Practical knowledge about hydraulic and pneumatic components and systems

Written by  Bud Trinkel, Certified Fluid Power Engineer

Edited by Mary Gannon and Richard Schneider, Hydraulics & Pneumatics magazine.

Table of Contents

Preface and Foreward

Answers to Ebook 1 Quizes

Chapter 1: Fluid Power Basics
What is fluid power? • Who uses fluid power? • Who is responsible for fluid power? • What are the advantages and disadvantages of fluid power? • Which should be used: hydraulics or pneumatics? • Basic laws governing fluid power • Pascal's law • Work and power • Compressibility of fluids • Boyle's law • Charles' law • Static head pressure • Vacuum • Quiz

Chapter 2: Hydraulic Fluids
What hydraulic fluids do: transmit energy, lubricate, seal • How are hydraulic fluids specified? • Viscosity • Viscosity index • Pour point • Common fluid additives • Lubricity adders • Viscosity modifiers • Oxidation resistance • Wear inhibitors • Anti-foam inhibitors • Rust inhibitors • Fluid temperature considerations • Fire resistant fluids • Water • High water content fluid • Water-in-oil emulsions • Water glycol • Synthetics • Handling new fluid • Storing hydraulic fluids • Quiz

Chapter 3: Plumbing
Plumbing a fluid power system • Considerations when plumbing a plant air system • Pipe types and size • Standard pipe layouts • Grid piping system • Loop piping system • Unit distribution system • Typical air piping side view • Pneumatic machine plumbing • Considerations when plumbing a hydraulic system • Pipe types and sizes • Hose • General rules for sizing hydraulic lines • Pump inlet line • Return lines • Working pressure lines • Quiz

Chapter 4: ISO Symbols
Part 1 : Part 2
Fluid power schematic symbols • Basic information • Reservoirs • Filters • Pumps • Flow meters • Relief and unloading valves • Directional control valves • Check valves • Spool valves • Proportional valves • Servovalves • Cartridge valves • Prefill valves • Flow-control valves • Sequence valves • Counterbalance valves • Cylinders • Rotary actuators • Pneumatic and hydraulic motors • Flow dividers • Accumulators • Accessory items • Measuring devices • Air Logic

Chapter 5: Basic Fluid Power Systems
Two types of fluid power circuits in this manual • Schematic drawing of a pneumatic circuit • Physical drawing of a pneumatic circuit • Typical pneumatic circuit • Why a schematic drawing • Parts of a typical pneumatic system • Air logic controls • Directional control valve • Limit valve • Air drills • Schematic drawing of a hydraulic circuit • Physical drawing of a hydraulic circuit • Typical hydraulic circuit • Parts of a typical hydraulic system • Directional control valve • Pressure control valves • Accumulators • Conventions used in this publication • Quiz

Chapter 6: Reservoirs, Heaters, and Coolers
Air receivers • Hydraulic reservoir types • Hydraulic reservoir function • Reservoir cooling formula • Tank heating types and formula • Air and water type fluid coolers • Quiz

Chapter 7: Air and Hydraulic Filters, Air Dryers and Lubricators --
Air filters and lubricators • Compressed air dryers: deliquescent, refrigerant, and desiccant • Filter rating • Beta Ratio • Filter placement • Quiz

Chapter 8: Pumps and Accessories
Part 1 : Part 2
Pneumatic pumps • Air compressor installation • Hydraulic pumps • Non-positive displacement pumps • Positive displacement pumps • Gear, internal, external, gerotor • Screw • Vane, un-balanced, balanced • Circuits for fixed volume pumps • Variable-volume pumps: vane, piston, radial, wobble-plate, in-line or axial, bent-axis • Load sensing, horsepower, or torque limiting • Circuits for pressure-compensated pumps • Air bleed valves • Closed-loop bi-directional pumps • Bi-rotational pumps • Cavitation causes and damage • Pump mounting and alignment • Pump testing • Quiz

Chapter 9: Relief Valves and Unloading Valves
Why a relief valve? • Relief valve terms • Direct acting relief valve • Pilot-operated relief valve • Remote controlled relief valve • Solenoid Operated relief valve • Proportional relief valve • Direct acting unloading valve • Pilot operated unloading valve • High-low pump circuit • Modular relief valve symbols • Quiz

Chapter 10: Directional Control Valves
Part 1 : Part 2 : Part 3 : Part 4 : Part 5
Check valves • Pilot-to-open and pilot-to-close check valves • Pilot-operated check valve circuits • Pre-fill valves • Pre-fill valve circuit • 2-Way, 3-Way, 4-Way, 5-Way valve flow functions and symbols • Valve types • Sliding plate, poppet, and spool types • Spool valve advantages and disadvantages • Hydraulic 4-Way spool valves • All ports open, all ports closed, float center, and tandem center • Solenoid pilot operated valves • Crossover conditions • NFPA and ISO Sub-plate layouts • Schematic bar manifold drawing example • Quiz

Chapter 11: Slip-In Cartridges (Logic Valves)
What are slip-in cartridge valves? • Advantages of slip-in cartridge valves • 1:2 slip-in cartridge valves and how they are used • 1:1.1 slip-in cartridge valves and their uses • 1:1 slip-in cartridge valves and their function as pressure controls • Slip-in cartridge valves as a reducing valve • Quiz

Chapter 12: Infinitely Variable Directional Control Valves
What are infinitely variable directional control valves? • Where are infinitely variable directional control valves used? • Proportional infinitely variable directional control valves • Servo infinitely variable directional control valves • Quiz

Chapter 13: Flow Controls and Flow Dividers
Fixed orifices • Needle valves • Non-compensated flow controls • Pressure-compensated flow controls • Proportional flow controls • Three port flow controls • Priority flow controls • Spool type flow dividers • Motor type flow dividers • Quiz

Chapter 14: Pressure Control Valves (Except Relief and Unloading Valves)
Sequence valves • Sequence valve circuit • Counterbalance Valves • Counterbalance valve circuits • Reducing valves • Reducing valve circuits • Reducing relieving valves • Reducing relieving valve circuits • Air line regulators • Quiz

Chapter 15: Fluid Power Actuators
Part 1 : Part 2
Single-acting rams, push type and pull type • Telescoping cylinders single and double acting • Spring-return or spring-extend cylinders • Diaphragm actuators • Bellows actuators • Rotary actuators • Hydraulic motor • Quiz

Chapter 16: Accumulators
Non-separated, bladder, piston, and weight-loaded types • Accumulator applications by type • Accumulator dump and unloading valves • Common accumulator circuits • Quiz

Chapter 17: Air-Oil Systems & Intensifiers
Air-oil tanks and circuits • Tandem cylinders and circuits • Air to hydraulic and hydraulic to hydraulic single stroke intensifiers • Air-to-air, air-to-hydraulic, and hydraulic-to-hydraulic reciprocating intensifiers • Quiz

Chapter 18: Miscellaneous Fluid Power Items
Electric motors, heat engines, non-electric motors • Shuttle valves • Quick exhaust valves • Quick-acting couplings • Pressure gauges • Temperature gauges • Rotating unions • Flow meters • Pressure switches • Shutoff valves • Mufflers • Machine members • Quiz

Chapter 19: Moving Part Air Logic Controls
And • Or • Not • Yes • Flip-Flop • Memory • Differentiator • On-delay timer • Off-delay timer • Frequency generator • Pressure decay sensor • Sequencer • Sequence valve • Anti-tiedown • Counters • Air logic circuits • Quiz

Chapter 20: Brain Teasers

Chapter 21: Practice Circuits

Chapter 22: Useful Formulas

Fluid Power Basics, Copyright 2013 by Penton Media Inc. No part of this document may be used or reproduced in any manner without written permission except for brief quotations embodied in critical articles or reviews.

Preface

Fluid Power Basics starts with background information about simple air and hydraulic circuits, principles of fluid power operation and physical laws governing fluid power. Subsequent chapters cover different types of hydraulic fluids, fluid rating, operating parameters, and how to apply them. Next, a discussion on plumbing of fluid power systems covers tubing, pipe, and hose installations. A short section on vacuum and its applications is followed by basic circuit information. Coverage then shifts to discussing different components that make up a complete hydraulic or pneumatic system: reservoirs, filters, pumps, flow meters, gauges, and relief valves.

A detailed discussion of directional control valves covers check and prefill valves; decompression circuits; sliding plate, spool and poppet designs; in-line and sub-plate mounted valves, as well as screw-in and slip-in cartridge valves. One chapter is dedicated to an explanation of proportional and servovalves.

Subsequent chapters cover all types of flow controls and their use in a circuit. Next are pressure controls except relief and unloading valves. This chapter includes sequence, counterbalance, and reducing valves. Shuttle valves, quick exhaust valves, and other special-purpose valves are explained. There is a chapter on accumulators that shows and explains how the different types work and common applications.

The book also covers all types of actuators, including cylinders, rams, motors, and rotary actuators. Application of these components in different circuits gives a general overall view of how they are used.

Circuit diagrams are intended to show the function of the components and do not necessarily show all the components to make a safe and reliable system. Drawing practices and symbols according to ISO standards have been used when possible.

FOREWORD

I began my fluid power profession as a salesman of cylinders and valves. I had used the same cylinders as a designer of plastic injection molds, diecast dies, and related tooling. In my ten years in design, I never had to think about what it took to make the cylinders operate. The people who built the fixtures or ran the tooling took care of getting the valves and hooking them to the cylinders. Not until I started my sales career did I realize the tool shop depended on fluid power salesmen to tell them what they needed. It may surprise you to know that salesman design more than 90% of all fluid power circuits in the United States.

The first two cylinder air circuits I designed took several hours and some of these didn't even work. Fortunately, the company I started with was committed to training, so design time decreased and working circuits increased. In a few years air logic controls and a full hydraulic pump and valve line brought more challenges and more knowledge about fluid power.

I quickly found teachers of fluid power were in demand at schools with adult evening classes. Many people worked with fluid power but had little or no training in how it worked. I can testify to the fact that teachers always learn more than their students. I often think about some of the early classes and questions I could not answer.

This material is the text for a basic fluid power course I have been teaching for several years. It is dedicated to people who bought fluid power products even if the circuit didn't work the first time. It is also dedicated to many students who asked questions I'd never thought of, and to students who came up with ideas I might have passed over because I was sure it could not be done.

Bud Trinkel, CFPE
Fluid Power Consultant

***Editor's Note: The fluid power industry lost an icon in 2009 when Bud Trinkel passed away, but we are glad to keep his work alive with his technical training manuals published as eBooks. Please see his obituary below, or read a tribute to him from H&P Editor Alan Hitchcox here. Edgar W. “Bud” Trinkel Jr., of Evansville, Ind., died suddenly on August 12. He was best known in the fluid power industry for his years spent working as a hydraulic pneumatic specialist in sales and later for starting his own consulting business, Hydra-Pneu Consulting. He wrote several books on fluid power, including Fluid Power Basics, Fluid Power Circuits Explained, and others to aid his training endeavors. They became so well-accepted that he began producing them as stand-alone books. As president of Hydra-Pneu, Trinkel designed fluid power circuits, provided training, and performed troubleshooting for industrial clients. Prior to founding Hydra-Pneu Consulting in 1984 as a part-time fluid power consulting firm — which became a full-time endeavor in 1988 — Trinkel worked as a technical sales and service representative for a fluid power distributor. Prior to that he served as a sales and service representative for Miller Fluid Power for 14 years. Earlier in his career he worked as an industrial designer in the plastics industry. A veteran of the United States Air Force, he is survived by his wife of 56 years, Sharon; son, Charles (Michelle); daughter, Julie Woodson (Russ); and six grandchildren.

Chapter 1: Fluid Power Basics

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.

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.

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.

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.


 

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.

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.

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.

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.

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. Hg 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.

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.

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.


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.


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.

Quiz

Chapter 2: Hydraulic Fluids

For long service life, safety reasons, and reliable operation of hydraulic circuits, it is very important to use the correct fluid for the application. The most common fluid is based on mineral oil, but some systems require fire resistance because of their proximity to a heat source or other fire hazard. (Water is also making its return to some hydraulic systems because it is inexpensive, fireproof, and does not harm the environment.

Transmit energy.
The main purpose of the fluid in any system is to transmit energy. Electric, internal combustion, steam powered, or other prime movers drive a pump that sends oil through lines to valves that control actuators. The fluid in these lines must transmit the prime movers energy to the actuator so it can perform work. The fluid must flow easily to reduce power losses and make the circuit respond quickly.

Lubricate.
In most hydraulic systems, the fluid must have good lubrication qualities. Pumps, motors, and cylinders need ample lubrication to make them efficient and extend their service life. Mineral oils with anti-wear additives work well and are available from most suppliers. Some fluids may need special considerations in component design to overcome their lack of lubricity.

Seal.
Fluid thickness can be important also because one of its requirements is for sealing. Almost all pumps and many valves have metal to metal sealing fits that have minimal clearance but can leak at elevated pressures. Thin watery fluid can flow through these clearances, reducing efficiency and eroding the mating surfaces. Thicker fluids keep leakage to a minimum and efficiency high.

There are several areas that apply to specifying fluids for a hydraulic circuit. Viscosity is the measure of the fluids thickness. Hydraulic oils thickness is specified by a SUS or SSU designation, similar to the SAE designation used for automotive fluids. SUS stands for Saybolt Universal Seconds (or as some put it, Saybolt Seconds Universal). It is a measuring system set up by a man named Saybolt. Simply stated, the system takes a sample of fluid, heats it to 100° F, and them measures how much fluid passes through a specific orifice in a certain number of seconds.

Viscosity is most important as it applies to pumps. Most manufacturers specify viscosity limits for their pumps and it is best to stay within the limits they suggest. The prime reason for specifying a maximum viscosity is that pressure drop in the pump suction line typically is low and if the oil is too thick, the pump will be damaged due to cavitation. A pump can move fluid of any viscosity if the inlet is amply supplied. On the other end, if fluids are too thin, pump bypass wastes energy and generates extra heat. All other components in the circuit could operate on any viscosity fluid because they only use what is fed to them. However, thicker fluids waste energy because they are hard to move. Thin fluids waste energy because they allow too much bypass.

Viscosity index (or VI) is a measure of viscosity change from one temperature to another. It is common knowledge that heating any oil makes it thinner. A normal industrial hydraulic circuit runs at temperatures between 100° and 130° F. Cold starts could be as low as 40° to 50° F. Using an oil with a low VI number might start well but wind up with excessive leakage and wear or cause cavitation damage at startup and run well at temperature. Most industrial hydraulic oils run in the 90- to 105-VI range and are satisfactory for most applications.

Pour point is the lowest temperature at which a fluid still flows. It should be at least lower than the lowest temperature to which the system will be exposed so the pump can always have some lubrication. Consider installing a reservoir heater and a circulation loop on circuits that start or operate below 60° F.

Refined mineral oil does not have enough lubricating qualities to meet the needs of modern day hydraulic systems. Several lubricity additives to enhance that property are added to mineral oil as a specific manufacturers package. These additives are formulated to work together and should not be mixed with others additives because some components may be incompatible.

Refined mineral oil also is very much affected by temperature change. In its raw state it not only has low lubricity but also would thin out noticeably with only a small increase in temperature. Viscosity modifiers enhance the oils ability to remain at a workable viscosity through a broad temperature range.

There are several causes of hydraulic oil oxidation. These include contamination, air, and heat. The interaction of these outside influences cause sludge and acids to form. Oxidation inhibitors slow or stop the fluids degradation and allow it to perform as intended.

Wear inhibitors are additives that bond with metal parts inside a hydraulic system and leave a thin film that reduces metal-to-metal contact. When these additives are working, they extend part life by reducing wear.

In most hydraulic systems, fast and turbulent fluid flow can lead to foaming. Anti-foaming agents make the fluid less likely to form bubbles and allow those that do form to dissipate more rapidly.

Moisture in the air can condense in a hydraulic reservoir and mix with the fluid. Rust inhibitors negate the effect of this unwanted water and protect the surfaces of the systems metal components. All of these additives are necessary to extend system life and improve reliability.

Overheating the fluid can counteract the additives and decrease system efficiency. Overheating also thins the oil and reduces efficiency because of internal bypassing. Clearances in pump and valve spools let fluid pass as pressure increases, causing more heating until the fluid breaks down. External leaks through fittings and seals also increase as fluid temperatures rise. Another problem caused by overheating is a breakdown of some seal materials. Most rubber compounds are cured by controlled heat over a specific period of time. Continued heating inside the hydraulic system over long periods keeps the curing process going until the seals lose their resiliency and their ability to seal. It is best if hydraulic oil never exceeds 130° F for any extended period. Installing heat exchangers is the most common cure for overheating but designing heat out of a circuit is the better way.

Cold oil is not a problem as far as the oil is concerned but cooling does increase viscosity. When viscosity gets too high, it can cause a pump to cavitate and damage itself internally. Thermostatically controlled reservoir heaters easily eliminate this problem in most cases.

Fire-resistant fluids
Certain applications must operate near a heat source with elevated temperatures or even open flames or electrical heating units. Mineral oil is very flammable. It not only catches fire easily but will continue to burn even after removing the heat source. This fire hazard situation can be eliminated by several different choices of fluids. These fluids are not fireproof, only fire-resistant, which means they will burn if heated past a certain temperature but they will not continue to burn after removing them from the heat source.

Generally, the fire-resistant fluids do not have the same specifications as mineral oil-based fluids. Pumps often must be down rated because the fluids lubricity or specific gravity is different and would shorten the pumps service life drastically at elevated pressures or high rotary speeds. Some fire-resistant fluids are not compatible with standard seal materials so seals must be changed. Always check with the pump manufacturer and fluid supplier before using or changing to a fire-resistant fluid.

Water
Originally, hydraulic circuits used water to transmit energy (hence the word hydraulics). The main problem with water-filled circuits was either low-pressure operation or very expensive pumps and valves to operate with this low viscosity fluid above 500 to 600 psi. When huge oil deposits were discovered, mineral oil replaced water because of its additional benefits. Water made a brief comeback during an oil shortage crisis but quickly succumbed when oil flowed freely again.

In the late 90s, water again made inroads into oil-hydraulic systems. Several companies have developed reliable pumps and valves for water that operate at 1500 to 2000 psi. There are still limitations (such as freezing) to using water, but in certain applications it has many benefits. One big advantage is that there are fewer environmental problems during operation or in disposing of the fluid. Price also is a factor because water costs so little and is readily available almost anywhere.

Some suppliers are making equipment that operates on seawater to eliminate possible contamination of the earths potable water sources. These systems operate at elevated pressures without performance loss.

High water-content fluids
Some types of manufacturing still use water as a base and add some soluble oil for lubrication. This type of fluid is known as high water-content fluid (or HWCF). The common mixture is 95% water and 5% soluble oil. This mixture takes care of most of the lubricity problems but does not address low viscosity concerns. Therefore, systems using HWCF still need expensive pumps and valves to make them efficient and extend their life.

Rolling mills and other applications with molten metals are one area where HWCF is prevalent. Often the soluble oil is the same compound used for coolant in the metal-rolling process. This eliminates concerns about cross-contamination of fluids and the problems it can cause.

Water-in-oil emulsions
Some systems use around 40% water for fire resistance and 60% oil for lubrication and viscosity considerations. Again, these are not common fluids because they require special oil and continuous maintenance to keep them mixed well and their ratio within limits. Most manufacturers do not want the problems associated with water-in-oil emulsions so their use is very limited.

Water glycol
A very common fire-resistant fluid is water glycol. This fluid uses water for fire resistance and a product like ethylene glycol (permanent anti-freeze) for lubricity, along with thickeners to enhance viscosity. Ethylene glycol will burn, but the energy it takes to vaporize the water present quickly quells the fire once it leaves the heat source. This means a fire would not spread to other parts of the plant. Always remember fire-resistant not fireproof.

Water glycol fluids are heavier than mineral oil and do not have its lubricating qualities, so most pump manufacturers specify reduced rpm and lower operating pressures for water glycol. In addition, the water in this fluid can evaporate, especially at elevated temperatures, so it must be tested regularly for the correct mixture.

Cost is also a consideration. Water glycol is more expensive than oil and requires most of the same considerations when disposing of it.

Always check with the pump manufacturer before specifying water glycol fluid to see what changes are necessary to run the pump with this fluid. Seal compatibility is usually not a problem, but always check each manufacturers specifications before implementing this fluid. In addition, it is imperative to completely flush a system of any other fluids before refilling with water glycol.

Synthetics
The other main fire-resistant fluids are synthetic types. They are made from mineral oil, but have been processed and contain additives to obtain a much higher flash point. It takes more heat to start them burning but there is not enough volatile materials in them to sustain burning. These fluids may catch fire from a pot of hot metal but quickly self-extinguish after leaving the heat source.

Synthetic fluids retain most of the qualities of the mineral oil from which they are derived, so most hydraulic components specify no operating restrictions. However, most of these fluids are not compatible with common seal materials so seal specification changes are usually necessary. Special consideration must be given to handling of synthetics because they can cause skin irritation and other health hazards. Also most synthetic fluids require protective epoxy paint for all components in contact with them.

Of all the fluids discussed, synthetics are the most expensive. They can cost up to five times more than mineral oil.

No matter which fluid is chosen, design the circuit to work in a reasonable temperature range; install good filters and maintain them; and check the fluids regularly to see if they are within specification limits.

A good operating temperature range is between 70° and 130° F with the optimum being around 110° F. A rule of thumb would be: warm enough to feel hot to the touch but cool enough to hold tightly for an extended period. Overheating hydraulic fluids is second only to contamination when it comes to reasons for fluid failure.

Continuous filtration of any hydraulic system is necessary for long component life. Fluids seldom wear out but they can become so contaminated that the parts they drive can fail. (The filter section of this book offers some good recommendations on keeping fluid clean.)

Even with the best of care, any hydraulic fluid should be checked at least twice a year. Systems located in dirty atmospheres may need to be checked more often to see if a pattern exists that requires special consideration. Pay close attention to the sampling process and packaging procedures recommended by the test facility that will process the sample. Expect a report on the level of contamination plus an analysis of the additive contents, water content, ferrous and non-ferrous material amounts, and any other problem areas the test facility finds. Use this information to know when to change fluids and to check for abnormal part wear problems.

Fig. 2-1. Filter cart (used to transfer hydraulic fluids) and its circuit schematic diagram

New oil or other fluids from the supplier are not necessarily clean. The fluids are shipped in drums or by bulk, and there is no way of knowing how clean these containers are. Some suppliers offer filtered oil with a guaranteed contamination level at added cost. Otherwise, about the lowest level of contamination from most manufacturers is 25 microns.

Anytime a system needs new fluid, it is best to use a transfer unit, Figure 2-1, with a 10-micron or finer filter in the loop. Another way of filtering new or refill fluid is with a filter permanently attached to the reservoir, Figure 2-2. In this arrangement, the breather or other possible fill points should be made inaccessible.



Fig. 2-2 Hydraulic power unit and circuit diagram of its filter arrangement

The filter cart shown in Figure 2-1 can also be used to filter any hydraulic unit in the plant. Instead of this filter unit sitting idle except when filling systems, set it up at a machines power unit for a timed run. Place the suction hose in one end of the reservoir and the return hose in the opposite end. This adds a continual filtration loop to any machine even when the machines main pump is shut off. Run the cart until the fluid is clean and then move is to another power unit. Repeating this process on a regular schedule can assist the hydraulic units filters and add extra life to the fluid and the hydraulic components. This process may also show a pattern on machines that have a contamination problem.
 

Hydraulic fluids should be stored in a clean dry atmosphere. Keep all containers closed tightly and reinstall covers on any partially used drums.

Never mix fluids in any hydraulic system. Make sure all containers are clearly marked and segregated so fluids will not be mixed with one another. Mixing fluids can result in damage to components and some combinations are very difficult to clean up. Be especially careful when mineral oils and synthetic or water-glycol fluids are used in different parts of the same plant.

Fluids are the lifeblood of any hydraulic system and should be given the utmost care.

Quiz

Chapter 3: Plumbing

Poor plumbing practices can permanently cripple a fluid power circuit even if it was designed with the best engineering practices and assembled with the most up-to-date components. Undersized lines, elbows instead of bends, incorrect component placement, and long piping runs are a few of the items that strangle fluid flow.

Other problems, such as using tapered pipe threads or lines with thin walls, can make a circuit a maintenance nightmare that requires daily attention. Fortunately, there are numerous publications that assist in specifying correct line size and conductor thickness to give low pressure drop and safe working-pressure limits.

Because pneumatic circuits are less complicated and operate at lower pressures, they are not as vulnerable to plumbing problems. One very important aspect that often is overlooked is the length and size of lines between the valves and actuators. Piping between the valve and actuator should be as short as possible and of the minimum diameter to carry the required flow. The reason for this is that all the air in the pipes between the actuator and valve is wasted every cycle. These runs must be filled to make the device move but the air it takes to fill them does no work. During each cycle, air in the actuator lines exhausts to atmosphere without helping cycle time or force. For this reason, always mount the valve close to the actuator ports.

Another aspect of plumbing a pneumatic system is the in-plant pipe installation procedure. To get the required amount of compressed air to the point of usage requires some planning -- or the site may be starved at times.

Fig. 3-1. Pipe size selection chart (in feet) for plant-air systems

Pipe materials and size: Air systems are normally plumbed with Schedule 40 black iron pipe. (Galvanized pipe is not recommended because some galvanizing material may flake off and get into moving parts.) Several other available plumbing materials could be used for air piping because pressure is relatively low. Some mechanics recommend plastic pipe, but be aware a few synthetic compressor lubricants attack plastic and cause it to lose strength. This type of damage weakens the plastic until it can burst, sending shards of plastic flying everywhere in the plant. Never use any piping material not specifically designated by code.

To help select pipe size, the chart in Figure 3-1 shows flow (in cfm) down the left-hand side, length of run (in feet) across the top, and minimum Schedule 40 pipe size in the body at the intersection of these two.

This chart is based on a 1-psi pressure drop for the run lengths given. The right-hand column shows approximate compressor horsepower for the flow figures on the left. Using larger than specified pipe is of little help in reducing pressure drop, but provides more storage volume to handle short brief-flow needs. This chart does not consider fittings and valves, but they also must be considered when figuring the length of a run. Add 5 to 7 feet of pipe length for each fitting or valve -- to be on the safe side.

Not having enough air to run the equipment is expensive, so never try to save a few cents at installation by skimping on pipe size. One or two pipe sizes over minimum add little to cost up front, but can make a big difference later. It is less expensive to run oversize pipe initially than to have to add a line later.

Fig. 3-2. Typical grid system layout for plant air

There are three basic compressed-air piping layouts that meet the requirements of most industrial plants. Some facilities may have two or more of these systems to handle special needs. In general, smaller plants use a modified grid system, especially when the facility is growing. A unit distribution system offers flexibility, but can be expensive up front. A loop system is best suited to new construction; it provides extra storage capacity and dual supply for short bursts of high flow.

Figure 3-2 shows a typical grid-system layout using a centrally located air compressor. All air from the receiver goes to a large header pipe that runs down the center of the plant or department. Branch lines from the header go to separate areas where working drops come down to specific machines. With preplanning for future working drops, this arrangement is very flexible.



Fig. 3-3. Typical loop-piping system layout for plant air

Figure 3-3 shows a typical loop piping system for compressed air. Again, the compressor and receiver are at a central location. The oversized loop around the periphery of the plant -- or department -- adds storage and allows flow with low pressure drop. It also allows for short bursts of high-volume flow to any section because flow in the loop is bi-directional. (Another way to get short high-volume flows with any of these piping systems is to install extra receiver tanks at or near areas that need such flow.)



Fig. 3-4. Typical unit distribution layout for plant air

Figure 3-4 illustrates a unit distribution layout that works well in plants that run departments on different days or shifts -- or plants that started out small and added compressors as business grew. It is the most expensive configuration of the three for a new installation, so is not often used there. One advantage of the multiple compressors is that there is always backup air available for critical operations should a single compressor fail. The disadvantage . . . besides higher price . . . is that some compressors might be neglected by maintenance personnel because they are spread throughout the facility.



Fig. 3-5. Side view of typical compressed-air header and drop arrangement

Figure 3-5 shows a typical pipe run layout for optimum performance from a compressed air system. Strict attention to the details shown here assures a smooth-operating and trouble-free air system.


Pneumatic-machine plumbing

Machines can be plumbed with any of the materials recommended for plant piping. However, because piping at the machine is usually much smaller, polyethylene, nylon, or vinyl tubing with push-to-connect fittings will work very well. Such tubing and fittings come in a variety of sizes (and colors) and require only a few tools to install. The push-to-connect fittings also release easily for troubleshooting checks or rework.

Pipe materials and sizes

Fig. 3-6. O-ring sealed straight-thread fittings
Fig. 3-7. Two types of O-ring sealed flange fittings

Though many hydraulic circuits are plumbed using black-iron pipe with tapered pipe threads, this is not the recommended way. It is nearly impossible to maintain leak-free operation of a 1000- to 3000-psi hydraulic system for any period of time with tapered pipe threads. Even if pipe-connection compound is used, expansion and shock soon loosen the taper interference and fluid weeps through the resulting openings. Another problem with tapered pipe threads occurs on circuits that must be routinely dismantled. Every time a tapered pipe thread is unscrewed, it must be tightened past where it was originally to get a good seal at reassembly. This can only happen so often before the pipe and/or valve must be replaced.

The recommended plumbing material is steel tubing with straight-thread O-ring fittings up to 2-in. OD, Figure 3-6. In sizes larger than 2 in., use steel pipe with welded SAE O-ring-sealed flange fittings on each end, Figure 3-7. For flexible connections, reinforced rubber hose is most common; however, some prefer sealed steel swivel joints.

A good reason for using steel tubing is that it is easily formed to allow for direction changes. Instead of installing fittings that can cause turbulence, use a tube bender to make sweeping turns that eliminate most of the pressure drop associated with elbows. This produces less pressure drop and less heat. Tubing is designated by its outside diameter (OD). As wall thickness increases, inside diameter (ID) decreases. (Black-iron pipe is measured by its nominal ID, but also has a decrease in ID with increase in wall thickness.)

Hydraulic hose

There are places on many machines where rigid pipe or tubing cannot be used because of their inflexibility. Rigid lines can cause problems at cylinders with pivot mountings, pumps on noise-isolation mounts, or connections between separate units. Hose avoids these problems.

However, wholesale use of hose in place of rigid lines it is not generally recommended. Hose is expensive, must be replaced on a regular basis, and flexes and stretches under pressure surges. This flexing produces extra volume and adds to cycle time. It is never recommended to use hose in a servo circuit (although there are times it can't be avoided). Servo circuits are for actuators that need precise control and flexing of hose lines can cause these valves to respond slowly and then go into high frequency oscillations.

Hose is specified by its ID and, unlike pipe and tube, this dimension does not change. Thicker walls for higher pressures make the outside diameter (OD) of hose greater. Pressure is specified in working and burst values (similar to pipe). Working pressure should always be equal to or higher than maximum system operating pressure. Flow rates of hose are slightly higher than pipe and about the same as steel tubing due to hose's smooth bore. However, many of the end connectors for hose are restrictive because they always go inside the inner liner. These fittings are only short restrictions, but can raise pressure drop noticeably in some cases.

Several factors influence hose service life and each one is controllable by some up-front fact finding and planning. First: never go under the manufacturers recommendation for minimum bend radius. Bending hose always causes stress but flexibility is the main reason for using it. Standard hose construction entails wire- or fiber-braided material laid down when the product is straight. Bending these braided materials puts extra stress on the outside of the bend and bunches up those on the inside. Add the constant expansion and contraction from pressure fluctuations during operation and it is easy to see the adverse effect.

Second: don't use hose above its rated working pressure. While maximum pressure might be set at or below the hose rating, higher shock pressures could be damaging during every cycle. Make sure the pressure rating of the hose on all machines is at or above operating pressure -- and design out system shock to protect the hose and other hydraulic components.

Third: avoid operating at temperatures above the rating of the hose. Most hose manufacturers make hoses in different temperature ratings. Of course, the higher the temperature rating, the more expensive the hose is, but it is false economy to use the wrong hose to save a few pennies.

Fourth: don't install hose where it must twist during each cycle or make it operate in a twist because of poor tightening procedures. Always hold the hose straight while tightening a connection. Either case stresses the hose and causes premature failure and its accompanying extra expense.

Hose distributors know of these pitfalls and can help with installation suggestions, as well as troubleshooting hose problems. The causes of hose problems are usually quite evident to someone who works regularly with hose, even when all he or she sees is the damaged part.

Sizing hydraulic lines

Fluid flow is measured in feet per second (fps), so the type of conduit is irrelevant. Many books have charts that relate gpm to fps for all standard piping systems. Use these charts to pick out the correct size fluid carrier for the required flow.

Pump inlet line (suction line)

Fluid velocity should not exceed 2 to 4 fps. The reason for this recommendation is that the highest possible pressure drop in the pump inlet line is one atmosphere. Actually, no type of hydraulic pump can even come close to this, so most inlet lines never see much more than 3- to 4-psi vacuum. Using velocity higher than 2 to 4 fps dramatically increases pressure -- causing cavitation and pump damage. It is best to use a suction line equal to or larger than the size of the pump inlet being plumbed. There are circumstances when a smaller suction line is satisfactory, but only do this when absolutely necessary and with the supplier's approval.

The suction line should be full size; as straight as possible; have no or the minimum number of fittings; never include a standard pipe union; and be completely sealed. Using hose in place of pipe or tube can overcome many possible suction-line problems. Hose is a viable alternative and is quite satisfactory if certain precautions are addressed. Always use hose designed and specified for suction (vacuum) use. Hose normally used for pressure may be rated at 3000 psi but is not suited for suction lines. The reason for this is pressure hose uses an inner lining like a tube in a tube-type tire. The outer layers are strong but they are porous and would leak high-pressure fluid except for the inner tube. High-pressure hose as a suction line sees constant negative pressure trying to collapse the inner tube. After some time, it is possible for the inner liner to be drawn in, restricting flow and causing pump cavitation. This phenomenon may not happen immediately, but usually does cause problems in time.

Return lines

Fluid velocity in return lines should be held between 10 and 15 fps. The pump can push oil returning to tank, but any backpressure in these lines must be overcome by extra pressure at the pump outlet. To maintain a high-efficiency circuit, it is important to keep pressure drop in all lines as low as possible. All energy used to push oil through the lines is wasted and converts to heat.

Working-pressure lines

Medium pressure lines (up to 2000 psi) should not exceed 15 to 20 fps. Flow in systems that operate above 2000 psi can go as high as 30 fps. Unlike air systems, there is usually excess pressure capacity in hydraulic circuits when actuators are in motion. Typically, high pressure only comes into play when the actuators near the end of stroke. In an effort to keep line and valve sizes small, it is common practice to use these higher velocities -- but keep in mind this practice wastes energy.

Several fluid power handbooks are available with excellent charts showing tubing and pipe in all different wall thickness, along with flow in gallons per minute (gpm) for all standard sizes. Remember each fitting or valve in the circuit has its own pressure-drop adders and they must be taken into consideration as part of the overall pressure-loss picture.

Quiz

Chapter 4: ISO Symbols and Glossary

A family of graphic symbols has been developed to represent fluid power components and systems on schematic drawings. In the United States, the American National Standards Institute (ANSI) is responsible for symbol information. The Institute controls the make-up of symbols and makes changes or additions as required. The International Standards Organization (ISO) has the same control over symbols used internationally. Both systems have almost the same format (especially since ANSI changed its symbols in 1966 to eliminate all written information).

Standard symbols allow fluid power schematic diagrams to be read and understood by persons in many different countries, even when they don't speak the same language. Either symbol set (ANSI or ISO) may be -- and is -- used in the United States. However, many companies today use the ISO symbols as their standard for work with foreign suppliers and customers.

The following pages go through all standard ISO symbol information as it applies to hydraulic and pneumatic schematics. There are still many plants that modify the standards to suit some individual's taste. This widespread practice may be confusing to novices. Symbols have been developed to represent most of the available fluid power components. However, some parts must be made up of combinations of different symbols to show how they function. Other times there is no standard symbol and one must be made up. In such cases, look first in the supplier's catalog for the symbol they show. If the supplier did not make a symbol, the only other option is design one for the new part. Try to design the new symbol using standard practices shown here.

As the phrase fluid power implies, these symbols cover both hydraulic and pneumatic components. Any exceptions are noted.

Read more about hydraulic symbology in a series from author Josh Cosford.

Chapter 5: Basic Fluid Power Systems

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 17 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.

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.)

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

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 with this manual and can be seen in Chapter 4.

Quiz

Chapter 6: Reservoirs, Heaters, and Coolers

Fluid power reservoirs

Fluid power systems require air or a liquid fluid to transmit energy. Pneumatic systems use the atmosphere -- the air we breathe -- as the source or reservoir for their fluid. A compressor takes in atmospheric air at 14.7 psia, compresses it to between 90 and 125 psig, and then stores it in a receiver tank. A receiver tank is similar to a hydraulic system’s accumulator. A receiver tank, Figure 6-1, stores energy for future use similar to a hydraulic accumulator. This is possible because air is a gas and thus is compressible. A receiver tank is a pressure vessel and is constructed to pressure vessel standards. At the end of the work cycle the air is simply returned to the atmosphere.

Hydraulic reservoirs

Figure 6-1. Simple pneumatic power unit.

Hydraulic systems, on the other hand, need a finite amount of liquid fluid that must be stored and reused continually as the circuit works. Therefore, part of any hydraulic circuit is a storage reservoir or tank. This tank may be part of the machine framework or a separate stand-alone unit. In either case, reservoir design and implementation is very important. The efficiency of a well-designed hydraulic circuit can be greatly reduced by poor tank design. A hydraulic reservoir does much more than just provide a place to put fluid. A well-designed reservoir also dissipates heat, allows time for contamination to drop out of the fluid, and allows air bubbles to come to the surface and dissipate. It may give a positive pressure to the pump inlet and makes a convenient mounting place for the pump and its motor, and valves.

Some standard reservoir layouts

Pump on top. Figure 6-2 shows this common reservoir/pump layout -- used by many suppliers. The flat top surface of a standard reservoir makes a perfect place to mount the pump and motor.

Figure 6-2. Pump and motor mounted on top of tank.

The main disadvantage to this configuration is that the pump must create enough vacuum to raise and accelerate the fluid into the pump inlet. For most pumps, this is not a big problem, but it is not the best situation for any of them. Axial or in-line piston pump life can be adversely affected by medium to high vacuum at its inlet when using this layout. The piping in this configuration must be sealed, should be as short as possible, and have few or no bends.

Pump alongside tank. Figure 6-3 shows another design that is satisfactory for any type pump. (Many suppliers prefer this layout.) This arrangement is sometime called a flooded suction, because the pump inlet always is filled with fluid.

Although the pump inlet always has fluid, there will be some vacuum in the inlet line when the pump is running. A pump with its inlet below fluid level no longer has to raise the fluid, but it does have to accelerate and move it. However, this design is far better than the pump on top and can extend the service life of any type pump.

Figure 6-3. Pump and motor mounted alongside tank.

Notice the shutoff valve in the inlet line. This valve allows maintenance work to be done on the pump without draining the tank. Some precautions: install a free-flowing valve (such as a quarter-turn ball type) and use a valve with a limit switch to indicate full open. Wire this limit switch in parallel with the pump motor starter, so the pump cannot start until the shutoff valve is open.

Pump under tank. Figure 6-4 shows the very best pump/tank layout. This design puts the pump below the reservoir to take advantage of static head pressure. As explained in Chapter 1, there is pressure at the bottom of any column of fluid (about .4 psi per foot of elevation). With the tank above, the pump not only has fluid at its inlet all the time, but this fluid also could be at 2- to 4-psi positive pressure. (Note that this arrangement can be difficult to work on without ample headroom for the mechanic.) The same shutoff valve precaution goes for this layout as mentioned for the pump-alongside design.

Tank functions

The main reason the reservoir exists is to store fluid. The accepted rule for sizing a tank is: the tank volume should be two to four times the pump flow in gpm. This is only a general rule. Some circuits may require more volume, while less fluid may be adequate for other circuits. A 25-gpm pump would work well with a 50- to 75-gallon reservoir for most circuits. With this general rule, the returned fluid theoretically will have two to three minutes in the tank before it circulates again. As Figure 6-5 shows, a baffle separates the return line from the pump inlet line, forcing the fluid to take the longest possible path through the reservoir before returning to the pump inlet. This arrangement also mixes the fluid well and provides more time to drop contaminates and de-aerate. In addition, the fluid spends more time in contact with the outer walls of the reservoir to dissipate heat.

Figure 6-4. Tank located above pump and motor.

When a circuit has single-acting cylinders or cylinders with large rods, the volume of fluid returned on the extend stroke is greatly reduced – or even non-existent. In these cases, the tank must be larger than the general rule states to keep fluid level from falling below the pump inlet line.

Another situation where a tank may need to be larger is if the circuit has accumulators. Accumulators need fluid to fill them at start up and space into which to discharge this fluid at shut down. An undersized reservoir may not have enough fluid to keep the pump inlet covered at all times.

Another case for making the tank larger than the general rule is to add cooling capacity. All of the tank’s exterior walls can radiate heat to the atmosphere, so the larger the tank the greater the heat dissipation. Use the formula in Figure 6-6 to figure tank-cooling capacity. An example problem is shown later in this chapter. Several data books include formulas and charts showing tank-cooling capacity. These also can be used in conjunction with this manual.

Heat dissipation is the main reason for having the tank bottom off the floor and why it is important not to stop free air flow around the tank. It is not good practice to enclose a power unit to reduce noise.

Tank components

The filler-breather cap should include a filter media to block contaminants as the fluid level lowers and rises during a cycle. If the cap is used for filling, it should have a filter screen in its neck to keep large particles out. It is best to pre-filter any fluid entering the tank . . . either by a filter cart transfer unit or by a filter fill unit (as shown in Chapter 2, Figure 2-2.)

Figure 6-5. Standard features of non-pressurized reservoir designed for pump to be mounted on top.

Remove the drain plug to empty the tank when the fluid needs be changed. At the same time, the clean-out covers should be removed to provide access to clean out all residue, rust, and flaking paint that may have accumulated in the tank (and doesn’t flow out with the fluid). If this is not done, the new fluid gets dirty immediately — defeating the purpose of the fluid change. In the design in Figure 6-5, the clean-out covers and internal baffle are assembled together, with some brackets to keep the baffle upright. Rubber gaskets seal the clean-out covers to prevent leaks.

If the system is badly contaminated, it is wise to flush all pipes and actuators while changing the tank fluid. This can be done satisfactorily by disconnecting the return line and placing its end in a drum, then cycling the machine. Do not over-fill the drum during this operation or it may rupture and spill fluid.

Sight glasses make it easy to visually check fluid level. Calibrated sight gauges provide even more accuracy. If the sight glass or gauge is difficult to see or is damaged, find another way to check the fluid level.

Many sight gauges include a fluid-temperature gauge. Tanks that feel hot to the touch may actually be within operating range. The temperature gauge gives a more specific indication. On older systems where the temperature gauge may have stopped working, it’s best to check fluid temperature with some other method.

Pump connections

The inlet line to the pump should be at the same end of the reservoir as the return line, with the baffle between them forcing returning fluid to travel to the opposite end of the tank and back. The inlet must be below the fluid level and may include a strainer. If the inlet line is just a straight piece of pipe installed vertically, it is best to cut the line’s open end at 45° so it is impossible to butt it against the floor of the tank, which could block flow. Outside the tank, this line should lead as directly to the pump as possible, with no unnecessary bends or connections. Never use a pipe union in the inlet line; unions are almost impossible to seal against air leaks. Even a minute leak in the inlet line can cause pump cavitation and all of its problems.

The return line should be located in the same end of the tank as the inlet line and on the opposite side of the baffle (as shown in Figure 6-5). The return line must terminate below fluid level to reduce turbulence and aeration. The open end of the return line also should be cut at 45° to eliminate the chance of stopping flow if it gets pushed to the bottom. Another good practice is to point the opening toward the side wall to get the most heat-transfer surface contact possible.

When a hydraulic reservoir is part of the machine base or body, it may not be possible to incorporate some of the features discussed in this chapter. However, keep in mind the different functions mentioned to try to eliminate ongoing problems.

Non-pressurized and pressurized reservoirs

Reservoirs are seldom pressurized because that feature is not required under most circumstances. One reason for using a pressurized reservoir is to provide the positive inlet pressure required by some pumps -- usually in line piston types. Another reason is to force fluid into a cylinder through an undersized prefill valve. Both of these reasons may require pressures between 5 and 25 psi and could not use a conventional rectangular reservoir design.

Figure 6-6. Formula for estimating how much heat a tank of a given size can dissipate.

Another reason for pressurizing a tank is to keep out contaminates. If the reservoir always has a positive pressure in it there is no way for atmospheric air with its contaminants to enter. Pressure for this application is very low – 0.1 to 1.0 psi -- and may be all right even in a rectangular design tank.

A pressurized reservoir would be built like any pressure vessel, but with the baffling and other features shown in Figure 6-5. Note however that the reservoir pictured in Figure 6-5 is non-pressurized. The symbol for this type tank is shown at the left. The symbol also indicates how lines that terminate above and below fluid level are shown. If a drain line comes from an area that might have suction part of the time, it might not be best to terminate it below fluid level. If this type line terminated below fluid level, it could suck oil into the unit and possibly cause sluggish action. The drain line from the case drain of a pressure-compensated piston pump and an air bleed valve should terminate below fluid level at all times. This keeps air from being sucked in and causing problems.

Heat in hydraulic systems

All heat in a hydraulic circuit comes from wasted energy. Any horsepower put into the circuit that does not do useful work wastes energy.

Any circuit has inefficiencies that can be up to 15% of input power. This is bypass fluid in pumps, valves or other components and pressure drop through these components and the flow lines. These losses can be reduced, but never completely eliminated in a typical hydraulic circuit. Some ways to reduce inefficiency losses is to correctly size piping and valves, keep working pressure at or only slightly higher than required for all operations, and never allow fluid to relieve to tank. Flow controls also generate heat because they restrict flow. Reducing valves, counterbalance valves, and sequence valves also waste energy... especially if they are not set correctly. A pressure sequence or a counterbalance valve will do its job even when set too high, but will waste more energy at an unnecessarily high setting.

Each lost horsepower generates 2,545 BTU/hour of heat. To put this in perspective, 10 hp would heat a three-bedroom home when the outside air temperature is 30°F. Thus it is obvious what that much heat would do to the temperature of a 20-gallon tank of hydraulic fluid.

In a hydraulic circuit, you must calculate wasted horsepower to determine heat generation. In a highly efficient circuit this figure could be low enough to use the reservoir’s cooling capacity to keep maximum operating temperature below 130°F. If heat generation is slightly higher than a standard tank will handle, it may be best to oversize the tank rather than adding a heat exchanger. An oversize tank is less expensive than a heat exchanger; and avoids the cost of installing water lines.

It is easy to figure heat generation by figuring horsepower input and subtracting horsepower output. With the gauge at a fixed-volume pump’s outlet reading 150 psi and a gauge at the working cylinder reading 125 psi, there is a 25-psi pressure difference between energy in and work out. To figure horsepower loss, multiply (0.000583)(gpm)(psi). For this example, assume a 40-gpm pump. Then, lost horsepower = (0.000583)(40)(25) = .58. To determine actual heat loss, this figure must be divided by the percentage of the cycle that it occurs. If this figure is from the cylinder extending and the extend time was four seconds during a total cycle of 12 seconds, then figure 1/3 of the .58 hp or .19 hp as waste. Do this operation for each actuator in both directions of travel to determine total wasted horsepower. After all actuator losses are calculated, add them together to determine total wasted horsepower. (Note that when this horsepower total is less than the answer from the formula in Figure 6-6, no heat exchanger is required.)

The example just cited would be a straightforward circuit without any flow controls or other added restrictions. With a flow control in the circuit, pressure drop would be much higher and energy waste would increase drastically. Most circuits using a pressure-compensated pump would have flow controls so the pump would be at compensator setting while the actuator would be at whatever pressure is required. A circuit without flow controls or other restrictive plumbing usually has low energy losses. This type system may get by without a heat exchanger when ambient temperature stays below 110°F.

Tank cooling capacity

Use the formula in Figure 6-6 to estimate how much heat a given tank can dissipate. This formula assumes the tank is open on all sides with free air movement around it. (Remember, pipes, cylinders, and valves also have surface areas that can dissipate heat but those areas are usually not included when calculating cooling capacity.) For a 100-gal tank and a 30°F. temperature difference, this amounts to about 1.4 hp. A single cylinder circuit wasting only 0.7 hp would not require a heat exchanger on a 100°F. day while maintaining a maximum fluid temperature of 130°F.

Heating and cooling devices

Tank heaters: Most industrial hydraulic units operate in a warm indoor atmosphere so low temperature is not a problem. For circuits that see temperatures below 65° to 70°F., some sort of fluid heater is recommended. The most common tank heater is an electric-powered immersion type. These units consist of resistive wire in a steel housing with a mounting option. They often have an integral thermostatic control. The heating element on most units contacts the fluid directly, similar to a hot water heating system. Some immersion heaters in a sealed housing heat the air in the enclosure. The air then transfers heat to the fluid. This heater-in-tube arrangement allows the heating unit to be changed without draining the reservoir.

Figure 6-7. Typical electric-powered in-tank heater.

Figure 6-7 shows an electric immersion heater and its ISO symbol for schematic drawings. (This symbol may be replaced by some sort of pictorial rendition of the heating unit on many schematics.)

Electric heating units should not have concentrated heat like those used to heat water. Oil viscosity at low temperature is thick enough to retard movement. This could allow fluid next to the heating rods to overheat and possibly breakdown. The usual recommendation is for the heating rods not to have over 8 to 10 watts per square inch density. This limit may require multiple heating rods to meet the heat requirement of some systems.

Another way to electrically heat a tank is with a mat that has heating elements similar to those in an electric blanket. This mat is attached to the outside of the bottom of the reservoir and adds heat during low temperature conditions. This type heater requires no ports in the tank for insertion. It also evenly heats the fluid even during times of low or no fluid circulation.

Figure 6-8. Formula for estimating heater capacity to increase fluid to a minimum temperature.

Heat can be introduced through a heat exchanger by using hot water or steam in place of the cooling water. The exchanger becomes a temperature controller when it also uses cooling water to take away heat at other times. Figure 6-10 depicts the symbol for this type heat exchanger. (Chapter 7 shows an alternate way to add heat while filtering system fluid.) Temperature controllers are not a common option in most climates because the majority of industrial applications operate in controlled environments.

The formula for estimating how many kilowatts are needed to heat a certain size tank from an expected minimum ambient to a nominal working temperature of 50° to 70°F. is shown in Figure 6-8. Use it to size a heater when the tank is exposed to low temperatures.

Heat exchangers

Cooling hydraulic systems is necessary more often than heating them due to wasted energy from inefficiency and/or poor circuit design. A well-designed circuit eliminates most heat generation and may not need a heat exchanger. Use the same method to estimate how much heat a system generates as was used for the previous tank-cooling example.

When figuring wasted horsepower, see if there is any way to reduce or eliminate it so it does not have to be paid for twice. It costs money to produce the unused heat and it is expensive to get rid of it after it enters the system. Heat exchangers are expensive, the water that through them is not free, and maintenance of this cooling system can run high.

Items such as flow controls, sequence valves, reducing valves, and undersized directional control valves can add heat to any circuit. Are these items absolutely necessary? Can they be replaced with another valve or part that does the same thing with less pressure drop? Anytime these questions can be answered with a yes, the circuit is not ready to build.

Air-cooled heat exchangers

After calculating wasted horsepower, review heat exchanger manufacturers’ catalog to pick out a unit that will dissipate that amount of energy. Most catalogs include charts for given size heat exchangers that show the amount of horsepower and/or BTU they can remove at different flows, oil temperatures, and ambient air temperatures.

Figure 6-9. Typical radiator-and-fan oil cooler.

Figure 6-9 shows a typical air-cooled heat exchanger that may be used in place of a water-cooled unit in some applications. Air-cooled heat exchangers are not as efficient as water-cooled heat exchangers, but they require only an electrical outside hookup. They work well in cool atmospheres or when the amount of heat to be removed is low. Note that airborne contaminants such as heavy dust or water and coolant vapors can quickly reduce an air-cooled heat exchangers low efficiency to almost nothing. Some manufacturers offer a filter pack to take out airborne contamination before it clogs the heat exchanger’s radiator fins and tubes.

On circuits with pressure-compensated pumps, a small air-cooled heat exchanger often is used to cool case drain flow. On this type system, most of the heat generation is from internal leakage and control oil that flows to tank through the pump case drain. One type heat exchanger -- called a coupling cooler -- is a finned tube formed into a circle and wrapped around a blower that is driven by the motor turning the pump. A similar arrangement uses a small flat radiator attached to the intake end of the fan-cooled electric motor that drives the pump. Both units are low-flow, low-backpressure devices and dissipate only a small amount of heat.

Some systems use a water-cooled heat exchanger in the summer and an air-cooled one in the winter. This arrangement eliminates plant heating in summer weather and saves heating expense in the winter.

Water-cooled heat exchangers

Two popular water-cooled heat exchanger designs. As before, manufacturers’ catalogs will assist you in picking out a unit to dissipate the amount of wasted heat energy. For water-cooled heat exchangers, catalogs ask for information such as how much and what temperature water is available, how many horsepower or BTU of energy must be dissipated, what is the fluid flow in gpm, and how many passes will the water makes to get through the body. The more passes -- up to four maximum usually -- the greater the heat dissipation per gallon of water flow. Charts that use this information make it easy to pick the correct size heat exchanger.

Figure 6-10. Two popular water-cooled heat exchanger designs.

Figure 6-10 shows two types of water-cooled heat exchangers commonly used for hydraulic systems. The shell-and-tube design is the most common one at present, but the plate-and-frame or brazed-plate type are coming along because they are much smaller and easier to maintain. It is important to use clean water in either type unit to keep from building up insulating deposits or corroding the tubes until they leak water into the oil. Treated water from a cooling tower works best.

Figure 6-11. Temperature-controlled water valve.

Either type heat exchanger should have a thermostatic control to turn on the water or fan only when the fluid temperature rises above its normal operating range. Without a thermostatic control, fluid could be too cold and thick, while wasting the energy to operate the heat exchanger. Figure 6-11 shows one type of water-control valve that requires no electrical hookup. Heat-sensitive liquid in a thermometer-type probe in the tank expands and opens a water valve upon reaching a preset temperature. The temperature is fully adjustable to meet any requirement and it operates in all types of fluid. Another option is an electrically operated temperature sensor that controls a solenoid-operated water valve. This installation requires an electrical hookup but is able to maintain a fluid-temperature range at any desired setting.

Quiz

Chapter 7: Air and Hydraulic Filters

Fluid power filters

Moving parts in fluid power systems are subject to wear from contamination. Neither atmospheric air nor hydraulic fluids are clean enough as supplied to avoid this and they both become more contaminated with use. Therefore all fluid power systems require filters to remove contamination and thereby increase component life.

Pneumatic filters and lubricators

A pneumatic filter should be the first component at the inlet of most air circuits. This unit usually is one part of a combination of components that filters the air, regulates its pressure, and adds lubricants for moving parts in the circuit. The air filter and lubricator are covered in this section. (An air line regulator performs the same function as a hydraulic reducing valve and is covered in that section.)

Air from the compressor contains dust from the ambient atmosphere, condensed water, and rust and oil sludge that bypass the compressor rings. These by-products of compressing and transmitting air must be removed to keep moving parts of the machine working properly. Most filters clean the air and separate condensed water from it before the air enters the circuit.

7-1. Cross-section of typical air filter, with ISO symbols.

Figure 7-1 shows a simple air filter (and the ISO symbols that represent it). These units usually require little attention if the compressor has an air dryer at its outlet. Air enters at the left and is channeled into the bowl with a downward circular motion. The centrifugal force of this swirling action slings water droplets outward. They collect and fall to the bottom of the bowl below the baffle into a quiet zone for draining either manually or automatically. The air then flows through a porous filter element to the outlet. These units typically remove particles of 40-micron (40-µ) size or larger but they also are available for particles as small as 5 µ if required.

7-2. Cross-section of coalescing air filter, with ISO symbols.

The coalescing filter shown in Figure 7-2 removes water and oil vapors as well as condensed moisture from an air line. To accomplish this, coalescing filters are not only designed differently but they have reverse flow in relation to a standard filter. These filters will remove particles as small as 0.3 to 0.6 µ. They use a coarse mat of very fine fibers that are small enough to catch and hold these very fine particles. As the vapor collects into droplets, they are channeled to the quiet zone and drained.

Coalescing filters must be applied according to manufacturers’ specification to keep collected liquid from being re-entrained into the air stream. Also, always make sure the flow direction is according to information on the filter housing. Several companies make these high-efficiency air filters though they are not often applied to every day circuits.

7-3. Cross-section of typical air-line lubricator, with ISO symbol.

Figure 7-3 shows a cutaway view of an air-line lubricator. After the combination unit filters and regulates air pressure, some downstream system components may require a small amount of lubrication. (For example: air motors are one item that needs a constant supply of oil to extend their life and maintain torque.) Some cylinders are pre-lubed and most valves require little if any lubrication, so keep oil supply to these units at minimum. A general rule: a ½-pint bowl of oil in a lubricator should last three weeks to a month in most situations.

When air passes through the lubricator’s venturi section, pressure drop across it gives a negative pressure in the area below the adjustable orifice. Vacuum in this area draws oil from the bowl as fast as the adjustable orifice will allow. These droplets then mix with the air as it passes through. This arrangement means that oil flows only when there is air flow and only as fast as the adjusting screw allows.

Air line lubricators are designed to send a mist of oil to the parts in the downstream circuit. However, the physical size of some circuits makes it impossible for the mist to stay in suspension long enough to reach some parts. In this case -- and for some air motor applications -- it is necessary to inject oil at the components inlet. There are electric and air-driven lubrication units to meet the needs of these applications.

Compressed-air dryers

Why are air dryers necessary? All air compressors take in atmospheric and compress it eight or ten times to a pressure between 110 and 125 psig. All the moisture vapor and heat in the atmospheric air is also compressed and concentrated, so air at the compressor outlet is hot and wet. The temperature of the air at the compressor outlet can be as high as 350ºF. The air also can be saturated with water vapor. As this air cools in the receiver and plant piping, the water vapor in it condenses into water droplets.

A 25-hp compressor running at 75% capacity pumps 18 gallons of water into a plant air system on a day with average ambient humidity. An aftercooler condenses and removes about 11 gallons of this liquid, but that still leaves 7 gallons to collect in low spots, retard valve movement, damage production parts, and cause problems in general.

Types of air dryers

Figure 7-4 shows a cutaway view of (and the symbol for) a deliquescent air dryer. Wet air enters the dryer (which is a pressure vessel), passes up through a bed of hygroscopic chemicals, and flows on to the outlet. The chemicals (often a form of sodium) absorb moisture from the air as it passes through the bed. As they dry the air, the chemicals break down into a slurry of water and chemical drops that falls to the bottom of the tank. A manual or automatic drain keeps the water mixture from rising too high and mixing with inlet air flow.

7-4. Absorbent-type deliquescent air dryer.

A typical deliquescent dryer removes moisture to a dew point of about 40ºF. Air that has passed through this dryer must be cooled below 40ºF before any more moisture will condense. This dew point is satisfactory for most plants -- even during winter cold. However, the hotter the air passing through the chemicals, the less the amount of moisture the chemicals will collect. Thus it is important to keep the incoming air at or below 100ºF. This usually requires an upstream aftercooler to lower the temperature of the air being delivered from the compressor. Air at higher temperatures at the inlet results in higher dew points at the outlet.

Deliquescent dryers are the least expensive of the three types mentioned in this section, but they might cause problems in some installations. There is always a chance of the chemicals or their vapors being picked up by the air stream and sent into the pneumatic system. These chemicals are corrosive and can damage internal parts. Also chemicals must be replenished on a regular basis. This means shutting down the compressor or bypassing the dryer when chemicals get low. Finally, slurry must be removed.

7-5. Condensing-type refrigeration dryer.

Figure 7-5 shows a cutaway view of a refrigeration-type dryer. This unit condenses water vapor by cooling compressed air to lower its dew point. There are no chemicals to replace or break down, so the system can run as long as required. A refrigeration dryer is more complex and more expensive to purchase than chemical dryers, but has a lower operating cost. The main operating cost is electricity. Maintenance cost is minimal and life expectancy is high, so this type dryer can be cost effective over the long haul.

To keep operating cost down these systems usually have an upstream water-cooled aftercooler to cool compressor air and take out the bulk of the moisture. Normally a maximum incoming temperature of 100ºF is recommended. When a higher inbound temperature is present, oversize the refrigerated dryer to handle the extra energy removal. (Of course, this adds extra cost to the purchase of the unit, as well as increasing operating cost over the life of the system.)

Wet air enters the unit through an internal air-to-air heat exchanger. This unit is piped to pre-cool the hot inbound air and re-warm the cold dry air before it exits the dryer to enter the plant piping. This arrangement saves energy by removing some of the heat in the incoming air. It also keeps the plant piping from condensing water vapor from ambient air on its cold exterior surfaces and dripping water on the production floor.

The pre-cooled air then passes through a Freon-to-air heat exchanger that reduces its temperature to approximately 35ºF. This procedure condenses more water vapor to achieve a 35ºF dew. If plant temperature stays above 35ºF (as it does in most plants), there will be no more condensation inside the air piping. The condensed water drains from the dryer through a water separator.

The main potential problem with refrigerated dryers is they will freeze up if temperature is set too low or if the amount of air passing through them is low and intermittent.

Another approach to refrigerated drying is drying the air before it enters the compressor. One company offers chillers that take atmospheric air down to –40ºF. and feed it to the compressor. After it is compressed to 100 psi, the air has a pressure dew point of approximately 35ºF. Because the compressor takes in air that is denser and at such a low temperature, the heat of compression is negligible. Also, most of the airborne contaminants are removed during the cooling, condensing, and freezing process. These inlet air dryers use dual refrigeration units. While one is drying input air, the wasted heat from the drying unit defrosts the other.

7-6. Adsorbing-type desiccant air dryer.

Figure 7-6 pictures a twin-tank desiccant dryer that uses a hygroscopic material (such as silica gel or activated alumina) that collects water vapor but is not broken down by it. This type dryer is called an adsorber because it collects water vapor but once the moisture is removed by heat or other methods, the chemical is ready to work again. This desiccant dryer may achieve dew points of –40ºF or lower, so the air can be used in most outdoor circuits without fear of freeze ups.

Electric heaters and purge air from the opposite tank handle the drying process. Other methods are steam heat, dried purge air only, and desiccant replacement.

As wet compressed air enters the control valve, it is channeled to one of the desiccant tanks. (The control valves are usually set up to shift automatically, triggered by a signal from a device that monitors the output air’s dew point.) Wet air is forced through the desiccant material to take out water vapor, then sent on to the plant. Some of the dried air is diverted through an orifice to the spent tank, where it is heated and then passed through the wet desiccant in the other tank. Water again vaporizes and exhausts to atmosphere. This drying process continues at the rate necessary to maintain the required dew point.

Desiccant dryers are the most expensive type to operate and maintain. In particular, they are subject to failure if the incoming air contains carry-over compressor oil. The oil can coat the desiccant and make it incapable of collecting moisture. The main expense of operating these dryers is the energy used to dry out the idle tank of desiccant.

Hydraulic filters

In Figure 7-7, a schematic drawing of a hydraulic circuit shows filters in the standard locations, with typical filtration ratings listed next to them. Note that most circuits would not have all of these filters, but every circuit should have adequate filtration to protect the pump, valves, and actuators from contamination.

7-7. Relative size of contamination particles at 500X magnification.

There are several sources of contamination in and around hydraulic units. Normal component wear, contamination in new oil, sloppy filling practices, poor plumbing installation, and dirt carried in on piston rods are the main ones. Some of these areas are simple to address ahead of time, while others can only be handled by filtration.

New oil from the supplier is not as clean as most hydraulic circuits require. At best about 25-µ cleanliness is all most suppliers will offer. One reason: the drums that new oil comes in probably were used before. Although they were cleaned, they are not contaminant-free. In addition, a standard drum pump draws fluid from the bottom of the drum where all the residue has settled.

Figure 7-8 shows a manually operated 3-way valve installed in the suction line of an offline filter pump. When this valve is shifted, all new fluid has to go through the offline filter before it enters the tank.

7-8. Schematic drawing of hydraulic circuit showing typical locations for filters.

Another item in Figure 7-8 is a tank-fill filter with no bypass, which filters all new oil entering the tank. Another option is to use a purchased filter cart to fill the tank through its normal fill port.

Whatever the method, it is important to keep filling practices from introducing contamination. While the fill port offers one of the simplest places to address contamination problems, it often is overlooked.

Another way that contamination enters a hydraulic circuit – even before startup -- is through poor plumbing practices. All pipe, tube, and hoses should be inspected for contamination before installation and cleaned if necessary. It’s good practice to seal clean conductors with caps until installation time. Use care when cutting and preparing pipe or tube ends to make sure no metal chips or filings stay in the conduits. If the system includes servovalves, flush it with filtered oil through flushing covers for the time recommended by the valve manufacturer before startup. Use every possible precaution on a new or replacement plumbing system to make startup go smoothly.

In a running circuit, one of the most common sources of ingressed contamination is the cylinder piston rods. Every time a piston rod extends, it is damp with system oil. In a dusty atmosphere this oil-dampened piston rod attracts and holds fine particles. Many of these are dragged back into the cylinder and washed off. Most cylinders have a rod wiper to help keep out contamination but this wiper only catches large pieces. Everything smaller passes by. This type contamination must be filtered out continuously to protect system components. Some suggest flexible boots or bellows over the rod end of the cylinder to eliminate rod contamination ingression. A flexible bellows does a good job unless it gets a tear or other type hole. At this point it actually sucks in ambient air with its contamination and holds it closely to the rod.

A similar situation takes place at the tank breather. During each cycle, the fluid level in the reservoir changes -- either drawing dirty atmospheric air in or discharging air through the breather. The filler breather should be capable of trapping the inbound contaminants in this air flow.

The other main cause of contamination is normal component wear. Dynamic pumps, motors, and cylinders have constantly rubbing metal-to-metal contact areas. Even with good lubrication, small eroded particles get into the fluid. This contamination must be constantly captured by filters to eliminate a damaging buildup of residue.

Condensed water is another form of contaminant that should be addressed. Water in hydraulic fluid can cause corrosion, break down fluid additives, and make viscosity vary. Water vapor often enters the system through the breather and condenses to liquid form when the tank cools. Using a breather with a hygroscopic media can eliminate most water contamination. Most of these hygroscopic breathers must be changed when they become saturated. They often use silica gel that is blue tinted when dry and turns a pink tint as it gets wet, signaling that a change is needed.

Cleanliness levels

The circles in Figure 7-7 indicate relative sizes of contamination particles. Not all contamination is nice round marble like pieces as the two overlaid examples show. Particle size is measured by enclosing it in a circle until it touches at two or more places. The 150-micron (150-µ) particle may only be 20 to 25 µ thick but is considered 150 µ because of its length.

Most pump manufacturers specify at least 10-µ clean fluid to protect their pumps from premature failure.

ISO has set up cleanliness level standards for hydraulic and lubrication fluids. The system most used for hydraulics is based on the ISO 4406 standard. This standard covers the number of particles of a given size that can be present in a fluid sample in three different micron ranges. It is designated ISO Code XXX/XXX/XXX, where the numbers relate to the minimum and maximum number of particles of a given size that can be present in a 100 milliliter sample. The first number indicates how many particles of 2-µ size can be present; the second number is for 5-µ particles; and the third number is for 15-µ particles. It might be written ISO Code 18/16/13. The numerals always descend in value from left to right. This code would mean that there could be between 1300 and 2500 particles of 2-µ size, 320 to 640 particles of 5-µ size, and 40 to 80 particles of 25 µ.

Figure 7-7 indicates that, even with good eyesight, people cannot see a particle smaller than 40µ. Thus it is impossible to look at a sample of fluid and determine whether it is clean. Of course, it is possible to tell it is contaminated when large particles are plainly visible in the sample.

The ISO 4406 chart shows the range number and the number of particles that it represents. From this chart it is easy to set up or pick out any ISO Code cleanliness level.

Some typical fluid cleanliness level ISO Code’s are shown in the following chart:

Beta ratios

Another rating applied to hydraulic filters is the Beta ratio (also known as the Filtration ratio). It is a measure of the particle-capture efficiency of a filter element. The ISO 4572 Multipass Test passes fluid through the circuit shown in Figure 7-7 to check for contaminant retention. A measured amount of contaminant is injected upstream of the filter. Laser particle counters record the number of particles into and out of the filter. When 100,000 particles are measured upstream of a 10-µ filter and 10,000 downstream, it would have a Beta ratio of 10 (100,000/10,000 = 10).

Component ISO chart

A Beta Ratio number is of no use alone, but it is required to find the filter’s efficiency rating. Efficiency of a filter element is what counts when comparing one filter to another. The higher the efficiency, the fewer contaminants will pass through it. Efficiency coupled with the volume of contaminant retention can make a more expensive filter cost less due to its longer useful life.

Efficiency is calculated by the formula:

Efficiency10 = (1–1/10) X 100
Efficiency10 = (0.9) X 100
Efficiency10 = 90%

Always make sure the filter meets or exceeds the desired cleanliness level of the system it is protecting.

Filter locations

Figure 7-8 shows most of the locations where filters might be found in any hydraulic circuit. Note that all of these filters are seldom found in a single circuit but some circuits might have filters in two places. Two other types of filtration -- off-line and filter fill -- also are shown in Figure 7-8. Any of these filters could be the dual, change-on-the-run type when required. Dual filters are more expensive but can reduce downtime.

Suction strainers

A suction strainer (as shown at the lower left in Figure 7-8) often is found on the pump inlet line. Strainer is a common term for filters with openings of 75 m1 or larger. Strainers on the pump inlet line protect the pump from large, damaging contaminant particles that could cause catastrophic failure. These particles might be startup debris left in the tank and piping or large contamination introduced to the system from external sources or from internal part failure.

Pumps without supercharged inlets can only tolerate a portion of one atmosphere pressure drop without affecting inlet flow. With this low pressure drop, (14.7 psi maximum at sea level on an average day), a restriction such as a low-micron filter can cause the pump to cavitate. Cavitation causes pump failure faster than dirty oil; so avoid it in every situation. (See the write up on pump inlet conditions and cavitation in Chapter 8.)

Suction strainers are available with 75- to 150-µ openings. Some manufacturers have inlet filters with ratings as low as 25µ. A low-micron element needs large filtering surfaces to keep pressure drop low. When a pump is force fed by another pump -- sometimes called a supercharging pump -- a low-micron rated element can be used. The supercharging pump forces fluid through a very fine filter to the working pump, thus keeping it from cavitating.

A suction strainer or filter should have a bypass relief valve. Set the bypass to open at a pressure of 1 to 3 psi if the strainer clogs. The reasoning behind this is that the pump will run many hours on contaminated oil, but will fail in a short time with little or no oil. Suction strainers may be located inside or outside the reservoir. Internal strainers are less expensive, but their condition is more difficult to monitor. External strainers are easy to service and often have an indicator to show when the filter starts bypassing. The indicator can be as simple as a vacuum gauge or it might be a vacuum-operated electrical output to a warning light, alarm, or shut-down controller.

Many older circuits have nothing but a suction strainer for filtration. Retrofitting these systems with off-line or kidney-loop filters (discussed later) is advisable.

Return-line filters

Another common location for filters is in the return line. (Figure 7-8 indicates this location at the right center.). A return-line filter keeps most contamination that is caused by part wear or ingestion from getting into the tank, this protecting the whole system. Return-line filter protection ratings typically range from 3 to 25 µ. Obviously, you should select a return-line filter rated for the desired system-cleanliness level or less.

Return-line filters should have integral bypass check valves. If the filter becomes loaded, return oil needs an open flow path to tank until it is convenient to change the filter. Without a bypass, the filter element probably will collapse, or the element housing or seal may rupture.

Typical bypass checks require 10 to 50 psi to open. The bypass pressure should be high enough to stop fluid from going around the filter during normal conditions, but low enough to avoid damaging filter element and its housing seal.

Some designers size return-line filters just large enough to handle the pump’s rated flow. This can cause problems, especially if cylinders in the circuit have oversized rods, or if one cylinder must return one or more other cylinders. For example, if a cylinder has a 2:1 rod diameter, flow to tank while the cylinder is retracting is double the pump flow. Sizing the filter just for pump flow in this case allows contaminated oil to bypass the filter, and may damage the housing or seals. Paper filters can collapse, have holes blown through the element, stop filtering, and never indicate they need replaced. On pleated elements, the pleats can collapse, giving a “loaded element” indication prematurely.

Even with a correctly sized return-line filter, flow through it changes constantly. Steady flow through the element gives the most efficient filtering. If a filter passes constant flow, the bypass valve will not open until the filter fills with contaminants. This means only clean fluid leaves the filter. Visual and electrical indicators are available to show when a return filter is bypassing.

Pressure filters

Another location for filters is in the pressure line (as shown at the left middle of Figure 7-8). These filters are mandatory in systems using servovalves. Servovalves have low contamination tolerance. They have small internal orifices, very close tolerance fits, and must shift rapidly at low pilot pressure differential. A servovalve can stop functioning in as little as two minutes when supplied with oil that is clean enough for a typical hydraulic system. Even when a 3-µ return-line filter is in place, contamination generated by the pump is enough to shut down a servovalve in a short time. (Note that fluid for a proportional-valve circuit often requires the same cleanliness level as a servovalve circuit to maintain fast response and consistent operation.)

Actually, pressure-line filters would be an added advantage for any hydraulic circuit, but high initial and replacement cost limits their use. Pressure-line filter housings must be strong enough to withstand full system pressure. When there is a high pressure drop across the filter, the element must not collapse. These requirements make filter housings and elements much more expensive than other type filters. Pressure-line filters usually have elements with 1- to 5-µ openings. The pressure-line filter should have an absolute rating, or have a Beta Ratio of 50 or higher. A pressure-line filter should not have a bypass. If the filter element clogs, it is better to stop flow to a servovalve than to contaminate it. Visual and electrical clogging indicators are available for most pressure-line filters. They warn of potential clogging so that elements can be replaced well before production speed is affected.

Off-line filtration

Off-line filtration systems -- sometimes called kidney loops or bypass filters -- consist of a separate pump, motor, and filter circuit that takes oil from the reservoir and re-circulates it. The system pumps oil from one end of the tank, passes it through a filter, and returns it to the opposite end of the tank. Figure 7-8 shows an off-line filtration system at the lower right. This arrangement is a good way to provide high-micron continuous filtration. The systems are easy to retrofit to existing hydraulic circuits and offer an excellent way for new installations to get high cleanliness levels.

Off-line filter circuits are usually rated in the 3- to 10-µ range and should sized to filter the volume of fluid in the reservoir every 1 to 3 hours minimum. This low, constant flow rate makes the filter very efficient, never opens the bypass, never causes media channeling, and never blows holes in the element.

When the filter indicator shows a clogged element in the off-line system, the main hydraulic circuit can continue to operate during filter change. Conversely, this off-line filter system can continue to run when the main hydraulic circuit is off overnight or weekends.

System-fill filters

New oil is not as clean as most hydraulic systems require so it should be filtered before use. To do this, introduce new oil to the tank through a pair of shut off valves, or a 3-way ball valve (as shown in Figure7-8 at the lower right) in the suction line of the off-line filter pump. Rotate the 3-way ball valve 180 degrees, hooking the off-line pump’s suction to a flexible hose from the oil drum or fluid container. This setup filters all oil from the fluid container as it fills the reservoir.

Another way to make sure all fluid is filtered before use is through the use of a tank-fill filter (as shown in Figure 7-8 at the lower middle). Here a low-micron pressure filter is installed in the tank wall and provides the only way to fill it. The filling process can only introduce clean fluid to the reservoir.

Additions to a filter loop

An off-line filter circuit also can provide heating or cooling functions. Figure 7-8 shows a bypass circuit with a normally open solenoid relief valve, a high-horsepower motor, a temperature switch, a heat exchanger, and a temperature-controlled water valve or switch. These additions can effectively control temperature while filtering the fluid. To only filter the oil, leave the water or fan turned off and the solenoid relief valve de-energized (or open).

If oil temperature drops, a temperature switch energizes the solenoid on the relief valve, and pressure rises. All electric motor horsepower converts to heat until the temperature switch indicates correct oil temperature. Unlike an immersion-type electric tank heater, the fluid is being circulated, so there are no hot spots.

For every electric horsepower, there will be 2544 Btu/hour heating capacity. After figuring the Btu/hour to heat or maintain minimum temperature, divide by 2544 to calculate the horsepower needed. (The formula for calculating tank heating appeared in Chapter 6.)

If the tank fluid temperature goes over a preset limit, a temperature-controlled water valve opens to send water through a heat exchanger or a temperature switch turns on the fan of an air-cooled heat exchanger. All filtered flow is cooled when the temperature-control device indicates elevated temperatures.

When installed in an off-line filtration loop, the heat exchanger receives constant flow, so it needs no bypass valve. Also, the heat exchanger sees flow even when the system uses a pressure-compensated pump.

Bidirectional pressure filters

The only difference between a bidirectional pressure filter, Figure 7-9, and a standard pressure filter is the four check valves in the housing. These check valves cause oil flow to pass through filter element in the same direction regardless of the direction fluid enters the housing. Another name for a bidirectional filter is last chance filter. It is installed in a working line to an actuator so it must withstand maximum system pressure. Bidirectional filters with 3- to 10-µ ratings are adequate for most circuits.

7-9. Schematic drawing of hydrostatic drive circuit with bi-directional filter.

Closed-loop hydrostatic transmission circuits are one place to use bidirectional filters. The oil can stay in the loop between the pump and motor for long periods. Any contamination in this closed loop continues to cause damage, even with ample filtering of oil in the tank.

Quiz

Chapter 8: Pumps and Accessories

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.

  1. Set the relief valve at maximum pressure.
  2. Set the pump compensator at a pressure that is 200 to 300 psi higher than final system pressure.
  3. Set the relief valve 100 to 150 higher than the final compensator setting.
  4. 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.

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.

  1. Set the relief valve at maximum pressure.
  2. Set the pump compensator at a pressure that is 200 to 300 psi higher than final system pressure.
  3. Set the relief valve 100 to 150 higher than the final compensator setting.
  4. 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.

Quiz

 

Chapter 9: Relief and Unloading Valves

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.

Quiz

 

Chapter 10: Directional Control Valves

Directional control valves are the most widely used . . . and the least understood . . . valves in fluid power circuits. Many people are confused by the schematic symbol representations, and have difficulty understanding the terms ways, positions, and operators. Learning to read schematic drawings is similar to learning a foreign language. To the trained eye, a symbol speaks volumes even when no words are present. This chapter attempts to take away some of the confusion and apparent magic of fluid power schematic drawings, and thus help make designing and maintaining fluid power systems easier.
Directional control valves can only perform three functions:

  • stop or block fluid flow
  • allow fluid flow, and
  • change direction of fluid flow.

This seems to simplify a seemingly complex subject, but remember, many valves may combine these functions. This makes them a little more complicated, but still not rocket-science material.

Check valves

At first glance, the valve type shown in Figure 10-1 does not appear to be a directional control valve. However, check valves do allow flow in one direction and block flow the opposite direction. Use a check valve in any line where back flow cannot be tolerated. Also pilot-operated check valves (discussed in the next section) can be shifted by an external source to allow reverse flow or stop free flow.

Fig. 10-1. Cross-sectional views and symbols for two types of check valves

The cross-sectional views show the standard poppet design used in most check valves. As in most early designs, the symbol still pictures a ball on a seat. Ball check valves work well until they are disassembled for repair or when troubleshooting. As these valves operate, they wear a groove where the ball contacts the seat. If this wear groove is not reinstalled exactly where it was, the valve is no longer leak free. On the other hand, a guided poppet always goes back in the same relationship to the seat and seals easily after reassembly.

It is easy to understand the function of a check valve. Fluid entering opposite a spring pushes against the poppet and spring to move it out of the way. The inline valve has holes around the angled seat face above the body seat to allow flow to pass. The right-angle design pushes the poppet out of the way and fluid flows by with little restriction.

Check valves are almost trouble-free devices. Seldom is one the cause of a problem. Potential problems can be minimized further if the check valves are: right-angle types, screw-in cartridges, or subplate mounted. Note that an inline check valve’s plumbing must be disassembled before the valve can be checked.

Check valves also can control pressure. Almost all check valves use a spring to return the poppet. In most valves, this spring has very light force, because any spring force results in an energy loss and heat. The light springs from most suppliers require about 5 psi to move the poppet against them (some go as low as 1 psi). Some large check valves, when they are mounted vertically, may require no spring because the weight of the poppet causes it to fall onto its seat.

Strong springs give extra resistance to flow so a check valve could replace a relief valve when low-pressure bypass is required. Many manufacturers have check valves with springs that require as much as 125 psi to push their poppets back. These valves work for low-pressure circuits such as a bypass around a low-pressure filter or heat exchanger, or to maintain minimum pilot pressure for pilot-operated directional control valves. When the spring functions as a backpressure or relief valve, the symbol usually shows the spring as part of the symbol.

Fig. 10-2. In-line check valve with orifice drilled through poppet

Another lesser-known use for check valves is as a fixed-orifice flow-control function. Figure 10-2 shows an inline check valve with an orifice drilled through the poppet. The orifice allows free flow in one direction and measured flow the opposite way. The orifice is non-adjustable, so this component is tamper proof. The only way to change actuator speed is to physically change the orifice size. This orificed check valve could protect an actuator that might run away if a line broke or a valve malfunctioned. It will not affect speed in the opposite direction. For this application it should be flange fitted or hard piped directly to the actuator port.

Pilot-operated check valves

The check valves in Figure 10-3 operate like standard check valves, but can permit reverse flow when required. They are called pilot-to-open check valves because they are normally closed but can be opened for reverse flow by a signal from an external pilot supply.

Fig. 10-3. Three types of pilot-to-open check valves with symbols

The first cutaway view of a pilot-to-open check valve in Figure 10-3 is a standard design using a pilot piston with a stem to unseat the check valve poppet for reverse flow. The pilot piston has an area three to four times that of the poppet seat. This produces enough force to open the poppet against backpressure. Some pilot-operated check valves have area ratios up to 100:1, allowing a very low pilot pressure to open the valve against high backpressure.

The second valve in Figure 10-3 shows a pilot-to-open with decompression function. It has a small, inner decompression poppet that allows low pilot pressure to open a small flow passage to reduce backpressure. After releasing high backpressure, the pilot piston can easily open the main poppet for full flow to tank. (This arrangement does not work when the high backpressure is load-induced or generated by other continuous forces.)

The third valve, pilot-operated with external drain, isolates the stem side of the pilot piston from the in free-flow port backpressure that would resist pilot pressure trying to open the poppet. Notice that in the other two cutaway views, any pressure in the in free-flow port pushes against the pilot piston stem side and resists pilot pressure’s attempt to open the poppet. Backpressure could be from a downstream flow control or counterbalance valve in some circuits.

The external-drain port also can be used to make the pilot piston return when using the valve for a pilot-operated 2-way function.

Fig. 10-4. Typical circuit incorporating pilot-operated check valve

The circuit in Figure 10-4 shows a typical application for pilot-operated check valves. Spool-type directional control valves cannot keep a cylinder from moving from a mid-stroke position for any length of time. All spool valves allow some bypass, so a cylinder with an outside force working against it slowly moves out of position when stopped. Installing pilot-operated check valves in the cylinder lines and connecting the directional valve’s A and B ports to tank in center position assures that the cylinder will stay where it stops (unless the piston seals leak).

Fig. 10-5. Circuit with pilot-operated check valve that prevents load from running away if pilot pressure is lost

The circuit in Figure 10-5 shows a pilot-operated check valve holding a load on the rod end of a vertically mounted cylinder. Pilot-operated check valves can hold potential runaway loads in place without creep, but this circuit usually has problems on the extend stroke. This is because a pilot-operated check valve opens the rod end of the cylinder to tank, letting it run away. When the cylinder moves faster than the pump can fill it, pressure in the cap end and pilot pressure to the pilot-operated check valve’s pilot port drops and the valve closes quickly. This can generate high-pressure spikes that may cause pipe and part damage. Almost immediately, pressure to the pilot-operated check valve’s pilot port builds again and the runaway/stop scenario repeats until the cylinder meets resistance or something fails. The best valve to control runaway loads is the counterbalance valve explained in Chapter 14.

Fig. 10-6. Circuit with vertically mounted cylinder that is unable to extend

Figure 10-6 illustrates another problem with using a pilot-operated check valve to hold back a runaway load: a pilot-operated check valve may not open when signaled to let a cylinder with an oversize rod and heavy load extend. When the directional valve shifts to extend the cylinder, load-induced pressure can hold the pilot-operated check valve poppet closed. It may take 300 to 400 psi to force the poppet open, even with its 3:1 or 4:1 area difference. Pressure builds at the pilot port, but at the same time it increases in the cylinder cap end. With a 2:1 rod-differential cylinder, it can add 600 to 800 psi to the load-induced pressure. The additional downward force causes pilot pressure to increase, which causes more downward force, which causes more pilot pressure -- until the circuit reaches maximum pressure. At that point, the relief valve bypasses or the pump compensator kicks in to stop flow. The cylinder simply cannot start to extend . . . and even if it could, the action would be erratic, as in Figure 10-5.

Pilot-to-close check valves

There is also a pilot-to-close check valve, but it is seldom used. It is rarely necessary to have a valve that always stops flow in one direction and also is capable of stopping it the opposite direction.

Fig. 10-7. Pilot-to-close check valve and symbol

Notice in the cutaway view in Figure 10-7 that the spring-loaded poppet does not have communicating holes through it to the spring chamber. Flow passes freely from inlet to outlet until a pilot signal is fed to the pilot port. Because the pilot port side of the main-flow poppet has more area than the inlet side, this valve can be closed against free flow.

Poppet-type pre-fill valves

Pre-fill valves operate similarly to pilot-operated check valves, but they are usually much larger. Some pre-fill valves can handle flows in excess of 6000 gpm at pressure drops of less than 4 to 8 psi. Their normal function is to fill and exhaust a large bore cylinder as it travels to and from contact with the work piece. Large, high-tonnage presses -- both vertical and horizontal -- use pre-fill valves to reduce pump size while maintaining cycle time.

The cutaway view and symbol in Figure 10-8 show the construction of a typical poppet-type pre-fill valve. A large main-flow poppet seals the path between the tank and the cylinder ports. As the piston advances, vacuum in the void behind it allows atmospheric pressure to push the main-flow poppet open so fluid from the tank can fill this void. On the retraction stroke, a signal to the pilot piston pushes the main-flow poppet open so fluid can return to tank. While a pilot-operated check valve’s pilot piston is larger than the poppet it opens, the main-flow poppet in a pre-fill is much larger in diameter than the pilot piston. Thus it is impossible to open the main-flow poppet against high backpressure. This keeps decompression shock from damaging pipes and components.

Fig. 10-8. Two models of pre-fill valves with symbols

Decompression shock occurs when large volumes of fluid at high pressure are released suddenly. Because all hydraulic oil has some entrained air (bubbles so small they cannot be seen without magnification), there is a 0.5 to 1% compressibility that must be dealt with when using large-bore cylinders. On top of fluid compressibility, the cylinder tube may stretch diametrically and longitudinally. In addition, the framework that is resisting the tonnage produced also can stretch. Summing all these factors, a 50-in. bore cylinder with a 72-in. stroke can contain more than 25 gal of extra fluid at 3000 psi. If this trapped fluid suddenly has a large open path to atmosphere, its velocity at first release is such that it can break fittings, blow hoses, straighten tubes or pipe bends with relative ease. Releasing this same trapped fluid in a controlled manner over a few seconds dissipates the excess energy and no damage is seen.

The plain pre-fill valve might be used on smaller cylinders or circuits that have other means for decompressing. The pre-fill valve with decompression has a small poppet in the large poppet that is easy to open at high pressure but will not allow the high flow that causes decompression shock. This decompression poppet usually has a means to adjust how fast the cylinder decompresses.

Another pre-fill valve design is the sleeve type that must be externally shifted open and closed. Both designs give the same results even though their operation is different. (See Chapter 4 for a cutaway view and symbol of a sleeve type pre-fill valve.)

Typical decompression circuit

The circuit in Figure 10-9 operates a vertical single-acting hydraulic ram press with pullback cylinders for the retraction stroke. The press has a poppet-type pre-fill and gets a fast stroke from only filling the pullback cylinders during the approach stroke. A sequence valve keeps pump flow from going to the ram until pressure reaches a preset level.

Fig. 10-9. Typical vertical ram circuit with pre-fill valve

During the approach part of the stroke, atmospheric pressure pushes fluid into the large-bore ram through the pre-fill valve because there is vacuum behind the extending ram. When it contacts the work, the ram stops and the pre-fill valve closes. Pressure starts to rise and when it is high enough to open the sequence valve, pump flow goes to the pullback cylinders and the ram. Extension speed slows and tonnage increases to do the work required.

A signal that the work is complete shifts the directional control valve to send pump flow to the rod ends of the pullback cylinders and to the pilot signal of the pre-fill valve. The pre-fill valve’s pilot piston moves forward and contacts the decompression poppet. This lets trapped fluid flow out at a controlled rate. Pressure in the ram drops quickly and smoothly. When pressure is low enough, the pilot piston opens the main poppet to let fluid from the ram return to tank. When the ram loses pressure, the pullback cylinders can raise the platen and push fluid from the ram back to tank.

General directional control valve terminology

Directional control valves are specified generally by the number of ports or ways (lines attached to the symbol’s box) and the number of positions (boxes or envelopes in the symbol) they have. Other information about them includes whether they are normally closed (not passing fluid), normally open (passing fluid), how they are operated (solenoid, manual, or spring) and other features such as manual overrides, drain ports, pilot ports, etc.
Some general rules for drawing symbols are:

  • only draw flow lines to one box of the symbol
  • always see that flow paths and direction of flow in each box is compatible
  • on 4-way hydraulic valves, pipe the A port to the cap end of the cylinder and the B port to the rod end
  • draw all symbols in their at-rest position. Show valves that are held actuated by a machine member in their shifted condition, and
  • provide information such as pressure settings, flow rates, orifice sizes, horsepower and rpm where applicable.

(According to this method of specifying, check valves and pre-fill valves would be 2-way valves because they have two ports. However, because these valves are basically single function and have infinitely variable flow paths, their symbols and terminology do not follow general directional control valve rules.)

Figure 10-10 shows the symbol for a 2-way directional control valve and how it could function in a circuit. Notice the symbol has two boxes (or envelopes) to indicate two positions. Each position is a flow path. The box with flow lines coming to it is the normal or at-rest position of the valve. The normal or at-rest position is usually at the spring end of a spring-return valve as seen in the figure.

Fig. 10-10. Circuits in which 2-position, 2-way valves operate cylinders

The circuit at rest in Figure 10-10 illustrates how a schematic drawing shows the component symbols for the system builder or troubleshooter. Valves, actuators, flow paths and line connections are all shown according to the ANSI or ISO graphic symbols that were explained in Chapter 4. To understand how the circuit operates, a person must be able to read the symbols and know how they represent a piece of hardware. The valve in this circuit is 2-way, 2-position, direct solenoid-operated, spring return, normally closed. The diagrams to the right of the circuit at rest show how the directional control valve shifts to its second position and ports fluid to the cylinder. In the real world, this is done in a person’s imagination . . . and can be confusing when several valves are working simultaneously. In the diagram it is easy to see that with the solenoid energized, the normally open box moves in line with the input flow and sends fluid to the cylinder. The arrow in the normally open box shows flow from inlet to cylinder port, causing the piston to extend. If the solenoid is de-energized, the spring returns the valve to the circuit at rest condition and the cylinder stops in its last position.

Two-way valves cannot have more than two positions because they can only stop or allow fluid flow. It is easy to see that a 2-way directional control valve will not operate a single-acting cylinder. These valves are only good for operations that require an on-off supply. As shown in the bottom half of Figure 10-10, two 2-way valves are needed to control a single-acting cylinder. A double-acting cylinder needs four 2-way valves to control it. There are both normally closed and normally open valves in these circuits.

Figure 10-11 shows how 3-way valves can replace 2-way valves and make a machine simpler. This circuit at rest has a cylinder powered by a 3-way, 2-position, solenoid pilot-operated, spring-return, normally closed directional control valve. Because this valve has a flow path from the pressure port to the cylinder port and from the cylinder port to atmosphere, it can control a single-acting cylinder. The diagrams to the right show that when the solenoid is energized, the cylinder extends under power. The next schematic diagram shows the cylinder retracting from external forces with the solenoid de-energized.

Fig. 10-11. Circuits in which 2- and 3-position, 3-way valves operate cylinders

Two 3-way valves are needed to power a double-acting cylinder as shown in Figure 10-11. The double-acting palm button activates this circuit. The valve on the cap end is normally closed and the valve on the head end is normally open. This is a simple anti-tie down circuit, but is not OSHA safe because one palm button can be depressed before the second one and the cylinder will move. OSHA requires that both buttons be operated concurrently to make the cylinder extend. It does meet the anti-tie down requirement because the cylinder will not retract until both palm buttons are released.

The double-acting inching circuit in Figure 10-11 uses two 3-way, 3-position, spring-centered valves to make it possible to stop the cylinder at any point in its stroke. A 3-way valve can have a third position to perform another function. The pictured center condition has all ports blocked, which stops flow at all ports. This is the center condition normally found on a 3-way valve.

Note: pneumatic inching circuits cannot stop and hold a load consistently. Any change in speed, load, or pressure can produce a different stopping position. About plus or minus one inch would be the best position accuracy an air cylinder would achieve, unless it is moving very slowly. Air leaks in the plumbing or valves also interfere with trying to stop and hold an intermediate position. Leaks may let one end of the cylinder bleed off and allow air from the opposite end to expand and move the cylinder out of position.

Using two 3-way valves attached directly to each cylinder port will save air. By eliminating all piping between the valve and the actuator, less air is consumed during each cycle. The air savings per cycle may not be great, but it can add up on fast-cycling equipment with multiple cylinders.

A 3-way valve can be used as a 2-way function when an on-off condition is needed.

The 4-way valve in Figure 10-12 makes it possible to operate a double-acting cylinder with a single valve. The four ports on hydraulic valves are marked P for pump, T for tank, and A and B for cylinder or outlet ports. Most valve manufacturers follow this universal marking system. Most air valves are configured as 5-way functions with two exhaust ports. This works well for air valves because atmosphere is the tank. Return piping is not required.

Fig. 10-12. Circuits in which 2- and 3-position, 4-way valves operate cylinders

The circuit at rest in Figure 10-12 shows a 4-way, 2-position, direct solenoid-operated, spring-return directional control valve. In at-rest condition, pump flow holds the cylinder in the retracted position while the cap end is ported to tank.

In the solenoid-energized, cylinder-extending condition, pump flow connects to the cylinder cap end while the head end is connected to tank. The cylinder is extending under power at this time. In the solenoid-de-energized, cylinder-retracting condition, the valve returns to normal and the cylinder retracts under power.

A single 4-way directional control valve can power an actuator in both directions. At the bottom of Figure 10-12, a 4-way, 3-position, double direct solenoid-operated, spring-centered, tandem-center directional control valve powers a double-acting cylinder in the vertical position with its rod up. As shown in the at-rest condition, pump flow goes to tank and the cylinder ports are blocked. Energizing the extend solenoid sends pump flow to the cylinder cap end to make it extend. Energizing the retract solenoid sends pump flow to the cylinder head end, making it retract. With both solenoids de-energized, the cylinder stops and holds position for some time. Because most directional control valves use a metal-to-metal fit spool, there is some bypass, so the cylinder might drift when it has external forces acting on it. Note: this double-acting inching circuit may need a counterbalance valve to stop it from running away as it retracts.

Some manufacturers offer 4-way valves in special 4-position configurations. The fourth position is often a regeneration path to move the cylinder more rapidly at reduced force.

The 5-way valve in Figure 10-13 is found most often in pneumatic circuits. Although most hydraulic valve designs are 5-ported, the tank ports are connected internally by cored passages so only one external tank connection is needed. Air valves exhaust to atmosphere so having two exhaust ports is not a problem.

Fig. 10-13. Circuits in which 2- and 3-position, 5-way valves operate cylinders

Notice the speed-control mufflers in these circuits. They reduce exhaust noise and act as meter-out flow controls. A 5-ported valve, especially if it’s spool type, can offer advantages when piping certain pneumatic circuits.

When this circuit is at rest, air pressure ported to the cylinder’s head end holds the cylinder in its retracted position. Meanwhile the cap end is exhausted to atmosphere. With the solenoid energized, cylinder-extending air is ported to the cylinder cap end while the head end exhausts. With the solenoid de-energized, the return spring shifts the valve back to normal and the cylinder retracts under power.

A 5-way spool-type valve also can be piped with dual inlets at different pressures -- to conserve energy, to smooth stroke times and speed, or to cause a cylinder to stroke at high speed. (See Chapter 13 on Flow Controls and Chapter 17 on Quick-Exhaust Valves for circuits to do these.)

At the bottom of Figure 10-13, the double-acting inching circuit uses a 3-position, 5-way valve with all ports blocked in center condition to cycle a cylinder. Within reasonable limits, the cylinder can be stopped and held for short periods. (See the note on 3-way valves from Figure 10-11 on the reasons for poor results in pneumatic inching circuits.)

Another center condition for a 5-way air valve is pressure blocked and cylinder ports open to atmosphere. This center condition can be used for mid-stroke stopping of a horizontally mounted cylinder.

Both 4- and 5-way valves can replace 2- and 3-way valves by plugging or not using certain ports to produce the desired function. This can save money in inventory and time when troubleshooting. Only one spare valve of a given size takes care of many problems on the floor.

Types of directional control valves

Directional control valve designs generally fall into three categories:

  • sliding-plate valves
  • poppet valves, and
  • spool valves.

Sliding-plate valves use linear or rotary action to open and close ports to change flow paths. Figure 10-14 has a cutaway representation of each type. A linear sliding-plate valve usually is pilot- or solenoid pilot-operated to generate enough force to reliably move the lap-fitted linear sliding plate. As the hollow linear sliding plate passes over openings in the body, fluid is channeled to a working port or to exhaust through the hollowed out cavity or through the body. (Linear sliding-plate valves are used only in pneumatic circuits.)

Fig. 10-14. Two types of sliding-plate valves, with their symbols

Rotary sliding-plate valves are often manually operated. Some manufacturers offer a pneumatic or electrically powered rotary actuator for automatic operation as well. These valves are used in pneumatic and hydraulic circuits as control and/or isolation valves -- when high shifting speed is not needed. The seal between the rotary sliding-plate and the body can be lap-fitted or may have spring- and pressure-loaded seals to eliminate leakage.

Poppet-type directional control valves

Poppet-type directional control valves are similar to pilot-to-close check valves. The cutaway view in Figure 10-15 shows the construction of a hydraulic 2-way, normally closed, poppet-type directional control valve. Fluid at the inlet port passes through the control orifice to the backside of the poppet. The tip of the spring-loaded armature closes off the outflow orifice to trap fluid behind the poppet. As the symbol shows, the valve is a check valve that stops flow from inlet to outlet in the normal condition. This design will not stop flow from outlet to inlet, although flow in this direction may be at a reduced rate. If using this valve for reverse free-flow in its normally closed condition, make sure to choose one with free-flow capability.

Fig. 10-15. Poppet-type, 2-way hydraulic directional control valve

Energizing the solenoid coil creates a magnetic field that raises the armature to open the outflow orifice. This orifice is larger than the control orifice, so the greater flow through it causes a pressure drop behind the poppet. Now, inlet pressure pushing on the poppet’s annulus area outside the seat diameter unseats it to allow fluid flow to the outlet. De-energizing the solenoid coil lets spring force reseat the armature tip to again trap fluid behind the poppet and close it.

Unlike spool valves whose lands overlap, poppet valves open a flow path to outlet immediately. This means response time of whatever the valve controls is very fast. Also, when a spool valve shifts open it goes to the end of its stroke regardless of the amount of flow. On the other hand, a poppet only opens as much as the flow going through it needs. This means the poppet has less distance to move to stop flow, so again its response is faster.

Chapter 11 covers poppet-type slip-in cartridge valves used for directional control. These valves have the same characteristics as just explained here and they work well in circuits that require fast response. Chapter 12 covers infinitely variable spool valves that also offer very fast response.

The 4-way poppet valve in Figure 10-16 is a typical design for pneumatic service. Poppet design valves are very tolerant of contamination and many plants use them for this reason. They are also very responsive and provide a positive seal when their poppets seat. (Many poppet valves are built with resilient materials on the poppets where they contact the seats.

Fig. 10-16. Poppet-type, 5-way pneumatic directional control valve

One drawback to this design is that air is free to go any direction as the poppets shift from one flow path to the other. In valve terminology this is called open crossover (and can be helpful with hydraulic valves as explained later). The cutaway view in Figure 10-16 shows how flow can go to both cylinder ports and to both exhausts as the poppets move to the opposite seat.

Another possible problem with poppet valves is that they usually only operate in one manner. When you purchase a 2-way, normally closed poppet valve, it cannot be changed to normally open. The port marked In is always the supply line. Air piped to the Out or Cyl port usually blows through the valve with little resistance. Spool-type valves (discussed next) overcome these problems in most cases.

The poppet valve in Figure 10-16 shifts to its second condition when the coil of the solenoid operator receives an electrical signal and pulls the armature up. This action lets supply air into the large pilot piston to move the poppets to the second valve position. Even though the small pilot piston has supply air against it all the time, it has less force. De-energizing the solenoid operator exhausts the large pilot piston and the poppets return to the normal position.

This is a very reliable design because there are no springs to rust, weaken, or break. Usually the area ratio is 2:1; so shifting force is equal in both directions. Some manufacturers also use a spring in the return end to keep the poppets in place when there is no air supply. Valves with this type of shifting arrangement usually require a minimum pressure of 25 to 40 psi or an external pilot supply at least that high.

Spool-type directional control valves

For circuits with flows less than 100 gpm, the most common hydraulic directional control valves use a spool-like internal member to direct flow. (Many air valves also use a spool, due to the advantages offered by this design.) The cutaway views in Figure 10-17 show some simplified spool arrangements and terms associated with this valve. Notice that counting the number of ports that carry working fluid on the cutaway or symbol gives the number of ways the valve has. A 2-ported valve is a 2-way valve and a 5-ported valve is a 5-way valve.

All valves in Figure 10-17 are two position as shown by two boxes in the symbol. As stated before, a 2-way valve can have only two positions because it can only stop or allow flow. All other valves are able to have three positions, while 4-way valves can have four positions in special cases. A 5-way valve is a special case mainly used in pneumatic applications where an extra exhaust port is not a problem. Notice that a 4-way valve has five ports but its tank ports are internally connected to eliminate an extra port in the body. This is important in hydraulic valves because it reduces piping and potential leak points.

Fig. 10-17. Views of a variety of valve spool configurations, with their symbols. (All have palm-button operators.)

Spool valve advantages

The main advantage of spool valves is that fluid entering the valve from any working port does not affect spool movement. The poppet in a poppet valve can have pressure on one side and only a light spring on the other. This can result in premature movement of the poppet when pressure enters a port. In a spool valve, pressure always is applied to two equal opposing areas or the edge of a land. Thus pressure forces that could move the spool are cancelled or non-existent. This means that a spool valve can be shifted manually, electrically, mechanically, pneumatically, or hydraulically with the same force regardless of the operating pressure. Low-force solenoids can be used because the most they need to overcome is mechanical friction and light springs.

Spool valve disadvantages

Many spool valves are designed with metal-to-metal sliding fits. As a result, some fluid may bypass these seals. If this happens, an actuator may not hold its position if outside forces are applied. It also means wasted energy and resulting heat. (Many pneumatic valves use some sort of resilient seal in the body and/or on the spool to eliminate air leaks.) To reduce bypass, spool valves have land overlap, so as they start shifting to open a flow path, there is a delay before fluid starts flowing. The delay only lasts for milliseconds and does not cause a problem -- unless the cycle is very fast and/or there are several valve shifts per cycle.

Another time delay occurs when a spool shifts to the end of its stroke. There is often more movement than required for the flow needed. When the spool shifts back to center or to the opposite flow path, it consumes more time to travel the extra distance. This slows the cycle, especially when several valves are involved. Stroke limiters that control maximum spool movement can eliminate this delay, but are seldom seen in actual practice. The common fix for these problems is to speed up traverse time by installing a larger pump. However, faster actuator movement can add shock and heat due to higher energy input.

Hydraulic 4-way spool valves

Most manufacturers of hydraulic valves only build a basic 4-way function. When they offer a 2-way function, it is usually a 4-way valve with a different spool and the unused ports plugged or piped to tank. Any 4-way valve can perform 2- or 3-way functions in a normally closed or normally open configuration by using the right ports and plugging or draining unused ones.

Hydraulic 4-way valves usually come in 2- or 3-position configurations. They may be 2-position, single-solenoid, spring-return; double-solenoid, detented; or 3-position, double-solenoid, spring-centered. Some manufacturers offer 4-position valves with a float or regeneration center position for special circuits, but they are rare.

The majority of hydraulic circuits use a 4-way, 3-position directional valve even though it complicates the electrical circuit. One reason may be to provide the ability to stop an actuator in mid cycle -- either for manufacturing or setup functions. Other reasons are to port pump flow to tank while the machine is idling or to let external forces move an actuator.

Figures 10-18 through 10-21 show typical circuits in schematic form with valve cutaways for the four commonly used center conditions in hydraulic 4-way directional control valves. (Symbols for other spool center conditions are shown in Chapter 4.) Each center condition offers a flow path to meet a specific circuit need that should be obvious when reading a schematic. Note that these typical circuits are not the only way to apply these valves.

Fig. 10-18. Typical hydraulic circuit for a 4-way valve with all-ports-open center condition

The circuit in Figure 10-18 uses an all-ports-open center-condition valve that allows flow to and from all ports. Notice how the spool lands are too narrow to block the fluid ports. This means fluid is free to go to other ports while the spool is centered. The symbol for the valve plainly shows this open-center condition. A circuit with this type of valve normally has a fixed-volume pump. The open center lets all pump flow return to tank at little or no backpressure. This saves energy and reduces heat to the point that a heat exchanger is not necessary on most circuits.

The cylinder in Figure 10-18 is free to float when acted on by outside forces. Otherwise it sits still. Normally this circuit only has one valve and actuator. Other valves and actuators trying to use this pump would not receive fluid due to the free path to tank.

Fig. 10-19a. Typical hydraulic circuit for a 4-way valve with all-ports-closed center condition

The valve in Figure 10-19a has an all-ports-closed center-condition that blocks pump flow. This valve appears to be able to stop and hold a cylinder in place. Notice how the spool lands are wide enough to completely cover the A and B ports. This blocks flow to and from them, and also stops flow at the P port. This circuit normally has a pressure-compensated pump. System pressure is at the pump compensator setting until all pump flow is going to the actuators at their working pressures. The pump in a closed-center circuit can supply other circuits one at a time or simultaneously with low to medium energy loss -- even when operating at less than maximum flow.

Because all metal-to-metal fit valves have some spool bypass, a closed-center valve will not stop and hold a single-rod cylinder except for a short period. Figure 10-19b shows how bypass fluid at the spool lands leaks directly into the A and B ports and pressurizes both ends of the cylinder at roughly half system pressure. Equal pressure at both ends of a single-rod cylinder always causes it to extend due to unequal forces on unequal areas. The cylinder will not move rapidly because some fluid must go to tank across another leak path. This cylinder action is called regeneration, and will be explained under cylinders in Chapter 15.

Fig. 10-19b. With all-ports-closed valve center condition, bypass flow will extend cylinder

In a new circuit, bypassing fluid may not affect the cycle, but it can cause problems later on. Also, cylinders with small rods and/or heavy loads may not have enough force to move -- especially when machine fits are new and tight. The actual force in this regeneration mode is calculated by multiplying the rod area by pressure at the cylinder cap end.

The circuit in Figure 10-20 shows a float-center valve. The P port to the pump is blocked, and ports A, B, and T are interconnected so that both cylinder ports are open to tank. Notice that the spool lands are wide enough to block the P port but still allow flow to or from A and B ports flow to or from each other or tank. A pressure-compensated pump normally supplies a circuit with this valve. System pressure is the pump compensator setting until all pump flow is going to the actuators at their working pressures. The pump in a float-center circuit can supply other circuits one at a time or simultaneously with low to medium energy loss, even when operating at less than maximum flow.

Fig. 10-20. Typical hydraulic circuit for a 4-way valve with float center condition

When a cylinder in a multi-actuator circuit must be positively locked in place, select a valve with a float-center condition, and add pilot-operated check valves or a counterbalance valve. (Adding these blocking valves to an all-ports-closed-center directional control valve often does not prevent cylinder movement.)

With a float-center valve a single-rod cylinder may extend at any speed when the circuit has high tank backpressure. High tank backpressure goes to both ends of the cylinder and causes it to extend if the load is low and/or the cylinder has an oversize rod. To overcome this situation, install a check valve at the T port to stop back flow from the tank line.

On horizontally mounted cylinders, use pilot-operated check valves in the cylinder lines to positively lock the cylinder from moving due to external forces. On vertically mounted cylinders, use a counterbalance valve to hold the load and keep it from running away while cycling. (Counterbalance valves are explained in Chapter 14.)

Fig. 10-21. Typical hydraulic circuit for a 4-way valve with tandem center condition

The circuit in Figure 10-21 uses a valve center condition with the A and B ports blocked, and the P connected to T. This valve lets pump flow go to tank and blocks both cylinder ports. This configuration is often referred to as a tandem-center valve. Notice that the spool lands are wide enough to block the A, B, and P ports – the same as an all-ports-closed valve. However, this valve has a hollow spool and cross-drilled ports at P and both ends at T. The drilled passages provide a path for all pump flow to go directly to tank in the center position. Because the drilled passages also introduce extra backpressure, most suppliers’ catalogs show a lower nominal flow or higher pressure drop for tandem-center valves.

Circuits with tandem-center valves normally have fixed-volume pumps. The pump-to-tank path lets all flow return to tank at little or no backpressure. This saves energy and reduces heat to the point that a heat exchanger is unnecessary in most circuits. Be aware of the reduced flow or higher backpressure when specifying or using tandem-center valves. A circuit may look good on paper, but can run hot because of wasted energy. This is especially true when using tandem-center valves in series. The backpressure for each valve is additive. A three-valve circuit can easily require more than 300 psi to unload the pump.

Fig. 10-22. Views of double-solenoid detented and single-solenoid, spring-return valves, with symbols

Figure 10-22 presents symbols and cutaway views for standard 2-position valves. Notice that the double-solenoid detented valve has the same outward appearance as a 3-position valve. The only way to tell the difference in these two configurations is to know the part number designation or to probe the manual overrides. A 3-position valve shifts easily and follows the probe as it is withdrawn. A detented valve shifts hard when breaking out of the detent notches but stays shifted when the probe is withdrawn.

Most 2-position valves are found in circuits with pressure-compensated pumps because they block flow when an actuator stalls. When actuators need pressure continuously, a 2-position valve simplifies the electrical control circuit. A momentary signal shifts a double-solenoid detented valve to its other position and the valve stays there until it receives the opposite signal. A single-solenoid spring-return valve must have a maintained electrical signal to stay shifted. This solenoid setup causes all actuators to return to home position at the loss of control circuit power or after an emergency-stop signal.

Single-solenoid spring-return or double-solenoid detented valves are commonly used in air circuits because pressure is always available and blocking flow does not cause overheating.

Solenoid pilot-operated valves

All the foregoing valves are operated directly by a solenoid plunger pushing against a spool. All D02-and D03-, and most D05-size valves are direct operated. This arrangement works well for small, low-flow valves, but is not a good setup for high-flow valves with large spools or poppets. Most direct solenoid-operated valves are rated at 20 gpm or less nominal flow. (Nominal flow is usually considered as the maximum flow a valve can pass at 35- to 50-psi pressure drop.) Most manufacturers’ literature shows flow and pressure-drop information for all valve sizes and flow paths. Keeping pressure drop low saves energy, reduces shock, makes for a quieter system, and minimizes leakage potential.

For systems with flows higher than 10 to 20 gpm, use solenoid pilot-operated valves with high flow capacity. They offer low pressure drop by incorporating small solenoid-operated valves that hydraulically shift large, high-flow spools. These valves look different physically and also have additional ports. The valves require a minimum pressure of 50 to 75 psi to shift the working spool, especially on spring-return or spring-centered models. A D02, D03, or D05 directly operated, 3-position valve that is spring returned, detented, or has a float center is the control choice. (See Chapter 4 for NFPA, ISO, CETOP, and NG valve size designations, relative physical size, port diameters, port configurations, and nominal flow ratings. For actual flow information and dimensions, always check the supplier’s catalog.)

Fig. 10-23. Single-solenoid, 2-position, pilot-operated, spring-return 4-way directional control valve

Figure 10-23 shows a cutaway view and the symbol for a single-solenoid, spring-return, solenoid pilot-operated directional valve. In the normal condition, a D03 single-solenoid spring-return, pilot-control valve receives its pilot oil from the X port or the internal pilot-supply port, and sends it to the right end of the working spool. When the pump runs, pilot pressure holds the working spool in place to open flow paths from P to A and B to T. Because system fluid controls the working spool’s position, a minimum of 50- to 75-psi pressure must be available at all times. As system pressure rises, pilot pressure also rises but this has little affect on spool shifting time because of flow restrictions and the distance the spool must travel.

The simplified symbol typically is found on schematic drawings because internal details usually are not important. Notice that the symbol in Figure 10-23 has a solenoid slash and an energy triangle in the operator box to indicate solenoid pilot operation. The arrows in the symbol show the flow configuration of the working spool because it makes the actuator move. The only indication of the solenoid-operated valve is the solenoid slash mark in the operator box. The pilot control valve must have the right configuration to make the working spool operate, but otherwise is unimportant. In the case of this 2-position, spring-return valve, the pilot-control valve must also be a 2-position, spring-return model.

The complete symbol includes directional controls, internal porting, and internal connections that are plugged or open. (Later in this text, other features of the complete symbol will be shown and explained.) The enclosure outline around the two valves indicates they are a single piece of hardware. Lines passing through the enclosure outline are external connections and must be plugged if not connected.

Fig. 10-24. Double-solenoid, 2-position, pilot-operated, detented 4-way directional control valve

Figure 10-24 has a cutaway view and symbol for a detented, solenoid pilot-operated directional control valve. The only difference between this valve and symbol and those in Figure 10-23 is that the pilot control is a double-solenoid, detented valve instead of having a return spring. A spring-return valve can be converted to a detented valve merely by changing the pilot control. It is also easy to replace this double-solenoid, detented valve with a double-solenoid, 3-position, spring-centered valve that could allow the working spool to float when both solenoids are de-energized.

These first two types of solenoid pilot-operated valves must always be mounted so the working spool is horizontal. This keeps the working spool in place when there is no pilot pressure. The working spool could drift when mounted vertically because there are no springs to maintain its position. It is good practice to mount all spool valves horizontally to keep spools in place, especially during shutdown.

The Y (external drain) and X (external pilot supply) ports must be used in some circuits but their use is optional and up to the designer in others. A general rule is to drain pilot oil through the Y port when backpressure at T is (or can be) equal to or higher than pressure at P or X. Backpressure at T is seen at the tank port of the pilot-control valve. It resists pressure from the P or X port that tries to move the spool.

The X port is seldom used except where a constant pilot pressure is necessary. Another possible circuit is where a backpressure check valve in the T port can cause cylinder regeneration when the spool shifts through the crossover or transition position.

The solenoid pilot-operated 3-position, spring-centered valve in Figure 10-25 is the most common configuration for this type of hydraulic directional control valve. The center condition is often used to unload a fixed-volume pump, stop an actuator or allow it to float, and hold a cylinder at mid stroke while installing or removing tooling. More than 80% of hydraulic directional control valves are 3-position spring-centered for the above reasons and others. All directional valves have a spool that is capable of shifting to a center condition, but 2-position valves never stop in center position when working properly.

Fig. 10-25. Double-solenoid, 3-position, pilot-operated, spring-centered, 4-way directional control valve

The cutaway view in Figure 10-25 represents a solenoid pilot-operated valve with an all-ports-open center condition. These larger, high-flow valves operate in the same way and perform the same functions as the direct solenoid-operated valves discussed earlier. (Refer to Figure’s. 10-18 through 10-20 for typical spring-centered circuits.)

The main difference in the circuits for solenoid pilot-operated valves is that the models with a pump-to-tank center condition need some method to maintain a continuous minimum pilot pressure to the pilot-control valve. Figure 10-25 shows the standard methods for maintaining such minimum pilot pressure.

These valves may also need a way to provide a free path for pilot oil to flow to tank (port Y) under some conditions. Adding or removing plugs – or installing an orificed plug -- can stop or allow flow to the internal-drain port or from the internal pilot-supply port.

The cutaway in Figure 10-25 shows springs and centering washers at both ends holding the working spool in its center position. The centering washers prevent a stronger spring from pushing the spool past center in case one spring is weaker. The spool in the cutaway view and the complete symbol represent an all-ports-open center condition. Other spools and their simplified symbols are shown below the cutaway.

The solenoid-control valve is a double-solenoid, spring-centered model with a float-center spool. All 3-position, solenoid pilot-operated, directional control valves must have a float-center spool to work properly. Any other spool center condition (except one with T blocked, and A, B, and P connected) will either not let the working spool center or will generate excessive heat. Always make sure the solenoid-control valve is 3 position and has a float-center spool when the pilot-operated directional valve is 3 position.

When the a solenoid of the solenoid-control valve is energized, it sends pilot oil to the right end of the working spool. This shifts the spool to the left. Pump flow at the P port is directed to the A port, and flow from the B port goes to tank through the T port. It is easy to follow flow paths on the complete symbol to see how any manufacturer’s valve functions, even when its construction is different. Energizing the b solenoid shifts the working spool to the right, producing the opposite flow condition and reversing flow to the actuator.

With the valve in Figure 10-25 at rest, it is obvious pressure would be low. When the valve and lines are sized for low pressure drop, pressure would probably be below 50 psi. This pressure would not generate enough force to shift the working spool against the centering springs, so no fluid can flow to or from the actuator. To overcome low or no pilot pressure the following options are available:

Option 1: Use a 75-psi backpressure check valve out of the T port, use an orificed plug in the internal pilot-supply port, and drain the pilot-control valve externally through the Y (external-drain) port. When the pump runs, the backpressure check valve maintains at least 75-psi pilot pressure, so the working spool can shift when signaled. When the circuit operates, pilot pressure may go higher, but it never drops below 75 psi. With an open-center valve, this option can make the cylinder regenerate if the cylinder has low resistance and/or an oversized rod. If this situation is suspect, use Option 2.

Option 2: Install the 75-psi backpressure check valve in the pump line and route a pilot signal from upstream of it to the X (external pilot-supply) port. This makes it possible to internally drain the pilot-control valve. An external drain could still be used, but would not be required. This option keeps pressure off a cylinder that could regenerate with the open-center valve in its center position. It also eliminates the possibility of regeneration when a tandem-center valve moves through an open crossover.

Option 3: Supply the X port with 75- to 3000-psi pilot pressure from another source. The external supply does not need to provide high flow. A constant pressure might be desirable to keep spool shifting and cycle time consistent.

Check the supplier’s catalog information to see what pilot pressure should be for a specific valve size and operating pressure. These figures change for different spools and are directly affected by tank line backpressure when using internal drains.

Option 4: Most suppliers offer an internal backpressure check valve option that operates almost the same as Option 1. The difference is that there is no pressure in the T tank line so an internal drain may be used. The main difficulty with this option is that this valve is not standard. Delivery probably is long and could affect machine downtime if the valve needs to be replaced.

Other solenoid pilot-operated valve options

The solenoid pilot-operated valve in Figure 10-26 shows some other options that manufacturers offer for special needs. All of these options may not be available from all suppliers, so check with the distributor before specifying a brand.

Fig. 10-26. Double-solenoid, 3-position, pilot-operated, spring-centered, 4-way directional control valve with all options

The spool-position indicator option helps to troubleshoot a circuit. When the pilot-control valve shifts electrically or manually but the actuator will not move, one reason could be that the working spool did not shift. On a valve without the indicator, the only positive way to know if this spool can move is to take the valve apart and check it.

The spool stroke limiter option imitates a simple flow control for different actuator speeds in both directions. This option limits spool travel, which restricts flow to or from the actuator -- similar to a flow control. This option should only be used where speed can fluctuate as pressure and force change.

The pilot-choke option installs a modular meter-out flow control between the pilot-control valve and the working spool to slow spool movement. Slowing the spool can give an actuator smoother acceleration and deceleration, thus reducing shock. The idea is great, but note that slower spool movement may increase cycle times beyond limits.

One thing that causes cycle time to increase is the fact that all solenoid-operated spool valve lands overlap the body lands. This overlap means the spool has to move some distance before fluid flow starts. When the spool moves slowly enough to give good control, the shift time out of overlap can be 0.5 to 1.0 second or greater. After the spool clears overlap, the actuator can accelerate very smoothly, but the extra time often cannot be tolerated.

Another addition to cycle time comes when the spool shifts to the end of its stroke. A spool can continue to move to the end of its stroke even though a partial stroke is passing all available flow. When reversing actuator motion or decelerating before the end of stroke, the spool may be shifted 1/16 in. or more past available flow. When the spool starts slowly moving to center, the actuator continues at full speed until the spool moves far enough to start restricting flow. From this point on, deceleration is very smooth, but time has been lost. Also remember: the speed at which the spool goes to center is the rate for accelerating the actuator in the opposite direction. This means that adjusting for acceleration both ways also affects deceleration in both ways. The spool stroke-limiter option can eliminate the time loss here, but will not help return shifting speed.

Another possible problem with the pilot-choke module is that it often has a flow rating of only 4 to 8 gpm. Adjusting that small amount of flow is very difficult, if not impossible. This problem limits the usefulness of the pilot-choke option. At present, the need for acceleration, deceleration, and flow variation can be handled better by proportional valves, which will be discussed in Chapter 12.

The integral backpressure check valve option was discussed as Option 4 on Figure 10-12.

Hydraulically centered valves

The cutaway view in Figure 10-27 represents a solenoid pilot-operated directional control valve that is hydraulically centered. A few designers prefer hydraulic centering to spring centering. The reasons given are: spring force changes over time, springs may break, response is slower with springs, and springs are relatively weak. Hydraulic centering has none of these faults, but is still specified on less than 2% of all hydraulic circuits. Part of the reason is lack of knowledge of many designers and users.

Fig. 10-27. Double-solenoid, 3-position, pilot-operated, hydraulically centered, 4-way directional control valve

Notice that in the complete symbol, the pilot-control valve has port T to tank blocked in the center condition, with ports P, A, and B connected. With the pump running, the pilot-control valve sends pilot oil to both ends of the working spool, centering it. The working spool can center because the differential-area sleeves with centering washers can only move until they contact the valve body. With pressure at both ends, these items give a difference in area that causes the working spool to move until it centers. Other than the way the working spool centers, this design valve works the same as other solenoid pilot-operated directional valves.

Valve operators

Figure 10-28 shows all the operators for directional control valves. Prior to 1966, the operator box on the symbol had letter abbreviations for the method of operating the valve written in them. (See Chapter 4, where old operator symbols are shown across from present day symbols.).

Fig. 10-28. Symbols for valve operators

When ISO standards were adopted, all writing was eliminated from the picture-like symbols. The abbreviation MAN, for a manual operator, changed to extended lines at the operator box outer end or to a stick drawing of a palm button, hand lever, foot pedal or treadle.

With picture-like drawings there was no language barrier when schematic diagrams went from one country to another. ANSI adopted the new standards (with a few exceptions) and the fluid power industry changed soon thereafter. Old machines with pre-1966 schematics can confuse newcomers, but the drawing usually can be deciphered with a little effort.

Spool-valve transition conditions

When a spool valve shifts from one flow condition to another, it can pass through different flow conditions than those indicated by straight or crossing arrows in the symbol. These transition or crossover conditions are unimportant in most circuits, but can cause shock, drifting, or cylinder regeneration under certain circumstances. Figure 10-29 shows typical transition conditions for open- and closed-center, 2- and 3-position valves. For transition conditions of other valve center types, see Chapter 4.

Fig. 10-29. Symbols for typical valve transition or crossover conditions

An example of how an open transition or crossover can eliminate shock is the case of a fast-moving actuator that must reverse direction before the end of its stroke. With a closed-crossover valve in this circuit, all flow to and from the actuator is blocked for a brief period as the spool shifts from extend to retract. The pump side of the valve always has a relief valve (or a pressure-compensated pump) to protect it from over pressure. However, the actuator side has no protection of any kind. When flow from the extending actuator is blocked, pressure can build to levels well above the pressure ratings of pipes, seals, and hardware. This pressure spike is very brief, but it happens every cycle -- and soon shows up as cracked fittings, blown hoses, leaking seals, or broken parts. Changing to an open-crossover spool in this application would connect the impending spike to the relief valve or pressure-compensated pump. There is still a pressure spike but its intensity is now below the damage level of the components.

An example of where a closed transition spool helps is when a vertical cylinder with a heavy platen must be reversed in mid stroke. With an open transition, the cylinder will continue its forward travel after the valve receives the reverse signal. With a closed transition, the cylinder will stop almost immediately and start reversing shortly thereafter. Also, when the platen must retract from a stop position, such as during set up, it can drop before being powered up with an open-crossover valve in the circuit.

Crossover problems usually show up when a 3-position valve shifts to or from center condition. Cylinders may move in the opposite direction or move when signaled to stop for no apparent reason. Most suppliers show crossover flow in their catalog so check it out if the problem is an unusual movement.

Mounting pneumatic valves

Pneumatic valves can be line-, subplate-, or manifold-mounted. Inline or bar-type manifolds make it convenient to stack valves with a common inlet and/or exhaust. Pneumatic circuits seldom, if ever, require custom manifolds as hydraulic circuits do. The graphics in Figure 10-30 depict some pneumatic-valve mounting styles. In-line mounting types have the whole valve assembly in a body with ports out the sides. This style is usually less expensive, but is more trouble if it has to be replaced.

Fig. 10-30. Mounting arrangements for pneumatic valves

The typical pneumatic subplate assembly is often a subplate with end covers bolted to it. Fasteners hold the parts together and molded seals eliminate leaks. Valves with seals mount to the subplate and all piping connects to it. Some manufacturers have wiring troughs in the subplate and use plug-in connectors on solenoid-operated valves.

The typical pneumatic manifold assembly consists of two or more subplates connected to make a valve stack with a common inlet and exhaust. These assemblies eliminate many connections and make valve installation replacement easy. These units also are available with wiring troughs and plug-in valves for solenoid operation. Most air valves use unique mountings and port arrangements that are not inter-changeable. However, there is an ISO standard subplate mounting that several companies offer. The valves match each other’s mounting patterns but otherwise do not have interchangeable parts. This assembly is physically large and thus more expensive, but it makes it easy to combine valves from several suppliers.

Mounting hydraulic valves

In-line and subplate-mounted hydraulic valves are common. However, most in-line valves are screw-in cartridge type with aluminum or steel bodies. Figure 10-31 shows an example of an in-line cartridge valve. (This valve also could be screwed into the custom manifolds discussed later.)

Fig. 10-31. Cross-sectional view of in-line mounted hydraulic valve

Screw-in cartridge valves perform all directional and flow control, relief, sequence, counterbalance and reducing functions -- the same as in-line and subplate valves. Only their physical makeup is different. They normally handle flows less than 40 gpm, but some manufacturers offer sizes with up to 120-gpm capacity.

There are worldwide interface standards for subplate-mounted hydraulic directional control valves. The information in Figure 10-32 shows port and bolt locations and relative sizes for all standard sizes. The figure also lists the numbering systems for U. S. National Fluid Power Association (NFPA), worldwide International Standards Organization (ISO), European Committee for Oilhydraulics and Pneumatic Control (CETOP), and the NG part of the German DIN Standard, which relates only to port size in metrics. (Actually, the NG port size can be for any type valve.)

Fig. 10-32. NFPA and ISO standard interface layouts for hydraulic directional control valves

Interface standards cover both size and location for ports and bolts. A directional valve from any country or manufacturer using this standard is interchangeable with all other valves of the same size. The only difference should be whether the bolts have SAE or metric threads.

Figure 10-33 depicts typical subplate and bar-manifold mounting arrangements for subplate-mounted directional control valves. Subplates mount a single valve and are used in simple single-cylinder applications. They come in bottom- and side-ported models for piping convenience. Some side-ported models have bottom ports as well.

Fig. 10-33. Subplates and bar manifold for mounting hydraulic valves

Subplates can be used for multi-cylinder circuits but require a lot of pipe connections that can restrict flow and may be potential leakage points. Many multi-cylinder circuits work well with the bar manifold shown in Figure 10-33.

Subplates are available for all the valves listed in Figure 10-34. Porting for larger valves usually involves SAE flanges on the valve body. When they have subplate mounts, a special subplate must be made or they are mounted on a custom manifold.

Bar manifolds come with series and parallel porting related to pump and tank connections. Series manifolds usually are limited to three stations or three valves maximum, while parallel manifolds can have as many as 16 stations.

When circuits are not too complex, bar manifolds and modular accessory valves can eliminate most pipe connections and put everything in one location. Symbols for these modular valves are shown for most of the types available. They are always at the end of a section for a particular accessory valve.

Bar manifolds are only offered in sizes D02, D03, D05, D05H, D07 and D08. Size D10 valves use subplates or specially made bar manifolds, or are mounted on a custom manifold.

Custom manifolds

Figure 10-34 depicts a custom manifold that can eliminate many plumbing connections and make valve replacement easy. Such manifolds are usually more expensive than simple plumbing, but can save many times the extra first cost by minimizing fluid loss through leakage and the cost of cleanup.

Fig. 10-34. Custom hydraulic manifold

A number of companies specialize in designing and building custom manifolds. All they require is a schematic of the circuit with the valves to be included inside chain lines; plus other information such as preferred pipe connection type, connection locations and size, which face or faces to leave clear and manifold material. If there is a preferred mounting arrangement, it should be noted along with mounting bolt locations.

Remember: manifolds are difficult to modify, so they should only be applied to proven working circuits. If in doubt about the viability of a new design, work out the bugs in standard piping before buying a manifold.

Fig. 10-35. Symbols for most common modular check valves

Most check valve functions are available as modular or sandwich valves that mount between the directional control valve and subplate. Figure 10-35 shows most of the common configurations presently offered by fluid power suppliers. These modules are usually available in all valve port sizes up to D08 (3/4 in.).

Quiz

 

Chapter 11: Slip-in Cartridges

Cartridge valves

The term cartridge valves commonly refers to pressure, directional, and flow control valves that screw into a threaded cavity. These valves are mostly rated for low flows - 40 gpm or less, although some manufacturers have units that will flow more than 100 gpm. Compact screw-in cartridges help build inexpensive circuits that are reliable and easy to maintain. Screw-in cartridges are most often part of a drilled manifold but can be purchased in individual bodies. The performance and function of screw-in cartridge valves is similar to the in-line or sub-plate-mounted valves discussed in Chapter 10.

Slip-in cartridge valves (sometime referred to as logic 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 go as high as 3000 gpm. (Several companies make screw-in logic valves in sizes as small as 15 gpm for use when the special features of logic valves are required at lower flows.) Slip-in cartridges also are compact, have low pressure drop, and operate at pressures up to 5000 psi. Slip-in cartridges can function as pressure, flow, and directional valves.

The symbols that illustrate this chapter are the preferred type as first used by the manufacturers. Chapter 4 shows cartridge valve symbols that follow the using ISO rules.

Why use slip-in cartridge valves?

The main reason for using slip-in cartridge valves in high-flow circuits is economy. Large spool valves are available with high flow capacity but few are manufactured, making them expensive with long delivery times. A better choice is a slip-in cartridge valve in a manifold body, piloted by D03- or D05-size directional control valves. (There is at least one supplier of valves that uses logic elements instead of spools that bolt directly to D08- and D10-size sub-plates.)

One important feature of slip-in cartridge valves is that there is almost zero bypass through the A port. Leakage for this port is the same as any good check valve. Flow past the B port is very low because the leak path is long and has a close fit. So when applying this type of valve, always use the A port for the connection that must block flow completely.

Another feature is fast response on opening. Because there is no overlap, flow is almost instantaneous after the valve receives a start signal. Even when controlling poppet-opening speed, there is no lag in flow response that increases cycle time. Response also is fast on closing because the poppet only opens far enough to pass the flow going through it. This means it does not have to move any extra distance to start to restrict flow and shut it off.

Logic valve circuits are also very versatile when set up with multiple control valves. Many center or crossover conditions can be duplicated by which control valves are open and how they are signaled during a cycle. This means that when designers want to try different valve configurations, they only have to change control valves or control circuitry. This versatility can be applied anytime with the right design parameters up front.

How do slip-in cartridge valves work?

As directional control valves, the most common slip-in cartridge valve has the 1:2 poppet-area ratio shown in Figure 11-1. This is a pilot-to-close check valve with the pilot area and the areas at the A and Bports all equal. There are no communicating holes through the poppet to allow fluid from the A or B ports to get behind it. Fluid entering the A or B port pushes the poppet open, so flow can go either way, restricted only by the light spring that holds the poppet in place during shutdown. Spring force choices from most suppliers are usually 25, 50, or 75 psi.

It is easy to see that flow through the valve in either direction can be blocked by pressure on the pilot area. Such pressure must be equal to or greater than the pressure at the A or B port. When there is equal pressure at the A and B ports, pressure on the pilot area must be equal to or greater than that which is trying to push the poppet open.

From the foregoing description it should be obvious that when the pump is off and pilot pressure is gone, any load-induced pressure will push the poppets open and running-away loads will run away. (Note that this does not happen with spool valves.) Circuit design is different with slip-in cartridge valves and some of the pitfalls will be discussed later.

Slip-in cartridge valves are held in place by a cover that also contains passages for pilot oil. In addition, the covers may have an interface for directional or pressure control functions. Covers can also have control orifice inserts to retard poppet movement for better control. The plain cover shown in Figure 11-1 would receive a signal from another slip-in cartridge valve with a solenoid-operated valve interface. (A plain cover may serve as a check valve as shown in Figure 11-4.)

Fig. 11-1. 1:2 poppet-type slip-in cartridge valve

The symbol and cutaway view in Figure 11-2 is for a 1:2 slip-in cartridge valve with a single-solenoid directional control valve supplying pilot pressure to the spring side of the poppet. Pilot pressure would be the same as system pressure when the pump is running, so this valve would be normally closed - even with pressure at both ports. Pilot pressure could be blocked and the spring side of the poppet could be open to tank by changing the selector plug location or by using a directional control valve with a P-to-A, B-to-T at-rest condition. The P and T ports connect to pressure and tank in the manifold block through the cover. The A and B ports also go to the manifold block to be connected to other valves as required.

Fig. 11-2. 1:2 poppet-type cartridge valve with directional control valve interface

It would take four separate spool-type directional control valves to do what four of the slip-in cartridge valves shown in Figure 11-2 can accomplish. All that is required to change a cartridge valve circuit is to use different selector plug locations or different directional control valves and/or a change in the electric control circuit. Figure 11-3 shows the equivalent spool valve conditions that are possible with different solenoids energized or de-energized.

Fig. 11-3. 1:2 poppet-type cartridge valves with separate directional control valves -- showing different spool configurations possible

The directional control valve operator could also be a 2-position detented or a 3-position spring-centered valve according to circuit needs. (Chapter 4 presents symbols of other control valve configurations.)

The slip-in cartridge valve in Figure 11-4 has a drilled pilot line that intersects the A port of the cartridge. Fluid is free to flow from port B to port A, but is blocked when it tries to reverse. Changing the pilot line to communicate with the B port can change this check valve to flow freely from port A to port B.

Fig. 11-4. 1:2 poppet-type cartridge valve as a check valve (with free flow from B to A, and checked flow from A to B)

A single-function check valve can be made with an available orificed poppet. The orifice always communicates with the A port so fluid is free to flow from B to A, but blocked from A to B. The orifice is drilled through the poppet as indicated by the phrase: Orifice in Poppet Here Can Replace Pilot Line.

An orificed poppet can serve other functions. For example, it can be used as a blocking valve controlled by a 2-way pilot valve. The symbol and cutaway view in Figure 11-5 show a slip-in cartridge valve set up as a pilot-operated check valve. A special cover with an integral pilot-operated 3-way valve either delivers fluid to or exhausts fluid from the pilot area. The pilot signal comes from other sources -- such as the circuits shown in Chapter 10.

Fig. 11-5. 1:2 poppet-type cartridge valve as a pilot-operated check valve (with free flow from B to A, and controlled flow from A to B)

Fluid entering port A is always free to flow out of port B after its pressure overcomes the spring holding the poppet down. Flow from B to A is blocked until the valve receives a pilot signal. Without such external intervention, a pilot-operated check valve works like any other check valve.

The cover on a pilot-operated check valve has a ball check held on the left seat by a light spring. The ball check traps fluid from the B port, forcing it to hold the poppet down on its full diameter. When there is no pilot signal, there is no flow from port B to port A. The symbol plainly shows this function with the 2-position, 3-way, spring-and-pressure shifted valve held in a position to allow fluid from the B port to hold the poppet closed.

The pilot piston has an area that is three to four times that of the ball check. Thus, a pilot signal that is one quarter to one third the pressure holding the ball check on the left seat will shift it to the right seat to block fluid from the B port. At the same time, fluid trapped behind the poppet is free to go to tank through the drain port. At this point, fluid can freely flow from the B port to the A port with almost no restriction.

A slip-in cartridge valve operating as a pilot-operated check valve has all the circuit problems explained in Chapter 10. The only difference is its physical size and flow capacity.

Another feature that is available on slip-in cartridge valves is the poppet stroke limiter shown in Figure 11-6. A stroke limiter is simply an adjustable screw inside the poppet that can limit the poppet’s travel. The stroke limiter can be used as a non-compensated flow control or a maximum flow limiter on running away loads in certain applications. It works without problems as a meter-in device, but can cause unexpected regeneration as a meter-out device on the rod end of a cylinder with load-induced pressure and/or an oversized rod.

Fig. 11-6. 1:2 poppet-type cartridge valve with stroke limiter and directional control valve interface

Most slip-in cartridge valves with stroke limiters have poppets with extended noses. These may be tapered or have vee notches cut in them. This type poppet gives a better profile for flow reduction as the poppet moves toward shutoff. For best control of flow, the cartridge should be sized so the poppet travel is maximum.

The valve in Figure 11-6 has a solenoid-operated directional control valve interface. The stroke limiter also can come in a plain-cover model.

Because slip-in cartridge valves are pilot-to-close check valves, they must have pilot pressure to stay closed. A vertical cylinder holding a load, as in Figure 11-7, will not stay up when the pump is shut off without some means of maintaining pilot pressure. Slip-in cartridge valves are available with a shuttle-valve function so pilot pressure can be taken from more than one source. In the case of the loaded vertical cylinder, the second source is load-induced pressure.

Fig. 11-7. 1:2 slip-In cartridge valve with dual-pilot source pump (pump not running)

The cutaway and symbol in Figure 11-7 show how a shuttle valve works. Its simplest form contains a free-floating ball that can seal ports to the right or left to block flow from the opposite side. A pilot signal from the right or left always exits from the top -- never from the opposite port. If a shuttle valve receives two signals, it will always pass the highest one. Different pressures on equal areas always move the blocking device toward the lower pressure.

Figure 11-7 shows a slip-in cartridge valve cover with a built-in shuttle valve that can accept fluid from two sources and send it to the pilot area of the poppet. The cover may have an interface for a directional valve (as shown), or it can be a plain type.

Because pump pressure is normally higher than load-induced pressure, pilot pressure to the poppet would be from the pump at port In 1. The cylinder’s load in Figure 11-7 would drop when the pump was shut down if not for the load-induced pressure going to port In 2. At pump shutdown, the shuttle ball shifts to the left (as shown) and the load-induced pilot pressure holds the poppet shut and the cylinder stationery.

Pilot pressure could come from any source, but in the case of a loaded cylinder the most reliable place is the cylinder itself. This pilot pressure would not be suitable for other functions because the cylinder may be in a position where load-induced pressure does not exist.

The circuit in Figure11-8 shows why it is less expensive to use slip-in cartridge valves to power an actuator requiring high flow. This circuit uses a cylinder with a 2:1 rod-area differential. Assume the application calls for 450 gpm. With the cylinder extending with 450 gpm entering the cap end, only 225 gpm exits from the head end.

Fig. 11-8. Typical circuit for 1:2 slip-In cartridge valve

Conversely, while retracting at 450 gpm, 900 gpm exits from the cap end. Without special circuit design, a spool valve to cycle this cylinder would have to be capable of handling 900 gpm -- the high flow from the cap end. A spool valve with this flow capacity would be very large and expensive.

Note that the circuit in Figure 11-8 incorporates three different sizes of slip-in cartridge valves. Small cartridge CV at the cylinder head end handles the 225-gpm tank flow, medium-size cartridges CV2 and CV3 handle the 450 -gpm pump flow to both ends, and large cartridge CV4 at the cap end returns 900 gpm to tank. Each slip-in cartridge valve is sized to handle the flow it sees at a 50- to 75-psi pressure drop. These four standard cartridges, the manifold to contain them, and the directional control valve or valves to control them would cost less than half what a 900-gpm spool-type directional control valve would cost.

Fig. 11-9. 1:1 poppet-type slip-in cartridge valve

A 1:2 area ratio is the most common slip-in cartridge valve design and fits more than 90% of all circuits. To meet some special requirements, there is also a 1:1.1 area ratio poppet shown in Figure 11-9. With this cartridge valve, area at the A port is 90% and area at the B port is 10% of poppet area. With these area ratios, fluid entering the A port flows at a much lower pressure drop than fluid entering the B port. Another way of saying this is it takes just 10% as much pressure to flow from A to B as it does to flow from B to A.

Cartridge valves to control pressure

For pressure-control cartridge valves, the poppet has an area ratio of 1:1. This means flow can only go from the A port to the B port. A 1:1 area ratio makes the valve stable when controlling flow at pressure. Notice that the poppet is straight sided and sits on a tapered seat.

The symbol and cutaway view in Figure 11-10 illustrate a relief valve, but the design would function as a sequence or counterbalance valve as well. A relief valve never needs to have reverse flow, while a sequence or counterbalance valve usually requires a reverse-flow check valve piped around it. Also, as a sequence valve, it would always need a separate drain line to tank. These valves and an unloading valve are shown and discussed next.

Fig. 11-10. 1:1 poppet-type slip-in cartridge valve as a pilot-operated relief valve

Figure 11-10 shows a direct-acting relief valve symbol in the cover with pilot lines and orifices connecting it to a 1:1 area-ratio poppet. The A port is always the inlet, and as a relief valve, the B port is always the outlet to tank. The operation of a slip-in cartridge valve as a relief valve is identical to the poppet-type relief valve discussed in Chapter 9. The main attraction of slip-in cartridge valves is that they come in larger sizes for high-flow applications. Most relief valves for everyday circuits handle less that 200 gpm. Cartridge relief valves can handle flows in excess of 1500 gpm.

Slip-in cartridge relief valves can be set up with all the features of the pilot-operated relief valves that were shown in Chapter 7. Solenoid venting and multiple pressure selection work in the same manner except for higher flow capabilities. Look at the symbols of the different types of arrangements for slip-in cartridge relief valves in Chapter 4.

