motors are larger and more expensive, but they produce high torque at low speeds. A less expensive, more-compact gerotor-design air motor that operates in the lower-speed range also is available.

Air motors operate safely in most environments and save space at the work site. Air motor efficiency is very low — only in the 15 to 30% range. Minimizing run time by starting and stopping the motor helps offset low efficiency in many applications.

Air motor speed varies with load changes. Power level and torque also vary widely with speed change. Without a good silencer or muffler, air motors are loud.

Fig. 12-1. Uni-directional hydraulic motor.

Fig. 12-2. Bi-directional hydraulic motor.

Hydraulic motors, on the other hand, can produce speeds from 1 to 5000 rpm. With proper valving they produce little shock, and their efficiency is in the 80 to 95% range. They also operate safely in most environments and save space at the work site. They are very rugged. Unlike cylinders with resilient seals, hydraulic motors always leak or bypass internally. To stop and hold over-running loads, use a brake or other external device.

Fig. 12-3. Variable-displacement, bi-directional hydraulic motor.

Fig. 12-4. Uni-directional air motor.

Figure 12-1 and Figure 12-4 show the symbols for a uni-directional hydraulic and air motor. Figures 12-2, 12-3, and 12-5 show motors designed for bi-directional operation. Figure 12-3 shows a variable-displacement bi-directional hydraulic motor symbol.

Fig. 12-5. Bi-directional air motor.

Sizing a hydraulic motor system

To size a hydraulic motor system, the following information must be known:

1. Maximum torque required (usually in in.- or ft-lb)

2. Maximum speed required (usually in revolutions per minute, rpm)

3. How motor will be stopped (coast, braked, decelerated, other)

4. Maximum pressure allowed (in psi, arbitrarily decided by the engineer or designer)

Fig. 12-6. Bi-directional hydraulic motor at rest with pump running.

Fig. 12-7. Bi-directional hydraulic motor running clockwise.

Sample problem – see Figures 12-6 through 12-9

Maximum torque = 5400 in.-lb Maximum pressure = 3000 psi

Maximum speed = 220 rpm Motor will coast to a stop

Fig. 12-8. Bi-directional hydraulic motor running counterclockwise.

Fig. 12-9. Bi-directional hydraulic motor stopping.

Sizing an air motor system

To size an air motor system the following information must be supplied:

1.	Maximum torque required	Usually in foot-pounds (or horsepower could be substituted)
2.	Maximum speed required	Usually in revolutions per minute, rpm
3.	How motor is stopped	Coast, braked, decelerated, other
4.	Maximum pressure allowed	Usually 80 to100 psi

Fig. 12-10. Bi-directional air motor at rest.

Fig. 12-11. Bi-directional air motor just starting to rotate.

Sample problem – see Figures 12-10 through 12-13

1.	Maximum torque required	10 ft-lb
2.	Maximum speed required	150 rpm
3.	How to stop motor	Retarded
4.	Maximum pressure available	90 psi

Fig. 12-12. Bi-directional air motor running at full torque.

Fig. 12-13. Bi-directional air motor stopping with backpressure.

Directional controls for hydraulic motors

Figure 12-14 shows a solenoid-operated control circuit for a small, uni-directional hydraulic motor. Energizing the solenoid starts and runs the motor. Deenergizing the solenoid allows the motor to coast to a stop. Depending on the type and amount of load, starting and stopping the motor may be anywhere between smooth and abrupt.

Fig. 12-14. Small hydraulic motor circuit with single direction of rotation.

Figure 12-15 shows how to control a large uni-directional hydraulic motor via a normally open, solenoid-operated relief valve. This circuit is less expensive than one that uses a large directional valve and a relief valve. Solenoid-operated relief valves discharge pump flow to tank at about 20 to 50 psi, normally keeping the motor stopped. Energizing the solenoid on the relief valve causes it to start closing. The closing of the relief valve builds pressure and the motor starts to turn. If pressure tries to go higher than the relief setting, this valve stays partially open, which gives the motor time to accelerate. When the motor is at maximum speed, pressure drops, and the relief valve closes completely, directing all pump flow to the motor inlet. If the system pressure tries to go higher than the relief setting, the valve opens to protect the circuit. De-energizing the solenoid on the relief valve allows the motor to coast to a stop without cavitation.

Fig. 12-15. Large hydraulic motor circuit with single direction of rotation.

Note that if the motor can turn at the low unloading pressure (20 to 50 psi), it may never stop completely. Use a brake valve at the motor outlet to keep it from turning if this situation occurs.

Figure 12-16 shows two hydraulic motors in a parallel circuit. Supplying two or more motors from a single valve lets the fluid follow the path of least resistance. Synchronizing the motors with a flow divider (Chapter 11) or by a mechanical linkage would keep them together. With parallel circuits, all motors have full torque but only get a portion of pump flow. In other words, the motors have equal power at reduced speed.

Fig. 12-16. Parallel hydraulic motor circuit with bi-directional rotation.

The series circuit in Figure 12-17 has the outlet of the first motor piped to the inlet of the second. This series circuit gives nearly perfect speed synchronization, but reduces each motor’s power.

The series circuit works well with hydraulic motors because outlet flow is almost identical to inlet flow. This near-equal flow provides tolerable synchronization, but generates backpressure on the inlet and outlet ports of the leading motor. All hydraulic motors have some internal leakage that normally goes to the low-pressure or tank-side outlet through internal check valves. With motors in series, backpressure to run the downstream motors can cause excessive case pressure that may blow out the shaft seal. Some hydraulic motors are available with high-pressure seals that eliminate the blowout problem. Another option is to select motors with external case drains (as shown in the Figure) that allow leakage oil to return to tank at little or no backpressure. When external case drains, remember that leakage in the leading motors will cause the downstream motors to run more slowly.

Fig. 12-17. Series hydraulic motor circuit with bi-directional rotation.

All of these circuits show open-centered valves for smooth stopping or coasting of the motors. This works well if the motor has little or no tendency to coast or if coasting is allowable. Next, we’ll look at some circuits for rapidly and smoothly stopping hydraulic motors with over-running loads.

Circuits for hydraulic motors with over-running loads

Figure 12-18 depicts the most common way to slow a hydraulic motor with an over-running load. Notice that the directional valve has blocked ports A and B. This valve’s center condition makes a motor stop abruptly, which could cause high shock and physical damage. Shock occurs because the motor becomes a fixed-displacement pump without a relief valve to protect it from over pressure. The cross-port relief valve shown in the Figure allows fluid from the over-pressured port to go to the opposite motor port. The reason for piping the outlet of the reliefs to the opposite motor port is to keep it from cavitating as the motor decelerates. Cross-port reliefs are available with both relief valves in a common housing to save piping time and reduce potential leakage points.

Fig. 12-18. Dual cross-port relief valves.

Set the pressure for the cross-port relief valve the same or higher than system pressure. This gives full starting torque and smooth controlled stopping. When the cross-port relief valve pressure setting is lower than system pressure, it reduces starting and maximum running torque. The only reason for a lower setting at the cross-port reliefs is for longer coasting time. The higher the pressure setting of the cross-port reliefs, the more quickly the motor stops.

By deenergizing the directional valve with the motor at full speed, outlet flow from the motor is blocked and pressure increases. When pressure reaches the setting of the cross-port relief valve, the valve opens and allows flow to the opposite motor port. Backpressure equal to or greater than what it took to start the motor now holds back against the load. The energy of the over-running load dissipates over a short period of time to eliminate shock.

Fig. 12-19. Single cross-port relief valve with check valves.

Dual cross-port relief valves allow different stopping times for the two directions of rotation. If this feature is not required, use the single relief and check valves shown in Figure 12-19. With this circuit, pressure is set only once for both directions of rotation. Notice that the check valves direct flow through the relief from either motor port to the opposite motor port. This check valve arrangement is available in a manifold to save piping time.

Fig. 12-20. Sequence valve for cross-port relief.

