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FLUID POWER
FUNDAMENTALS
DECEMBER 2019
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FLUID POWER FUNDAMENTALS
DECEMBER 2019
CONTENTS
02
05
08
12
15
18
Basics of
Pneumatic Logic
Electrohydraulic
Valves
Electrohydraulic
Motion Control
Hydraulic Fittings
and Flanges
Hydraulic System
Flushing Procedures
Hydraulic
Power Units
22
25
28
31
34
37
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Reducing Noise from
Hydraulic Systems
Pneumatic Quick-Acting
Couplings
Seals for
Hydraulic Cylinders
Rotary Unions
& Swivels
Water
Hydraulics
More Resources from
Hydraulics & Pneumatics
FLUID POWER FUNDAMENTALS
Basics of Pneumatic Logic
In this age of digital electronics, air logic can still
provide a simple, effective, and reliable means of
machine control.
E
lectrical and electronic devices
control most fluid power circuits.
Relay logic circuits, programmable controllers, or computers are common control methods. But another way
to control pneumatic systems is with
air logic. Air logic controls can perform
any function normally handled by relays, pressure or vacuum switches, time
delays, limit switches, or counters. The
circuitry is similar, but compressed air is
the control medium instead of electrical
current.
Environments with high levels of
moisture or dust are excellent places for
air logic controls. No danger from explosion or electrical shock is presented
by the air-logic system. Water can splash
on the controls without affecting their
operation. If there is an external explosion, the control media—clean compressed air—cannot ignite.
Some designers of pneumatic equipment prefer to use air-logic controls because only one utility service is needed
to operate it. No electricity is necessary.
This arrangement can be a selling point
in user facilities where electrical and
mechanical maintenance must be handled by different labor trades. Because
there are no electrical devices involved,
one craft works on both the air-logic circuit and machine parts.
Two basic roadblocks to using airlogic control are a lack of understanding of how the components work and
an inability to read the special schematic drawings. If an air-controlled
machine fails, few people know how
troubleshoot it. Also, air-logic circuits with long conductor lines—more
than 10 to 15 ft—may cycle noticeably
slower than similar electrical controls.
They cannot fill with and exhaust air as
quickly as electrical signals travel. Finally, air quality must be above average
to ensure consistent performance and
long component life.
The components used for air logic
controls are basically miniaturized
3- and 4-way air valves. The actions
of these valves turn functions on or
off, just as relays or switches do, then
exhaust the spent signal. The symbols
that were developed for air logic are
similar to electronic symbols. In fact,
some manufacturers use modified electrical symbols and ladder diagrams to
show circuitry.
Figures 1 and 2 show two types of
AND elements. Some manufacturers
supply passive and active types of elements but designate the passive AND
simply AND, whereas the active AND is
designated YES. The difference in the elements is that the passive AND element
uses the lower of the two inlet ports as
an output. In contrast, the active AND
element has two inputs to achieve an
output, but the designer has the choice
of which input goes to the output. Using
this feature can amplify a weak signal.
The weak signal pilots the valve open
while the through signal comes from a
full-pressure supply. The YES is an active element.
Basic Logic Elements
Following are text explanations of the
functions of basic logic components,
with illustrations using standard ANSI
logic symbols and ISO graphic symbols of a comparable directional control
valve.
An AND element must receive two
input signals simultaneously before it
passes an output signal. This ensures
that two upstream functions are complete before there is a command to a
downstream function. In other words,
inputs A and B must both be present
before an output action occurs. When
using more than two inputs, AND elements are connected in series. The
first AND receives signals 1 and 2, and
the output of this element interfaces
with an input to the second AND.
The other input of the second AND
receives a third signal, making three
inputs necessary before an output action can occur.
1. Passive AND element.
2. Active AND element (sometimes designated YES).
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FLUID POWER FUNDAMENTALS
A signal at either input port of an
OR element (Fig. 3) produces an output signal. Another way of saying this is
that either signal A or B will produce an
output. A shuttle valve serves the same
purpose as an OR element. Pilot signals from two different sources can pass
through the OR to start the next function. 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. OR elements can be stacked to accommodate
more than two inputs. Use an extra OR
for each input after the first two signals.
A NOT logic element (Fig. 4) is a normally open 3-way valve. An input signal
to the Supply port will pass through the
valve until there is a pilot signal at port
A. Pressurizing port A blocks supply
and exhausts the output signal to atmosphere through port B. As long as there
is pilot pressure on the A port, NOT elements will block a signal or supply. NOT
elements always return to a normally
open condition when the pilot signal is
removed.
3. OR element.
4. NOT element.
A NOT element can simulate a limit
switch to indicate that a cylinder is at the
end of stroke. Pressure from the cylinder
port goes to port A of the NOT, holding it closed. As the cylinder extends
toward the work, pressure is maintained
because of the meter-out flow control.
When the cylinder contacts the work,
the signal at port A exhausts, and the
NOT opens to pass an output signal to
start the next operation.
Note that the cylinder can stop at any
position and the NOT’s output signal
will indicate that motion has stopped—
whether the cylinder stopped where it
should be or is stalled by some other
means. Because of this, take care when
using a NOT to replace a limit switch.
CAUTION!
ANY PRESSURE-CONTROL VALVE
only responds to a pressure buildup.
When a positive location must be identified, use limit valves.
On the other hand, this phenomenon
can be advantageous when clamping
different sized parts. Use a NOT element
for applications where different work
locations stop the cylinder.
Most manufacturers supply a different pilot ratio for a NOT element used
as a limit switch. The valve function
is the same but the pressure that shifts
it is much lower. Some manufacturers
build special NOT elements that mount
directly in a cylinder port. A portmounted meter-out flow control used
in conjunction with this special NOT
makes a compact installation.
A FLIP FLOP (Fig. 5) is a double-piloted 5-way valve that directs supply air
to either outlet port in response to signals at pilot ports S or R. (Supply air can
be system pressure or a signal from another logic element.) The main purpose
of a FLIP FLOP is to exhaust the first pilot signal to a directional control valve.
Then a second signal to the valve’s opposite pilot port can shift it back. FLIP
FLOPs are sometimes called MEMORY
elements because they stay in the last
shifted position even with no air supply. Whether the signal is maintained or
drops out, output from the FLIP FLOP
stays the same.
5. FLIP-FLOP element.
The S and R signal designations stand
for SET and RESET. The SET signal
shifts the FLIP FLOP for a function, and
whether S is maintained or not, the element stays shifted. The RESET signal
returns the FLIP FLOP to its original
position until the next cycle.
Another use for a FLIP FLOP is to
set up a new cycle allowing the operator
to momentarily push the start buttons.
This same FLIP FLOP can be installed
to block unwanted signals and set up the
circuit for cycle completion as required.
Figure 6 shows another valve actually called a MEMORY element. A
MEMORY element is a normally closed
3-way valve with an integral shuttle
valve. The MEMORY’s output air hold
the shuttle valve shifted once it receives
a SET signal. A momentary SET signal
gives continuous pilot output. A RESET
6. MEMORY element.
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FLUID POWER FUNDAMENTALS
signal shifts the MEMORY element to
normally closed and exhausts output
air. Also, turning SUPPLY pressure off
returns a MEMORY element to the start
position.
In air logic control there are three
different types of time delays. Fixed
or adjustable time delays are common
in both normally closed and normally
open configurations. Some time delays
use an orifice and accumulator chamber for delays as long as one minute.
Some manufacturers use air-actuated
diaphragms and orifices that eliminate
inaccuracies due to system pressure
fluctuation.