Figure 11-11 shows an internally piloted, externally drained sequence valve. This is the same valve illustrated in Figure 11-10, except the drain fluid from the direct-acting relief valve section must be ported directly to tank. If fluid is drained to the B port, backpressure at that port would add to the spring set pressure. In some cases, the internal pilot is changed to an external pilot as the circuit dictates. As mentioned previously, a sequence valve circuit often needs reverse flow so a cartridge-type check valve would be piped around it to provide free flow in reverse.

Fig. 11-11. 1:1 poppet-type slip-in cartridge valve as an internally piloted sequence valve or counterbalance valve

The valve pictured in Figure 11-11 also can perform as a counterbalance valve. Counterbalance valves can be internally or externally piloted but they usually are internally drained. Chapter 14 discusses sequence and counterbalance valves in detail and shows why pilot and drain functions are not always the same. Without exception, counterbalance valve circuits must have reverse flow so they always need a bypass check valve.

To add an unloading valve function to a cartridge valve, a special cover is required. The symbol and cutaway view in Figure 11-12 show the setup for an unloading valve that could be used for a hi-lo pump circuit. Chapter 9 provides a complete explanation of unloading valves.

Fig. 11-12. 1:1 poppet-type slip-in cartridge valve as an unloading valve

An unloading valve is similar to -- and functions just like -- a relief valve if the unloading port were not connected. Pressure would build enough to push the ball against the adjustable spring; then all excess pump flow would go to tank at set pressure. The addition of the unloading piston makes it possible to move the ball back far enough so that all pressure on top of the 1:1 poppet drops off. The valve then opens pump flow to tank at 30 to 50 psi. Pressure to the unloading port usually comes from a high-pressure pump. This keeps a high-volume pump unloaded during the work stroke.

The last of the pressure-control valves is the reducing valve. Unlike the other four pressure controls, a reducing valve is normally open instead of normally closed. The valve also is the one slip-in cartridge valve design that does not use a poppet to block fluid. Slip-in cartridge reducing valves are similar to in-line and subplate-mounted valves except that they have higher flow capabilities.

Fig. 11-13. Slip-in cartridge valve as a reducing valve

The symbol and cutaway view in Figure11-13 show a common construction for pilot-operated slip-in cartridge reducing valves. Notice the symbol for the cartridge is the standard ANSI-ISO drawing for any reducing valve. A cover containing a direct-acting relief valve sets the pressure. As fluid enters port A, it is free to go directly out port B . When outlet flow meets resistance, pressure starts to build and goes to the inlet of the direct-acting relief valve. The flow path is directly from ports A and B through the check valve. When pressure reaches the direct-acting relief valve setting, fluid starts to flow slowly through the valve. Flow increases as pressure continues to increase. At set pressure, flow through the direct-acting relief valve exceeds flow through the orifices. At that point, pressure on top of the flow-through spool drops. When this pressure drops enough, the spool raises and restricts outlet flow at set pressure. Flow never shuts off completely because flow from the reduced pressure is always going to tank through the direct-acting relief valve. When pressure at the outlet drops, the direct-acting relief valve closes and forces the passage through the spool to reopen.

Quiz

 

Chapter 12: Infinitely Variable Directional Control Valves

Proportional and servovalves
Infinitely variable directional control valves

The directional control valves discussed so far in this series have all been configured to either pass full flow or completely block flow. The only way to decrease flow through these valves is by adding flow controls or by mechanically limiting movement of an internal part.

The first infinitely variable valve available was the servovalve. Internal flow-modifying parts could be moved to any position at any rate, so output from any port could be varied at will. (Some call these valves infinitely variable 4-way flow controls.) The main problem with servovalves was (and still is) that they require very clean fluid to keep them operating effectively. Fluid from a standard well-maintained hydraulic circuit contains enough contamination to cause most servovalves to fail in a matter of minutes or only last a few hours at best. This meant that the original servovalves were tried and removed from many machines that needed precise control but not at the perceived cost of cleaning up the hydraulic oil.

Why use infinitely variable valves?

Some actuators must move at a precise speed, stop at a close-tolerance position, or produce a very accurate force to perform the work for which they were designed. With the proper input signals and feedback devices, proportional or servovalves can make an actuator perform any or all these functions flawlessly.

Rolling mills turn out sheet consistently to a tolerance of ±0.0005 in. at sheet speeds of 2000 to 5000 feet per minute. Hydraulic cylinders controlled by servovalves maintain the proper force and position the rolls precisely from feedback signals sent by sensors that measure metal thickness, cylinder force, and position. Airline pilots train in simulators moved by hydraulic cylinders so precisely that the pilots get the feel of landing gear raising and locking in position. Even entertainment rides use servovalves to make passengers think they are in 20-ft waves when they are actually in an enclosed articulated room in a shopping mall.

For less precise movement, there are proportional valves that mimic the output of servovalves but respond more slowly. They are less expensive than servovalves and more contamination tolerant, so they have replaced cam valves and other mechanical devices used to get smooth motions.

Hydraulic proportional directional control valves

The symbol and cutaway view in Figure 12-1 represent a direct-acting proportional valve that handles flows as high as 10 to 30 gpm in D03- or D05-size valve interfaces. Proportional valves use the same interface standards as NFPA and ISO directional valves so they can be installed in a circuit without having to change the piping.

Fig. 12-1. Direct-acting proportional valve

Physically, proportional valves appear the same as their on/off solenoid counterparts. The big difference is in the way their solenoid coils perform. Proportional coils operate on DC current and produce varying force with varying voltage. The symbol shows the solenoid slash in the operator box with a sloping arrow through the slash. This indicates the solenoid has variable force that moves the spool more or less as voltage increases and falls. The other indication on the symbol that shows the spool is infinitely variable is the parallel lines down both sides of the boxes. Proportional valves operate similarly to manual valves, but they use electronics instead of hand power.

To eliminate flow lag from spool overlap, most manufacturers cut vee notches or use some similar method that allows some flow to pass as soon as the spool moves. Vee notches also give smooth flow buildup until the spool moves through the land overlap.

Proportional valves only have two center configurations (as shown by the symbols in Figure 12-1). This means that pressure-compensated pumps with accumulators normally power circuits with proportional valves. The circuits are pressurized at all times to produce fast response from an actuator when motion is called for. (A pressure-compensated circuit also wastes the least amount of energy when throttling flow.)

The valve in Figure 12-1 depends on a certain voltage to move the spool a certain distance to pass a certain flow. This works reasonably well, but is not accurate over a broad range of pressures, flows, and temperatures. Most valves of this design are used to smoothly accelerate and/or decelerate an actuator. The spool is electronically controlled to shift over a period of time to increase flow at a controlled rate. Spool-shift speed can be controlled electronically as it opens and closes to give smooth acceleration and deceleration. Spool-shift distance can also be limited electronically to set a maximum speed when required.

Fig. 12-2. Direct-acting proportional directional control valve with spool-position feedback transducer

To give better spool control, a linear variable-displacement transducer (LVDT) is added to the basic valve. The cutaway view and symbol in Figure 12-2 represent a direct-acting valve with an LVDT. Feedback from the LVDT tells the electronic controller the spool’s position and makes sure it goes to the same place when it receives the same signal. With this arrangement, the spool always shifts to an exact location and opens the same size orifice so it can pass the same flow when pressure drop and viscosity stay the same. Control of flow is more accurate with an LVDT but pressure drop does not stay constant and viscosity often changes throughout the day so speed variations are still apparent. Such flow variations caused by system pressure fluctuation can almost be eliminated by the addition of a hydrostat module in port P as shown in Figure 12-3.

Fig. 12-3. Direct-acting proportional directional control valve with LVDT and pressure-compensating hydrostat module

A hydrostat is simply a pressure-reducing valve set to hold downstream pressure in a 100- to 150-psi range. However, a hydrostat has a pilot line from a shuttle valve that reads downstream pressure at ports A and B, then feeds it back to the bias-spring end of the spool that controls the 100 to 150 psi. The function of a hydrostat is to maintain a constant pressure drop across the spool orifice so flow stays constant regardless of changes in system pressure. (For a thorough understanding of how pressure compensation works, see the pressure-compensated flow control valve section in Chapter 13.)

With the addition of a hydrostat, actuator speed is controlled as accurately as possible without a closed-loop electronic circuit that reads speed and modifies spool position. Closed-loop electronic circuits are used with proportional valve systems, but they only give nominal control. When accurate control is required, use servovalves with closed-loop electronics.

The symbol and cutaway view in Figure 12-4 is for a proportional valve that only controls flow. Such valves are commonly called throttle valves because they are not pressure compensated unless a hydrostat module is added.

Fig. 12-4. Direct-acting proportional throttle valve with spool-position feedback transducer

Basic operation is identical to the proportional control valves just discussed. The only difference is they have a single solenoid and may be piped with dual flow (as shown). Dual-flow piping allows a given size valve to pass twice the volume at the same pressure drop. It can be used with any 4-way directional control valve with one precaution: the valve must be capable of handling maximum system pressure in its tank port. Many wet-armature valves will not operate at full rated pressure at their tank port. Check the supplier’s catalog to see what maximum tank line pressure is allowed. Air-gap solenoid valves and solenoid pilot-operated valves with external drains normally allow full rated pressure in the tank port.

The circuit in Figure 12-5 shows a possible use for a proportional throttle valve. The vertical-down acting cylinder with a platen needs speed, acceleration, and deceleration control. This could be done with a 4-way proportional valve, but the circuit uses an inline or screw-in cartridge valve that is not directly replaceable. Adding a proportional throttle valve to the tank line of the present 4-way circuit can give the required control without extensive piping changes. The circuit is shown using a single flow path for low volume. A dual flow path setup (like the one in Figure 12-4) would allow as much as twice the flow. As stated earlier, make sure the 4-way directional control valve can accept tank-line backpressure without damaging it or causing a malfunction.

Fig. 12-5. Typical circuit using a direct-acting proportional throttle valve with spool-position feedback transducers

If the cylinder had a resistive load, the proportional throttle valve could be placed in the pump line of the 4-way as a meter-in flow-control circuit. A meter-in circuit would not damage the directional control valve or cause it to malfunction. The circuit in Figure 12-5 could have a counterbalance valve in the rod-end line to make it resistive. (See Chapter 14 for counterbalance valve operation and applications.)

The platen would start to move at a controlled rate when the 4-way valve shifts and the proportional throttle valve is signaled to open slowly. The shift time for the throttle valve determines the acceleration time, while shift travel distance determines maximum cylinder speed. Cylinder speed would be infinitely variable to match any production need.

Near the end of the stroke, a slowdown limit switch would signal the proportional throttle valve to start shifting back to its closed position. The proportional throttle valve would close at a controlled rate and flow from the cylinder would be retarded smoothly. When the cylinder slows sufficiently, it contacts the end-of-stroke limit switch and the 4-way directional control valve shifts to center to stop it.

The circuit in Figure 12-5 would only hold position if some retract signal was applied when stopped. This is due to internal leakage by the spool of the 4-way directional control valve and the proportional throttle valve. Another option would be to use a float-center directional control valve and a counterbalance valve.

Proportional valves for flows higher than 25-to-30 gpm use solenoid pilot-operation similar to conventional directional control valves. A small pilot-operated valve receives a signal and then sends hydraulic oil to proportionally move a larger control spool that controls actuator movement.

The cutaway view and symbol in Figure 12-6 depict a typical solenoid-pilot valve arrangement. A reducing valve module, between the pilot operator and the pilot-operated valve, keeps maximum pilot pressure below 200 psi. A proportional Input to one of the coils on the pilot operator directs flow to the spool of the pilot-operated valve and shifts it against a spring. As pressure against the spool increases, it shifts farther and sends more flow to the actuator. Feedback signals from both spools tell the electronic controls that the command has been carried out. A vee-notched spool allows flow to increase at a smooth rate so actuator speed is consistent throughout the speed range.

Fig. 12-6. Solenoid pilot-operated proportional directional control valve with spool-position feedback transducers

Ports for internal or external pilot X or drain Y provide options for these control lines to meet a particular requirement.

The complete symbol is shown in Figure 12-6. (For the simplified symbol, see Chapter 4.)

Flows up to 200 gpm are common for a D10-size proportional valve. For higher flows, use slip-in cartridge valves (discussed in Chapter 11) with proportional operators.

Hydraulic servo directional control valves

Proportional control valves are infinitely variable but they are neither highly responsive nor capable of handling minute flow changes rapidly and accurately. On the other hand, servovalves easily meet both of these requirements . . . but at a cost. They are more expensive than proportional valves, they require super-clean fluid, and they need extra electronics to exploit their full capabilities.

The three common servovalve types are flapper, jet pipe, and mechanical. Each design has advantages as far as operation accuracy, leakage, contamination tolerance, and price. They range in flow capacity from less than 1 gpm to more than 1000 gpm. Most manufacturers make valves that operate at 3000 psi, but some offer valves at 5000 psi.

The main difference between proportional and servovalve circuit design is that servo systems have a method of feedback that assures that the actuator is doing what the controller tells it to do. A super-simple form of servo control would be a backhoe operator moving manual valves to cause a bucket to move toward him at a given rate. Feedback from the operator’s eyes would tell his hands when and how far to move the levers to give more or less flow to maintain the action he wants. Other familiar mechanical feedback examples are hydraulic driven power steering and hydraulic power brakes on a vehicle.

The circuit in Figure 12-7 is an example of a working circuit with mechanical feedback that controls a hydraulic press. The operator needs to have a feel for the motion of a platen as it cuts through some tubing. Originally this was done with an arbor press, but it was hard for the operator to keep up with production due to the physical exertion. Now all the operator has to do physically is overcome the spring force of the manually controlled mobile directional valve to make the platen move. As the platen moves, the directional valve body also moves, so the operator has to keep moving the lever to advance or retract the platen. Notice the mechanical link between the valve body and the cylinder rod that moves the valve body at the same rate the cylinder rod moves. The operator now has a hydraulic force multiplier that gives some feel to what he is doing.

Fig. 12-7. Simple mechanical servo system for force multiplication

The reason for using a mobile-type valve is because those valves have less spool overlap and the spool has notches cut in it. The notches pass a small flow almost immediately when the spool moves. That flow increases in proportion to spool movement.

Most industrial applications use feedback from electronic linear, rotary, or force transducers. A transducer is a device that produces an electrical signal in direct relation to a position, force, or speed.

Linear potentiometers work for short strokes (12 in. or less). Longer strokes require a device such as a Temposonics transducer. In either case, these devices feed a precise position or speed indication back to an electronic controller.

For rotary motion, an encoder or similar device that produces multiple pulses per revolution sends a signal about rpm or angle of rotation to the controller.

When information about force is required, a load cell sends the data to the controller.

With these very accurate feedback devices and a fast-response servovalve, an actuator’s position, speed, and/or force can be repeatedly established within an extremely close range. Electronics provides the accuracy while hydraulics provides the force via a super-responsive servovalve.

The cutaway view and symbol in Figure 12-8 show a less-responsive but more contamination-tolerant servovalve. There are other mechanical ways of driving the spool. The valve in Figure 12-9 uses a rotary drive and an eccentric to move the spool left or right to an infinite number of positions. Because the drive is quite strong and there are no orifices to clog, this valve can operate with fluid that meets ISO Code 4406 20/16/13.

Fig. 12-8. Rotary-drive servovalve

Notice the difference in design between spools in proportional valves and servovalves. Most proportional valves use spools with overlap and some sort of notches that pass flow while moving out of overlap. A servovalve has no overlap or underlap of the spool lands to the body lands. (One manufacturer calls it “Critical lap” because all points blocking fluid cannot move without passing flow.) This spool design makes the valve very responsive (as well as very expensive and prone to above average bypass). Servovalve spools and bodies always come in matched sets because of their close fit and four points of land-to-land match.

The rotary-drive eccentric valve pictured in Figure 12-8 has fast, controllable spool movement from a rotary drive that incorporates a feedback loop. When the drive receives a signal to move the spool to pass a certain flow, a position feedback output sends a signal back to the controller when the motion is complete. There is still feedback from the actuator that what was commanded is happening, so spool position can be changed via the electronic feedback and controller as necessary.

Several factors determine when a given input will not produce the desired actuator output. The main factor is actuator load. As load changes, input force must change -- by allowing more or less fluid into the circuit. Fluid viscosity also has an effect, so the flow path must be reduced as viscosity lowers and enlarged as viscosity rises. Then there is system pressure. As pressure fluctuates, flow across the spool orifice changes. The higher the pressure drop, the greater the flow. Because a servovalve circuit has feedback from the actuator, it can adjust flow or pressure to match system changes continuously.

Figure 12-9 shows another mechanically driven servovalve. This setup works for cylinders and hydraulic motors, but must be directly attached to the actuator (as shown) or driven by it with a toothed belt and pulleys. This is a very contamination-tolerant valve arrangement because the stepper motor is quite strong and there are no small orifices to clog. The spool has no overlap or underlap so any movement immediately initiates fluid flow to and from the cylinder. The piston and rod cannot rotate so the feedback screw turns as the piston extends or retracts.

Fig. 12-9. Stepper-motor-driven servovalve with mechanical feedback

As the stepper motor turns, a threaded rod inside the threaded spool moves the spool to direct fluid to extend or retract the piston. When the cylinder piston moves, the feedback screw turns the spool back on the threaded rod to counteract the stepper motor shift. When the stepper motor turns, the piston moves at a speed proportional to the stepper motor’s rpm. When the stepper motor stops, the piston catches up and stops also. Manufacturers claim a tolerance of ±0.001 in. repeatability. If an external force tries to push or pull the piston out of place, the feedback screw shifts the spool and fluid starts resisting movement.

From the foregoing explanation, it is easy to see how this valve – when attached to a hydraulic motor -- would give an exact number of turns and repeatedly cause the motor to stop at exactly the same place. This will happens even if the hydraulic motor has internal leakage. The only time the actuator gets out of place is when it cannot overcome the load and stalls. The stepper motor continues to receive pulses, but it also stalls when the spool shifts all the way. The stepper motor received a signal that should have placed the actuator at a certain distance but the actuator did not get there because of insufficient force. When the valve reverses, it starts its motion from the wrong point and will overshoot home position as it returns. A limit switch at the home position can alert the controller that there is a problem when the actuator overshoots -- and prevent the machine from producing scrap.

The jet-pipe servovalve pictured in Figure 12-10 also tolerates contamination due to a control orifice that is large enough to pass large particles. Pilot oil is tapped off the system fluid inlet, sent through a coarse filter, and on to the jet pipe that terminates in an orifice. The orifice outlet is centered over the inlet of two passages that terminate at each end of a critical-lap spool. Flow into these passages puts equal pressure on both ends of the spool as the feedback wire holds it centered. Current signals to the coils cause the armature to rotate and shift more of the output of the jet pipe to one passageway than the other. Pressure increases on one end of the spool and decreases on the other end. As the spool shifts, it starts to pass flow to the actuator at a rate set by the input electrical signal.

Fig. 12-10. Jet-pipe servovalve

When the jet pipe shifts to the left, the spool moves to the right. At the same time, the feedback wire also moves to the right, pulling the jet pipe nozzle back to center and stopping spool movement. A given input to the coils electromechanically shifts the armature that moves the jet pipe. This moves the spool hydraulically and forces the jet pipe back to center mechanically through the feedback wire.

A measured electrical input to a servovalve produces a fixed flow output, similar to a proportional valve. This control alone does not give much better control than a proportional circuit even though the valve is more responsive. There is still no compensation for viscosity or pressure changes that can cause the actuator’s speed to fluctuate. To overcome this problem, some sort of electronic feedback from the actuator is necessary. The feedback signal through an electronic circuit board modifies the signal to the servovalve to make the actuator perform as planned. Actually, the electronics do the work as long as the valve can respond quickly enough to keep everything working at the correct rate. This means the spool must be free enough to move easily without excessive bypass. Anytime the spool moves, it should pass flow to and from the actuator.

The valve in Figure 12-10 is considered a 2-stage valve. The first stage is electronic and receives an electronic input signal, while the second stage is fluid powered by a hydraulic signal.

The jet-pipe servovalve depends on clean oil for long trouble-free operation -- not as clean as the requirement for the flapper-valve design discussed next, but clean enough to prevent the jet-pipe nozzle from clogging. It is obvious that once nozzle flow is retarded enough or stops, the valve loses all ability to control flow to the actuator.

The most responsive and accurate servovalve design is the flapper valve, shown in Figure 12-11. This design is the least tolerant of contamination because it depends on very small orifices for fast response with minimal wasted energy. It is called a flapper valve because the element that holds equal pressure on both ends of the spool at rest reminds one of a flapping device. It is a 2-stage valve with an electronically controlled torque motor as the first stage and a pilot-operated spool as the second stage. As in all servovalves, the spool has no overlap or underlap that would make it sluggish or bypass a lot of fluid unnecessarily.

Fig. 12-11. Flapper-design servovalve

Fluid from the pump inlet is tapped off through rather-coarse filter elements, passes through orifices past both ends of the spool, goes on to nozzles, and out to the return line. The orifice diameters are slightly larger than the nozzle diameters, so there is a pressure buildup at both ends of the spool. A feedback wire attached to the flapper terminates in a ball end that sits in a very close-fit slot in the spool. A sleeve around the spool can be moved left or right by a null adjustment to align the spool and body lands perfectly when the valve is first installed. (Usually null adjustment is only required at startup of the valve.)

The null adjustment usually is a hexagonal wrench fitting attached to an eccentric pin located in the sleeve slot. With the null adjustment centered, turning it one round moves the sleeve from center to full right, back to center to full left, and back to center. If the valve cannot be nulled within one rotation of the null adjustment, replace it and send it in for repair. This usually indicates a clogged orifice or nozzle controlling one end of the spool.

Unplug the electrical supply to the valve before setting null. Start the pump and watch for actuator movement. If the actuator moves, loosen the null lock screw and carefully turn it. Observe whether the actuator slows or picks up speed. A nulled valve stops actuator movement because the forces on both sides are equal. High-flow 3-stage valves cannot be nulled to the point of stopping an actuator due to the piloted spool slipping by the stop-flow position as the pilot operator is adjusted. When null is set, lock the null screw and reattach the electrical plug.

Turning the null screw with the electric plug detached is one way of moving an actuator manually. This might be done to prove the valve is working properly and the problem is electrical.

When the torque-motor coils receive a current signal, the armature rotates clockwise or counter-clockwise and pushes the flapper closer to one nozzle and farther away from the opposite one. This allows pressure to increase at one end of the spool and decrease at the other. The spool then starts to move away from the higher pressure. If the armature turns clockwise, pressure builds on the left end of the spool and it moves to the right, as shown in the left cutaway view of Figure 12-12.

Fig. 12-12. Flapper-type servovalve shifting from an electrical input signal

As the spool moves to the right, it also drives the feedback wire to the right. The feedback wire is strong enough to overcome armature force and pull the flapper back to center. After the flapper centers, pressure is equal on both ends of the spool and it stops. More current to the coils causes more rotation and additional spool shift until the feedback wire again centers the flapper.

From the foregoing explanation, it is obvious why this valve needs clean oil. If an orifice or nozzle clogs, the spool shifts all the way to one end and the actuator moves until it runs into a resistance it can’t overcome. Also, the spool must start shifting at a very low pressure drop across it to keep response high. Contaminated fluid can cause sticking and require high differential shifting pressure that makes spool movement erratic.

For flows above 60 to 80 gpm, a 3-stage servovalve is required. It consists of a small 2-stage pilot operating a large pilot-operated spool, as depicted in Figure 12-13. The 2-stage valve operates as just explained, but its output goes to move a pilot-operated spool in the third stage to a precise position to control high flow to large actuators.

Fig. 12-13. Three-stage flapper-type servovalve

An LVDT signals the electronic control circuit that the pilot-operated spool is where it was signaled to go. After receiving that position signal, the 2-stage valve shifts to no flow or whatever flow it takes to keep the pilot-operated spool in place.

A 3-stage valve also depends on feedback signals from the actuator to modify the input signal when the action is not in compliance with the command. This makes 3-stage valves very accurate controllers of large cylinders.

Pneumatic proportional and servo directional control valves

Since the late ‘80s, several companies have been controlling air cylinders with open- and closed-loop proportional or servovalve circuits. The difference between the air valves in these circuits is how fast they respond. Most proportional valves have a sealed spool that controls direction and flow so the valves tend to hang up and jump. Pneumatic servovalves often have spools with metal-to-metal fits that float on bypass air.

The valve shown in Figure 12-8 is sold as a proportional or servo directional control valve for hydraulic or air circuits. Controlling the amount of air and which direction it goes is not a problem but the compressibility of air creates some giant hurdles to overcome.

Proportional valves usually control only acceleration, deceleration and/or speed because these circuits do not include feedback transducers. It is very easy to get smooth acceleration and deceleration with high speed in between without other controls or shock absorbers to stop the load mechanically.

Adding feedback transducers to a proportional air circuit can provide servo-like control for light loads -- such as those found in pick-and-place applications. However, a proportional valve is usually not responsive enough for exacting part placement or speed control. For very accurate control, a servovalve with feedback transducers can give close-tolerance positioning (with light loads), repeatable velocity control, and very accurate holding force.

Pneumatic proportional and servovalves are not a replacement for electromechanical or servo hydraulics, but they have price advantages over both systems. When the loads are light and cost is a factor, they are worth a look.

General information for hydraulic infinitely variable valves

  • The symbol for proportional and servovalves shows a 4-way, 3-position function and the valve can move to each of the positions. However, the parallel lines along the sides of the symbol indicate the valve does not have to shift all the way all at once. These valves can shift into straight or crossed arrows in any proportion from 0 to 100%. They are infinitely variable and can pass any flow desired.
  • servovalves are always 3-position, all-ports-blocked center condition, as shown by the symbols in Figures 12-8 through 12-11.
  • always size proportional or servovalves for high pressure drop. Proportional valves should have 200- to 500-psi pressure drop at full flow. Most servovalve manufacturers rate their valves at 1000-psi pressure drop at full flow. This means the valve may look physically small for a given flow in relation to conventional valves. It also means most servo and proportional valve circuits require a heat exchanger to deal with excess wasted energy.
  • always mount the valve as close as possible to the actuator ports. Any piping between the valve and the actuator holds extra fluid that can make the system softer and less responsive. This is especially important on air-powered circuits.
  • never use hose between the valve and the actuator. If isolation is necessary, mount the valve on the moving part and use flexible lines for supply and return.
  • use in-line pressure filters at the supply to each servovalve or bank of valves to protect the valve from contamination in the pump and piping.

Specific Information for pneumatic infinitely variable valves

  • use air at the highest pressure possible that does not exceed component or plumbing limits. This is usually as high as 250 psi.
  • size the valve to flow just enough to produce the maximum desired actuator speed. Oversize valves produce erratic control because a small spool movement gives more flow than required.
  • use an actuator with the largest area practicable so the load moves with a low pressure difference. Note: the larger the cylinder, the more air it consumes so operating cost escalates.
  • pneumatic servo circuits do not work well when outside forces push against the actuator. The actuator tries to resist, but force buildup is slow in comparison to electromechanical or electrohydraulic systems.

Typical servo circuits

Figure 12-14 shows schematic drawings of three typical servo circuits. In the figure, each type circuit controls a different actuator, but any actuator could have more than one type of control.

Fig. 12-14. Servovalve and closed-loop electronic circuit for accurate position, force, and speed control

The typical power unit for a servo system is a pressure-compensated pump with an accumulator or accumulators. A servovalve must always have a ready supply of fluid because no matter how fast it reacts, without an immediate supply of fluid the system will be sluggish. The pressure-compensated pump may be at full pressure when no actuators are moving, but its flow is zero. Adding accumulators assures that there is no wait for the pump to come on stroke before the actuator receives flow. (A full discussion of how accumulators work and how they are applied is in Chapter 16.)

Again, pressure filters in the lines to the servovalves make sure they receive clean oil. One filter could be sufficient for multiple valves when the valves are close to each other. The reason for pressure filters in the valve lines is the pump constantly produces contamination particles that will shut down flapper-type servovalves.

Cylinder 1 is in a position loop. It can be placed at a precise location repeatedly within ±0.0005 in. A programmable logic controller (PLC) sends a signal to the summing amplifier’s control card. The signal passes on to the valve-driver card and then to the valve coils. This signal shifts the servovalve to start cylinder movement at a set rate. The linear potentiometer sends position feedback to the summing amplifier and modifies valve position to find and maintain a certain position. Often, position control is paired with speed control to accelerate the actuator to a certain speed, then decelerate and stop it at the desired position.

Cylinder 2 is in a force loop. A certain size cylinder operating at a given pressure produces a given force. This force can be calculated by multiplying area times pressure, but the result is not exact. Friction from seals and between external machine members can reduce this force by a few pounds or more on an operating machine. When an exact force calculation is required, a servovalve-controlled cylinder that has a load cell for feedback can keep forces within 1/2% with ease. The summing amplifier sends a signal from the PLC and feedback from the load cell modifies the valve position to exactly match the input signal to generate the desired force.

The hydraulic motor is in a speed loop that maintains the motor’s rpm when the fluid viscosity, pressure, or load changes. A rotary device called an encoder constantly sends rpm information back to the summing amplifier to open or close the servovalve as needed. Just as a cylinder does, a hydraulic motor will slow when the load increases, when fluid gets thinner due to temperature increases, or when system pressure fluctuates as other actuators move. If the encoder sends a reduced-rpm signal back, the servovalve opens to let more fluid in. If the hydraulic motor tries to speed up, the servovalve closes enough to maintain the set speed.

Other infinitely variable valve applications

Proportional and servovalves can be applied to variable-volume pump controls to accurately set flow from the pump in relation to feedback from an actuator. Most pressure-compensated pumps available today have this feature as an option.

An example might be a closed-loop hydrostatic pump and motor that must stay at a constant speed. Feedback from an encoder signals the pump to increase or decrease flow as the motor overspeeds or underspeeds.

Proportional coils make excellent infinitely variable flow controls that can be controlled electrically. They can act as simple throttle valves or perform as full-blown pressure-compensated types that adjust to pressure fluctuations.

Common circuits for proportional flow controls include feed-speed changes to deal with load fluctuations; hydraulic motor rpm changes to control product backup; or maintaining constant speed as loads vary. (Pressure-compensated proportional flow control valves are covered in Chapter 13.)

Proportional coils also make excellent infinitely variable pressure-control valves. Pressure settings from remote locations or a PLC at anytime from minimum to maximum almost instantaneously. Most of these pressure controls are for relief and reducing functions but could be used on any of the pressure controls. (Proportional pressure-relief valves were covered in Chapter 9.)

Quiz

 

Chapter 13: Flow Controls and Flow Dividers

Speed control of hydraulic and pneumatic actuators

In some applications, there are times when it is necessary to vary the speed of an actuator. One method of controlling an actuator’s speed is by using a variable-volume pump. This works well for a circuit with a single actuator or in multi-actuator circuits where only one actuator moves at a time. However, most circuits that need actuator-speed control have multiple actuators and some of them operate simultaneously. For most circuits, a variable orifice called a needle valve or flow control is common. Fixed orifices may be used in some cases.

Non-compensated flow control valves

Figure 13-1 shows non-compensated flow devices in symbol and cutaway form. At the top are non-compensated fixed-orifice in-line flow controls for tamper-proof applications. These can be purchased as in-line valves or they could be a drilled plug or insert located in a pipe fitting or valve port.

Fig. 13-1. Non-compensated flow devices

Flow through standard orifices is affected by viscosity changes in the fluid, while flow through knife-edge (or sharp-edge) orifices changes very little when fluid viscosity changes from thin to thick. A knife-edge orifice is the style used on most valves that are designated as temperature compensated. (A classic example of a non-compensated fixed orifice with a bypass check is the orificed check valve shown in Figure 10-2.)

Pressure-compensated flow control valves

The pressure-compensated flow control cutaway view and symbols depicted in Figure 13-2 are the component used with actuators that must move at a constant rate. A non-compensated flow control passes more or less fluid as pressure raises and lowers. This is because more fluid can pass through a certain size orifice when pressure drop across the orifice increases.

Fig. 13-2. Pressure- and temperature-compensated flow control

The needle valve section of a pressure-compensated flow control is the same as any flow control. The difference is the addition of a compensator spool that can move to restrict Inlet flow at the compensating orifice. The compensator spool is held open by a 100- to 150-psi bias spring that sets pressure drop across the knife-edge orifice.

Flow from the inlet goes through the compensating orifice, past the compensator spool, and out through the knife-edge orifice. A drilled passage ports Inlet fluid to the right end of the compensator spool, which forces the spool to the left when pressure tries to go above 100 to 150 psi at gauge PG01. After pressure reaches or goes above 100 to 150 psi, the compensator spool moves to the left and restricts flow to the knife-edge orifice flow control. Pressure at gauge PG01 never goes above 100 to 150 psi (plus any backpressure at the outlet). Pressure at the outlet is ported to the bias-spring chamber and increases the spring force. The compensator spool assures that pressure drop across the knife-edge orifice flow control stays at a constant 100 to 150 psi. With a constant pressure drop, flow stays the same regardless of inlet or outlet fluctuations.

Pressure-compensated flow controls are four to eight times more expensive than standard controls so they should only be applied to actuators that must move consistently.

The no-jump option is an adjusting screw that holds the compensator spool within a few tenths of an inch of its operating position. This is an especially important option when the valve is oversize for the present flow setting. A compensator spool without a stroke limiter may close and open violently until it stabilizes and sets pressure drop for the orifice. During this time the actuator also moves erratically.

The two symbols represent the American National Standards Institute (ANSI) and the International Standards Organization (ISO) way of indicating that the valve is pressure compensated. The arrow indicating pressure compensation is easier to distinguish in the ANSI symbol -- especially when the schematic drawing has been reduced to fit into a machine's documentation book.

Three-port flow control valve

Three-port flow controls are mainly used in fixed-volume pump circuits to save energy. (See the load-sensing pump circuit explained in Chapter 8.) If 20 gpm of fluid enters the Inlet and the flow control is set at 12 gpm, 8 gpm goes to tank as wasted energy. With a conventional relief valve setup, pressure between the pump and flow control would be maximum. With the 3-port flow control, pressure in this portion of the circuit is whatever it takes to move the actuator plus bias-spring force. (Bias-spring force is usually 70 to 125 lb.) An outlet pressure of 200 psi gives a pressure of 270 psi between the pump and the flow control. All fluid going to tank is discharged at 270 psi, not 2000 psi. This takes place because the sensing line sends feedback to the pressure-control side of the relief valve, allowing it to open at load pressure plus bias-spring force. Pressure between the pump and flow control constantly changes with load variations. When the load requires more than the maximum-pressure adjustment setting, the relief valve opens and sends all pump flow to tank at maximum pressure.

Fig. 13-3. Three-port flow control

A 3-port flow control is only effective with one actuator -- or one actuator at a time. It would not be useful on a pressure-compensated pump circuit because a load-sensing circuit for this type pump would save even more energy. (See Chapter 8 for a load-sensing circuit with a pressure-compensated pump.)

Proportional flow control valves

Figures 13-4 and 13-5 show cutaways and symbols for proportional flow control valves that can electronically remotely control flow through a PLC or other controller. There are many different designs of valves and controllers that control pneumatic or hydraulic fluid. The design in Figure 13-4 uses a modified 2-way pilot-to-close poppet with a drilled pilot passage to send inlet fluid behind it. A light spring holds the poppet closed when there is no pressurized fluid at the Inlet.

Fig. 13-4. Proportional flow control valve without feedback

The armature controls a small normally closed poppet and shifts the signaled amount to let fluid behind the pilot-to-close poppet leave faster than the pilot passage can supply it. This causes a pressure imbalance that lets the pilot-to-close poppet open enough to give the correct fluid flow. The flow rate is infinitely variable and can be controlled from a variety of inputs.

Fig. 13-5. Proportional flow control valve with feedback

The valve in Figure 13-4 opens from a given signal but may not always repeat a set flow from the same input. The feedback LVDT added to the valve in Figure 13-5 assures that the pilot-to-close poppet always shifts the same amount so it has the same size flow opening. However, pressure or viscosity changes still affect actual flow, so a hydrostat is necessary when exact flow repeatability is required. Many manufacturers make valves with a built-in hydrostat for pressure compensation.

Meter-in flow control circuits

Figure 13-6 provides a schematic drawing of a meter-In flow control circuit restricting fluid as it enters an actuator port. Meter-in circuits work well with hydraulic fluids, but can give erratic action with air. Note that the cylinder is horizontally mounted, which makes it a resistive load. Meter-in flow controls only work on resistive loads because a running-away load can move the actuator faster than the circuit can fill it with fluid.

Fig. 13-6. Meter-in flow control circuit

The left-hand circuit in Figure 13-6 is shown at rest with the pump running. Notice that the check valves in the flow controls force fluid through the orifices as it enters the cylinder and lets fluid bypass them as it leaves.

The right-hand circuit depicts conditions as the cylinder extends. The directional control valve shifts to straight arrows and pump flow passes through the left-hand flow control to the cylinder cap end at a controlled rate. Fluid leaving the cylinder rod end flows to tank without restriction. The cylinder extends at a reduced speed (in a hydraulic circuit) until it meets a resistance it can’t overcome or it bottoms out. With the non-compensated valve shown, speed can vary as pressure fluctuates or viscosity changes.

While the cylinder is in motion, pressure at PG1 reads the setting of the relief valve or pump compensator. The pressure at PG2 reads whatever it takes to move the load at any point in the cycle. Pressures at PG3 and PG4 only read tank-line backpressure as the cylinder extends.

It is obvious that if the cylinder had an external force pulling on it, it would extend rapidly. Because fluid enters the cap end at a reduced flow rate, a vacuum void would form there until the pump had time to fill it.

Meter-in flow controls can have a problem in pneumatic circuits. When fluid is directed to the cylinder cap end, pressure at PG1 immediately rises to the regulator setting. However, pressure at PG2 starts at zero and increases slowly. Until pressure at PG2 rises enough to generate breakaway force, the cylinder does not move. At breakaway pressure, the cylinder extends quickly and expanding air may cause it to lunge. Often, the lunge forward moves the piston ahead of the incoming air and pressure drops back below the breakaway level so the piston stops. Pressure starts to build again and the lunge/stop scenario continues to the end of stroke. The meter-out circuit discussed next is always the best choice to control air cylinders.

The circuits in Figure 13-7 show applications where a meter-in circuit is the only choice for both pneumatics and hydraulics. On the left in Figure 13-7, a single-acting pneumatic cylinder is mounted with the rod vertically up. The only way to control extension speed is via a meter-in flow control. When retraction speed must be controlled as well, a meter-out flow control also is necessary.

Fig. 13-7. Circuits where meter-in flow control is required

The cylinder pictured on the right in Figure 13-7 is extending to perform an operation prior to retracting or starting the cycle of another actuator. A signal to continue the cycle can come from a pressure switch or a sequence valve. Either of these devices can be set to give an output at any pressure. Usually they are set 50 to 150 psi below system operating pressure for hydraulics, or 5 to 15 psi lower for air. The reason for meter-in flow control is that pressure between the flow control and the cylinder normally stays low until the cylinder contacts the workpiece. At work contact, the resulting pressure buildup switches these pressure-actuated devices and starts the next sequence. Always remember: a pressure switch or sequence valve does not directly indicate that the actuator has reached a physical position. They only indicate that pressure has reached a predetermined setting . . . not why it has.

Other circuits that require meter-in flow controls are the load-sensing pump circuits in Chapter 8.

Meter-out flow control circuits

Figure 13-8 shows a schematic drawing of a meter-out flow control circuit that restricts fluid as it leaves an actuator port. Meter-out circuits work well with both hydraulic and pneumatic actuators. Cylinder-mounting attitude is not important because outlet flow is restricted and an actuator cannot run away. Meter-out flow controls work on resistive loads or running away loads because the actuator can never move faster than the fluid leaving it allows.

Fig. 13-8. Meter-out flow control circuit

The left-hand circuit in Figure 13-8 is shown at rest with the pump running. Notice how the check valves in the flow controls allow fluid to bypass the orifices and freely enter the cylinder. As fluid leaves the cylinder, it is forced through the orifices at a set rate. The only gauge showing pressure is PG3 because the load on the cylinder rod is inducing pressure at the valve’s blocked port.

The right-hand circuit shows conditions when the cylinder is extending. The directional control valve shifts to straight arrows and pump flow bypasses the upper flow control to go to the cylinder cap end. Fluid leaving the cylinder rod end is held back before it goes to tank -- even with an external load trying to move it. The cylinder extends at a reduced speed in both hydraulic and pneumatic circuits until it meets a resistance it can’t overcome or it bottoms out. With the non-compensated valve shown, speed can vary as pressure fluctuates or viscosity changes in a hydraulic system. (There are no pressure-compensated flow controls for pneumatic circuits.)

While the cylinder is in motion, gauges PG1 and PG2 read the relief valve or pump compensator setting. Gauge PG4 reads tank backpressure. Gauge PG3 reads load-induced pressure plus the pressure from cap-area-to-rod-area intensification. This intensified pressure could be 1.2 to 2 times the cap-end pressure, or higher, depending on the rod size.

Meter-out flow controls work equally well in pneumatic circuits when the load is constant. Changing loads can cause the actuator to stop and/or lunge under certain circumstances. (For a more extensive coverage of flow control circuits and situations that can arise with them, see our second e-book entitled "Fluid Power Circuits Explained," which will be launched on hydraulicspneumatics.com in the coming months.

Bleed-off flow control circuits

Bleed-off flow control circuits are found only in hydraulic systems and normally only in those with fixed-volume pumps. There is little or no advantage to using this type flow control with pressure-compensated pumps. Figure 13-9 shows a bleed-off circuit at rest with the pump running. A needle valve’s inlet is teed into a line going to the cylinder and its outlet is connected to tank. The circuit only works with one actuator moving at a time because all pump flow goes to the presently operating function. Like a meter-in circuit, it only works with resistive loads because it controls fluid into the actuator. The main plus for this type speed control is it saves energy while using a fixed-volume pump with low-pressure travel forces.

Fig. 13-9. Bleed-off flow control circuit

When the directional valve in Figure 13-9 shifts, all pump flow passes through it and toward the actuator. On the way to the actuator, part of the flow is bled off to tank, so the actuator does not reach full speed. Pressure at PG1 only rises to whatever it takes to move the actuator and its load, so excess flow goes to tank at low pressure. (When using a fixed-volume pump and a meter-in or meter-out circuit, excess flow also goes to tank, but at relief valve pressure.) Many circuits only perform work at the end of stroke so this flow control system saves energy while the actuator moves to and from the work position, yet still gives good speed control.

Some words of caution:

  • Pressure in the actuator during traverse time must be higher than the pressure in the path to tank, so fluid will flow to tank.
  • Because pressure may change during traverse time (especially when the actuator contacts the workpiece), use a pressure-compensated needle valve so flow to tank remains constant.
  • Even with a pressure-compensated needle valve, actuator speed will be inconsistent. Pump and/or actuator efficiency allows bypass that directly affects flow to the actuator not bleed-off to tank.

Pressure-compensated flow control valve applications

When pressure drop across an orifice changes, flow through the orifice also changes. As pressure drop increases, flow increases, and as pressure drop decreases, flow decreases. Because of this fact, if pressure drop across an orifice were constant, regardless of upstream and downstream pressure fluctuations, then flow through it would stay the same. A pressure-compensated flow control valve (such as the one shown in Figure 13-2) automatically maintains a constant pressure drop across the orifice. There is a short discussion on pressure-compensated flow control valves on page 13-1, but a valve in cutaway form is applied to a bleed-off circuit in Figure 13-10.

Fig. 13-10. Bleed-off flow control circuit with positive-displacement pump and pressure-compensated flow control valve

In the bleed-off circuit, fluid from the directional control valve is sent to the cylinder to start it extending. Because the circuit has a fixed-volume pump and needs speed control, a bleed-off flow control is used to save energy. Instead of controlling flow to or from the actuator, excess flow is bled to tank across a pressure-compensated flow control at whatever pressure it takes to move the fluid. A meter-in or meter-out flow control circuit would send excess flow to tank across the relief valve at maximum pressure – wasting a lot more energy.

The reason for using a pressure-compensated flow control is that pressure will fluctuate as the actuator moves toward the workpiece and the flow to tank from a non-compensated flow control would change continuously. As a result, actuator speed could vary considerably while it moves. With a pressure-compensated flow control, flow to tank is constant, but actuator speed could still change due to pump efficiency as pressure increases or decreases. Any speed change from pump efficiency is present but practically imperceptible.

In the Figure 10-13 circuit, a 10-gpm pump sends 7 gpm to the cylinder and 3 gpm to tank. Fluid entering the pressure-compensated flow control passes by the compensator spool and flows on to the variable knife-edge orifice, which is set at 3 gpm. The variable knife-edge orifice restricts flow and creates backpressure in the incoming fluid. When backpressure reaches (and attempts to exceed) 125 psi, fluid in the inlet-pressure pilot line forces the compensator spool to the right. This restricts flow at the compensating orifice. After the compensator spool settles in at its 125-psi bias-spring setting, pressure at PG3 reaches 125 psi and stays there. This means that pressure drop across the variable knife-edge orifice is 125 psi. As the cylinder continues to move and pressure at PG1 and PG2 increases or decreases, pressure at PG4 stays at 125 psi and flow is constant. The cylinder moves at the same speed whether pressure is at or above 125 psi, and as much as 125 psi below the maximum pressure setting.

Fig. 13-11. Meter-in flow control circuit with pressure-compensated pump and pressure-compensated flow control valve

Figure 13-11 shows a pressure-compensated flow control in a meter-in circuit. Fluid from the valve enters the flow control and is restricted. Backpressure from restricted flow goes through the inlet-pressure pilot line and shifts the compensator spool to the right, restricting flow to the variable knife-edge orifice. Backpressure from cylinder resistance acts on the right end of the compensator spool through the outlet-pressure pilot line and adds to the 125-psi bias-spring force. This action and interaction always keeps pressure 125 psi higher at PG5 than at PG2. A constant pressure drop across the orifice maintains a constant flow to the cylinder.

Fig. 13-12. Meter-out flow control circuit with pressure-compensated pump and pressure-compensated flow control valve

Figure 13-12 shows a pressure-compensated flow control in a meter-out circuit. Fluid from the cylinder rod end enters the pressure-compensated flow control and is restricted at the variable knife-edge orifice. Backpressure through the inlet-pressure pilot line shifts the compensator spool to the right and restricts flow to the variable knife-edge orifice. Pressure at PG5 settles in at 125 psi and flow stays the same across the variable knife-edge orifice. Any backpressure from tank flow adds to the 125-psi bias-spring force and increases pressure at PG5 so it always stays 125 psi above PG4.

Pressure-compensated flow control valves are as much as five times more expensive than non-compensated models, so they should not be specified when accurate flow control is not required.

Changes in fluid viscosity also cause flow fluctuations. Thick fluid flows more slowly than thin fluid. A flow control valve without temperature compensation allows varying flow from cool oil at startup to oil running at normal or high temperature. The most common fix for viscosity variations is to use a knife-edge orifice. Knife-edge orifices have no flats to slow fluid flow, so they produce little change in flow between thick and thin fluids. Other devices to obtain constant flow with viscosity variations are available, but they can be complex and may cause malfunctions.

A flow control in a hydraulic circuit always generates heat. Some pump and flow control combinations produce a lot more heat and should be avoided if possible. The following examples show different pump and flow control combinations and suggest how much heat can be expected.

Fig. 13-13. Heat generation in fixed-volume pump circuits with meter-in and meter-out flow controls

The fixed-volume pump and meter-in or meter-out flow control combination in Figure 13-13 is the worst-case situation. The example shows a cylinder stroking to the workpiece with flow controls set at 3 gpm. A 10-gpm pump driven by a 5-hp electric motor powers the circuit. Because it only takes 100 psi to move the cylinder while traversing, a lot of heat-generating energy is wasted. This example is somewhat exaggerated, but is not at all unheard of. Note the example only shows energy wasted on the extension stroke. With a reduced-speed retraction stroke, heat generation could almost double the figures shown.

The main generator of heat is the excess pump flow going across the relief valve at 1000 psi. The two circuits in Figure 13-14 show how to eliminate such wasted energy with a different flow control circuit or a different pump. While the energy wasted across the flow control valve is much less at these low flows, it still adds heat to a system. Also, the amount of pressure drop may be lower than indicated here because some actuators require more pressure to move them to and from the workpiece. Energy loss across a flow control cannot be eliminated. The amount of loss depends on pressure drop and flow rate across the orifice.

Fig. 13-14. Two flow control circuits that reduce heat generation

The circuits in Figure 13-14 show a fixed-volume pump with a bleed-off circuit and a pressure-compensated pump with a meter-in circuit. Both of these combinations save a lot of energy (although not as much as the load-sensing circuit that was shown in Figure 8-27). This type of flow control circuit wastes the least energy possible when using flow controls for speed control.

Fluid flow dividers

The flow divider in Figure 13-15 is called a priority flow divider because it splits pump flow into a fixed controlled-flow (CF) outlet and sends excess fluid out an excess flow (EF) port. Volume orifices (drilled as specified by the purchaser) preset fluid flow out of the CF port. EF flow is any flow the pump produces over and above the controlled flow. This type flow divider is often used on vehicle power steering, where an engine-driven pump’s output may vary as rpm changes or as its flow is used for other functions. A priority flow divider assures that the power steering always has ample fluid at any engine speed or when other functions are active.

Fig. 13-15. Priority flow divider with relief valve in priority leg

As fluid enters the valve, the path of least resistance leads through the controlled-flow-volume orifices and out port CF. If pump flow is more than the volume orifices can pass, pressure builds on the right end of the flow-control spool through the excess-flow pilot line. When pressure rises enough to overcome the bias spring and any backpressure from the steering circuit, the flow-control spool moves to the left, just enough to let excess flow exit through port EF. Excess flow changes as pump flow varies, but flow to port CF takes priority. A relief valve in port CF can be set for any pressure and has no affect on pressure at port EF. The controlled-flow relief valve is required even when maximum pressure is the same for both outlets.

Notice that controlled flow is pressure compensated. As pressure builds at port CF, it pushes back against the excess-flow pilot-pressure pilot to maintain a constant pressure drop across the volume orifices.

Priority flow dividers are also manufactured with adjustable flow for the priority port and without a relief valve for circuits that already have one. (The symbol shown is borrowed from a manufacturer's catalog because there is no standard symbol in ANSI or ISO literature.)

The flow divider in Figure 13-16 is a spool-type divider that splits flow at any predetermined rate according to the sizes of the drilled orifices. It is usually set up with identical orifice sizes for a 50-50 split. This particular design does not allow reverse flow, so bypass check valves are required when flow must return the same way it entered.

Fig. 13-16. Spool-type flow divider for 50-50 split

Fluid entering the Inlet port goes left and right through orifices, then out outlets 1 and 2. When either outlet encounters more backpressure than the other does, the high-pressure side forces the spool towards the low-pressure side until pressures on both sides equalize. Equal pressure drop across both orifices produces equal flow. (Most manufacturers specify flow equality at ±5%.) Pressure differences at the two outlets should be low because Inlet pressure always equals the highest outlet pressure -- which means pressure drop across the low-pressure outlet wastes energy.

Spool-type flow dividers only split flow. When more than two outlets are required, dividers must be used in series. A 50-50 split divider flowing into two more 50-50 dividers gives four equal outlets. A 66-33 divider into a 50-50 divider gives three equal outlets. The flow divider/combiner in Figure 13-17 equalizes flow in both directions. It can be used with double-acting actuators to synchronize speed in both directions of travel. The spool in this divider is made in two sections with a connecting link that allows the sections to move together in the closed condition (as shown) for combining, or be spread by Inlet pressure when they are dividing. Springs at both ends of the spool keep the sections together when pressure equalizes or is not present. Inlet orifices set nominal flow, while outlet orifices control flow to or from an actuator.

Fig. 13-17. Spool-type flow divider/combiner with 50-50 split

Flow to the inlet-return port goes through the inlet orifices to split into two equal parts. Pressure drop across the orifices causes the split spool to separate so the outlet orifices are working at the outer edge of the outlet-return ports. When unequal pressures on its ends shift the spool, flow is retarded to the low-pressure outlet port to keep it from receiving too much fluid. When the actuator reverses, flow into the outlet-return ports goes through the outlet orifices and on through the inlet orifices, causing the spool sections to come together. Now the outlet orifices control return flow on the inner edge of the outlet-return ports. They will retard flow from any actuator port that is trying to run ahead.

Motor-type flow dividers

A motor flow divider is constructed from two or more hydraulic motors -- in a common housing -- with a common shaft running through one set of gears on all motor sets. There is a common Inlet to all motors and separate outlets. The motors are usually gear-on-gear or gerotor design. Flow split is commonly 50-50 but many outlet flow combinations are possible by changing gear or gerotor widths.

The cutaway view and symbol in Figure 13-18 pictures a 2-outlet 50-50 split gear-motor-type flow divider. (There is no ISO or ANSI symbol for a motor flow divider so the one shown in the figure is from a supplier’s catalog.) One gear from each motor set is keyed to the common shaft, so both motors must turn at the same rate. If one motor stalls, they both stop because of the common-shaft arrangement. Due to internal clearances in the motor elements, there is some bypass flow that does not turn the motors. As a result, the outlet flows are not always exactly equal . . . especially at high outlet-pressure differences.

Fig. 13-18. Motor-type flow divider with 50-50 split

From Figure 13-18, it should be obvious that this flow divider does not have a priority side like a spool-type flow divider does. Thus, when Inlet flow changes, it is always split equally. The main advantage of motor-type over spool-type flow dividers is there is less wasted energy when the outlets are not at or near the same pressure. If pressure at the right outlet was 1500 psi and pressure at the left outlet was 300 psi, pressure at the inlet would be 900 psi. Pressure at the inlet is always the average of the sum of the outlets.

This feature can be an asset or a problem. If one outlet meets resistance while the other is flowing to tank, an inlet pressure of 2000 psi can result in the pressurized outlet intensifying to 4000 psi. If pressure that high cannot be tolerated, a relief valve must be installed at the outlets. On the other hand, intensification can allow a 1000-psi system to produce 2000 psi to perform work -- similar to a hi-lo pump circuit. Note that while pressure doubles, flow is halved through the high-pressure outlet.

Looking at Figure 13-18, it appears the motor flow divider is also a combiner. This is partially true. The circuit in Figure 13-19 shows a motor flow divider synchronizing two hydraulic motors. As the motors turn in right-hand rotation, they stay almost perfectly synchronized. Pressure to each motor may vary but flow from each flow-divider outlet remains near constant. If the directional control valve shifts to turn the motors in left-hand rotation, the flow divider may get equal flow and the hydraulic motors may stay synchronized. However, if one hydraulic motor meets more resistance than it can overcome and stalls, all pump flow goes to the running hydraulic motor. The second motor then turns twice as fast. During this scenario, one flow-divider motor overspeeds while the opposite one cavitates. The only way to make sure both hydraulic motors stay synchronized in both directions of rotation is to install motor flow dividers at both valve ports.

Fig 13-19. Synchronizing circuit for 50-50 flow divider

Spool and motor flow dividers work reasonably well to synchronize circuits with hydraulic motors and cylinders. However, because both devices do not divide flow perfectly, the actuators they control will not stay perfectly synchronized. A high-pressure difference at the divider's outlets is the worst problem; it can allow a 5 to 10% lag in actuator position. This means that synchronizing circuits using flow dividers often require some type of re-synchronizing valving to realign the actuators more exactly when they stop at home position. (Due to internal bypass, actuators with short cycles may re-synchronize themselves because the error is small.)

Fig. 13-20. Motor-type flow-divider circuit with 50-50 split

Another design consideration is the intensification of pressure at the outlets of a motor flow divider. The circuit in Figure 13-20 has two cylinders that are synchronized by a motor flow divider. Because this circuit operates at 2000 psi, it is possible that pressure at one cylinder could reach as much as 4000 psi due to intensification. Intensification occurs when one cylinder is lightly loaded or has no load and the other one is loaded heavily. In Figure 13-19, the load is shifted to one side of the platen -- making the right-hand cylinder do all the work. Inlet pressure is at 2000 psi and the cylinders are stalled. Pressure at the lightly loaded left-hand cylinder is 250 psi, so pressure at the right-hand cylinder is 3750 psi. The intensification is due to energy transfer through the motors in the flow divider. Because inlet pressure for both motors is 2000 psi, the unused 1750 psi from the left side is transmitted through the common shaft and drives the opposite motor to 3750 psi. (For other flow-divider circuits. see the author’s book, “Fluid Power Circuits Explained,” available through the same outlet for this manual.)

Fig. 13-21. Symbols for modular flow controls and flow dividers

Most flow control functions are available as modular or sandwich valves that mount between directional control valves and a subplate. Figure 13-21 shows most of the common configurations presently offered by fluid power suppliers. Although the symbols show non-compensated flow controls, most configurations also are available with pressure-compensated flow controls. Where a needle valve is shown, a flow control with bypass may actually be installed. This is not a problem because there is never a reason for flow reversal. Figure 13-21 also shows two modular flow dividers that are available from one supplier. These modules are usually available in all valve sizes up to D08 (3/4-in. ports).

Quiz

 

Chapter 14: Pressure Control Valves (Except Relief and Unloading Valves)

Pressure controls (other than relief and unloading valves)

There are some parts of fluid power circuits that need pressure control. (Chapter 9 covered relief and unloading valves that control pressure in pump circuits.) Other types of pressure controls include sequence valves, counterbalance valves, and reducing valves. Though the internal works (and the symbols) are similar, these three pressure controls perform entirely different functions. Sequence valves and counterbalance valves are normally closed -- like relief valves and unloading valves -- but they usually allow bi-directional flow, so they need a bypass check valve in their bodies. Sequence valves always have an external drain connected directly to tank. Counterbalance valves are internally drained, except when used in some regeneration circuits.

Reducing valves are normally open and respond to outlet pressure to keep outlet flow from going above their set pressure. They also can have a bypass check valve. Reducing valves always have an external drain connected directly to tank. Any backpressure in this drain line adds to the valve’s spring setting.

Relief valves, unloading valves, sequence valves, counterbalance valves, and reducing valves are the most difficult to discern on a schematic drawing because their symbols are so similar. Take extra care when diagnosing a problem to make sure these valves are correctly identified and their function understood.

Sequence valves

There are times when two or more actuators, operating in a parallel circuit, must move in sequence. The only positive way to do this is with separate directional control valves and limit switches or limit valves. This setup assures the first actuator has reached a specific location before the next operation commences. If there is no safety concern or possibility of product damage if the first actuator does not complete its cycle before the second starts, a sequence valve can be a simple way to control the actuators’ actions.

The symbols and cutaways in Figure 14-1 are for hydraulic and pneumatic sequence valves. The main difference between these valves is that most hydraulic sequence valves are single purpose and must be used in series with a directional control valve, while many air sequence valves are pilot-operated directional control valves with an adjustable spring return. In either case, a preset pressure must be reached before the valves allow fluid to pass or change flow paths. Many manufacturers offer a direct-acting internally piloted hydraulic sequence valve like the design shown in Figure 14-1. This valve can be changed to external pilot in the field if required.

Figure 14-1. Hydraulic and pneumatic sequence valves

Several manufacturers offer pilot-operated sequence valves also. Pilot-operated sequence valves stay closed to within 50 psi or less of their set pressure. Direct-acting sequence valves may partially open at pressures that are 100 to150 psi below set pressure -- and thus allow premature actuator creep.

A balanced spool -- held in place by an adjustable-force spring -- blocks fluid at the hydraulic sequence valve’s inlet. When pressure at the inlet reaches the spring setting, pressure in the internal pilot line pushes the spool up to allow enough flow to the outlet to keep pressure from going higher. Pressure at the inlet never drops below set pressure when there is flow to the outlet. When outlet pressure exceeds set pressure, the valve opens fully and pressure at both ports equalizes. Notice that the drain port hooked to tank must be at no pressure or constant pressure because any pressure in this line adds to spring setting. (Remember that a sequence valve must always have an external drain.)

A bypass check valve allows reverse flow when the valve is used in a line with bi-directional flow. In some applications a sequence valve may be externally piloted from another operation. Most valves can be converted in the field. (The designer should always change the part number to reflect the conversion.)

Pneumatic sequence valves typically are 5-way directional control valves with adjustable springs to set their shifting pressure. They are used to start a second operation after the preceding one finishes. Some older machines have one solenoid valve to start the cycle and several sequence valves to extend and retract all other actuators. Some precautions: • A sequence valve shifts on a pressure build-up and may start a second operation prematurely if an actuator stalls or is stopped for any reason. If personnel safety or product damage can occur due to an incomplete stroke, don’t use sequence valves. Instead, use limit switches or limit valves and directional control valves for each operation sequence. • When flow controls are required they must be meter-in types. Take the signal to the sequence valve from the line downstream from the flow control because pressure at this point will be whatever is required to move the actuator and its load.

The circuit in Figure 14-2 is typical for air-powered machines. Cyl. 1 extends to clamp a part when an electrical input signal shifts the solenoid pilot-operated valve. As Cyl. 1 extends, pressure beyond the meter-in flow control at its cap end becomes as high as necessary to move the cylinder and its load. With the sequence valve set to shift at 70 psi, Cyl. 2 should not move until Cyl. 1 has extended and securely clamped the part. If the clamp does not make a full stroke for any reason, the Cyl. 2 extending prematurely will not damage the part or be unsafe. When the clamp is at 70 psi or higher, the sequence valve shifts to extend Cyl. 2. Both cylinders can return simultaneously without causing any problems.

Figure 14-2. Typical pneumatic sequence valve circuit

One great feature of a sequence-operated circuit is it does not matter how far the first cylinder must move before the next operation takes place. Thick or thin parts are clamped at the same force before the next operation starts because pressure must build to the same level to trigger the next sequence.

Cyl. 2 has meter-out flow controls to retard its movement and hold pressure on Cyl. 1 during the stamping operation. De-energizing the solenoid pilot-operated valve allows both cylinders to return home at the same time.

The hydraulic sequence circuit in Figure 14-3 is typical for a machine that must clamp and hold pressure while a second operation takes place. Sequence valve 1 is set at 550 psi; pressure at clamp Cyl. 1 must be at least 550 psi before punch Cyl. 2 can extend. While punch Cyl. 2 is extending, pressure in the circuit never drops below 550 psi. If the punching operation requires more than 550 psi, the pressure in the whole circuit increases -- up to the relief valve setting.

Figure 14-3. Typical hydraulic sequence valve circuit

Sequence valve 2 (set at 450 psi) keeps Cyl. 1 from getting a retract signal until Cyl. 2 has returned and pressure increases. A pilot-operated check valve maintains clamp force while the punch cylinder retracts. The signal to open the pilot-operated check valve comes from the line between Sequence Valve 2 and Cyl. 1, so there is no signal until Cyl. 2 fully retracts. (This circuit is not safe if pressure buildup comes from some source other than clamp contact or the end of stroke so that the punch cylinder operates prematurely.)

Sequence valves often generate a great deal of heat because the first actuator to move takes higher pressure than the subsequent actuators. This means there is usually a high pressure drop across a sequence valve that results in wasted energy. In some circuits, a kick-down sequence valve can reduce the energy loss. The cutaway view and symbol in Figure 14-4 show the inner workings of a kick-down sequence valve to explain how it controls opening pressure and then unloads it.

Figure 14-4. Kick-down sequence valve

Fluid from the inlet flows through the control orifice and up to the adjustable poppet where it is blocked. The resulting pressure tries to open the poppet while equal pressure and a light spring acting on the opposite side hold it shut. When pressure increases enough to unseat the adjustable poppet and more flow starts passing the poppet than going through the control orifice, the pressure imbalance lets the poppet raise. When the poppet moves enough to let trapped fluid go through the bypass orifice, pressure on top of the poppet drops off -- because the bypass orifice is larger than the control orifice. At this point, the only force acting to hold the poppet shut is spring force and backpressure at the outlet port. When flow stops, the poppet closes again due to pressure equalization and spring force on the poppet.

The circuit in Figure 14-5 is the same as in 14-3 except it incorporates kick-down sequence valves in place of standard sequence valves. Cyl. 2 will not extend in this circuit until pressure on Cyl. 1 has reached 750 psi. The difference is when a kick-down sequence valve opens at its pressure setting, it allows fluid to pass at 50 psi plus whatever it takes to overcome downstream resistance. This means the whole circuit from the pump to all actuators is 50 psi plus Cyl. 2’s resistance. The pilot-operated check valve at Cyl. 1’s cap-end port keeps it pressurized at near full force, while Cyl. 2 extends at low force. Energy waste is very low so heat buildup is minimal. (Other sequence valve circuits can be found in the e-book Fluid Power Circuits Explained by the author of this manual, which will be launched in the next few months.)

Figure 14-5. Hydraulic circuit with kick-down sequence valves

Counterbalance valves

The fourth and last normally closed pressure control valve found in hydraulic circuits is the counterbalance valve. Cylinders with external forces -- such as weight from a platen, machine members, or tooling -- acting against them will overrun when cycled if oil flowing out of them is not restricted. A meter-out flow control circuit is one way to control overrunning loads but it has one main drawback. A flow control’s speed is fixed except for manual adjustment or when using an infinitely variable proportional type. Because flow is fixed, the actuator will continue at the same speed – even when working flow to it increases or decreases. Thus, control is minimal and there could be high energy waste. (Figure 13-8 shows a meter-out flow control circuit for running away loads.)

A counterbalance valve keeps an actuator from running away regardless of flow changes because it responds to pressure signals, not flow. A counterbalance valve is almost the same as a sequence valve except it normally does not have an external drain connection. The cutaways and symbols in Figure 14-6 depict the physical makeup of three different counterbalance valves and how they are represented on a schematic drawing.

Figure 14-6. Three types of counterbalance valves

The two cutaways and symbols on the left are spool designs with internal and external pilots. The valve on the right is a poppet design that is both Internally and externally piloted. Each valve type has advantages in different circuit arrangements that will be discussed later. A counterbalance valve usually has a bypass check valve for reverse flow because its most common use is in controlling actuators with running away or overrunning loads.

An internal pilot-operated counterbalance valve shifts to allow excess fluid to flow to the outlet when pressure at the inlet increases to the pressure set by the pressure adjustment. Pressure at the inlet never drops below set pressure when there is flow at the outlet. Flow from the inlet to the outlet is just enough so that backpressure on the actuator never drops below set pressure. This means the actuator moves only as fast as it is supplied and stops when Inlet flow ceases.

Pressure adjustment on the Internal-piloted counterbalance valve is usually made by first screwing the pressure adjustment all the way in. To assure that the valve is capable of high enough pressure, start the pump and raise the load a small amount. Then center the directional valve -- which connects the cylinder rod-end port to tank -- to see if it holds. If the load holds, next raise the load in increments -- checking for load stop every few inches. With the load suspended, start reducing set pressure on the counterbalance valve slowly until the load creeps forward. When the load starts drifting down slowly, increase pressure until movement stops, then turn the pressure adjustment another quarter to half turn higher. This method of adjusting usually wastes less energy while it always stops and holds the load.

The main disadvantage of an internal pilot-operated counterbalance valve is that backpressure is constant and it holds back even when the actuator needs maximum force. Another disadvantage is that to maintain optimum performance, an Internal-piloted counterbalance valve must be readjusted every time the load changes. The valve’s main advantage is that it produces smooth cylinder action while advancing to the work.

An external pilot-operated counterbalance valve shifts to allow excess fluid flow to the outlet when pressure at the opposite cylinder port reaches the pressure set by the pressure adjustment. Pressure at the inlet never drops below load-induced pressure plus pressure set on the pressure adjustment when there is flow at the outlet. Flow from inlet to outlet is just enough that the actuator moves only as fast as it is supplied and stops when flow to the actuator ceases.

Pressure adjustment on the external pilot-operated counterbalance valve can be made on a test stand by setting the pressure adjustment at 100 to 200 psi. If pressure must be set on the machine, set the pressure adjustment higher than 200 psi and lift the load a small distance to make sure it stops and holds. If it holds, continue to raise the load high enough to have some time for the next step. Now, power the load down and observe pump pressure. Pump pressure while lowering the load should not exceed 200 psi. Continue this action until pump pressure is between 100 and 200 psi while the load is lowering. This method of adjusting usually wastes less energy while always stopping and holding the load.

The main disadvantage to an external pilot-operated counterbalance valve is that it may cause lunging or even stop cylinder action while advancing to the work. The main advantage is that backpressure is only present when the actuator is advancing to the work. At work contact, pressure at the actuator inlet increases and forces the counterbalance valve wide open, thus eliminating all backpressure. Another advantage is that an external pilot-operated counterbalance valve does not need to be readjusted when the load changes.

Internal and external pilot-operated counterbalance valves shift when pressure at the internal pilot area reaches the pressure set on the pressure adjustment and allows excess flow to go to the outlet. Pressure at the Inlet never drops below set pressure when there is flow at the outlet. Flow from the inlet to the outlet is just enough that backpressure on the actuator never drops below set pressure. This means the actuator moves only as fast as it is supplied and stops when Inlet flow ceases.

Pressure adjustment on an internal and external pilot-operated counterbalance valve is usually made by first screwing the pressure adjustment all the way in. To assure that the valve is capable of high enough pressure, start the pump and raise the load a small amount. Then center the directional valve that has the cylinder rod-end port connected to tank -- to see if it holds. If the load holds, then raise the load in increments -- checking for load stop every few inches. With the load suspended, start reducing set pressure slowly until the load creeps forward. When the load starts drifting down slowly, increase pressure until movement stops, then turn the pressure adjustment another quarter to half turn higher. This method of adjusting usually wastes less energy while always stopping and holding the load.

An internal and external pilot-operated counterbalance valve lowers loads smoothly and opens fully when pressure at the actuator inlet increases upon contact with the work. The valve does need to be readjusted when loads change, but this is a small price to pay for good control.

Figure 14-7 depicts a vertically oriented cylinder with rod facing down and a load trying to extend it. To keep the cylinder from running away, the counterbalance valve must resist the load-induced pressure from the weight. The load-induced pressure can be calculated and the counterbalance valve could be preset at 100 to 150 psi higher on a test stand, but pressure adjustment is usually done at the machine (as mentioned earlier).

Figure 14-7. Internally pilot-operated counterbalance valve circuit

Notice that the directional control valve has ports A and B connected to tank in the center condition. There is no chance of extra pressure buildup in the pilot line while the circuit is at rest. If ports A or B were blocked, pressure could build and pilot the counterbalance valve open, allowing the cylinder to drift.

Energizing solenoid A1 sends pump flow to the cylinder cap end. As pressure builds there, pressure also increases in the rod end. When pressure at the cylinder rod end reaches 100 to 150 psi above the load-induced pressure, the cylinder starts to extend as fast as the pump fills the cap end. When flow increases, cylinder speed increases and when flow decreases, cylinder speed decreases.

As stated in the counterbalance valve explanation, backpressure at the cylinder rod end is present during the entire extend stroke. As a result, at work contact cylinder force is reduced by counterbalance pressure times the cylinder’s rod-end area. The total weight of the platen and tooling on a press plus the amount of added pressure at the counterbalance valve cannot be used to do work. Energy is expended to raise the weight but it is not recouped during the work cycle. Energizing solenoid B1 sends fluid around the counterbalance valve through the bypass check valve and on to the cylinder rod end to retract it.

The circuit in Figure 14-8 shows the same cylinder with an external pilot-operated counterbalance valve. An externally piloted valve can be set at approximately 100 to 200 psi regardless of load-induced pressure in the cylinder. This is especially convenient in applications where loads constantly change. It is also the best use of energy because the counterbalance valve opens fully when the cylinder meets resistance so the weight is able to do some work. Because backpressure on the cylinder rod end is zero, more force is available.

Figure 14-8. Externally pilot-operated counterbalance valve circuit

Energizing solenoid A1 sends fluid to the cylinder’s cap end to start it extending. As pressure builds in the cylinder cap end, it pressurizes the external pilot and opens the counterbalance valve The valve only opens enough to let fluid out when the cap end is at pilot pressure. If pilot pressure is set too low, the counterbalance valve may quickly open too far -- allowing the cylinder to run away and pilot pressure to drop. At this point, the counterbalance valve shuts abruptly and the cylinder stops. Almost immediately, pressure again builds at the cylinder cap end, the counterbalance valve reopens, and the same scenario repeats until the cylinder meets resistance. A meter-in flow control in the external pilot line can help, but is very difficult to set. Energizing solenoid B1 sends fluid around the counterbalance valve through the bypass check valve and on to the cylinder rod end to retract it.

The internal and external pilot-operated counterbalance valve in Figure 14-9 incorporates the best features of both valves. The internal pilot provides a smooth advance stroke at low force, while the external pilot opens the valve fully to eliminate backpressure from the cylinder rod end when it contacts the workpiece. (Like the internally piloted valve. this version must be reset at each load change to maintain its efficiency and keep energy losses low.)

Figure 14-9. Internally and externally pilot-operated counterbalance valve circuit

The symbols in these example circuits show a direct-acting pressure control valve. Several suppliers offer a pilot-operated version that is more stable and has less pressure differential between cracking and full flow operation.

The circuits shown here work equally well with hydraulic motors, except that a counterbalance valve will not stop and hold a running away load on a motor without creep. All hydraulic motors have internal leakage that increases as the motor wears. The counterbalance valve may not have any bypass but fluid will slip by the motor parts no matter what its design.

There are no counterbalance valves for air circuits. Air circuits depend on meter-out flow controls to keep an actuator from running away. Usually an air circuit uses a 2-position valve that keeps pressure on the retract side at rest so it stays in place at end of stroke. When a load must be stopped in mid-stroke, a 3-position valve with cylinder ports blocked in center is the common method of trying to do this. There also is available a pilot-operated check valve for air service that gives some control for stopping and holding a pneumatic cylinder in mid-stroke.

Air line regulators

Most plant air systems produce pressures between 90 and 125 psi, while most air circuits are designed to operate at 75 to 85 psi. Other systems may operate at pressures as low as 15 to 20 psi. To accommodate these ranges, some method is needed to reduce the system pressure without wasting energy. A relief valve that would release plant air to atmosphere and try to lower the whole system it is not a good solution. The air line regulator shown in Figure 14-10 reduces outlet pressure by shutting off flow when downstream pressure tries to go above the regulator’s setting. There is very little energy loss because air merely expands from its elevated pressure to meet the lower pressure requirement. In other words, an air compressor operating at 120 psi only has to run about a third as often when regulated or reduced to 40 psi.

Figure 14-10. Air line regulators (or reducing valves)

This points out the main reason why an air line regulator should be set just high enough to do the job at hand. Without a regulator, not only does it cost more to operate a machine, but the machine tries to run a repetitive cycle with fluctuating pressure, therefore different forces and speeds.

The cutaway views and symbols in Figure 14-10 show two common direct-acting air line regulators. They are normally available in sizes from 1/8-in. through 2-in. pipe thread. (Larger sizes are built but they usually are pilot-operated from a small direct-acting regulator.) Air flows freely from inlet to outlet until the outlet pressure reaches the set pressure. The adjustable spring holds the shut-off valve off its seat by extending the diaphragm during free flow. As pressure at the outlet continues to build, it passes through the pilot passage to the underside of the diaphragm. At set pressure, the diaphragm pushes the adjustable spring back, allowing the shut-off valve to seat. The light spring pushes the shut-off valve closed. Pressure at the outlet now is stable at its reduced setting -- as long as the inlet pressure is equal to or higher than the outlet. Any pressure drop at the outlet reduces pressure under the diaphragm and the adjustable spring again pushes the shut-off valve open to let more air in.

If there is a possibility of the reduced pressure line seeing excess pressure for any reason, use the relieving-type regulator shown on the right in Figure 14-10. This valve closes a hole through the diaphragm’s center section with the shut-off valve’s stem. After reaching set pressure, the shut-off valve cannot move up. Any extra pressure buildup under the diaphragm raises its center section off the shut-off valve’s stem and allows air to flow to atmosphere through the vent hole. This feature should not be used as a relief valve function where pressure increases during every cycle -- it is only for occasional overpressure situations.

Every pneumatically powered machine should have a regulator set for the lowest pressure that will produce good products. Costly overpressure should be eliminated in every case. Use an air line regulator anytime a job can be done at a pressure lower than plant air supply.

Another application for air line regulators that can save compressor output is reducing pressure on the return stroke of actuators that can use low power to retract. Many cylinders need high force to extend and do work, but the retract portion of the cycle needs very low force. An air line regulator positioned as shown in Figure 14-11 can save air during part of every cycle on many cylinder operations in most circuits.

Figure 14-11. Pneumatic circuit with air-saving regulator

A 5-way spool valve, piped with a dual-pressure inlet as shown, can give normal cycle time while conserving plant compressed air. Return pressure is set on the regulator supplying the cylinder rod end at the lowest possible pressure that maintains cycle integrity. A reduction as small as 20 psi below working pressure can pay for the regulator in a short time. Shifting the 5-way valve starts the cylinder extending. (There will be a brief lunge as the lower-pressure air in the rod end compresses to hold back against the higher pressure in the cap end.) To control cycle time, adjust cylinder speed with the rod-end meter-out flow control. When the 5-way valve shifts again to return the cylinder, the meter-out flow control on the cap end must be adjusted for a faster rate because return power is limited.

Pressure-reducing and reducing//relieving valves

There are times in multi-actuator hydraulic circuits when system pressure is too high for some actuators while others need maximum force. One suggested remedy is the circuit in Figure 14-12. Cylinder 1 needs 2000 psi to maintain force, while Cylinder 2 can damage the product when pressure exceeds 800 psI. Adding Relief Valve 2 (set at 800 psi) takes care of Cylinder 2’s overpressure, but limits the entire circuit to 800 psi. Pressure in a circuit with more than one relief valve will never be higher than the setting of the lowest valve. The correct way to have two or more pressures in a single circuit is to incorporate reducing valves. (Figure 14-14 diagrams a circuit using a reducing valve to give two pressures.)

Figure 14-12. Dual-pressure hydraulic circuit with two relief valves

The cutaway and symbol in Figure 14-13 depicts a pilot-operated reducing valve that allows flow from the inlet to the reduced-pressure outlet until pressure reaches the setting on the direct-acting relief valve in the pilot section. Unlike the other four pressure controls (relief, unloading, sequence, and counterbalance valves), a reducing valve is normally open and blocks flow at set pressure.

Figure 14-13. Pilot-operated reducing valve

The normally closed direct-acting relief valve in the pilot section traps fluid from the reduced-pressure outlet port through the control orifice on top of the spool when pressure is below its setting. The spool stays in the normally open position because pressure on both ends balances it hydraulically while the light spring keeps it pushed down. As pressure at the reduced-pressure outlet port continues to increase, it finally starts to open the direct-acting relief valve in the pilot section. Some fluid then flows to tank through the drain port. When flow through the direct-acting relief valve is more that the control orifice can handle, pressure on top of the spool drops and pressure on the bottom of the spool pushes it closed. The spool never closes completely because there is flow through the drain port anytime pressure at the outlet is lower than at the inlet. Drain port flow amounts to about 60 to 90 cim. This flow is all wasted energy and it can cause a system to overheat if more reducing valves are installed than necessary. When pressure drops below the direct-acting relief valve’s setting in the pilot section, the valve closes and forces the spool to the open position.

A reducing valve is normally open so it appears reverse flow should not be a problem. However, when the valve is working, it is almost closed -- and it can be held closed by back flow when the actuator starts to return. Anytime a reducing valve must pass reverse flow, select a valve with an integral bypass check valve to eliminate the possibility of blocked return flow.

It also is very important to have a free-flow drain port with very low (or even no) backpressure. Backpressure in the drain port adds to the setting of the direct-acting relief valve and can cause erratic results when drain pressure fluctuates. (Our next e-book, Fluid Power Circuits Explained, discusses how this drain port can be used advantageously in a dual-pressure circuit. This book will be launched in the next few months.)

The modified circuit in Fig, 14-14 allows two pressures without lowering system pressure (as happened in Figure 14-13). A pressure-reducing valve in place of Relief Valve 2 makes it possible to set pressure for Cylinder 2 without affecting pressure at Cylinder 1. This reducing valve never has reverse flow so a bypass check valve is not required.

Figure 14-14. Dual-pressure hydraulic circuit using a relief valve

When it is working, a reducing valve is nearly closed and will pass very little reverse drain flow unless it has a bypass check valve. Even then, reverse flow must be at a pressure greater than that at the inlet. If this much pressure cannot be tolerated, use the reducing/relieving valve depicted in Figure 14-15.

Figure 14-15. Pilot-operated reducing/relieving valves

Reducing/relieving valves function exactly like reducing valves -- until an external force starts to increase pressure at the reduced-pressure outlet above the pressure set by the pilot section. When outlet pressure is 4 to 6% above set pressure, the spool moves up until the outlet is connected to tank. Any fluid at pressure above set pressure returns to tank, so outlet pressure does not continue to climb. Tank flow comes only from the reduced-pressure outlet, not from the pump through the inlet. When excess pressure at the outlet drops, the reducing/relieving valve continues to perform its reducing function.

Note that the left cutaway view has an internal drain for the pilot section. This saves connecting a separate drain line for pilot flow. However, when backpressure in the tank line is high or may fluctuate due to other return functions, it adds to the pilot-section setting and can elevate pressure at the reduced-pressure outlet above allowable rates. When tank-line backpressure may be high or when pressure fluctuations cannot be tolerated, use a valve with an external drain. When reverse flow is necessary, specify a model with an integral bypass check valve for piping convenience.

Figure 14-16. Modular sequence, counterbalance, and reducing/relieving valves

Figure 14-16 shows most of the modular valve configurations for sequence, counterbalance, and reducing valves. Modular valves simplify piping and eliminate many connections that can generate backpressure or add potential leakage points.

Quiz

 

Chapter 15: Fluid Power Actuators

Cylinders, fluid power motors, and rotary actuators

Fluid power actuators

Fluid power actuators receive fluid from a pump (typically driven by an electric motor). After the fluid has been pressure, flow, and directionally controlled, the actuator converts its energy into rotary or linear motion to do useful work. Cylinders account for more than 90% of the actuators used in fluid power systems for work output. Of the approximately 10% of actuators that produce rotary output, more than 90% are hydraulic motors, while the rest are some form of rotary actuator.

Single-acting ram cylinders

The symbols and cutaway views in Figure 15-1 show single-acting ram cylinders in push and pull types. Rams can be as small and simple as a service station lift operated by air over oil, or as big and complex as a 100,000-ton extrusion press.

Figure 15-1. Single-acting ram cylinders for push and pull applications

Single-acting rams often are mounted vertically up and are weight returned. When a ram cylinder is mounted vertically down or horizontally, it must have some method of retracting it to the home position. Figure 15-1 shows one method. Small single-acting pull rams -- mounted alongside the large working ram -- raise and hold it in the up position with a counterbalance valve (not shown). A directional valve or a bi-directional pump directs fluid to the push or pull rams to make them cycle. Another retraction method uses single-acting push rams that oppose the platen movement from the opposite side. (For a circuit that uses a large-diameter vertical down acting ram cylinder, see Figure 10-9.) Small ram cylinders may be returned manually or via a spring.

Ram cylinders only have seals where the ram passes through the body. Anytime a ram cylinder drifts from its stopped position, the cause is valve or pipe leakage if no fluid is coming out around the ram seal.

As the ram moves, stops and guide protrusions on it keep it aligned and indicate maximum stroke. Usually on large-area rams, the stops tear off the packing gland and bushing retainers if the ram is not stopped some other way. Most machines using rams have other methods to keep them from overstroking. (Some only have warning placards about problems if the ram is powered beyond certain limits.) The guide protrusions and bushing align the ram in its housing so it runs true.

Figure 15-2 shows another type of ram cylinder. When there is a need for a long-stroke actuator with a short retracted length, one option is a telescoping cylinder. Although the majority of telescoping cylinders are single acting, double-acting models are available. Most telescoping cylinders stroke slowly and cycle infrequently because their construction is not robust enough for high production applications.

Figure 15-2. Single- and double-acting telescoping cylinders

Telescoping ram cylinders

The cutaway views and symbols in Figure 15-2 depict typical multi-stage telescoping cylinders. The one on the left is single acting; the one on the right is double acting.

Single-acting telescoping cylinders are usually mounted vertically with the small ram up. The cylinder then can be weight returned. This arrangement leaves the large ram with its ports attached to a stationary machine member.

Double-acting telescoping cylinders can be mounted vertically with the small ram down or horizontally when required. The best mounting position for any double-acting telescoping cylinder is with the small ram attached to a stationery machine member so the ports do not move. Long-stroke double-acting telescoping mounted horizontally need some sort of carrier to support the center section during extension so they will not sag and wear out seals and bushings prematurely. Also note: the return area may only be only 10% of the extend area, so the return force is not capable of doing much work. This small area also requires very little fluid to give maximum retraction speed without excessive backpressure at the extend port. Another possible problem with double-acting telescoping cylinders occurs if the retract port is blocked while the cylinder is trying to extend. Up to a 10:1 intensification can result and the high pressure may damage the housing or rams. Installing a safety relief valve at the retract port may be necessary if this port can be blocked or restricted for any reason.

For all telescoping cylinders, make sure the small ram can do the work required. As a telescoping cylinder starts to extend, the large ram always moves first at a lower pressure. When the first and subsequent rams bottom out, pressure and speed increase due to the decreased ram area. If the small ram produces insufficient force, the unit stops before making a full stroke.

Several suppliers build double-acting pneumatic telescoping cylinders in small sizes, with up to three stages.

One manufacturer makes a single-acting telescoping cylinder with internal porting and matching areas that cause all rams to move in unison as they extend and retract. These cylinders come in a maximum of three stages because the area staging would make any more rams into a vastly oversized package. An integral combination check and relief valve allows the rams to be filled and bled and to stroke fully in case of bypass at the seals. This design’s main advantage is smooth extension and retraction without the bumps of a typical telescoping unit.

Piston-and-rod cylinders

The cutaway view and symbol in Figure 15-3 is for a typical industrial-grade tie-rod cylinder. This cylinder includes all the standard features available from most manufacturers. The names of the parts are what most fluid power glossaries propose, while the names in brackets may be in common use.

Figure 15-3. Typical industrial-grade single-rod end tie-rod cylinder

The cap end and head end seal off the tube ends with tube-end seals. Tie rods hold the assembly together. The tie rods are tightened to a torque that will resist as much as five times the cylinder’s rated pressure. Tie-rod construction gives the package some flexibility or stretch without permanent deflection or damage. The piston provides the area for fluid to work against. The piston seals stop bypass that would waste energy. The piston rod transmits the force on the piston to the outside of the envelope and is attached to the work mechanism. The rod bushing and rod seal keep the rod aligned and stop fluid leaks to atmosphere. Cap end and rod end cushion plungers block high fluid flow near the end of stroke to allow smooth, no-shock stopping. Cushion-adjusting screws make it possible to adjust stopping speed, while cushion-bypass checks let the piston move rapidly as the cushion plungers are leaving their chambers.

The symbol on the left is the detailed symbol for a hydraulic cylinder with adjustable cushions on both ends. This cylinder also could be as shown as: non-cushioned, cushioned rod end only, or cushioned cap-end only. (When the energy triangles at the ports are blackened, the cylinder is pneumatic.)

Figure 15-4. Single-acting spring-return and extend cylinders

The simplified symbol shows less detail but represents the same unit. The 2:1 information over a single rod end cylinder indicates that the rod area is half that of the piston. (Cylinders with 2:1 area ratio will be discussed later in this chapter.)

The spring return and extend cylinders in Figure 15-4 illustrate another method of moving cylinder pistons and rods for some applications. The cutaway views show typical construction (using a tie-rod cylinder as the basic unit). Many other designs are available but essentially use similar parts. Notice that the pistons have mechanical stops to keep the spring from compressing enough to bottom out. Breather ports for air operation or connections for tank drains for hydraulic cylinders are commonly found at the spring end. Most manufacturers indicate that the spring is only capable of returning the piston and rod. It may not be capable of returning the external load. Springs can be less than reliable and difficult to monitor -- especially when they are internal. Because there usually is little savings in hookup or operation, use these cylinders with care.

Tandem cylinders

The tandem cylinder in Figure 15-5 can produce almost twice the force from the same diameter, but it is a little over twice the length. The two cylinders can be independently piped or drained to give extra force in one direction only or both directions. The center heads have guide bushings and seals for both sections so a different fluid can also be used in either end. (See Chapter 17 to learn how tandem cylinders allow oil to control speed and air as the power source. Circuits for matched and unmatched tandem cylinders can be found in the author’s upcoming e-book Fluid Power Circuits Explained.)

Figure 15-5. Tandem cylinders with attached rod

The tandem cylinder in Figure 15-5 has a common rod for both pistons. The tandem cylinder in Figure 15-6 has two separate pistons and rods and two different stroke lengths. This combination can be used to get three positive stops from an air or hydraulic cylinder with no special valves or controls. The stops are mechanically fixed, so the stop positions are in the same place every time. However, the stop positions only work for one situation. A four- or five-way directional valve at each cylinder plus flow controls are all that is normally required to operate this circuit.

Special consideration must be used in circuit design for the unattached tandem cylinder in Figure 15-6. If the long-stroke cylinder is not restrained while the short-stroke cylinder extends, it can overtravel and miss the exact position. This problem is exaggerated with horizontal or vertical rod-down applications. Meter-out flow controls or counterbalance valves can eliminate the problem, but could increase cycle time in some cases.

Figure 15-6. Tandem cylinders with attached rod for three positive stop positions (left) and cap-to-cap cylinders for four positive positions (right)

The cap-to-cap mounted cylinders in Figure 15-6 depict another way to use pneumatic or hydraulic cylinders to obtain positive positioning without special valves or equipment. Two four- or five-way valves and flow controls usually make this circuit operate smoothly.

Some designers specify double-rod end cylinders such as those shown in Figure 15-7. These cylinders cost about twice as much as single-rod cylinders and the design has a second place for fluid to leak. In most cases the reason for using them can be accomplished by other methods with equal or better results. If you must use a double-rod end cylinder, remember to allow space for the extra rod and the safety hazard it can cause. Also, the rod reduces the area on the working side of the piston, so a larger bore or higher pressure is necessary in many cases.

Figure 15-7. Double-rod end cylinders

A double-rod end cylinder might be specified so that the force and speed in both directions is the same when flow and pressure are equal. This may be true, but flow controls and a reducing valve can accomplish the same result at a reduced cost and in less space. Another alternative is a regeneration circuit, used when producing the exact speed and force in both directions is not critical. (Regeneration circuits are covered extensively in the author’s upcoming e-book Fluid Power Circuits Explained.)

It may appear that double-rod end cylinders reduce rod flexure when the cylinder is fully extended. The rods in their bushings and the piston in its bore provide snug bearing points -- but allow some play. As the piston nears the end of stroke, two of the bearing points get closer together, so lateral movement at the extended end of the rod can increase. It is supposed that the opposite rod will reduce lateral movement and hold the attached load closer to a centered position. However, from the cutaway it is obvious the distance between the piston bearing and the opposite rod bushing almost eliminates any centering effect of the piston. A better way to reduce lateral movement of the extended rod is to stop the piston short of full stroke – either by an internal stop tube or externally by machine members. This arrangement requires a longer cylinder but gives the desired results at a lower cost.

A main reason for using double-rod end cylinders is to mount limit switches to show cylinder position. A special bracket opposite the attachment end holds the limit switches and a doughnut-shaped protrusion on the rod contacts them as the piston strokes. For the same price (and consuming a lot less space), most cylinder manufacturers offer limit switches that attach to the head and/or cap and are activated by cushion plungers. Another signal indicator -- especially for pneumatics – is a Hall-effect switch and a magnetic piston to activate it.

All of the above cylinder-position indicators have one potential major flaw. If the part attached to the rod end gets disconnected for any reason, the machine still will cycle when the cylinder moves...even though the disconnected load may be in the way. If at all possible, mount limit switches on the machine member so its position is never misinterpreted.

Non-rotating rod cylinders

The cylinders in Figure 15-8 incorporate some method to keep the piston and rod from rotating as it strokes. A standard cylinder may try to turn as it extends and retracts, causing it to unscrew from its workpiece. In some applications, the cylinder is expected to orient the work piece it is driving and keep it aligned with mating parts. All the designs in Figure 15-8 attempt to accomplish this non-rotating function in different ways. At best they can keep the rod from turning but none can perfectly guide it when outside forces are acting to turn it. This is especially noticeable on long-stroke cylinders. It is always best to guide the workpiece externally and only use the cylinder to cycle it. Note that the oval-piston design offers the ability to mount cylinders side by side with minimum rod center distance between them and still produce ample force.

Figure 15-8. Non-rotating rod cylinders

Rodless cylinders

The cylinders in Figure 15-9 take up less space on long-stroke applications because they only need mounting space slightly longer than their stroke length. Conventional piston-and-rod cylinders require space more than twice their stroke length -- and can be difficult to conveniently place on many machines.

The earliest long-stroke design is the cable cylinder – shown at top left in Figure 15-9. A coated cable fitted with a work piece attachment, wrapped around two pulleys, and attached to a piston in a bore produces reciprocating motion as fluid -- usually air -- enters and exhausts through the ports. These cylinders are usually 4-in. bore or less, and may have strokes up to 30 ft (or more in certain configurations). Cushions may be specified when required. The cable is coated with nylon or Teflon so it can slide through seals with minimal damage to them. However, the coatings are prone to cracking and eventually will cut the seals until they leak. (The symbol for the cable cylinder is adapted from a manufacturer’s catalog because ISO does not show one.)

Figure 15-9. Cylinders for long-stroke applications

The rodless cylinder, top right in the figure, was introduced in the late 1970s. It is even more compact than a cable cylinder and avoids the coating wear problem. It consists of a piston in a bore that has a slot open to atmosphere along its whole length. A seal blocks air from escaping through the slot while the piston is not present. A second contamination seal keeps debris from filling the slot. The fluid seal and contamination seal pass through slots in the piston in slots as it reciprocates. The work piece attachment (connected to the piston) reciprocates to move machine members as fluid enters and exhausts the cylinder. Bores up to 2 12 in. and strokes as long as 33 ft are available from several manufacturers. Cushions may be specified when required.

The band cylinder is an alternative to the cable cylinder. Its smooth steel band passes through seals instead of a coated cable. The magnetic-drive cylinder uses magnetic attraction to keep the piston and workpiece attachment connected. It operates at pressures up to 120 psi and will maintain connection up to 180 psi.

One manufacturer has a modified rodless cylinder with a toothed belt and pulley arrangement to drive the workpiece attachment. It offers the option of an external output shaft to which a brake can be fitted to stop and hold position. This output shaft can also drive an encoder to show work piece position or can connect to another unit for synchronization. It also could act as a low power rotary actuator. It is available in 1- or 1 12-in. bores and up to 177-in. stroke.

Other types of linear actuators

The actuators in Figure 15-10 depict other ways of producing linear force. The rolling diaphragm, diaphragm, and bellows actuators are single acting. The rolling diaphragm is capable of long strokes but not long life at high cycles. The diaphragm is designed only for short-stroke applications only but can have high force due to large areas. All these single-acting devices use some internal or external method to retract them. Using vacuum and weight are two of the methods used for retracting bellows. The double-acting rolling diaphragm operates like a double-acting piston cylinder but is not designed for long life at high cycles.

Figure 15-10. Double-acting actuators

Sizing linear actuators

Because most cylinders are round, the formula for figuring their area that most remember is: π r2. Another formula for circle area is .7854 D2 which does not require changing to radius before figuring area. There are also data books that have charts for all standard size cylinders. To determine maximum cylinder force, multiply the known area by the operating pressure in psi. The standard formula is written F=PA where F is force, P is pressure and A is area.

For pneumatic cylinders, the usual working pressure is 80 psi even though most plants cycle their air compressor between 115 to 125 psi. The reason for the lower working pressure is because line losses, especially during high flow surges, make it impossible to maintain compressor output pressure over the entire air supply system.

Operating pressure for hydraulic circuits often falls between 1500 and 2500 psi because most components are rated at 3000 psi. Lower pressures require larger actuators at higher flows and over-sizes the entire system. Pressures over 3000 psi often require special components and are usually relegated to very high force applications.

A factor that is often overlooked when sizing cylinders, especially pneumatic ones, is the force figured by the aforementioned formula is not available until after reaching full pressure. A pneumatic circuit does not reach full pressure quickly like a hydraulic circuit does. For example, when tank pressure is high and tire pressure is low it is easy to rapidly fill a flat tire from a portable air tank. However, as tire pressure starts approaching tank pressure, pressure drop decreases and the transfer rate slows to a crawl as rising tire pressure gets closer to lowering tank pressure. Pressure equalization can take several minutes, adding to cycle time when the tire is an actuator. The 3 1/4-in. bore air cylinder in Figure 15-11 will not move the load attached to it because force required and force available do not allow enough pressure drop to move fluid from inlet to the piston. The 4-in. bore gives nominal speed while the 5-in. bore would give fast speed.

Figure 15-11. Air cylinder sizing

To get nominal cycle times, always size air cylinders at least 25% above workforce whether moving a load or applying force. When cycle time must be fast, oversizing up to 100% could be required. Anything above 100% will quickly lessen the effect and is not worth the expense.

Even with hydraulic cylinders, pressure buildup may be fast but is not immediate. Oversizing of 10% is usually enough to achieve reasonably fast cycles in most situations. Other things that slow hydraulic actuator cycle time and pressure buildup include trapped air, slow shifting valves or valves with long spool overlap, and using pressure-compensated pumps without an accumulator.

Typical cylinder seals

The cutaways in Figure 15-12 show the standard ways of sealing pistons and rods from fluid bypass. O-rings and U-cups are resilient and continuous seals that can stop all fluid bypass so a cylinder can be stopped and held almost indefinitely. Cast-iron piston rings do not seal completely but do a good job when system shock or high heat is part of the operation.

Figure 15-12. Standard options for sealing pistons and rods from fluid bypass

O-ring seals are inexpensive but do not have long life in most applications. They must be setup with interference so as they wear they continue sealing. This also adds to breakaway and running friction. They are bi-directional and should not be used in pairs. With two O-rings on a piston, fluid can bypass and get trapped between the seals causing a lot of friction. At pressures above 500 psi, always specify backup rings to keep the O-rings from extruding and being damaged. Always keep O-rings well-lubricated because they can stick and roll in their cavity. When they start rolling they wear rapidly.

U-cups are long-life, low-friction seals because their flexible lips can be heavily preloaded without much increased friction. These seals are pressure energized, which means pressure inside the “U” forces the seal material tightly against the sealing surfaces. U-cups usually use backup rings at pressures above 1500 psi to keep them from extruding.

Material for the resilient seals discussed here is commonly Buna N rubber for pneumatics and many hydraulic cylinders. When synthetic fluid or high heat is encountered, resilient seals often are made of Viton. Buna N is good for temperatures up to 250° F with mineral oil and water glycol. Viton works at temperatures of 400° to 450° F and any fluid used in air or hydraulic circuits. Other seal materials such as leather, Neoprene, and Teflon are available and are standard for some suppliers.

Automotive-type cast-iron piston rings in sets of two or four seal well but not completely. A general rule is that they will leak approximately one cubic inch per minute per inch of bore per one thousand psi. Different end effects have been designed to reduce leakage at the expansion joint. Placing the expansion joint 180° from each other for consecutive rings helps slow leakage.

Cylinder cushions

Cylinders that move at high speed need some sort of deceleration method to keep them from slamming when they reach end of stroke. Some applications with high speed and heavy loads may need valves and limit switches to give enough time to bring the load to a smooth stop. For most cases standard cylinder cushions work well. They are 3/4- to 11/16-in. long and can be in the head or cap end or both. Cushions add cost on most cylinders so should not be specified when unnecessary.

Hydraulic cylinder cushions

The cutaway in Figure 15-13 shows a cylinder with an oversize rod and standard cushions on both ends. The cushion-adjusting screw and cushion bypass check valve can only be installed in the cap end because the cushion plunger on the rod end is so large there is no room for them. The necessity of oversize rods is a problem for most manufacturers. It is possible to taper the cushion and get a smooth stop but it also slows the stroke as the cushion plunger leaves its chamber. The best option to control deceleration of oversize rod cylinders’ extend stroke is with external valves or proportional directional controls.

Figure 15-13. Operation of cushioned cylinders

The other problem with oversize rods and rod-end cushions is the high intensification pressure present when working pressure is high and/or there is a heavy overrunning load. Pressure in the rod end can easily reach two-to-four times rated pressure each time the cylinder extends. High pressure can damage tube end seals and piston seals, and stretch the cylinder tube past its tensile limits.

The top cutaway in Figure 15-13 shows the cylinder piston moving to end of stroke with full flow exiting through the cushion chambers almost unrestricted. This means the cylinder can travel fast until the cushion plungers enter their chambers. The bottom cutaway is what happens after the cushion plungers have entered the cushion chambers. Trapped fluid decelerates the piston quickly to a speed set by the cushion-adjusting screws. Deceleration is sudden on the straight cushion plungers shown here because hydraulic fluid is almost non-compressible. The piston continues to end-of-stroke at a preset controlled speed without damage to itself or the machine.

On cylinders with standard and some oversize rod sizes, the rod-end cushion would function the same as one on the cap end. There is always some pressure intensification on the rod end but this is normally not a problem. When the cylinder starts to extend again, the cushion bypass check valve opens to allow fluid to the full piston area so it can extend quickly at full force. Without the bypass valves, takeoff speed would be as slow as deceleration on retract.

The cutaways in Figure 15-13 show a way of making cushions work. In actual cylinder design, cushions are built in many ways but the general function is the same. One company offers a self compensating non-adjustable cushion option without adjusting screws or bypass checks. Another supplier offers tapered or tapered slots in the plungers that give smooth deceleration instead of the sudden slowdown. Tapered cushions only work for a given pressure, load, and mounting position. They must be figured from information collected prior to building the cylinder.

Pneumatic cylinder cushions

Pneumatic cylinder cushions have similar designs but operate differently because the fluid is compressible. When the cushion plunger enters a cushion chamber, the trapped air starts compressing according to Boyles’ Law discussed in Chapter 1. When the load is light and the starting trapped pressure is high enough, the cylinder stops smoothly without slamming. When the load is heavy and/or starting trapped pressure is low, the cylinder slows but still may bang the end of stroke. This situation requires the addition of external deceleration in the form of valves and/or shock absorbers. Often the rod cushion is ineffective due to less area for the trapped fluid to work against. As with hydraulic cylinders, oversize rods exaggerate the problem.

Most suppliers offer standard cushions on both air and hydraulic cylinders. This option could smooth the operation of certain heavy loads that need extra deceleration distance. These longer cushions only work well for a fixed load and speed condition. If the machine has changing loads and/or speeds a cushion is not the way to go. Using external shock absorbers or proportional valves makes control easy to setup.

Cylinder rod column strength and buckling

Most cylinders are designed to have ample rod strength regardless of mounting style to rated pressure and strokes to 20 in. Above 20-in. strokes, standard rod size may be too small to keep it from bending in compression. Some mounting styles are more prone to buckling problems and need special attention. When an order is placed for cylinders with long strokes that operate above certain pressures, cylinder suppliers usually offer engineering help on design changes and/or rod modifications. Most cylinder catalogs have formulas and/or charts that show how to figure rod column strength for different mounting styles. In many cases the life of a cylinder is dependent on cylinder mounting type and rod size.

Another problem encountered in cylinder design is with bearing loads. A cylinder has bearing points at the rod and its bushing and the piston in the bore. Neither area is strong enough to carry external loads such as a machine member or other parts. Always use the cylinder to push and pull -- not as a guide.

Figure 15-14 shows column strength and bearing load problems on long stroke cylinders and how to overcome them. The vertical cylinder at top left with the heavy load does not have a large enough rod to lift it. When pressure is applied the undersized rod will bend and the load will not move. If the load does move, rod flex soon wears the bushing and allows leakage. Matching rod size to load and pressure makes a workable design.

Figure 15-14. Cylinders with greater than 20-in. strokes

The lower horizontally mounted long-stroke cylinder with a standard rod has two problems -- column strength and high bearing loads. The weight of the cylinder as the cylinder extends wears on the bearing points and high force will finally bend the rod instead of moving the work. The oversize rod on the lower, horizontally maintained cylinder is sized to handle the load at maximum force but does little to help bearing load problems.

Adding a stop tube to the piston rod keeps the bearing points farther apart and reduces wear. Notice the overall package is longer now because the stop tube takes up space inside the cylinder. Always check out bearing side-loading problems before building the machine. It can be difficult to fit the longer stroke cylinder or more costly to constantly change rod bushings and seals.

Cylinder mounting styles

The mounting styles shown in Figure 15-15 depict the standard NFPA-approved ways to mount cylinders. Up to 25 different companies make cylinders that match these styles in every dimension. This means no supplier has to be the sole source for any cylinder on a machine. Starting at top left are the least expensive mounting styles. Tie-rod mounts are extensions of the tie-rod threaded section. These extensions go through a machine member with nuts installed and tightened to hold the cylinder in place.

Figure 15-15. NFPA-approved mounting styles

The standard offerings are designated two tie rods extended both ends, MX4, four both ends, MX1, four cap end, MX2, four head end, MX3. At best, these mounting styles hold a cylinder in place but not necessarily in a precise location. They should only be considered when what they are attached to does not travel a rigid path.

The tapped mount, top right, is another inexpensive mount but may be hard to locate and hold cylinder position. It can create problems on long-stroke cylinders that need to stretch when pressure is applied. With both ends tied down, the cylinder is not free to grow from pressure or external force without binding. They are designated by the MS heading.

The side and end lug-mounted cylinders are another way to rigidly mount a cylinder but also can be a problem on long strokes. The lugs are extra long so dowel pins can be installed after a precise position is determined. It is not good practice to have dowels in both ends of the cylinder or on opposite corners especially on long strokes. They are designated by the MS heading.

Other rigid mountings are flange mounts in rectangular and square style. They are designated by the MF and ME headings. The bolt-on flanges are for light-duty operation or where the flange is in compression. The extended head and cap flanges are extremely heavy duty and directly replace the bolt on flanges.

The most versatile mounting styles are pivot mounts. They allow freedom of movement in one or both planes and keep the cylinder in perfect alignment. The intermediate trunion is very good on long stroke applications since cylinder weight can be balanced and reduce bearing loads. Care in choosing cylinder mounting style can add years to its life and eliminate rod seal leaks for long periods. Interchangeable NFPA tie-rod design cylinders come in bore sizes of 1 1/2, 2, 2 1/2, 3 1/4, 4, 5, 6, 7, 8, 10, 12, 14, 16, 18, and 20 in. Rod sizes available are 5/8, 3/4, 1, 1 3/8, 1 3/4, 2, 2 1/2, 3, 3 1/2, 4, 4 1/2, 5, 5 1/2, 7, 8, 9, and 10 in. The largest rod diameter for any bore is equal to or slightly under half the area of the piston. This is commonly known as a 2:1 area ratio cylinder. Oversize rods give fast return at low force, allow for regeneration of rod flow, and allow for higher push force capabilities due to greater column strength. Many manufacturers make cylinders in smaller and larger bores but they are not necessarily interchangeable.

Steel mills often use what are called mill cylinders that have bolted-on heads and caps. They are special to a machine in most cases with little or no interchangeability. Mobile equipment uses cylinders with welded heads and caps that are designed to be thrown away when they fail. However, many repair shops cut them apart and repair them. Mobile cylinders also have screwed-on heads or special snap ring arrangements in their assembly.

The cylinder in Figure 15-16 has such a small diameter and is so far from the directional valve that the lines to and from it hold more fluid than it does. In this situation fresh filtered fluid never cycles through the cylinder so it overheats and all wear particles and rod ingestion stay in it until it fails. Failure is always premature and the amount of contamination present is excessive when it is repaired. Changing the lines to the dual flow path shown in Figure 15-16 eliminates the contamination problem. Cylinder operation and longevity is greatly increased because filtered fluid constantly flushes it.

Figure 15-16. Cylinder flushing circuit

A tee is installed at A and B ports and check valves facing opposite directions are placed at each end. Separate lines run from the check valves to a tee at each cylinder port. Now all fluid must go to the cylinder through one line and return through the opposite one. Fresh fluid is continually being fed to the cylinder while used fluid returns to be cooled and filtered.

The dual flow path circuit is more important when the cylinder is mounted vertically with long lines. Contamination in this configuration lays on the piston and seals and can score cylinder tube and cause premature seal wear when allowed to collect.

Oversize rod cylinders

Cylinders with oversize rods can end up with dangerously high intensification pressures under certain conditions. The cutaway in Figure 15-17 shows a cylinder mounted vertically with its rod down. The rod area is approximately half the piston area so the annulus area around the rod is also approximately half the area of the piston. The left circuit shows the cylinder extending with a meter-out flow control circuit. For the 5-in. bore cylinder with a 3 1/2-in. rod and the pump compensator set at 3000 psi, pressure in the rod end would be approximately 5880 psi as the cylinder approaches the work. If this is a hydraulic press, middle circuit, with 5000 lb of platen, tooling a load-induced pressure of 499 psi would bring the rod end pressure to 6379 psi. This much pressure could damage the cylinder seals, over-pressure the flow control, and exceed the rating of pipe connections.

Figure 15-17. Pressure intensification on an oversize rod cylinder

The circuit on the right eliminates all of the above problems but still allows speed control of the cylinder. A counterbalance valve on the cylinder rod end set at 100 to 150 psi above load-induced pressure would keep the platen from falling while at rest or while it is approaching the work. A meter-in flow control sets cylinder speed and as fluid enters it, pressure in the cap end only rises to approximately half the counterbalance valve setting. The cylinder starts extending when pressure in the rod end reaches approximately 574 psi, which is well below the rating of all components.

Each of the circuits in Figure 15-17 would control cylinder speed but the counterbalance circuit is the best choice for the reasons given plus it has a lot less energy loss.

Another reason for using an oversize rod is for regeneration circuits like the ones in Figure 15-18. Any single rod cylinder will at least attempt to extend with equal pressure at both ports and this is called regeneration. Whether it actually extends depends on the load it must overcome, maximum system pressure available, and what the rod diameter is. This is because the maximum force during regeneration is pressure times the area of the rod. The piston is in balance during regeneration and serves no function during the process.

Figure 15-18. Single-rod cylinder regeneration

The standard rod cylinder in Figure 15-18 would extend at the rate of 12.25 in./sec at a maximum force of 3142 lb. Even if this is ample force to extend the present load, the amount of flow in regeneration is excessive. As the cylinder is regenerating forward with flow of 10 gpm from the pump, there would be 52 gpm coming from the head end for a total of 62 gpm to enter the cap port. This high flow would cause excessive back pressure and keep cylinder speed slow because the circuit relief valve would be bypassing at system pressure.

Another reason using a standard rod cylinder is not good practice is that its retract speed would only be 2.33 in./sec, so overall cycle time would not increase as much as first thought.

The above scenario is the prime reason for using 2:1 rod area ratios for regeneration. The lower cylinder in Figure 15-18 is the same bore but has a 31/2 in. oversize rod. This rod is not exactly 2:1 area ratio because it uses NFPA standard sizes for interchangeability. All NFPA cylinders have the largest standard rod that is up to but not over 2:1 area ratio.

The figures for the 2:1 ratio rod now show a net force on extend of 9621 lb. at a speed of 4 in./sec. During regeneration, flow from the head end is 10.4 gpm with a total flow to the cylinder of 20.4 gpm. This is a good measure of force and a reasonable flow rate that usually overcomes work resistance at easy-to-handle flow rates.

Retract speed would be 3.8 in./sec, making extend and retract speeds almost equal. When the rod diameter is exactly 2:1, extend and retract speed and force are identical -- the same as a double rod-end cylinder. However, getting exactly 2:1 area ratios requires an odd size bore or rod that may require special seals.

The circuits in Figure 15-19 show some standard regeneration setups used for particular needs. The full-time regeneration is a replacement for a double rod-end cylinder circuit. With a 2:1 area ratio rod, it will have identical speed and force in both directions. Even with standard rod diameter cylinders, force and speed are within 10% to 12% of the same, which is often satisfactory.

Figure 15-19. 2:1 rod cylinders in regeneration circuit

The full force at pressure buildup example uses a sequence valve to indicate work resistance and direct head-end oil to tank. A check valve in the regeneration flow line allows regeneration flow to the cap end and prevents pump flow to tank during the full force portion of the cycle. This circuit extends fast until work contacts, no matter the size of the part.

The full force at limit switch circuit uses a normally closed two-way directional control valve to send head-end flow to tank when a limit switch is made. This cuts cylinder speed in half so part contact is less abrupt. This circuit protects tooling, allows more time for visual inspection of alignment, and can give an operator more time to respond to unsafe conditions.

The circuits in Figure 15-19 may need a counterbalance valve to retard running-away conditions when the cylinders are vertically mounted. When this is necessary, the counterbalance valve must be externally drained to eliminate backpressure in the pressure adjustment chamber.

Adding a bleed-off flow control to the line between the head-end port and the check valve and after a counterbalance valve allows cylinder speed reduction when required. For complete coverage of regeneration circuits see the author’s upcoming e-book Fluid Power Circuits Explained.

Rotary output actuators

Air and hydraulic motors also are other types of actuators that turn fluid power energy into rotary output. They are pumps in reverse and can come in as many varieties. Unlike cylinders that are rated in pounds force thrust, motors produce twisting or torque that turns a shaft. Motors are rated in torque output usually stated in pound-inch (lb-in.), pound-foot (lb-ft.) American; or Newton-meters, (N-m) Metric.

Figure 15-20. Formula for figuring torque

The example in Figure 15-20 shows how torque is measured and applied to shafts. With a 500-lb weight hanging from the end of a 2-ft arm rigidly attached to a shaft, the shaft would have to resist 1000 lb-ft of torque (500 lb X 2 ft) to keep from turning. This is only true when the weight is at right angles to the shaft. Changing arm length or the amount of weight changes the amount of torque. As noted in the example, the metric figure is in Newton-meters and one Newton-meter equals 8.851 lb-in. and 0.7375 lb-ft.

Why use a fluid motor?

The main reason cited for using air or hydraulic motors is that they have high torque in a small package. Air drills of 1/2 to 3 hp for many rotation speeds and torqus require less than one-third of the space a comparable electric motor setup would use. This means space at the work is less cluttered and/or more units can be applied. Hydraulic motors are even more compact especially in higher horsepower units. The main reason for the small size is no reduction gearing is required for hydraulic motors and small planetary units work well with air motors.

  • Other reasons for choosing fluid motors are:
  • They have instant or almost instant reversing capabilities. Because these motors are so compact and have little inertia to overcome, they can reverse rotation quickly without damage. Some motors act as oscillators that never make a full revolution.
  • Variable speed capabilities of fluid motors are simple and result in little change in torque when it is a low-speed/high-torque hydraulic motor. A change in flow with flow controls or variable volume pumps requires little sophistication when minor speed changes can be tolerated. Sophisticated servo controls allow speed changes and can maintain tight control at any speed.
  • Overload protection is part of any fluid power system and motors have the same ability. When a motor circuit meets resistance it cannot overcome, it stalls and holds torque without damage to the circuit or machine. Fluid motors are capable of stalling for long periods without over heating or damaging themselves. There is some internal leakage while stalled but this can be minimized with the right motor selection.
  • Another place where fluid motors shine is in wet or explosive atmospheres. These motors have no electrical input so they pose no threat from sparks or overheating. A hydraulic motor can operate underwater with bio-degradable fluids without any of the problems electric motors have in this application.

Pneumatic motors

Air motors are often very inefficient and with air compressor inefficiency added there is often no more than 20% utilization of input power. This high inefficiency is offset somewhat in most applications because the air motor only has to run while work is taking place. This could mean the air motor must start and stop often but this is not a problem for fluid motors.

There are a variety of pneumatic motors in different configurations and with different attributes. The vane type shown in Figure 15-21 is one of the oldest designs. They are usually high to very high speed from 1000 to 30,000 rpm for everything from sanders and grinders to dentist drills. These motors may offer unidirectional or bidirectional rotation according to their design use.

Figure 15-21. Vane-type air motor

The cutaway shows vanes in a rotor that is off center in a cam ring. Air at the inlet acts against the vanes halfway around, forcing the rotor to turn while spent air exhausts during the other half revolution. Only one vane in this unbalanced design is producing torque. The remaining vanes on the inlet side have pressure on both sides so they have no force.

The vane motor in Figure 15-21 is for medium-to-high-speed applications at low torque. Several manufacturers add reduction gears to give low-speed/high-torque capabilities.

Figure 15-22. Radial-piston air motor

The radial piston motor in Figure 15-22 does not require reduction gearing for most low-speed/high-torque applications. Force from half the pistons is driving the crankshaft to turn it while the remaining pistons are exhausting. Inlet-outlet ports connected to a rotary valve driven by the crankshaft send 40 to 100 psi air through cored holes in the body and cylinder bores to and from the pistons. Most radial piston air motors run at less than 500 rpm due to air energy waste at faster speeds. These motors can be physically large and take a lot of mounting space. Some manufacturers make axial or in-line piston motors while a few have experimented with gerotor designs.

Hydraulic motors

Hydraulic motors come in the same variety as pumps. Many are low-speed/high-torque, some are high-speed/low-torque, and a few are low- or high-speed/high-torque. The main difference between pumps and motors is that a motor is usually capable of having either port pressurized.

High-speed motors can reach 3000 rpm continuous to drive fans, lawn mower blades, and grinders. They usually do not have high torque starting capabilities but most applications they are used on do not require this feature. Low-speed/high-torque motors usually have 75% to 90% of their maximum torque to start. They usually operate at 500 rpm or less. Piston motors of the in-line and bent axis design have high low speed torque and can run at 1500 to 2500 rpm without losing efficiency.

Hydraulic gear motors

The gear motor shown in Figure 15-23 is one of the oldest designs and is built for high-speed/low-torque needs. At first it appears fluid entering the lower port pushes against two teeth to start the gears turning. However, a closer examination shows the left gear has fluid pushing on opposing teeth as it comes out of mesh and only the right gear has any twisting action. After one tooth of revolution, the left gear drives while the right gear is balanced and so on as the motor turns.

Figure 15-23. Gear-on-gear hydraulic motor

Hydraulic gerotor motors

The high-speed gerotor motor in Figure 15-24 has similar characteristics to the gear-on-gear motor just mentioned. This is not a popular design but the gerotor concept with the idler gear held stationery shown next is made by many manufacturers and holds more than 50% of the small-to-medium high-torque/low-speed motor market.

Figure 15-24. High-speed gerotor hydraulic motor

The generated rotor hydraulic motor shown in Figure 15-25 is made high-torque/low-speed by holding the idler gear still and allowing the orbiting gerotor to cycle inside of it. This change causes the orbiting gerotor to make as many power strokes as it has teeth for every revolution of the output shaft. The seven-tooth gear shown makes seven power strokes while the output shaft turns once. A splined drive connection follows the orbiting gear and transmits the rotary motion to the output shaft.

Figure 15-25. Low-speed/high-torque, hydraulic-generated rotor motor

Generated rotor motors give at or near full torque from about 25 rpm and normally do not go higher than 250 to 300 rpm. Maximum output torque is directly related to the width of the gerotor element which may be as narrow as 1/4 in. to 2 in. Pressure ratings as high as 4000 psi are common from most manufacturers.

Gerotor motors can have a selector valve that changes the internal rotary valve output to feed only half the chambers, causing the motor to run at twice the speed and half the torque. The gerotor design is machined with too close tolerances but must have some clearance to allow the inner gear to move. This makes it less efficient and as the gears wear, internal leakage increases. The geroler design has rolling seal points and can be setup much closer and has less wear for longer life. Most of the geroler types also use a plate valve which has less leakage and is wear compensating as well.

Hydraulic vane motors

The hydraulic vane motor shown in Figure 15-26 is a very efficient design and works well for applications at 20 to 3000 rpm. Fluid entering one port pushes against two or four vanes as they extend in the cavities of the cam ring. Internal porting directs pressure and return fluid to the working and exhausting vanes. While half the vanes are being pushed by fluid, the other half are discharging spent oil to tank. The amount of torque is in direct relationship to the vane area exposed to pressure fluid and the distance the vanes are from shaft center. Speed is limited to how much displacement and what size ports the motor has.

Figure 15-26. Vane hydraulic motors

The high-speed/medium-torque design with an elliptical cam ring gets full torque at approximately 100 rpm and can go as high as 3000 rpm. Because it covers such a broad speed range it is suitable for many applications where other designs fall short on torque or speed. The low-speed/high-torque design is designed for approximately 10 to 400 rpm and usually eliminates any need for gear reduction. Using a direct drive eliminates maintenance problems and makes a smaller package at the work area.

Hydraulic piston motors

The most efficient and versatile hydraulic motors are piston type but they are also the most expensive. Inline or axial and bent axis types operate smoothly from 10 to 2000 rpm with high torque throughout their speed range. Radial piston types go as low as 1 rpm but usually not higher than 400 rpm. Their main use is very high-torque/low-speed applications.

The cutaway in Figure 15-27 shows typical construction of inline fixed- and variable-displacement hydraulic motors. Low-displacement variable motors may be controlled manually while larger motors need pistons to change displacement. An inline hydraulic motor shaft rotates as fluid pushes against a piston, forcing its shoe to slide up the angled swash plate. A small amount of pressure fluid goes through an orifice in the piston and behind the shoe to keep it from rubbing metal-to-metal while it is producing torque.

Figure 15-27. Axial or inline piston motors

With the swash plate at a steep angle, torque is high while speed is usually low. A shallow swash plate angle gives high speed but less torque. Most manufacturers recommend a minimum swash plate angle of 15° to 17° for best results. A maximum angle of 40° to 45° gives good torque and long motor life.

Figure 15-28. Fixed-volume, bent-axis hydraulic motor

The bent-axis piston motor in Figure 15-28 has the same operating characteristics as the inline motor but is more rugged and capable of higher operating pressures. Since there is no sliding action of piston shoes there is less friction and higher torque for a given energy input. The angle of the cylinder block to the input shaft determines torque and speed ranges. The greater the angle, the higher the torque and the lower the speed. Kidney-shaped openings in both inline and bent-axis motors port fluid to and from pistons as they rotate. Internal leakage is sent to tank through the case drain. Variable-displacement bent-axis motors are available but not commonplace due to expense and size.

Figure 15-29. Radial-piston hydraulic motor

The radial piston motor shown in Figure 15-29 uses pistons pushing against an eccentric to produce rotary motion. These motors usually have five or seven pistons with rods and shoes, with half of them pushing against the eccentric while the other half return oil to tank. The shoes have high pressure fluid fed to them from the piston through the rod to keep them from rubbing the eccentric during the power stroke. A rotary valve attached to the output shaft feeds and exhausts fluid to and from the pistons as they turn the eccentric.

Some radial piston motors are made with a moveable eccentric that allows different offset amounts. Usually the offset is full or one-half so a motor with this feature can have higher speed at lower torque for fast movement. Eccentric-type radial piston motors are one type of motor that cannot function as a pump without special inlet considerations.

Other radial piston motor designs are similar in action and torque output, but arrange the pistons in different configurations. One design has the pistons facing in and pushing outward against a cam-shaped housing. The shaft is connected to the machine and the housing rotates. It was originally designed as a wheel motor.

Rotary actuators

When rotary output is one or two turns or less, a hydraulic motor could be used but repeat stopping accuracy could be a problem. There are several designs of rotary actuators that give rotary output for limited numbers of revolutions (usually under one revolution).

Figure 15-30 shows one way to achieve rotary action when the motion is 90° or less. A clevis-mounted cylinder attached to a lever arm that is driving a shaft gives no less than half the cylinder force times the lever arm length. In the example shown, torque on the shaft would be approximately 4000 lb-in. when the cylinder starts and finishes and approximately 6600 lb-in. when the angle between the cylinder and lever arm is 90°. Often maximum torque is only required at the end of stroke so arrange cylinder mounting to give the greatest torque at that point. Remember, retract force is less than extend force which could cause the cylinder to stall when reversed.

Figure 15-30. Clevis-mounted cylinder for rotary action

The vane-type rotary actuator in Figure 15-31 is a common design for both pneumatic and hydraulic fluids. The vane has seals around the edges where it contacts the housing to control leakage. Fluid entering from the left, as shown, pushes the vane away and forces fluid out the opposite port as it turns the output shaft. A single-vane rotary actuator is usually limited to 270° rotation or less. A double-vane rotary actuator usually only turns 90° or less. Torque is equal to the vane area times input pressure times the radius from the center of the output shaft to halfway across the vane. The symbol shows a semi-circle to indicate rotary action that is not capable of being continuous in either direction.

Figure 15-31. Vane-type rotary actuator

Another common rotary actuator is the single-cylinder rack-and-pinion rotary actuator shown in Figure 15-32. Several companies make these units in single- and dual-cylinder models with torque outputs up to and above 300,000 lb-ft. They can be pneumatic- or hydraulic-actuated. This rotary actuator design can turn more than one revolution because turns are in direct relation to pinion gear size and rack gear length. It may also be supplied with cushions, and/or stroke limiters for smooth adjustable stops. Fluid entering ports at the cylinder ends forces the piston to move away and drive the rack gear against the pinion gear. The pinion gear continues to rotate until the opposite piston bottoms out. Reversing flow to the pistons reverses output shaft rotation. Torque is equal to piston area times input pressure times the radius of the pinion gear.

Figure 15-32. Single-cylinder, rack-and-pinion rotary actuator

The helical gear rotary actuator shown in Figure 15-33 shows another design that can have more than one revolution of the output shaft. The number of rotations is in direct relation to gear teeth angle and the non-rotating piston stroke. Fluid entering and pushing against one side of the non-rotating piston forces it to move and impart the turning action to the output shaft through the helical gear and nut arrangement. This rotary actuator design is able to have deceleration built in and may be purchased as a double-shaft model.

Figure 15-33. Helical gear rotary actuator

The chain-and-sprocket rotary actuator in Figure 15-34 depicts another way of achieving limited rotary output with more than one turn capability. The number of turns is in direct relation to the sprocket size and the working piston stroke. It can be used with pneumatic or hydraulic fluids. Fluid entering the CCW port pushes against the working piston and the isolating piston at the same pressure. Because the working piston has more area it will move left while the isolating piston moves right. The effective thrust is pressure times the area of the working piston minus pressure times the area of the isolating piston. The result of these calculations times the radius of the sprocket determines torque.

Figure 15-34. Chain-and-sprocket rotary actuator

There are some other rotary actuator designs, but in most cases, they are variations of the ones presented here. For circuits using rotary actuators see the authors upcoming e-book Fluid Power Circuits Explained.

Quiz

 

Chapter 16: Accumulators

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.

Quiz

 

Chapter 17: Air-Oil Systems 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.")

Quiz

 

Chapter 18: Miscellaneous Fluid Power Items

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.

Quiz

 

Chapter 19: Moving Part Air Logic Controls

Moving part air logic controls

Electrical and electronic devices normally control fluid power circuits. Relay logic circuits, programmable controllers, or computers are common control methods. Another way to control fluid power systems is with moving part air logic. Air logic controls perform any function normally handled by relays, pressure or vacuum switches, time delays, counters, and limit switches. The circuitry is similar, but compressed air is the control medium instead of electrical current.

Environments with high concentrations of dust or moisture are excellent places for air logic controls. There is no danger from explosion or electrical shock even in such environments. Water can splash on the controls with no effect on their operation. If there is danger of explosion, air controls cannot ignite the materials involved. Another place to use air logic is on machines that have cylinders or fluid motors but no electrical devices. Machines powered by air and controlled electrically must be supplied with both utilities -- and require two crafts to work on them. With air logic, a single craft works on the circuit and the machine parts.

A disadvantage of using air logic control is a general lack of understanding about how the components work and how to read the schematic drawings. If an air-controlled machine fails, very few people have the skill to work on it. Also, air logic with long control lines will have a noticeably slower cycle. Control lines longer than ten feet fill and exhaust slowly when compared to electrical signals. Another thing: control air quality must be above average for long trouble-free life.

What are air logic controls?

Air logic controls are basically miniaturized 3-way and 4-way air valves. The actions of the valves produce on or off functions -- like relays or switches do -- plus exhausting of the spent signal. The symbols used for air logic are similar to electronic symbols. Some manufacturers use modified electrical symbols and ladder diagrams to show circuitry.

Following are explanations of the basic air logic components with figures showing the ANSI logic symbol, an equivalent ISO graphic symbol, and a generic cross-sectional view of the element.

AND element

Figures 19-1 and -2 show two types of AND elements. An AND element must receive two inputs before there is an output. This assures that two functions have been completed before there is a command to continue the cycle. Another way of saying this is that there must be a signal at A and B before getting an output at C. For more than two inputs, connect AND elements in series. The first AND receives two signals and its output connects to one input of a second AND. The other input of the second AND receives the third signal making three inputs necessary before passing an output. The first AND takes two inputs but any additional input requires another AND. (See Figure 19-14 for an example.)

Fig. 19-1.Passive AND element

From the cross-sectional view it is seen that air entering the A or B port will push the dual-seat poppet onto a seat and block flow. However, when a signal is present at both the A and B ports, the lowest pressure signal will pass to the C port. This element may allow a small amount of air to pass when it receives the first signal if the dual-seat poppet is off the seat on the side from which the signal is coming. In most cases this is not enough of a signal to start the next function. If this is ever a problem, use the YES element discussed next.

YES element

Some manufacturers supply both types of elements -- calling the element in Figure 19-1 an AND, and designating the element in Figure 19-2 as a YES. The difference between the elements is that the AND in Figure 19-1 is passive because its output is always the lower of the two inputs. In contrast, B is always the output of the YES element's two inputs. Using this feature can amplify a weak signal because it pilots the valve open at the A port while the through signal at B comes from a full pressure supply. A YES element is called active because there is a choice of which signal passes to the output.

Fig. 19-2. Active AND element (or YES)

From the drawing in Figure 19-2 it is easy to see that a signal at the A or B port cannot pass. A signal at A only shifts the dual-seat poppet; a signal at B is blocked in the at-rest condition. Any air that was present at C exhausts through the exhaust port. The flexible diaphragm above the poppet keeps air that is entering A from exhausting or producing an output. This logic element can be used to amplify a signal because of the approximately 10:1 area ratio difference between the A and B ports. This means a 10-psi signal at A can shift against a 100-psi input at B.

OR element

OR element symbols and a cross-sectional view are shown in Figure 19-3. (A shuttle valve serves the same purpose as an OR element.) Either input to an OR element produces an output. Pilot signals from two different sources can pass through to start the next function. Another way of saying this is that a signal at A or B produces an output at C. An OR element differs from an in-line tee because an OR passes either input to the output but does not allow the inputs to pass to each other.

Fig. 19-3. OR element

From the cross-sectional view, it is easy to see how a signal entering the A or B ports can only pass out the C port. A slight puff of air may escape as the blocking wafer moves from seat to seat after an input signal, but this is usually not strong enough to start another function. Although an OR is a passive element, that is not a problem because it always passes the higher of the two signals it receives.

Stack OR elements to allow for more than two inputs. Use an extra OR for each new input after the first two signals -- with one input from the preceding OR and the other from the new signal.

NOT element

The NOT element symbols and cross-sectional view in Figure 19-4 designate an active, normally open logic element. A NOT logic element is a normally open 3-way valve. An input signal or pressure to the supply port passes through the valve until there is a pilot signal at port A. Pressurizing port A blocks supply and exhausts the output signal to atmosphere. Without a pilot signal, a NOT always returns to the normally open condition.

Fig. 19-4. NOT element

From the cross-sectional view you can see that supply is free to flow to output in the normal condition. A signal at the A port pushes the dual-seat poppet down to block supply and exhaust the output air. This element will block a signal as well as a supply when necessary. A diaphragm above the poppet prevents air entering port A from exhausting or passing to output. The area on top of the dual-seat poppet is approximately 10:1 in size over the shutoff seat at the supply port.

For some applications, a NOT element can replace a limit switch to indicate that a cylinder is at the end of stroke. Pressure from a cylinder port goes to port A of the NOT, holding it closed. As the cylinder strokes to the work, pressure stays the same when using meter-out flow controls. When the cylinder contacts the work, the signal on port A drops, the NOT then opens and sends a signal to start the next operation. (For a full explanation of this circuit see the author's book "Fluid Power Circuits Explained." Also see Figure 19-12 for a NOT as an end-of-stroke indicator.) Using a NOT element as a stroke-position indicator is not positive because an output signal is generated any place the cylinder stops for any reason. This always happens, whether the cylinder stopped where it should or stalled for some other reason. Because this is the case, take care using a NOT to replace a limit valve. Conversely, this feature can be advantageous when clamping different sized parts. Use a NOT element for applications where different work locations stop the cylinder and there is no safety hazard or possible part damage when the cylinder does not complete its full stroke.

Most manufacturers supply a different pilot ratio for a NOT element to be used as a limit switch. The valve function is the same but the pressure at which it shifts is lower. Some manufacturers make a special NOT element that mounts directly to a cylinder port. A port-mounted, meter-out flow control used in conjunction with this special NOT makes a compact installation.

Flip-flop element

The cross-sectional view and symbols in Figure 19-5 are for a flip-flop element. A flip-flop is a double-piloted 5-way valve that sends supply air to either the outlet port with a signal at pilot ports S or R. The S signal stands for set and the R signal stands for reset. The S signal shifts the flip-flop for a function and whether the signal continues or not, the element stays shifted. The R signal puts the flip-flop back to its original position for the next cycle. Supply can be system pressure or air from another logic element. Manual Overrides make it possible to check the valve function and operate the circuit manually. Flip-flops (sometimes called memory elements) stay in the last shifted position even with no air supply. Whether the signal is maintained or drops out, the output port of a flip-flop stays the same.

Fig. 19-5. Flip-flop element

The main use for a flip-flop is to eliminate dual pilot signals to a directional control valve powering an actuator. Applying two pilot signals to a directional control valve leaves the valve in the position it was when the second signal arrived. (In Figure 19-12 there is a circuit using a flip-flop to perform this blocking function.)

Another use for a flip-flop is to start a new cycle by allowing the operator to momentarily push the start buttons. This same flip-flop can then eliminate unwanted signals and set up the circuit for cycle completion as required.

A normally closed, double-pilot-operated 3-way element with only one output is also available to perform the same function as a flip-flop.

Fig. 19-6. Memory element

The symbol in Figure 19-6 shows how a flip-flop is represented schematically and the cross-sectional view shows one manufacturer's configuration. Note that this type uses differential pilot areas so the valve can be reset when there is a set signal present.

Time-delay elements

In air logic control there are several different types of time delays. Fixed or adjustable time delays are common in both on delay (or normally closed) and off delay (or normally open) configurations. Some time delays use a fixed or variable orifice and an accumulator chamber to produce delays as long as one minute. Some types use air-actuated diaphragms and orifices to eliminate inaccuracies due to supply pressure fluctuations. (Most types depend on guesses or stopwatch setting procedures.)

One-shot element

The symbol and cross-sectional view in Figure 19-7 is for a one-shot timer (also called an impulse generator.) A one-shot timer takes a signal at A and passes it on to the circuit. At the same time, input signal A goes through an orifice to an accumulator tank. The setting of the orifice and size of the accumulator tank gives a certain time delay before the normally open 3-way valve under the accumulator tank is piloted closed. After a one-shot times out and closes, it remains closed as long as the input signal at A stays on.

Fig. 19-7. One-shot element

The symbol in Figure 19-7 shows an adjustable time delay. Leaving off the sloping arrow in the symbol makes it a preset non-adjustable time delay. Times range from one half second to two or more seconds on valves with preset time delays.

Many circuit designs use one-shots to eliminate opposing signals. When a valve receives a signal to shift, the opposite pilot signal has no effect until loss of the first pilot signal. A one-shot element drops the first signal shortly after initiation, thus making the valve ready to accept the opposite signal. One problem is that if the short-duration signal meets a hard-to-shift valve, the time period may not be long enough to move the valve spool. When the valve does not have time to shift, the cycle stalls. For best results, add a flip-flop to drop an unwanted signal after it performs its task

Time-on delay element

Figure 19-8 shows an adjustable, normally closed time-on delay symbol and cross-sectional view .A time-on delay passes a signal through the element after timing stops. The symbol without the sloping arrow is a preset, fixed time-on delay. Most anti-tie-down circuits use a fixed time delay to force the operator to actuate two palm buttons concurrently.

Fig. 19-8. Time-on delay element

The symbol in Figure 19-8 shows supply air going to the blocked port of a 3-way directional valve. A signal also goes through a fixed or variable orifice to fill an accumulator tank. After the accumulator tank fills, pilot pressure shifts the 3-way valve to allow supply air to pass to the next operation. If the input signal stays on, the output stays on.

With an integral accumulator tank, time delay length is usually around one to one-and-a-half minutes. With added external accumulators, time delays up to five minutes are possible. The repeatability of long time delays using accumulators is poor. Diaphragm-type timers often go to three minutes with good repeatability.

Time-off delay element

With a normally open 3-way valve in place of a normally closed 3-way, valve, the timer becomes a time-off delay. Figure 19-9 shows the symbol and cross-sectional view for a time-off delay element. A continuous input to the supply port gives an output until a set time after receiving a signal. The input signal starts to fill the accumulator tank through a fixed or variable orifice and when it is full, it closes the normally open 3-way valve and exhausts the signal.

Fig. 19-9. Time-off delay element

Time-on and time-off delays often are identical in appearance. Checking the part number may be the only way to tell these units apart.

Air logic valve combinations

To get different functions, connect air logic elements together like the examples in Figures 19-10 and 11. They illustrate two common pairs that might be familiar to anyone using air or electronic logic.

Fig. 19-10. NAND output

A NAND output, Figure 19-10, uses an AND to signal a NOT. As long as there are no signals at A and B, air passes. If signals are present at A and B, the NOT closes and exhausts the output signal. The term NAND comes from the phrase: not A and B.

Fig. 19-11. NOR output

A NOR element, Figure 19-11, uses an OR to signal a NOT. When there is no signal at A or B, air passes through the NOT element. If a signal is present at either A or B, the NOT closes and exhausts the output signal. The term NOR comes from the phrase: not A or B.

Other air logic components:

Amplifiers detect low-pressure signals (as low as a 3-in. water column) and send them on as an 80-psi signal. They can be used with gap sensors to detect whether a part is breaking an air-bleed signal.

Pressure or vacuum sequence elements shift after reaching a set pressure or vacuum level. They can be used to indicate an actuator has completed its stroke. They should not be used when the actuator positively has to finish its task before the next operation starts.

Air-operated indicators show circuit conditions and/or functions. Several colors are available, but none emit light.

Example circuit

The drill circuit in Figure 19-12 uses most of the components discussed in this chapter -- plus other valves from other parts of this book. For safety, this circuit requires both of the operator's hands to be on the palm buttons simultaneously before a part can be clamped and drilled.

Fig. 19-12. Drill circuit with anti-tie-down start and anti-repeat

When the operator pushes Start 1 and Start 2 palm-button-operated, normally closed 3-way air valves, they send a signal to the AND1 and OR1 elements. One OR1 output goes to the N.C. time delay and it starts timing. At the same time, both signals from the palm-button valves satisfy AND1, so it sends a signal through NOT2 to the flip-flop to start the cycle. The output from NOT2 also goes to NOT1 and closes it. This keeps the N.C. time delay (which will time out in 0.5 to 1.0 sec) from sending a signal to block the start signal from AND1 to the flip-flop. When the flip-flop shifts, it sends a signal to A+ to start the clamp extending. The same signal also goes to LV1, where it is blocked. When the clamp reaches the part, it mechanically shifts LV1.

After the clamp cylinder extends fully, pressure in its rod end drops the inlet pressure to Start 1. If the operator still had the palm buttons depressed, the loss of air at Start 1 would be the same as releasing that button. Releasing either Start 1 or Start 2 causes AND1 to stop sending a signal to NOT2, which drops its signal to the flip-flop and to NOT1. When NOT1 loses its signal, the N.C. time delay sends a signal through NOT1 to close NOT2, thus preventing a later signal from the palm buttons through AND1 from giving another start signal. This same scenario also requires the operator to let up on both palm buttons anytime they are not depressed simultaneously within a 0.5- to 1.0-second delay. Now the anti-tie-down circuit also is anti-repeat.

LV1 sends a signal to start all three drills on their drill cycle. When the drills leave their home positions, they put out a signal that goes to AND2, AND3, OR2, and OR3. When AND2 and AND3 receive all three drill signals, they send a signal to the flip-flop that shifts it back to its home condition. Its output now goes to the inlet of NOT3. Any or all of the outputs of OR1 and OR2 hold NOT3 shut.

The three drills continue forward until they finish their operations. They then retract automatically (as commanded by their own internal valves). As each drill reaches home position, it drops its run signal to OR1 or OR2. When all drills have retracted fully, NOT3 opens and sends a signal to A- to unclamp the part.

Air logic controllers

To simplify circuit design and troubleshooting, most manufacturers of air logic components now offer a combination of elements in one control module. These units combine a flip-flop, an AND, and an OR element in one component. The symbols in Figure 19-13 show the parts and their arrangement in a single module. One module is required for each signal to and from a circuit. This means for a 3-actuator circuit, there must be at least six logic controllers. These controllers stack and lock together in a row and they have end closures with supply, start, and cycle-end connections.

Fig. 19-13. Air logic controller

The first controller in the row receives the start signal that shifts a memory element. The memory element sends a signal to one port of an AND element and an output to the first actuator's air-piloted directional control valve. It also goes to an OR element that sends a reset signal to the start circuit or to the previous controller The first actuator strokes and makes a limit valve at the end of its stroke that sends a return signal, indicating the action has taken place. The actuator's return signal satisfies the other port of the AND, which signals the next module to start the next sequence. This scenario repeats until the end of cycle when the last AND output indicates the controller is ready for another start signal.

The circuit in Figure 19-13 shows the simplicity of an air logic controller setup for the drill circuit in Figure 19-12. It also shows how other logic elements may still be needed to tie common functions together.

Drill circuit using air logic controllers

The circuit in Figure 19-14 pictures another way of controlling the drills using air logic controllers and multi-function modules in place of all air logic modules. The machine functions the same as previously described, but there are fewer connections to make and less circuit design skills involved. For this circuit, some logic elements still were required because the drills' feedback is not from switches. A NOT element passes a signal when the clamp cylinder is retracted at the end of the cycle.

Fig. 19-14. Drill circuit with anti-tie-down start and anti-repeat using logic controllers

When the operator depresses the Start 1 and Start 2 palm buttons simultaneously, an output from the anti-tie-down module starts the first logic controller and it sends a signal to the clamp cylinder's directional control valve to extend the clamp. When the clamp fully extends and makes limit valve LV1, LV1 sends a signal back to the controller's first section which drops the clamp extend signal and starts the controllers second section.

Output from the second section of the controller starts the drills extending. As they begin to move, they send signals to two ANDS and two OR elements. After all drills start, the two AND elements send an output to the second controller, dropping the drill-start signal and starting the third controller.

Output from the third controller goes to the NOT2 element, which is being held shut by signals from the drills through any or all of the OR elements. When the last drill has fully retracted, the signal blocking NOT2 drops out and the third controller sends a signal to the clamp's directional control valve to retract it.

Air from the cap end of the clamp cylinder has been holding NOT1 closed and will do so until the clamp fully retracts and pressure drops in its cap end. When pressure drops in the clamp cylinder cap end, NOT1 opens and signals the third controller to drop its output and send a signal back to the first controller that the cycle is complete. (A limit valve could have been used in place of NOT1.)

(For more air logic circuits, see the author's second Ebook: "Fluid Power Circuits Explained.")

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Chapter 20: Brain Teasers

BRAIN TEASERS


Problem #1

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Problem #2

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Problem #3

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Problem #4

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Problem #5

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Problem #6

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Problem #7

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Problem #8

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Problem #9

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Problem #10

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Problem #11

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Problem #12

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Problem #13

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Chapter 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.













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Chapter 22: Useful Fluid Power Formulas


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