Figure 12-20 shows an externally piloted and drained pressure control valve that slows the motor in both directions. A shuttle valve feeds the pilot port from either motor port to open the pressure control valve. The external drain line allows internal leakage to return to tank. As Figure 12-20 indicates, the pressure control must be set equal to or higher than the system pressure. For longer stopping times, use the piping arrangement in Figure 12-21.

Fig. 12-21. Externally piloted and drained sequence valve for cross-port relief.

In Figure 12-21, the drain line from the pressure-control valve is piped to the pump line of the directional valve. With the drain line piped this way, the setting of the pressure control is unimportant when the motor is running. Working pressure at the drain port of the pressure control valve adds to the set pressure. Shifting the directional valve keeps the cross-port relief valve from opening when the motor is running.

When the directional valve shifts to its center position, pump flow dumps to tank and the drain line pressure drops, allowing the cross-port relief to operate. This provides pressure lower than system pressure at the cross-port relief valve, so the over-running load takes longer to stop.

Brake valves to control hydraulic motors with over-running loads

Figure 12-22 depicts brake valves (sometimes called over-center valves) piped in the lines between the motor and the directional valve. To control the motor in both directions of rotation, install two brake valves as shown.

Fig. 12-22. Internally and externally piloted brake valve – at rest with pump running.

A brake valve is an internally and externally pilot-operated pressure-control valve. The internal pilot works on a small area and thus requires a high pilot pressure. The external pilot works on a larger area at a much lower pilot pressure. A common ratio for pilot areas is 8:1. With an 8:1 pilot-area difference, the valve allows fluid to pass when the internal pilot pressure is 1000 psi or the external pilot pressure is 125 psi. Always note the internal pilot pressure setting on the schematic diagram. The internal pilot setting indicates the amount of backpressure at the motor during deceleration.

Brake valve pressure setting is independent of system pressure. This means a lower or higher pressure on the brake valve does not affect the system’s maximum operating pressure.

Brake valves are necessary when a hydraulic motor’s load tries to make it go faster than the pump or control circuit feeds it. In a winch application, a directional valve shifts to lower a load. Without a brake valve, the winch load falls freely, the motor cavitates, and the circuit is unsafe. Most winch applications require only one brake valve.

Figure 12-22 shows a brake-valve circuit at rest. The brake valves stay closed because their pressure settings are high enough to stop the load. Notice the internal pilot lines on each brake valve that sense pressure in the motor outlet lines. Also, each brake valve has an external pilot line from the opposite motor flow line. Bypass check valves in each brake valve allow free-flow return during reverse flow.

NOTE: Never depend on any valves to stop and hold a hydraulic motor. Hydraulic motors always have internal leakage and will continue to turn slowly without some external braking arrangement. Spring-on, pilot-to-release, multiple-disk brakes are one method of holding a hydraulic motor stationary.

Figure 12-23 shows the brake-valve circuit after shifting the directional valve. Oil entering the left motor port must be at least 112 psi (1/8 of the 900-psi setting) to open the right brake valve. When the load is over-running, inlet pressure to the motor stays at 112 psi to hold the outlet brake valve open. When pressure at the inlet drops below 112 psi, the opposite brake valve closes to retard motor movement. During this portion of the cycle, backpressure keeps the motor from running away. When pressure tries to go above 112 psi, the opposite brake valve opens wide -- dropping all backpressure. As the motor is turning under power, the external pilot supply controls the brake valve.

Fig. 12-23. Internally and externally piloted brake valve – with hydraulic motor running.

When the directional valve centers, as depicted in Figure 12-24, external pilot supply to the brake valve drops. The brake valve tries to close, causing pressure to increase at the motor outlet port. When outlet pressure reaches 900 psi, it forces the brake valve open and holds back against the over-running load. Because 900 psi is more than enough to stop the load, the motor decelerates to a smooth controlled stop.

Fig. 12-24. Internally and externally piloted brake valve – with hydraulic motor stopped.

The brake valve only holds the motor back if it is trying to run away. This means little or no energy loss as the motor turns under load — and no running away when the load tries to turn the motor. This action is identical to that of a counterbalance circuit for a cylinder.

Because each brake valve is independent, a different pressure setting at each valve is possible.

Controlling the speed of hydraulic motors

When flow controls set the speed of a hydraulic motor, the result is nearly the same as with a cylinder. The main difference is that cylinders normally have positive seals while motors always have internal leakage. With meter-in and bleed-off circuits, internal leakage causes motor speed to fluctuation as pressure varies.

Figure 12-25 shows a motor circuit with meter-in flow controls. This is the preferred way to control the speed of hydraulic motors. Use a meter-in circuit whenever possible because the motor's internal leakage passes to the low-pressure port through internal check valves. This type circuit makes a case drain or high-pressure shaft seal unnecessary. There is little or no backpressure at the outlet of the motor to cause high pressure at the shaft seal.

Fig. 12-25. Meter-in flow-control circuit for a hydraulic motor.

When supplied with 3 gpm at 100 psi, the motor turns at low torque. Its speed is approximately 200 rpm. When the motor load increases and pressure drop across it climbs, increased internal leakage causes speed to slow by as much as 10 to 30%. (The lower the efficiency of the motor, the greater the change in speed.) Motors that turn very slowly may even stop as pressure drop increases. Fast-turning motors may lose speed but continue rotating. As discussed in the section on flow controls, a pressure-compensated valve keeps a constant flow to the motor but cannot offset its internal leakage.

Fig. 12-26. Bleed-off (or bypass) flow-control circuit for a hydraulic motor.

The bleed-off circuit shown in Figure 12-26 is even less effective than the meter-in circuit. With bleed-off flow control, the motor inefficiency combines with the pump inefficiency to produce an even greater loss of rotational speed as the motor loads. The circuit shows a pressure-compensated flow control bypassing 7 gpm. The motor operates at 3 gpm and 100 psi. As the motor loads and pressure increases, internal leakage in the motor and pump results in a greater drop in speed than with the meter-in circuit.

Fig. 12-27. Meter-out flow-control circuit with an external drain.

The meter-out circuit shown in Figure 12-27 provides the most accurate speed control. Note that heat generation is high, even with a pressure-compensated pump, but the resulting speed control is very accurate. Use pressure-compensated flow controls and motors with external case drains for the greatest accuracy and longest motor life. Oil enters the motor at low torque and the pressure drop across it is low. Even though high inlet pressure with high backpressure causes high internal leakage, pump input makes up for it. Accurately controlled fluid leaving the motor keeps speed constant because the pressure-compensated flow control maintains a steady flow. When the motor loads, pressure drop across it decreases. Internal leakage decreases but the amount of oil leaving the outlet port stays constant. As long as outlet flow does not change, speed stays the same.

Fig. 12-28. Meter-in flow-control circuit with a servo-controlled variable-displacement pump.

The circuit in Figure 12-28 generates only minimum heat, but speed fluctuates with load changes when using a variable-displacement pump. The only energy loss is from the inefficiency of the pump, valve, and motor. As the load increases, pressure at the motor inlet climbs. The increased pressure causes greater slip at the pump and motor, so motor speed slows. As shown in the Figure, a servo-controlled pump with electronic feedback from the hydraulic motor eliminates the speed change problem and gives extremely accurate speed control.

Simple torque limiter for air motors

Some circuits reduce the inlet pressure to limit air motor torque on a device. Reducing inlet pressure causes the motor to stall when the workload exceeds its torque. Limiting torque with this method is not reliable or repeatable. Torque levels vary by as much as 10 to 20% according to how fast the motor turns, the mass involved, and the pressure required. The faster the motor turns and the heavier the part being turned, the more the torque overrides. Adjusting inlet pressure reduces motor torque but will not overcome the variations that weight and speed cause. Also, adjusting inlet pressure to a very low level makes motor performance erratic, with an ever-changing torque. Finally, at lower pressures, reduced speed may increase cycle time.

Fig. 12-29. Simple unidirectional torque limiter for an air motor.

To overcome these problems, try the diagrammed circuit and simplified mechanical layout shown in Figure 12-29. With this circuit, the motor operates at full torque, while meter-in flow controls set its maximum speed. An adjustable spring force holds the motor in position by opposing its torque reaction. As the motor turns to tighten a device, torque is low until near the end of the operation. During this part of the cycle, the adjustable spring resists motor torque, keeping its housing stationary. When motor torque increases to match the amount set by the spring tension, the housing starts to rotate. When torque reaches spring setting, the motor housing moves the reaction arm against the limit switch, shifting the directional valve to center and stopping the motor. This mechanism — designed with minimal friction or binding — gives reliable and repeatable results. Air limit valves or electric limit switches work equally well with this arrangement. Use air limits with snap action shifting for good repeatability.

Hydraulic motor driven by a bi-directional pump

Figure 12-30 shows a variable-displacement bi-directional pump driving a bi-directional hydraulic motor in a closed-loop system. This circuit is commonly known as a hydrostatic transmission. The bi-directional hydraulic pump can produce up to full flow from either port, eliminating the need for directional or flow control valves. When the pump strokes to produce flow from one port, the motor starts turning a given direction. With controlled pump stroking, the motor starts and accelerates smoothly. The amount of pump stroke controls motor rpm and pressure is just enough to make the motor turn. With high-efficiency components, this system generates little heat, eliminating the need for large reservoirs while it gives infinitely variable speed and torque.

Fig. 12-30. Bidirectional hydraulic motor driven by a bidirectional pump (shown with motor at rest and pump running).

Due to internal leakage of the pump and motor, the circuit is not a true closed loop. Without leakage makeup, the circuit would quickly run dry, operate intermittently, and damage its parts. To keep the closed loop full, fixed-displacement charge pump A supplies flow through check valves C1 and C2 to the low-pressure side of the circuit. Charge oil keeps the pump inlet fully supplied, preventing cavitation. When the motor stops, excess oil from the charge pump goes to tank through relief valve B. Oil from the charge pump often operates the pump stroking mechanism, plus other circuits on the machine.

In Figure 12-31, the motor is running forward. Flow from the pump goes to the motor inlet, making it turn. Flow from the motor outlet returns to the opposite side of the pump. Check valve C1 stays closed (due to the working pressure) while check valve C2 opens to allow charge oil to make up for leakage and supply the closed-loop with cooled clean fluid. Shuttle valve E shifts when the motor runs, porting excess charge fluid through relief valve D, the motor case, the pump case, and back to tank. Charge flow sends spent oil through the motor and pump case to cool them, then through a heat exchanger to remove excess heat. Flow goes through relief valve D because its setting is 100 psi less than relief valve B.

Fig. 12-31. Bidirectional hydraulic motor driven by a bidirectional pump (shown with motor running forward).

While the motor runs, speed is infinitely variable and system pressure fluctuates with any load changes. The only heat generation in this circuit comes from the charge pump going across the relief valve, plus inefficiencies in the pump, motor, and valves.

When the motor meets a load large enough to stall it, relief valve F1 opens as shown in Figure 12-32. This directs pump flow around the stalled motor, protecting the system from overpressure. If relief valve F1 stays open for any length of time, the wasted energy quickly over-heats the closed-loop piping.

Fig. 12-32. Bidirectional hydraulic motor driven by a bidirectional pump (shown with motor stalled).

Relief valves F1 and F2 also protect the motor when the pump goes to zero stroke and an external force continues to drive the motor. When an external force drives a motor, the motor becomes a pump. Pressure at its outlet climbs until it reaches 5000 psi. At 5000 psi, one of the relief valves opens, allowing oil to flow to the opposite motor port. This protects the motor and circuit from excess pressure and possible cavitation. While the externally driven motor bypasses, relief valve backpressure decelerates it, stopping the load quickly and smoothly.

Closed-loop motor circuits give infinitely variable control of torque and speed with minimal shock and heat generation.