Figure 7 shows the symbol for a
ONE-SHOT timer, sometime called
an IMPULSE TIMER. A ONE-SHOT
timer takes a signal and passes it on to
the circuit. At the same time the input
signal is going through an orifice to an
accumulator chamber. The setting of
the orifice and size of the accumulator
gives a certain time delay before the
normally open 3-way valve closes. After a ONE SHOT times out and closes,
it remains closed as long as it has an
input signal. Figure 7 shows an adjustable time delay before losing the output. Omitting the sloping arrow in the
symbols makes it a preset time delay.
Times range from a half second to two
or more seconds on valves with preset
time delays.
7. ONE SHOT element.
Many circuits use ONE SHOTs to
eliminate opposing signals. When a
valve receives a signal to extend a cylinder, it resists a return pilot signal to
itself until loss of the first pilot. Using a
ONE SHOT element drops the extend
signal shortly after initiation. However,
when the short duration signal meets
a hard-to-shift valve, the time may not
always be long enough to move the
valve spool. When the valve does not
have time to shift, the cycle stalls. For
best results, use a FLIP FLOP to drop
an unwanted signal after it performs
its task.
8. TIME ON time delay element.
Figure 8 shows an adjustable, normally closed TIME-ON time delay
symbol. A TIME-ON delay passes a
signal through the element after timing stops.
A TIME-ON delay is a preset fixed
timer without the sloping arrow. Most
anti-tie down circuits use a fixed time
delay. This forces the operator to actuate
both palm buttons concurrently.
Assume an input signal is applied
to blocked port A of the 3-way directional valve in Fig. 8. The same signal
also passes through the meter-in flow
control to fill the accumulator. After
the accumulator is filled, pilot pressure
shifts the 3-way valve, allowing air to
pass on to the next operation. As long as
the input signal stays on, the time delay
stays open.
Some brands of TIME-ON delays use
shop air to the normally closed port A
of the 3-way valve while the signal to
the timing section comes from another
logic element or limit valve. This allows
a strong passing signal to travel long distances or to quickly shift several other
logic elements.
With an integral accumulator chamber, the time-delay length is usually
between a minute and a minute-anda-half. With additional external accumulators, time delays up to five minutes
are possible. The repeatability of long
time delays using accumulators is poor.
However, diaphragm-type timers often
can produce 3-minute delays with acceptable repeatability.
With a normally open 3-way valve
in place of the normally closed 3-way
valve, the function becomes a TIMEOFF delay timer (Fig. 9). An input to
the SUPPLY port passes continuously
to output until a set time after a pilot
signal is received at port A. When A
receives a signal, the time delay starts
and continues until the accumulator
fills and shifts the normally open 3-way
valve to block the signal at SUPPLY
and exhaust the downstream system.
Adjustable and preset, non-adjustable
TIME OFF delays are available. TIMEON and TIME-OFF delays often are
identical in appearance. The part number may be the only way to tell these
units apart.
9. TIME OFF time delay element.
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FLUID POWER FUNDAMENTALS
Electrohydraulic Valves
Servo and servo proportional valves control
pressure or flow—and ultimately, force or velocity.
Unlike simple directional valves, they can maintain
any position between fully open in one direction or
the other.
H
igh-performance valves are
usually classified as either
servo or proportional, a distinction that gives an indication of expected performance. Unfortunately,
this classification tends to generalize
and blur the true differences between
various valve styles. Selection depends
on the application, and each valve has
merit when it comes to controlling pressure or flow.
Traditionally, the term servovalve
describes valves that use closed-loop
control. They monitor and feed back
the main-stage spool position to a pilot
stage or driver either mechanically or
electronically. Proportional valves, on
the other hand, move the main-stage
spool in direct proportion to a com-
mand signal, but they usually do not
have any means of automatic error correction (feedback) within the valve.
Confusion often arises when a valve’s
construction resembles a proportional
valve, but the presence of a spool position feedback sensor (usually an LVDT)
boosts its performance to that rivalling a
servovalve. This reinforces the concept
that designers and suppliers should use
common terminology and focus on the
performance requirements of the particular application at hand.
Typically, proportional valves use one
or two proportional solenoids to move
the spool by driving it against a set of
balanced springs. The resultant spool
displacement is proportional to the current driving the solenoids. The springs
also center the main stage spool. Repeatability of the main-stage spool position
is a function of the springs’ symmetry
and ability of the design to minimize
nonlinear effects of spring hysteresis,
friction, and machining tolerance variations.
Servovalves
The term servovalve traditionally
leads engineers to think of mechanical
feedback valves, where a spring element
(feedback wire) connects a torque motor
to the main-stage spool. Spool displacement causes the wire to impart a torque
onto the pilot-stage motor. The spool
will hold position when torque from
the feedback wire’s deflection equals the
torque from an electromagnetic field
1. First-stage configurations for nozzle flapper and jet-pipe valves.
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FLUID POWER FUNDAMENTALS
2. The linear force motor often is used to drive the spool of high-performance valves directly. An alternative is to use one or two proportional solenoids to drive the spool.
induced by the current through the motor coil. These two-stage valves contain
a pilot stage or torque motor, and a main
or second stage. Sometimes the main
stage is referred to as the power stage.
These valves can be separated primarily
into two types: nozzle flapper and jet
pipe (Fig. 1).
The electromagnetic circuit of a
nozzle flapper or jet-pipe torque motor
is essentially the same. The differences
between the two lies in the hydraulic
bridge design. A hydraulic bridge controls the pilot flow which, in turn, controls the main-stage spool movement.
In a nozzle flapper, the torque produced
on the armature by the magnetic field
moves the flapper toward either nozzle
depending on command-signal polarity.
Flapper displacement induces a pressure imbalance on the spool ends which
moves the spool. In a jet pipe, the armature movement deflects the jet pipe and
asymmetrically imparts fluid between
the spool ends through the jet receiver.
This pressure imbalance remains until
the feedback wire returns the jet pipe or
flapper to neutral.
Historically, jet pipe and nozzleflapper servovalves have competed for
similar applications that require high
dynamics. Typically, better first-stage
dynamics gives the nozzle flapper bet-
ter overall response, whereas improved
pressure recovery of the jet/receiver
bridge design gives the jet-pipe motors
higher spool driving forces (chip-shearing capability). Both valves require low
command currents and therefore offer
a large mechanical advantage. Motor
current for these style valves is typically
less than 50.0 mA. Note that these servovalves are also proportional valves,
because spool displacement and flow
are directly proportional to the input
command.
Direct-Driven Valves
Direct-driven valves, unlike hydraulically piloted two-stage valves, displace
the spool by physically linking it to the
motor armature. These valves usually
come in two basic varieties: those driven
by linear force motors (LFM) and those
actuated by proportional solenoids.
Within these two general classifications, the valves can be separated into
proportional and servo-proportional.
The distinction is based on the use of
a position transducer to provide spool
position feedback.
Servo proportional valves must incorporate closed-loop spool position
feedback to increase repeatability and
accuracy necessary for high-control applications. Typically, servo proportional,
direct-driven valves have an overall
lower dynamic response than hydraulically piloted two-stage valves with the
same flow characteristics. This is usually
due to the large armature mass of the
LFM or solenoid and the large time constant associated with the coil, which is a
function of the induction and resistance
of the coil.
Unlike hydraulically piloted servovalves, direct-driven valve performance does not vary with changes in
supply pressure. This makes them ideal
for applications where pilot flow for
first- stage operation is not available.
Direct-driven valves also tend to be
viscosity insensitive devices, whereas
nozzle-flapper and jet-pipe valves work
best with oil viscosity below 6,000 SUS.
However, most direct-driven valves
cannot generate the high spool driving forces of their hydraulically piloted
counterparts.
Like the torque motor used in the
nozzle flapper/jet pipe servos, the LFM
allows for bidirectional movement
by adding permanent magnets to the
design and therefore making the armature motion sensitive to command
polarity. In the outstroke, the LFM
must overcome spring force plus external flow and friction forces. During the backstroke to center position,
however, the spring provides additional
spool-driving force which makes the
valve less contamination-sensitive.
Magnetic-field forces are balanced by
a bidirectional spring that lets the spool
remain centered without expending
any power.
Unlike the LFM, the proportional solenoid is a unidirectional device. Two
solenoids oppose each other to achieve
a centered, no power, fail-safe position.
When a single solenoid is used, holding
the spool at mid-stroke requires a continuous current to balance the load generated by the return spring. This makes
the design less energy-efficient than its
LFM or a dual-solenoids counterpart.
During a power loss, the LFM and dual
proportional solenoid designs fail to a
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FLUID POWER FUNDAMENTALS
4. The 4-way spool valve has
four individual lands that vary in
unison as the spool shifts—two
lands open while the other two
close. When drawn in schematic
form, it is clear that the four
lands constitute a bridge circuit,
and spool movement unbalances the bridge one way or the
other to cause a reversal in load
flow.
T
otal
force
requirements
must include
all static and
dynamic forces
acting on the
system.
3. When the flapper nozzle pilot section (a)
is drawn in schematic form, (b), it is obvious that a bridge circuit exists. By moving
the flapper, restrictions Ra and Rb change
in opposite directions. This unbalances
the bridge and causes the spool to move
against its centering springs.
neutral position and block flow to the
load—that is the piston. When a single
solenoid design loses power, the spool
must move through an open position
that tends to cause uncontrolled load
movements.
Multistage Valves
All the aforementioned designs can
be used to create a multistage hydraulic valve. The approach for each design
is specific to the application requirements. Usually, most designs do not exceed three stages. Mounting a nozzle
flapper, jet pipe, or direct-driven valve
onto a larger main stage satisfies most
requirements for dynamics and flow.
Sometimes, the jet-pipe valve is used
in a multistage configuration where the
mechanical feedback of a traditional
jet pipe is re-placed with electronic
feedback. This servo-jet style has pilot
characteristics of a typical jet pipe. Depending on the required control, many
multistage valves close a position loop
about the main stage using a linear variable differential transducer. This device
monitors the spool position. In case of
hydraulic power loss, springs on opposite sides of the main stage spool return
it to a neutral position.
Hydraulic System Design
To choose the proper hydraulic valve
for a specific application, designers must
consider specific application and system
configurations. Supply pressure, fluid
type, system force requirements, valve
dynamic response, and load resonant
frequency are examples of the various
factors affecting system operation.
Hydraulically piloted valves are sensitive to supply pressure disturbances,
whereas direct-driven valves are unaffected by supply pressure variation.
Fluid type is important when consider-
ing seal compatibility and viscosity effects on performance over the system’s
operating temperature range.
Total force requirements must include all static and dynamic forces acting on the system. Load forces can aid
or resist, depending on load orientation
and direction. Forces required to overcome inertia can be large in high-speed
applications and are critical to valve
sizing.
The load resonance frequency is a
function of the overall travel stiffness,
which is the combination of the hydraulic and structural stiffness. For optimum dynamic performance, a valve’s
90-deg. phase point should exceed the
load resonant frequency by a factor of
three or more.
The valve’s dynamic response is defined as the frequency where phase lag
between input current and output flow
is 90-deg. This 90-deg phase lag point
varies with input signal amplitude, supply pressure, and fluid temperature so
comparisons must use consistent conditions.
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FLUID POWER FUNDAMENTALS
Electrohydraulic
Motion Control
where the cylinder and load will eventu- cycle rate is 12 cycles per second. Furally stop. There is no means to back up ther, if we are given that cylinder stroke
should there be overtravel. Therefore, must be, say, 4 in., and that the cylinder
almost all the random factors that affect must remain extended for 0.5 sec., then
the stopping point of the discrete direc- we know that the actuator must fully extional system will be visited upon this, tend and retract within 1.5 sec. We can
add some reasonable acceleration times,
the proportional system.
account that
the cylinder
A system
that is proportional
for its plus take into
In
a departure
from conventional
wisdom,
motion
naturally want
to retract as fast
shock
control,
butbe
alsoimplemented
monitors and does
control
can
tonot
increase
production
as
it
extends.
(Yes,
this
statement
is corcontrols
position
on
a
full-time
basis,
rate and product quality and consistency—all at the
is called the electrohydraulic positional rect for the valve-controlled cylinder.)
same time.
servomechanism. It possesses all the de- To compensate for this, we can rob a few
sirable characteristics for effective mo- milliseconds from the extension time
tion control, but because it is a bona-fide and give it to the retraction time. This
otion control
the simul- allows us to reach conclusions about
feedback control
system, is
it introduces
taneous
control
of
acceleraits own set of challenges. However,
when the maximum speed needed in order to
tion, velocity,
and position
properly designed
and tuned,
this sys- achieve the required productivity rate.
for performing
a useful
task.solution
The critical
This example leads us to the conclutem
provides the
ultimate
to
concept
in
this
statement
is
that
all
three
the most demanding motion control ap- sion that the peak extend speed will have
variables—acceleration,
velocity, and to be about 10 in./sec. Once we are given
plication
challenges.
position—are
under
control. A review
The positional
servomechanism
is the maximum thrust force, we can deof some basic
electrohydraulic
designs
normally
considered
a positioning
sys- sign for a given supply pressure, which
will help
pave the
way the
to understandtem.
However,
when
concept of makes it possible to completely size the
ing the concepts
of motion control.
Two
profiling
is superimposed,
the result
is critical components in the system. Our
conventional
hydraulic
circuits
willmabe example introduces the idea of a momotion
control—total
control
of the
used toToillustrate
the point:
discrete,
or tion profile, which is a means of definchine.
be successful,
motion
control
so-called
directional
ing the speed, acceleration, and position
must
havebang-bang
three things
designed control,
into it:
and
control.
of the cylinder and load at any instant
• open-loop
a commandproportional
motion profile,
These
two
circuits
will
lead
us
to
an
• sufficient power to manage the within each cycle and throughout each
ultimate
the electrohydraulic
and every cycle for the entire life of the
load solution:
at the required
speeds, and
positional
servomechanism.
It will be machine.
• sufficient
closed-loop bandwidth
shown
that when
the electrohydraulic
The relationship between accelerato meet
accuracy
and stability reservoquirements.
is properly designed — and the tion, velocity, and position —which are
control
system is suitedprofile
to the task—the
The motion-control
is crucial drawn against a common time axis—is
result
is a true
motion
control
to
the motion
control
system,
andsystem
repre- represented in Fig. 5. Newton revealed
meeting
the
divergent
goals
of
increassents the most significant contribution of through the calculus that:
ing digital
machine
productivity
and product
• velocity is the integral of accelerathe
computer
to the motion-conquality,
whileThe
reducing
production
trol
process.
profile overall
is the means
by
valve
and a check
valve toof
prevent
tion
costs.
•
position
is the integral
velocitythe
which acceleration, velocity, and position
from having
cross over
designersProfiles
should pump
• velocity
is the to
derivative
ofcenter
posiareHydraulic
controlledsystem
simultaneously.
absorb any flow that might
realize
the design
that in order
tion,to
and
are
veryearly
easilyin
generated
by aprocess
digital combe forced
back
into the pump.
a power
unit providing
pres- otherwise
• acceleration
is the
derivative
of veputer
(motion
controller constant
card or circuit
Figure
2
also
shows
the
pressuresure
to
the
control
valve
is
best
suited
to
locity.
board), and are formulated at design time
characteristics
a conventional
thedetermine
motion control
task.
The that
ISO must
sym- flow
Even
if you never of
studied
calculus,
to
the peak
speeds
pressure-compensated
low
bolattained
for a constant
are mathematicalpump.
facts At
of our
be
by thepressure
actuatorsource,
basedalong
on a these
pressure,
it
behaves
as
a
fixed-displacewith
a
plot
of
its
behavior,
is
shown
in
physical world. What they mean to
stated productivity rate.
ment
pump,control
delivering
nearly
constant
FigIn1.determining
A real-world
constant-pressure
motion
system
designer
is
these
peak speeds, we the
flow,specifying
neglecting any
internal
source
is shown
in Fig. 2. that
one leakage.
of these When
three
must
first
collect schematically
data on the productivpressure reaches
value, P K ,
It normally
consists
of a pressure-comfor all the
timeknee
automatically
ity
needs of the
application.
For example, parameters
pensatedthe
pump
augmented
by large
ac- specifies
the pump’s
compensator
thepressure
other two
as well. Thisbeis
suppose
application
requires
stampcumulators
to accommodate
sudden
comes active,
andundeniable
the displacement
auof their
relationing
out 30 packets
of paper the
plates
per because
flow demands
of a servo
proportional
tomatically
withare
rising
presto eachdecreases
other—they
all inteminute.
Immediately,
weorknow
that the ship
M
grals or derivatives of each other.
This relationship makes it possible
to achieve the simultaneous control of
acceleration, velocity, and position, even
though the final system closes the loop
with position feedback only. It works
like this: The position profile, the plot
at the bottom of Fig. 5, is the actual
command to the motion-control servomechanism. The slope of the position profile represents the velocity of the
output at each and every instant in time.
The slope of the velocity profile is the
acceleration profile, which is actually
1. ISO symbol
of a position
constantcarried in the curvature
of the
pressure
source,
left,
and
profile. So the designer controls accelrepresentation
of how an of
eration by controlling
the curvature
ideal constant-pressure
the position command
profile.
sourceiswould
right.
The final result
this: operate,
For merely
pressure
wouldprohaving fed the Ideally,
position
command
remain
constant
regardless
file to the motion
control
servo
axis,
of flow
demand.
In implicit
reality,
all three dynamic
variables
are
fluctuates
with flow
in that profile. pressure
Furthermore,
designing
demand
and performance
of
a servo axis that
faithfully
follows the
components.
command profile,
controls all three parameters. This is the sum and substance
2. A real-world
constanton modern motion
control philosophy.
pressure
sourcemethodolconsists
At design time,
the design
of athat
pressure-compensated
ogy recommends
the entire motion
accumulator,
and
control profile,pump,
acceleration,
velocity,
check
valve. This based
is not an
and position, be
synthesized
on
idealother
constant-pressure
productivity and
motion needs.
source: pressure deviates
In contrast, at commissioning
time the
or below
a meanconpresprogrammer ofabove
a dedicated
motion
sureonly
based
on the frequency
troller card may
specify
that the
of the
compo-x to
motion must goresponse
from, say,
position
position y with nents.
maximum speed of v and
maximum acceleration of a. The digisure.
Thus,controller
at pressures
abovethese
P K, the
tal
motion
interprets
as
pressure-compensated
pumpthe
acts
very
instructions,
then generates
actual
nearly as
a constant-pressure
It
profile
commands
on the fly.source.
Consecertainlythe
is safe
to deadhead
pump
quently,
programmer
maythis
never
see
withentire
closed-center
valves without danthe
profile itself.
ger of overpressurization.
THIS INFORMATION was provided by
Discrete
ControlPE, an electrohydraulic
Jack L. Johnson,
In a conventional
electrohydraulic
specialist,
consultant, former
director of
control
(Fig. 3),ata the
simple,
solethe Fluidsystem
Power Institute
Milwaukee
noid-operated
directional
control
valve
School of Engineering, and contributing
can
operate
in only&three
discreteHe
states:
editor
to Hydraulics
Pneumatics.
can
centered
(off
shiftedattojack@idaseng.
the right to
be reached
via), e-mail
route
com. flow in one direction, or shifted
☞ LEARN MORE @ hydraulicspneumatics.com | 8
FLUID POWER FUNDAMENTALS
where the cylinder and load will eventually stop. There is no means to back up
should there be overtravel. Therefore,
almost all the random factors that affect
the stopping point of the discrete directional system will be visited upon this,
the proportional system.
A system that is proportional for its
shock control, but also monitors and
controls position on a full-time basis,
is called the electrohydraulic positional
servomechanism. It possesses all the deto the left
to route flowforineffective
the opposite
sirable
characteristics
modirection.
tion
control, but because it is a bona-fide
To understand
problems,
imagine
feedback
control the
system,
it introduces
that
limit
switch
LS1
in
Fig.
3
is
its own set of challenges. However,closed,
when
the
right-hand
solenoid
is energized,
properly
designed
and tuned,
this sysand
cylinder
moving solution
the load to
tem the
provides
theisultimate
the most
right demanding
at a speed determined
by the
motion control
apsupply
pressure
and the valve flow coefplication
challenges.
ficients.
At some point,
the load engages
The positional
servomechanism
is
the
limit switch,
whicha causes
the direcnormally
considered
positioning
systional
valve to center
block all
tem. However,
whenand
thethus
concept
of
ports.
The
will decelerate
rapidly,
profiling
is load
superimposed,
the result
is
subject
to its mass, the
cylinder’s
motion control—total
control
of thesize,
maand
how
the control
valvecontrol
shifts.
chine.
Toquickly
be successful,
motion
Severe
shock
and
vibration
caninto
occur
must have
three
things
designed
it:
before
the load actually
especially
• a command
motionstops,
profile,
if the
load mass ispower
great. to manage the
• sufficient
Inload
the event
large shocks,
at theof
required
speeds,machine
and
members
experience
high stress
levels
• sufficient
closed-loop
bandwidth
that can
shorten
their and
lives.stability
Very high
to meet
accuracy
repressure
peaks—or
spikes,
as
they
are
frequirements.
quently
called—also will
occur.
These
The motion-control
profile
is crucial
pressure
peakscontrol
can overstress
to the motion
system,the
andhydraureprelic
components,
including
the
cylinder
sents the most significant contribution of
tube
and seals,
leadingtotothe
premature
leakthe digital
computer
motion-conage
failures.
external
vibration
trol and
process.
The The
profile
is the means
by
can
move
the
entire
machine,
putting
which acceleration, velocity, and position
mechanical
stresses
on the plumbing,
are controlled
simultaneously.
Profiles
which
leadgenerated
to fitting failure.
are verycan
easily
by a digital comIn the
oscillatory
stopping
process,
puter
(motion
controller
card or
circuit
the
frequency
of
the
vibration
can
be
board), and are formulated at design time
measured
if one
pressure,
or
to determine
the has
peakforce,
speeds
that must
speed
sensors
whose
outputs
canonbea
be attained
by the
actuator
based
displayed
on an oscillographic
recording
stated productivity
rate.
instrument.
Withthese
the peak
recording,
In determining
speeds,the
we
frequency
of the
can be
must first collect
datavibration
on the productivmeasured.
frequency For
is called
the
ity needs of This
the application.
example,
hydromechanical
resonant
frequency
suppose the application
requires
stamp(HMRF)
and
can be calculated
using any
ing out 30
packets
of paper plates
per
of
several
methods. (Space
limitations
minute.
Immediately,
we know
that the
FLUIDPOWERFUNDAMENTALS
3. cycles
Simple directional,
or Furdiscycle rate is 12
per second.
moves
ther, if we are crete,
givenvalve
that control
cylinder
stroke
with that
rapidthe
acceleration
must be, say, 4loads
in., and
cylinder
and deceleration.
must remain extended
for 0.5 Results
sec., then
are inherent
we know that the
actuatormechanical
must fully exshock
to
the
load
andWe
prestend and retract within 1.5 sec.
can
sure spikes
in the hydraulic
add some reasonable
acceleration
times,
circuit
is not
plus take into system.
accountThis
that
the cylinder
intended
to
be
complete,
but
does not naturally want to retract as fast
illustrates
the shortcomings
as it extends. (Yes,
this statement
is corof simple directional
control.
rect for the valve-controlled
cylinder.)
To compensate for this, we can rob a few
prohibit
discussing
in detail,time
but
milliseconds
from HMRF
the extension
for
on this time.
important
andmore
give information
it to the retraction
This
topic,
article
from the author:
allowssee
us this
to reach
conclusions
about
http://hydraulicspneumatics.com/200/
the maximum speed needed in order to
TechZone/HydraulicValves/Article/
achieve the required productivity rate.
False/6495/TechZone-HydraulicValves)
This example leads us to the concluThe
self-destruction
is
sion thatpotential
the peakfor
extend
speed will have
probably
biggestOnce
shortcoming
of
to be aboutthe
10 in./sec.
we are given
the maximum
discrete control
However,
thrustsystem.
force, we
can depoor
repeatability
of thepressure,
stoppingwhich
point
sign for
a given supply
can
also
be
a
serious
drawback.
When
makes it possible to completely size the
system
designers useinathe
limit
switch
to
critical components
system.
Our
initiate
deceleration,
normally
it
example
introducesthey
the idea
of a do
mobecause
the
application
needs
to
stop
the
tion profile, which is a means of definload
at some
and
controlled
ing the
speed,predictable
acceleration,
and
position
position.
Unfortunately,
of the cylinder
and load in
at the
anydiscrete
instant
system,
several
effects cause
within each
cyclerandom
and throughout
each
considerable
variation
in thelife
ultimate
and every cycle
for the entire
of the
stopping
machine.point. The actual stopping time
is aThe
complex
functionbetween
of:
relationship
accelera•
magnitude
of
the
load’s
mass,
tion, velocity, and position —which are
• shift
time of
the valve,time axis—is
drawn
against
a common
• timing of
shut-offrevealed
within
represented
in the
Fig.land
5. Newton
the
valve,
through the calculus that:
• valve
internal
velocity
is theleakage,
integral of accelera• cylinder
internal leakage,
tion
• friction
and
load,
positionin
is the
the cylinder
integral of
velocity
• fluid
viscosity,
velocity
is theand
derivative of posi• scan
time of the digital controller
tion, and
if used.
• (PLC),
acceleration
is the derivative of veAC
solenoids
produce valve shift
locity.
times
are never
affected
by thecalculus,
instant
Eventhat
if you
studied
within
the mathematical
AC power line 60or of
50-Hz
these are
facts
our
sine
wave
when
switching
takes
place.
physical world. What they mean to
The
randomcontrol
variation
in shift
time of
the motion
system
designer
is
the
will be any
at least
onethatvalve
specifying
onethe
of time
theseofthree
half
a line cycle
msecautomatically
in the 60-Hz
parameters
for (8.3
all time
system
or the
10 msec
the as
50-Hz
specifies
otherintwo
well.system).
This is
This
addsoftotheir
the normal
random
variabecause
undeniable
relationtions
in the
valve
shift time.are all inteship to
each
other—they
Load
and cylinder
also congrals
or derivatives
of friction
each other.
tribute
variations.
These
are afThisrandom
relationship
makes
it possible
fected
by
system
temperature,
as
is the
to achieve the simultaneous control
of
viscosity
of the
process
fluid,
whicheven
imacceleration,
velocity,
and
position,
poses
itsthe
ownfinal
variations
the stopping
though
systeminto
closes
the loop
time.
If
a
digital
controller
is
used,
the
with position feedback only. It works
random
in stopping
time
will
be
like this:variation
The position
profile,
the
plot
at least
to one
full scan
The
the equal
bottom
of Fig.
5, isinterval.
the actual
slower
the digital
the greater
command
to the controller,
motion-control
serthe
variation. The
netslope
result of
is that
vomechanism.
The
the when
posithe
Fig. 3 is the
tested
for repeattionsystem
profile of
represents
velocity
of the
ability,
actual
stopping
point will
vary
output the
at each
and
every instant
in time.
considerably
from
trial
to
trial.
The slope of the velocity profile is the
acceleration profile, which is actually
Proportional
carried in the Control
curvature of the position
Figure
4
shows
the second
step toward
profile. So the designer
controls
accelmotion
control,
whichthe
uses
a proporeration by
controlling
curvature
of
tional
valve command
instead ofprofile.
the discrete dithe position
rectional
valve.
As with
Fig. For
2, the
scheThe final
result
is this:
merely
matic
not complete;
rather,
it
havingdiagram
fed the is
position
command
prois
intended
only to point
out the
obvious
file
to the motion
control
servo
axis,
advantages
of usingvariables
proportional
rather
all three dynamic
are implicit
than
on-off
control.
Note that
the cylin that
profile.
Furthermore,
designing
inder
is outfitted
a position
sensor
a servo
axis that with
faithfully
follows
the
that
measures
the position
piston
command
profile,
controls of
allthe
three
parelative
toThis
the cylinder
tube.
sensor
rameters.
is the sum
andThe
substance
works
full-time—it
is alwaysphilosophy.
sending an
on modern
motion control
analog
outputtime,
signal
the controller.
At design
thetodesign
methodolcontroller is
depicted
as amotion
digital
ogyThe
recommends
that
the entire
device
which
a deceleration
point
controlinto
profile,
acceleration,
velocity,
can
entered,beprobably
by means
and be
position,
synthesized
basedofona
conventional
keyboard.
Being a needs.
digital
productivity and
other motion
device,
it
can
give
its
attention
to
In contrast, at commissioning timeonly
the
one
item or task
at any given
instant.
programmer
of a dedicated
motion
conThat
is,card
whenmay
it “looks
at” the deceleratroller
only specify
that the
tion
set
point,
it
cannot
be
looking
at
motion must go from, say, position x to
the
position
sensor
output.
There
position
y with
maximum
speed
of vmust
and
necessarily
be a time lagof
between
maximum acceleration
a. The these
digitwo
events.controller
Additionally,
the PLC
probtal motion
interprets
these
as
ably
has manythen
other
tasks that
diinstructions,
generates
thewill
actual
vide
its commands
attention, such
as monitoring
profile
on the
fly. Conseand
controlling
temperature,
reservoir
quently,
the programmer
may never
see
level,
and
more.
The
resulting
delay
is
the entire profile itself.
the scan time of the controller, and it
dictates
how well the system
performs.by
THIS INFORMATION
was provided
Thus,
we
see
that
each
of
the tasks
Jack L. Johnson, PE, an electrohydraulic
is
actually serviced
regulardirector
intervals.
specialist,
consultant,at former
of
The
totalPower
time lag
between
instant
the Fluid
Institute
at thethe
Milwaukee
when
monitoring
the poSchoolone
of task—say,
Engineering,
and contributing
sition
output&and
the instantHe
when
editor sensor
to Hydraulics
Pneumatics.
can
itbedoes
so again—is
theattotal
scan time.
reached
via e-mail
jack@idaseng.
Furthermore,
the instant when an event
com.
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FLUID POWER FUNDAMENTALS
where the cylinder and load will eventually stop. There is no means to back up
should there be overtravel. Therefore,
almost all the random factors that affect
the stopping point of the discrete directional system will be visited upon this,
the proportional system.
A system that is proportional for its
shock control, but also monitors and
controls position on a full-time basis,
is called the electrohydraulic positional
servomechanism. It possesses all the desirable characteristics for effective motion control, but because it is a bona-fide
feedback control system, it introduces
its own set of challenges. However, when
properly designed and tuned, this system provides the ultimate solution to
the most demanding motion control application challenges.
The positional servomechanism is
normally considered a positioning system. However, when the concept of
profiling is superimposed, the result is
motion control—total control of the machine. To be successful, motion control
must have three things designed into it:
• a command motion profile,
• sufficient power to manage the
load at the required speeds, and
• sufficient closed-loop bandwidth
to meet accuracy and stability requirements.
The motion-control profile is crucial
to the motion control system, and represents the most significant contribution of
the digital computer to the motion-control process. The profile is the means by
which acceleration, velocity, and position
are controlled simultaneously. Profiles
are very easily generated by a digital computer (motion controller card or circuit
board), and are formulated at design time
to determine the peak speeds that must
be attained by the actuator based on a
stated productivity rate.
In determining these peak speeds, we
must first collect data on the productivity needs of the application. For example,
suppose the application requires stamping out 30 packets of paper plates per
minute. Immediately, we know that the
cycle rate is 12 cycles per second. Further, if we are given that cylinder stroke
must be, say, 4 in., and that the cylinder
must remain extended for 0.5 sec., then
we know that the actuator must fully extend and retract within 1.5 sec. We can
add some reasonable acceleration times,
plus take into account that the cylinder
does not naturally want to retract as fast
as it extends. (Yes, this statement is correct for the valve-controlled cylinder.)
To compensate for this, we can rob a few
milliseconds from the extension time
and give it to the retraction time. This
allows us to reach conclusions about
the maximum speed needed in order to
achieve the required productivity rate.
This example leads us to the conclusion that the peak extend speed will have
to be about 10 in./sec. Once we are given
the maximum thrust force, we can design for a given supply pressure, which
makes it possible to completely size the
critical components in the system. Our
example introduces the idea of a motion profile, which is a means of defining the speed, acceleration, and position
of the cylinder and load at any instant
within each cycle and throughout each
and every cycle for the entire life of the
machine.
The relationship between acceleration, velocity, and position —which are
drawn against a common time axis—is
represented in Fig. 5. Newton revealed
through the calculus that:
• velocity is the integral of acceleration
• position is the integral of velocity
• velocity is the derivative of position, and
• acceleration is the derivative of velocity.
Even if you never studied calculus,
these are mathematical facts of our
physical world. What they mean to
the motion control system designer is
that specifying any one of these three
parameters for all time automatically
specifies the other two as well. This is
because of their undeniable relationship to each other—they are all inte-
grals or derivatives of each other.
This relationship makes it possible
to achieve the simultaneous control of
acceleration, velocity, and position, even
though the final system closes the loop
with position feedback only. It works
like this: The position profile, the plot
at the bottom of Fig. 5, is the actual
command to the motion-control servomechanism. The slope of the position profile represents the velocity of the
output at each and every instant in time.
The slope of the velocity profile is the
acceleration profile, which is actually
carried in the curvature of the position
profile. So the designer controls acceleration by controlling the curvature of
the position command profile.
The final result is this: For merely
having fed the position command profile to the motion control servo axis,
all three dynamic variables are implicit
in that profile. Furthermore, designing
a servo axis that faithfully follows the
command profile, controls all three parameters. This is the sum and substance
on modern motion control philosophy.
At design time, the design methodology recommends that the entire motion
control profile, acceleration, velocity,
and position, be synthesized based on
productivity and other motion needs.
In contrast, at commissioning time the
programmer of a dedicated motion controller card may only specify that the
motion must go from, say, position x to
position y with maximum speed of v and
maximum acceleration of a. The digital motion controller interprets these as
instructions, then generates the actual
profile commands on the fly. Consequently, the programmer may never see
the entire profile itself.
THIS INFORMATION was provided by
Jack L. Johnson, PE, an electrohydraulic
specialist, consultant, former director of
the Fluid Power Institute at the Milwaukee
School of Engineering, and contributing
editor to Hydraulics & Pneumatics. He can
be reached via e-mail at jack@idaseng.
com.
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☞ Back to
Table of Contents
FLUID POWER FUNDAMENTALS
Hydraulic Fittings
and Flanges
Among the basic elements of virtually every
hydraulic system is a series of fittings flanges for
connecting tube, pipe, and hose to pumps, valves,
actuators, and other components.
I
f the components within hydraulic
systems never had to be removed,
connections could be brazed or
welded to maximize reliability. However, it is inevitable that connections
must be broken to allow servicing or
replacing components, so removable fittings are a necessity for all but the most
specialized hydraulic systems. To this
end, fitting designs have advanced considerably over the years to improve performance and installation convenience,
but the overall function of these components remains relatively unchanged.
Fittings seal fluid within the hydraulic system by one of two techniques:
all-metal fittings rely on metal-to-metal
contact, while O-ring type fittings contain pressurized fluid by compressing an
elastomeric seal. In either case, tightening threads between mating halves of the
fitting (or fitting and component port)
forces two mating surfaces together to
form a high-pressure seal.
All-Metal Fittings
Threads on pipe fittings are tapered
and rely on the stress generated by forcing the tapered threads of the male half
of the fitting into the female half or
component port (Fig. 1). Pipe threads
are prone to leakage because they are
torque-sensitive—over-tightening distorts the threads too much and creates
a path for leakage around the threads.
Moreover, pipe threads are prone to
loosening when exposed to vibration
and wide temperature variations—certainly no strangers to hydraulic systems.
Seepage around threads should be
expected when pipe fittings are used in
high-pressure hydraulic systems. Because pipe threads are tapered, repeated
assembly and disassembly only aggravates the leakage problem by distorting
threads, especially if a forged fitting is
used in a cast-iron port. Thread sealant
compound, a potential contaminant, is
recommended for pipe fittings, which is
still another reason why most designers
consider them to be obsolete for use in
hydraulic systems.
Flare-type fittings (Fig. 2) were developed as an improvement over pipe
fittings many years ago and probably
remain the design used most often in
hydraulic systems. Tightening the assembly’s nut draws the fitting into the
flared end of the tubing, resulting in a
positive seal between the flared tube
face and the fitting body. The 37-deg.
flare fittings are designed for use with
thin-wall to medium-thickness tubing
in systems with operating pressures to
3,000 psi. Because thick-wall tubing is
difficult to form to produce the flare, it
is not recommended for use with flare
fittings. The 37-deg. flare fitting is suitable for hydraulic systems operating at
temperatures from −65° to 400° F. It is
more compact than most other fittings
and can easily be adapted to metric tubing. It is readily available and one of the
most economical.
The flareless fitting (Fig. 3), gradually
is gaining wider acceptance in the U.S.
because it requires minimal tube preparation. It handles average fluid work-
1. Pipe fittings have given way to newer
fitting designs that simplify assembly and
maintenance, and reduce or eliminate leakage. Shown is a 90-deg. adapter elbow with
pipe threads at one end that mount permanently into the component port. The other
end of the fitting uses straight-thread flare
fitting for tubing connection.
2. Flare-type fittings offer several design
and performance improvements over pipe
fittings and are used with thin-walled and
medium-thickness tubing.
ing pressures to 3,000 psi and is more
tolerant of vibration than other types of
all-metal fittings. Tightening the fitting’s
nut onto the body draws a ferrule into
the body. This compresses the ferrule
around the tube, causing the ferrule to
contact, then penetrate the outer circumference of the tube, creating a positive seal. Because of this, flareless fittings
must be used with medium- or thickwalled tubing.
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FLUID POWER FUNDAMENTALS
4. Non-adjustable,
left, and adjustable
SAE straight-thread
O-ring fittings offer
ease of assembly
and high potential
for leak-tight
connections.
3. Flareless fittings offer advantages similar
to those of flare fittings and are used with
medium- to thick-walled tubing.
O-ring-Type Fittings
Surprising as it may seem, leakage
in hydraulic systems could have been
eliminated more than a couple generations ago. Although leak-free hydraulic
operation had always been desirable, the
need became more acute with higher
operating pressures that became necessary during World War II, primarily in the hydraulic systems of military
aircraft. Until then, common operating pressures had hovered around 800
to 1,000 psi. The post-war era ushered
in systems designed to operate at pressures to 1,500 psi and higher on applications where rapid cycling and high
shock pressures were present. It was not
long until pressures climbed to 2,500
and 3,000 psi—which certainly are not
uncommon today.
Fittings that use O-rings for leak-tight
connections continue to gain acceptance
by equipment designers around the
world. Three basic types now are available: SAE straight-thread O-ring boss fittings, face seal or flat-face O-ring (FFOR)
fittings, and O-ring flange fittings. The
choice between O-ring boss and FFOR
fittings usually depends on such factors
as fitting location, wrench clearance, or
individual preference. Flange connections generally are used with tubing that
has an outside diameter (OD) greater
than 7/8-in. or for applications involving
extremely high pressures.
O-ring boss fittings seat an O-ring
between threads and wrench flats
around the OD of the male half of the
connector (Fig. 4). A leak-tight seal is
5. A flat-face O-ring fitting uses an O-ring
in a recessed groove in the male half that
mates with a flat, smooth surface on the
female half.
formed against a machined seat on the
female port. O-ring boss fittings fall into
two general groups: adjustable and nonadjustable. Non-adjustable (or nonorientable) fittings include plugs and
connectors. These are simply screwed
into a port, and no alignment is needed.
Adjustable fittings, such as elbows and
tees, need to be oriented in a specific
direction.
The basic design difference between
the two types is that plugs and connectors have no locknuts and require
no back-up washer to effectively seal
a joint. They depend on their flanged
annular area to push the O-ring into the
port’s tapered seal cavity and squeeze the
O-ring to seal the connection. Adjustable fittings are screwed into the mating
member, oriented in the required direction, and locked in place when a locknut
is tightened. Tightening the locknut also
forces a captive backup washer onto the
O-ring, which forms the leak-tight seal.
Assembly is always predictable, because
technicians need only make sure that the
backup washer is firmly seated on the
port’s spot face surface when the assembly is completed and that it is tightened
properly.
The FFOR fitting forms a seal between a flat, finished surface on the female half and an O-ring held in a recessed circular groove in the male half
(Fig. 5). Turning a captive threaded nut
on the female half draws the two halves
together and compresses the O-ring.
Fittings with O-ring seals offer a
number of advantages over metal-tometal fittings. While under- or overtightening any fitting can allow leakage,
all-metal fittings are more susceptible to
leakage because they must be tightened
to within a higher, yet narrower torque
range. This makes it easier to strip
threads or crack or distort fitting components, which prevents proper sealing.
The rubber-to-metal seal in O-ring fittings does not distort any metal parts
and provides a tangible “feel” when the
connection is tight. All-metal fittings
tighten more gradually, so technicians
may have trouble detecting when a connection is tight enough but not too tight.
On the other hand, O-ring fittings
are more expensive than their allmetal counterparts, and care must be
exercised during installation to ensure
that the O-ring doesn’t fall out or get
damaged when the assemblies are connected. In addition, O-rings are not interchangeable among all couplings. Selecting the wrong O-ring or reusing one
that has been deformed or damaged can
invite leakage. Once an O-ring has been
used in a fitting, it is not reusable, even
though it may appear free of distortions.
Some manufacturers offer specially
designed, high-pressure fittings that
are equal in leak and weep resistance
to FFOR fittings and interchangeable
with a number of international fittings.
Testing has shown these new designs
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FLUID POWER FUNDAMENTALS
7. Properly designed and installed splitflange fitting has a uniform clearance of
0.010 to 0.030 in. between the port surface
and clamp halves.
6. Flanges come in a wide variety of standard configurations to suit most hydraulic applications.
to surpass all requirements with no evidence of leakage when exposed to vibrations up to 15 times more severe than
those experienced on a typical hydrostatic drive. These designs may appear
similar to standard fittings but should
not be mated with fittings from different
manufacturers.
Hydraulic Flanges
Fittings for tubing larger than 1-in.
OD have to be tightened with large
hex nuts which, in turn, require larger
wrenches to enable workers to apply
sufficient torque to tighten the fittings
properly. To install such large fittings,
system designers must provide the necessary space to give workers enough
room to swing large wrenches. In addition, worker strength and fatigue could
be factors affecting proper assembly.
Wrench extensions (cheater bars) might
be needed for some workers to exert an
applicable amount of torque.
Standard hydraulic flanges (Fig. 6)
overcome both of these problems.
Flanges use an O-ring to seal a joint
and contain pressurized fluid. An elastomeric O-ring rests in a groove on a
flange and mates with a flat surface on
a port—an arrangement similar to the
FFOR fitting. The O-ring flange is at-
tached to the port using mounting bolts
that tighten down onto flange clamps,
thus eliminating the need for a large
wrench when connecting large-diameter components. When installing flange
connections, it is important to apply
even torque on the four flange bolts to
avoid creating a gap through which the
O-ring can extrude under high pressure.
Manufacturers also offer split flanges,
which can be installed into existing systems. The basic split-flange fitting consists of four elements: a flanged head
connected permanently (generally
welded or brazed) to the tube, an O-ring
that fits into a groove machined into the
end face of the flange, and two mating
clamp halves with appropriate bolts to
connect the split-flange assembly to a
mating surface.
All mating surfaces must be clean
and smooth. Joints are more likely to
leak if either of the mating surfaces
are scratched, scored, or gouged. Additionally, wear tends to accelerate on
O-rings which are assembled against
rough surfaces. Where perpendicular
relationships are critical, all parts must
meet appropriate tolerances. While 64µin. surface finishes are acceptable, most
flange manufacturers prefer and recommend 32-µin. finishes on mating sur-
8. Unevenly tightened split-flange bolts
may cause the flange to tip up and damage
the O-ring, as shown at left, while overtightened bolts, right, can bend the flange
and bolts.
faces to ensure leak-free connections.
In a properly designed split-flange
assembly, the flange shoulder protrudes
approximately 0.010 to 0.030 in. beyond
the clamp face to ensure adequate contact and seal squeeze with the mating
face (Fig. 7). However, the clamp halves
do not actually contact the mating surface. The most critical operation during
assembly of a split-flange fitting to its
mating surface is to make certain that
the four fastening bolts are tightened
gradually and evenly in a cross pattern.
Air wrenches should not be used because they are difficult to control and
can easily over-tighten a bolt.
Fully tightening one of the bolts while
the others are still loose will tend to
cause the flange to tip upward (Fig. 8).
This action pinches the O-ring, and the
joint can then be expected to leak. When
the bolts are fully tightened, the flanges
sometimes bend downward until they
bottom on the port face, and the bolts
bend outward. In either case, the result
likely will be a leaking joint.
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FLUID POWER FUNDAMENTALS
Hydraulic System
Flushing Procedures
During flushing procedures, fluid velocity is
critical to successful removal of manufacturing and
installation debris from hydraulic systems.
T
olerances that exist in today’s
high-pressure hydraulic systems
demand tight control of system
contamination. Contamination that is
built into systems during manufacture
and assembly must be removed before
startup to ensure proper and predictable system performance throughout its
service life.
A new or rebuilt hydraulic system
should be flushed before it becomes
operational. The concept of flushing
is to loosen and remove contamination particles inside the system by forcing flushing fluid through it at high
velocity. In theory, this leaves the inside walls of the fluid conductors at the
same cleanliness level as the new fluid
to be installed. Then, during normal
operation, the system will experience
only externally and internally generated contamination that can be controlled with filtration.
Instructions for flushing usually
specify a level of system cleanliness that
must be achieved, and sometimes a fluid
velocity that must be maintained during
the flushing procedure. Typical instructions state that flushing must be accomplished at normal system fluid velocities for a certain period of time, with a
certain level of filtration. More stringent
specifications may call for a particular
fluid contamination level and require
documentation by fluid-contamination
analysis.
One shortcoming of all these flushing
methods is that they are based on procedures to clean the fluid, but ignore the
system’s interior cleanliness. Even if the
tubing and conductors were installed
with the greatest of visual care, the human eye can only see particles that are
larger than 40 µm—well below the needs
of even the crudest and most elementary
hydraulic system.
How High a Velocity?
The critical variable in flushing to
achieve acceptable fluid and conductor
cleanliness is fluid velocity. Traditional
flushing methods usually establish this
velocity in one of two ways:
• The velocity must be high enough to
achieve a Reynolds number (NR) of
3,000 or more.
• The velocity must meet or exceed
the system fluid’s normal operating
velocity as designed.
Experience has shown that neither of
these flushing velocities is sufficient to
assure the cleanliness of the inside diameter (ID) of the system’s conductors.
A short review of basic fluid dynamics
explains why.
Dimensionless Reynolds numbers
are used (along with other factors) to
classify fluid flow as either laminar, turbulent, or transitional (somewhere in
between). Reynolds numbers depend
on the fluid’s viscosity and velocity and
the ID of the pipe or tube. The flow condition that exists when NR is less than
2,000 is termed laminar, signifying orderly flow with parallel streamlines.
When the Reynolds number is greater
than 3,000, the flow becomes turbulent,
defined as the condition when fluid
streamlines are no longer orderly. Flow
exists in transition for Reynolds numbers between 2,000 and 3,000; this is
sometimes called the critical zone.
The hydraulic fluid velocity required
to achieve turbulent flow is well within
the recommended fluid-velocity guidelines for hydraulic fluid conductors.
This equation reinforces that statement:
NR = V × D/v
where V is the
fluid velocity in ft/
sec, D is the ID of
the fluid conductor in ft, and v is the
fluid kinematic viscosity in ft2/sec.
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This simplified sketch
outlines an experiment that Reynolds
used to study and
define the three
regimes of fluid flow.
FLUID POWER FUNDAMENTALS
First Example
Supp os e t he Re ynolds numb er
is 3,000, the conductor is a 1-in. tube
with a wall thickness of 0.049 in., and v
is 1.288 × 10-4 ft2/sec. Calculated fluid
velocity, then, is 5.14 ft/sec., which corresponds to a flow rate of 10.24 gpm.
Note that the viscosity (and therefore,
the Reynolds number) of a typical hydraulic fluid is influenced by temperature and pressure. That is, the hotter
the oil, the higher the Reynolds number
for the same fluid velocity and pressure.
And the higher the pressure, the lower
the Reynolds number for the same fluid
velocity and temperature. Thus, specifying that Reynolds number should be
3,000 is not a stringent requirement, but
is well within the normal operating fluid
velocities of a system. By definition, turbulent flow has been created because the
fluid-stream lines are no longer parallel. However, sufficient fluid motion to
clean the inside walls of the conductors
has not been generated.
Even at the recommended maximum
fluid velocities and Reynolds numbers
for hydraulic-system working conductors, fluid flow still is not turbulent
enough to greatly affect contamination
on conductor walls. Boundary-layer
fluid at the interior surfaces of the fluid
conductor remains undisturbed.
of viscous drag at the conductor wall is
known as the viscous sublayer.) A transition zone exists within the turbulent
flow range where flow resistance goes
from being governed by turbulence effects to being governed by the roughness
of the conductor’s inside wall.
This is shown clearly when inspecting
the Moody diagram which graphically
demonstrates the relationship between
Reynolds number, friction factor f, and
the roughness of the conductor’s inside
surface, e. Resistance to flow through a
fluid conductor, represented by the friction factor, is only affected by the surface
roughness of the fluid conductor when
the Reynolds number exceeds 4,000.
Thus, the majority of the resistance to
flow is created by turbulence effects.
Only when the Reynolds number is high
enough to cause surface projections of
the conductor walls to extend beyond
the viscous sublayer does the surface
come in contact with the turbulent flow
and affect the pressure drop in the conductor.
Surface Roughness
Average surface roughness for drawn
tubing e is 0.000005 ft. If the conductor
has the same 1-in. tubing with 0.049-in.
wall thickness, the ratio of roughness to
ID will be 0.000067. The Moody diagram indicates that the Reynolds number for this conductor must be at least
25,000 before the inside surface exposes
its resistance to fluid flow.
To ensure the inside wall of the conductor will be cleaned, the Reynolds
number must be greater than 25,000.
For flow to occur fully in the rough zone
of turbulent flow, the Reynolds number
must be greater than 3.25 × 107. Using
1.288 × 10-4 ft2/sec—the same fluid kinematic viscosity as in the first example—a Reynolds number of 25,000 corresponds to a fluid velocity of 42.8 ft/
sec (a flow rate of 85 gpm), still easily
attainable with conventional hydraulic
pumps.
Real-World Systems
It can be argued that if the walls of
a conductor are not greatly affected by
normal system fluid velocities, contaminants lodged there will have little chance
of entering the fluid stream. This may be
partially true, but the argument applies
only to smooth, straight conductors
at steady flows and pressures. It is not
representative of normal installations
that combine straight runs, bends, and
Second Example
The Reynolds number for flow at
normal system velocities next can be
calculated using the same conductor
size and kinematic viscosity as in the
first example, but with the velocity increased to 20 ft/sec. This higher velocity
results in a Reynolds Number of 11,671,
which corresponds to a flow rate of 39.8
gpm.
As the Reynolds number increases,
flow conditions go from laminar,
through the critical zone, to turbulent.
It has been proven empirically that once
the Reynolds number exceeds 3,000, resistance to fluid flow is a combination of
the effects of turbulence and of viscous
drag at the conductor wall. (This region
This modified Moody diagram relates friction factor f, Reynolds number, and conductor
surface roughness (e). (Courtesy: Ultra Clean Technologies Corp.)
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FLUID POWER FUNDAMENTALS
numerous fittings where flow patterns
are only predictable empirically, and
where pressure fluctuations and spikes
are commonplace.
Depending on the severity of service
that the system will experience, pressure spikes will dislodge contaminants
held in the walls of the conductors and
between fitting interfaces. In critical
systems, 3- to 25-μm particles can significantly impact system performance.
The only way to guarantee that conductor contamination (which can be
released at any time during operation)
does not affect system performance is
to protect each component with a filter,
an option so costly that it would not be
used in most systems. Although flushing hydraulic-system conductors at the
normal system operating fluid velocities
can provide fluid velocities higher than
flushing at a Reynolds number of 3,000,
the inside wall of the conductors still
will not be cleaned.
High-Velocity, High-Pressure
Flushing
Flows that produce Reynolds numbers of 25,000 are needed to ensure that
conductor walls are exposed to turbulent flow. Because system conductors
may consist of pipe, tube, and/or hose
and associated fittings, the specification
of a contractual number for the Reynolds number is difficult and still does
not guarantee that conductors will be
cleaned. The best you can do is establish
conditions that will maximize the Reynolds number. These conditions consist
of the highest possible velocity at the
lowest possible fluid viscosity. