3.2 Hydraulic Cylinder

3.2. Hydraulic cylinders and cushioning devices 2012

Hydraulic and pneumatic control lecture notes by Siraj K.

3.2. Hydraulic cylinders and cushioning devices

3.2.1. Introduction

Hydraulic cylinders convert fluid flowing under pressure to a linear motion. They are single-

acting and returned by either a spring or the force of gravity acting on the load, or they are

double-acting and have fluid piped to both ends of the piston. Single-acting cylinders with large

bore-to-diameter ratios are called rams. Figure 3.2.1 illustrates how cylinders and cushioning

devices are classified. The force exerted by a hydraulic cylinder is proportional to the cross

section area of the piston and the pressure of the fluid. Double-acting cylinders that have the rod

extending through both end caps exert the same force in both directions.

Cylinders with single-end rods exert greater force when the cylinder extends. The difference in

force between the two is accounted for by the difference in area between both sides of the piston.

Cushioning devices decelerate the load to prevent shocks that could damage cylinders and other

equipment. Cylinder cushioning devices use a tapered or stepped rod that enters a sleeve

mounted in the end caps of the cylinder as it nears the end of the stroke. A typical hydraulic

shock absorber cushions bumping loads with a spring-returned hydraulic cylinder that forces the

fluid through an adjustable orifice when the load strikes the end of the rod.

Figure 3.2.1 Actuator classification.



Cylinder actuators are defined by their ability to exert a linear force and hold it at any specified

position indefinitely. Several factors are used to select hydraulic cylinders, including:

• Purpose of the cylinder

• Construction

• Required force

• Temperature

• Load

• Duty cycle

• Mounting

• Acceleration and deceleration forces

The purpose of the cylinder describes its type. The cylinder may move a load resistance in one

direction only with a gravity or spring return, or move it in both directions under pressure, or

have equal velocities as it travels in both directions. Cylinders, then, may be of the single-acting

type, double-acting type or have a double-end rod, "DER," to equalize the areas on both side s of

the piston and add support. If the cylinder piston is connected to a single-end rod, it is considered

to be of a hydraulically unbalanced design. Double-end rod cylinders are considered to be

hydraulically balanced because both sides of the piston have the same area. Tandem balanced

cylinders exert twice the force available from single piston cylinders by providing twice the

effective piston area in two chambered spaces.

3.2.2. Hydraulic cylinders operating features

A cylinder is a hydraulic actuator that is constructed of a piston or plunger that operates in a

cylindrical housing by the action of liquid under pressure. Cylinder housing is a tube in which a

plunger (piston) operates. In a ram-type cylinder, a ram actuates a load directly. In a piston

cylinder, a piston rod is connected to a piston to actuate a load. At the end of a cylinder from

which a rod or plunger protrudes is a rod end. Its opposite end is the head end. The hydraulic

connections are a head-end port and a rod-end port (fluid supply).

Hydraulic cylinders are constructed of a cylinder barrel, piston and rod, end cap s, ports, and

seals (Fig.3.2.1). The piston provides the effective area against which fluid pressure is applied

and supports the piston end of the rod. The opposite end of the rod is attached to the load



resistance. The cylinder bore, end caps, port s, and seals maintain a fluid tight chamber into

which fluid energy is piped. Routing of the fluid determines whether the cylinder extends or

retracts. A common feature added to a cylinder is a mechanism called a cushioning device,

dashpot, snubber, or decelerator to reduce shock loads that would be caused by bottoming the

piston in either extreme stroke position. Cylinders referred to as rams use large cylinder rod s

approaching the size of the cylinder bore to give maximum support to the load end of the rod.

They are used extensively in such single-acting applications as car hoists , dump cylinders , and

hydraulic presses.

Figure 3.2.1 Cylinder components (courtesy of Tomkins-Johnson Company).

(a) Single-acting cylinder.

This cylinder has only a head-end port and is operated hydraulically in one direction. When oil is

pumped into a port, it pushes on a plunger, thus extending it. To return or retract a cylinder, oil



must be released to a reservoir. A plunger returns either because of the weight of a load or from

some mechanical force such as a spring. In mobile equipment, a reversing directional valve of a

singleacting type controls flow to and from a single-acting cylinder.

Figures 3.2.2 and 3.2.3 illustrate the construction of single-acting hydraulic cylinders. The

piston, or plunger, is driven hydraulically in one direction. In the other direction, the piston

moves under the action of an external force or a built-in spring.

Figure 3.2.2 Single-acting plunger-type cylinder, returned by external force.

Figure 3.2.3 Single-acting piston-type cylinders, spring returned.

A single-acting, piston-type cylinder uses fluid pressure to apply force in one direction. In some

designs, the force of gravity moves a piston in the opposite direction. However, most cylinders of

this type apply force in both directions. Fluid pressure provides force in one direction and spring

tension provides force in the opposite direction. Most piston-type cylinders are double acting,

which means that fluid under pressure can be applied to either side of a piston to provide

movement and apply force in a corresponding direction. This cylinder contains one piston and

piston rod assembly and operates from fluid flow in either direction. The two fluid ports, one

near each end of a cylinder, alternate as inlet and outlet, depending on the directional-control

valve flow direction. This is an unbalanced cylinder, which means that there is a difference in the

effective working area on the two sides of a piston. A cylinder is normally installed so that the



head end of a piston carries the greater load; that is, a cylinder carries the greater load during a

piston-rod extension stroke.

(b) Double-acting cylinder.

This cylinder must have ports at the head and rod ends. Pumping oil into the head end moves a

piston to extend a rod while any oil in the rod end is pushed out and returned to a reservoir. To

retract a rod, flow is reversed. Oil from a pump goes into the rod end, and the head-end port is

connected to allow return flow. The flow direction to and from a double-acting cylinder can be

controlled by a double-acting directional valve or by actuating control of a reversible pump.

The piston of a double-acting hydraulic cylinder is driven hydraulically in both directions of

motion. This cylinder may be single rod (Fig. 3.2.4), twin-rod symmetrical (Fig. 3.2.5), or twin-

rod nonsymmetrical (Fig. 3.2.6). The twin-rod cylinder is said to be symmetrical if the diameters

of the piston rods are equal. It is usually used in hydraulic servo systems.

Figure 3.2.4 Single rod cylinders.



Figure 3.2.5 Twin-rod symmetrical cylinders

Figure 3.2.6 Twin-rod nonsymmetrical cylinders.

c) Tandem Cylinders

The tandem cylinder duplicates the pressure force, for the same barrel diameter (see Figs. 3.2.7

and 3.2.8).

Figure 3.2.7 a differential tandem cylinder.

In a differential cylinder, the areas where pressure is applied on a piston are not equal. On a

head end, a full piston area is available for applying pressure. At a rod end, only an annular area

is available for applying pressure. The area of a rod is not a factor, and what space it does take up

reduces the volume of oil it will hold. Two general rules about a differential cylinder: with an



equal GPM delivery to either end, a cylinder will move faster when retracting because of a

reduced volume capacity. With equal pressure at either end, a cylinder can exert more force

when extending because of the greater piston area. In fact, if equal pressure is applied to both

ports at the same time, a cylinder will extend because of a higher resulting force on a head end.

Figure 3.2.8 a symmetrical tandem cylinder.

c) Three-Position Hydraulic Cylinders

Some of the operating organs may have three operational positions. In this case, the ordinary

double-acting hydraulic cylinder does not give the required controllability. Figure 3.2.9 shows

the typical construction of a three-position cylinder. The cylinder has two separate pistons and

piston rods. The three positions are obtained by pressurizing the cylinder chambers, as shown in

the table in Fig. 3.2.9.

Figure 3.2.9 Operation of a three-position cylinder. (Note: P and T are the pressure and tank

lines.)

d) Telescopic Cylinders

Telescopic cylinders are used in industrial and mobile equipment hydraulic systems. This class

of cylinders provides long cylinder strokes with relatively small installation space. The

telescopic cylinder may be either single acting (Fig. 3.2.10) or double acting (Fig. 3.2.11). If the



pressure affects the pistons via port A, they travel outward one after another. The double-acting

cylinder retracts by pressurizing the port B.

Figure 3.2.10 Single-acting telescopic cylinders. (Courtesy of Bosch Rexroth AG.)

Figure 3.2.11 Double-acting telescopic cylinders. (Courtesy of Bosch Rexroth AG.)

3.2.3. Cylinder mounting and mechanical linkages

Various types of cylinder mountings are in existence, as illustrated in figure 3.2.12. This permits

versatility in the anchoring of cylinders. The rods ends are usually threaded so that they can be

attached directly to the load, a clevis, a yoke or some other mating device. Through the use of

various linkages, the applications of hydraulic cylinders are limited only by the ingenuity of the

fluid power designer. As illustrated in figure 3.2.13, these linkages can transform a linear motion

into either an oscillating or rotary motion. In addition, linkages can also be employed to increase



or decrease the effective leverage and stroke of a cylinder. Much of efforts have been made by

manufacturers of hydraulic cylinders to reduce or eliminate the side loading of cylinders created

as a result of misalignment. It is almost impossible to achieve perfect alignment even through of

a hydraulic cylinder has a direct bearing on its life.

Figure 3.2.12. Various cylinder mountings



Figure 3.2.13 typical mechanical linkages that can be obtained with hydraulic cylinders.



Eye or Clevis Cylinder Mounting

Figure 3.2.14 illustrates the possible combinations of mounting a plane bearing and a spherical

bearing at the cylinder cap and rod eye.

Figure 3.2.14 Eye or clevis cylinder mounting. (Courtesy of Bosch Rexroth AG.)

Trunnion Mounting

The trunnion mounting allows angular movement of the cylinder and a shorter retraction length

(see Fig. 3.2.15).

Figure 3.2.15 Trunnion mounting. (Courtesy of Bosch Rexroth AG.)

Flange Mounting

The flange mounting is preferred for vertical cylinder mounting. When the cylinder is loaded

mainly by thrust force (tension or compression), the mounting bolts at the flange should be

unloaded. The mounting positions shown by Fig. 3.2.16 are recommended for this purpose.



Figure 3.2.16 Flange mounting. (Courtesy of Bosch Rexroth AG.)

Foot Mounting

The foot mounting of hydraulic cylinders is illustrated in Fig. 3.2.17. The mounting bolts should

be protected against shear stress. Therefore, thrust keys are provided to absorb the cylinder

forces as shown.

Figure 3.2.17 Foot mounting. (Courtesy of Bosch Rexroth AG.)

A universal alignment mounting accessory designed to reduce misalignment problems. By using

of these accessory components and a mating clevis at each end of the cylinder, the following

benefits are obtained:

• Freer range of mounting positions

• Reduced cylinder binding and side loading

• Allowance for universal swivel

• Reduced bearing and tube wear



• Elimination of piston blow-by caused by misalignment

3.2.4. Cylinder force, velocity, power and loads due to moving of weights

The out put forces (F) and piston velocity (v) of doubling acting cylinders are not the same for

extension and retraction strokes. This explained as follows (in conjunction with figure 3.2.3):

During the extension stroke, fluid enters the blank end of cylinder through the entire circular area

of the piston (Ap). However during the retraction stroke, fluid enters the rod end through the

smaller annular area between the rod and cylinder bore( )rp AA − , where pA equals the piston

area and rA equals the rod area. This difference in flow path cross-sectional area accounts for the

difference in piston velocities. Since pA is greater than( )rp AA − , the retraction velocity is

greater than the extension velocity for the same input flow rate.

Figure 3.2.18. The effective area of double-acting cylinders is greater for extension stroke than it

is for the retraction stroke



Similarly during the extension stroke, fluid pressure bears on the entire circular area of the

piston. However, during the retraction stroke, fluid pressure bears only on the smaller annular

area between the piston rod and cylinder bore. This difference in area accounts for the difference

in output forces. Since pA is greater than( )rp AA − , the extension force is greater than the

retraction force for the same operating pressure. Equations (3.2.1) through (3.2.4) allow for the

calculation of the output force and velocity the extension and retraction strokes of 100% efficient

double-acting cylinders.

Extension stroke

pext ApF ∗= 3.2.1

p

inext A

Qv = 3.2.2

Retraction stroke

( )rpret AApF −∗= 3.2.3

( )rp

inret AA

Qv

−= 3.2.4

Where AP = Piston area, m2

Ar = Rod-side area, m2

Fext = extension force, N

Fret = retraction force, N

p = Pressure, Pa

v = Piston speed, m/s

Qin = Flow rate, m3/s

The output force required from a hydraulic cylinder and the hydraulic pressure available for this

purpose determine the area and bore of the cylinder. For a required force and given pressure,

essure

ForceareaCylinder

Pr=

p

Fdb =4

2π



p

Fdb π

4=

Where the bore ( bd ) is computed in (in), (mm), the force (F) is in lbf (N), and the pressure (p) is

in lbf/in2 (Pa or bars).

The power developed by a hydraulic cylinder equals the product of its force and velocity during

a given a stroke. Using this relation and equations (3.2.1) and (3.2.2) for the extending stroke and

equations (3.2.3) and (3.2.4) for the retraction stroke, we arrive at the same result:

inQppower ∗= . Thus, we conclude that the power developed equals the product of pressure and

cylinder input volume flow-rate for both the extension and retraction strokes.

The horsepower developed by a hydraulic cylinder for either the extension or retraction stroke

can be found using equation (3.2.5).

( ) ( ) ( ) ( ) ( )1714550

/ psipgpmQlbFsftvkWPower inp ∗

=∗

= 3.2.5

Using metric units, the kW power developed for either the extension or retraction stroke can be

found using equation (3.2.5M).

( ) ( ) ( ) ( ) ( )kPapsmQkNFsmvkWPower inp ∗=∗= // 3 3.2.5M

Q = Flow rate, m3/s

QL = Internal leakage flow rate, m3/s

Note that the equating of the input hydraulic power and out put mechanical power in equations

(3.2.5) and (3.2.5M) assumes a 100% efficient hydraulic cylinder.

3.2.5. Cylinder loadings through mechanical linkages

In many applications, the load force that a hydraulic cylinder must overcome does not act along

the axis of the hydraulic cylinder. Because of this, the load force and the hydraulic cylinder force

are not equal. Therefore, it is necessary to analyze how to determine hydraulic cylinder forces



required to driven on-axial loads. Accordingly, we can classify the mechanical linkages into

first–class, second-class and third-class lever systems.

First-class Lever System

First-class lever system is characterized by the lever fixed-hinge pin being located between the

cylinder and the load rod pin. To determine the cylinder force Fcyl required to drive a load force

Fload, we equate moments about the fixed hinge, which is the pivot point of the lever. The

cylinder force attempts to rotate the lever counterclockwise about the pivot and this creates a

counterclockwise moment. Similarly, the load force creates a clockwise moment about the pivot.

At equilibrium, these two moments are equal in magnitude:

Therefore, counterclockwise moment = clockwise moment

( ) ( )θθ coscos 21 LFLF loadcyl =

Or

loadcyl FL

LF

1

2= 3.2.6

It should be noted that the cylinder is clevis-mounted to allow the rod

Figure 3.2.19. Use of first-class lever to drive a load



Since the cylinder is clevis-mounted, it allows the rod-pinned end to travel along a circular path

as it rotates about the fixed-hinge pin. If the centerline of the hydraulic cylinder becomes offset

by an angle ϕ from the vertical, we have,

( ) ( )θφθ coscoscos 21 LFLF loadcyl =∗

Or

loadcyl FL

LF

φcos1

2= 3.2.7

Examination of equation (3.2.6) shows that L1 (distance from cylinder rod to hinge pin) is greater

than L2, the cylinder force is less than the load force. Of course, this results in a load stroke that

is less than the cylinder stroke, as required by the conservation of energy law. When ϕ is 100 or

less, the value of cos ϕ is very nearly unity. (cos 00= 1, and cos 100= 0.985)and thus equations

(3.2.6) can be used instead of equation (3.2.7).

A length of moment arm for either the cylinder force or load force is the perpendicular distance

from the hinge pin to the line of action of the force. Thus, for the development of equation

(3.2.6), the moment arms are L1cos θ and L2 cos θ rather than simply L1 and L2. Similarly, for

the development of equation (3.2.7), the moment arm for the cylinder force L1cos θ x cos ϕ,

rather than simply L1 cosθ, which is based on the assumption that cos ϕ = 00.

Second-class Lever System

Figure 3.2.20 shows a second-lever system, which is characterized by the load rod pin being

located between the fixed-hinge pin and cylinder rod pin of the lever. The analysis is

accomplished by equating moments about the fixed-hinge pin, as follows:

( ) ( )θθφ coscoscos 221 LFLLF loadcyl =+

Or

3.2.8 ( ) loadcyl FLL

LF

φcos21

2

+=



Figure 3.2.20. Use of second-class lever to drive a load

Comparing equation (3.27) to equation (3.2.8) shows that a smaller cylinder force is required to

drive a given load force for a given lever length if a second-class lever is used instead of a first-

class lever. Thus, using a second-class lever rather than a first-class lever reduces the required

cylinder piston area for a given applications. Of course, using a second-class lever also results in

a smaller load stroke for a given cylinder stroke.

Third-class lever system

As shown in figure 3.2.21, for a third-class lever system the cylinder rod pin lies between the

load rod pin and fixed-hinge pin of the lever. Equating the moments about the fixed-hinge pin

yields

( ) ( ) θθφ coscoscos 212 LLFLF loadcyl +=

Or

loadcyl FL

LLF

φcos2

21 +=

3.2.9

An examination of equation (3.2.9) reveals that for a third-class lever, the cylinder force is

greater than the load force. The reason for using a third-class lever system would be produce a



load stroke that is greater than the cylinder stroke, at the expense of requiring a larger cylinder

diameter.

Figure 3.2.21. Use of third-class lever to drive a load

3.2.6. Hydraulic cylinder cushions and shock absorber

Double acting cylinders are of ten provided cylinder cushions at the two ends to slow down the

piston near the ends of stroke. This prevents excessive impact when the piston is stopped by the

end caps, as illustrated in figure 3.2.22. As shown, deceleration starts when the tapered plunger

enters the opening in the cap. This restricts the exhaust flow from the barrel to the port.

During the last small portion of the stroke, the oil must exhaust through an adjustable opening.

The cushion design also in corporate a check valve to allow free flow to the piston during

direction reversal. The maximum pressure developed by cushions at the ends of a cylinder must

be considered since excessive pressure build up would rupture the cylinder.



Figure 3.2.22. Operation of cylinder cushions

When cylinders reach the end of their stroke, the pressure rises quickly, creating a shock wave in

the hydraulic circuit. Cushioning is done to reduce this stock. The concept, shown in Fig. 3.2.23,

is quite simple. First, we consider the case in which the cylinder is retracting. The spear closes

off the large opening where the fluid is exiting the cap end of the cylinder. Fluid must now flow

out the small opening past a needle valve. This valve adjusts the orifice and sets the back

pressure that develops in the cap end. The resultant force slows the piston so that it “coasts” to a

stop. The resultant pressure shock in the main circuit is significantly reduced. The same

technique is used to cushion the cylinder when it is extending. In this case, a sleeve is mounted

on the rod to close the main opening so that flow goes through the orifice.



Figure 3.2.23. Schematic showing a technique for cushioning a hydraulic cylinder.

To understand cushioning, it is appropriate to review a basic principle of fluid power. Always

ask the question, what is happening at the relief valve? We will consider the case where a fixed

displacement pump is supplying the flow to extend the cylinder. When the spear closes the large

opening, the fluid must flow past the needle valve. The resultant pressure drop increases the back

pressure on the rod end. The pump must build a higher pressure on the cap end. This cap end

pressure must build to the point where the relief valve cracks open before the cylinder will slow.

Remember, a fixed displacement pump puts out a given volume of oil for each revolution.

Neglecting pump leakage, this oil either goes to the cylinder or through the relief valve. The

adjustments made to the needle valve on the cylinder interact with the characteristics of the relief

valve (and to a lesser degree with other components in the circuit) to produce a given

deceleration rate.

A number of techniques have been developed to cushion cylinders. Large cylinders are

cushioned with some type of stepped procedure that decelerates the piston in increments.

Manufacturer’s literature can be referenced for the back-pressure curves generated by these



techniques. The ANSI symbol for a cylinder cushioned on both ends is given in Fig. 3.2.24. The

arrow through the cushion block indicates that the cushioning is adjustable.

Figure 3.2.2 ANSI symbols for single and double rod cushioned cylinder

Cushioning devices are provided in the ends of hydraulic cylinders or as separate units when

loads must be decelerated. Hydraulic shocks to the cylinder and system are reduced by slowing

down the piston just before it contacts the cylinder end caps. The sleeve or spear-type cushioning

device is often used as the decelerating mechanism. The energy absorbed by decelerating a body

in motion is converted to heat and dissipated to the environment through the fluid and cylinder

body.

Refer to Fig. 3.2.25 as the piston completes its extension stroke, the slightly tapered end

cushioning sleeve enters the rod end cap blocking the normal flow of fluid. The flow of fluid is

then rerouted through the bypass port and needle valve at a controlled rate, decelerating the

piston. Cushioning at the head end of the cylinder during the retraction stroke is accomplished in

the same manner by the action of the tapered cushioning spear plunger as it enters the end cap.

This blocks the normal flow of fluid and causes fluid to be rerouted through the bypass port and

metering valve to decelerate the piston. The check valve, shown only on the blank end of the



cylinder, is used to direct fluid at system pressure to the full area of the piston during

acceleration.

Figure 3.2.25. Cylinder cushioning devices (courtesy of Carter Controls. Inc.).

Figure 3.2.26 shows a hydraulic cylinder with an adjustable cushioning element. The piston (1)

is fitted with a conical end-position cushioning bush (spear) (2). During the cylinder retraction

stroke, the returned oil flows freely to the return line. When approaching the end position, the

spear enters a cylindrical cave in the cylinder cap (4).

Figure 3.2.26 (a) Cushioning of hydraulic cylinders. (Courtesy of Bosch Rexroth AG.) (b)

Functional scheme of the cushioning element.



The oil returned from the piston chamber (5) is forced to flow through the passage (6) to the

throttle valve (7), and through the radial clearance between the spear and cylindrical cave. The

pressure in the piston chamber increases and acts to decelerate the moving piston. The check

valve (3) permits by-pass of the throttle valve at the start of the motion in the opposite direction.

When extending the cylinder, the oil flows to the piston chamber through the check valve. In this

way, the whole piston area is acted on by the pressurized oil.

The piston may be equipped with a cushioning element at one side or at both sides. The throttle

valve is either of fixed or of adjustable area. The illustrated construction has a throttle valve area

adjustable by a screw (see Fig. 3.2.26a). The kinetic energy of the moving parts should not

exceed the capacity of the cushioning element, defined as the work done during cushioning

period. The cushioning should produce a controlled deceleration of the cylinder, near one or both

end positions. This is done by creating a decelerating pressure force. In doing this, the pressure

must not exceed its limiting value. The kinetic energy of the moving parts is converted into heat

by throttling the out-going fluid in the cushioning zone. For an ideal operation, the piston will be

fully stopped at the end of the cushioning stroke, s. If the piston will be fully stopped at the end

of the cushioning stroke, the required deceleration, a, is

s

va

2

=

When the cylinder is installed horizontally, the decelerating (cushioning) force can be calculated,

considering the equilibrium of forces acting on the piston (see Fig. 3.2.26b), as follows:

pdd pAmaAp +=

Normally, the damping pressure may not exceed the nominal pressure of the cylinder. An

average value of the damping pressure, Pd, is given as follows:

+= p

dd pA

s

mv

Ap

2

1 2

Where =a Deceleration, m/s2

=dA Piston area subjected to pressuredp , m2

=pA Piston area subjected to pressure p, m2



=m Moving mass, kg

=p Driving pressure, Pa

=dp Mean pressure in the cushioning volume, Pa

=s Damping length, m

=v Piston velocity at the beginning of the cushioning stroke, m/s

For vertical mounting of the cylinder, the pressure generated by the weight of the moving parts

must be added or subtracted depending on the direction of motion. The cylinder friction is

ignored in these calculations. If the calculations give an unacceptably high mean damping

pressure, the damping length must be increased or the speed must be reduced.

Hydraulic shock absorber

External hydraulic energy absorbing devices also are employed many times to reduce shock

loading of machinery that would result in failure. Carriage stops on turntables, doors, transfer

machines, turnover s, and small crane s requiring controlled deceleration use the se devices (Fig.

3.2.27). In operation, fluid is metered through adjustable orifice s for a wide range of impact load

s and velocities. In Fig. 3.2.28, for example, liquid is metered through the orifice plate valve

along the tapered metering pin as the piston is forced into the cylinder. Adjustment is

accomplished by removing the protective plug and rotating the adjustment screw with a

hexagonal socket key for more or less damping. The hardened and chrome plated tubular piston

resists buckling and side load s under impact. The metered liquid is stored in the spring-loaded

separator chamber which also repositions the piston.



Figure 3.2.27. Typical deceleration application using shock absorbers.



Figure 3.2.28. Typical hydraulic shock absorber with full-range adjustability (courtesy of Enidine Corporation).

3.2 Hydraulic Cylinder

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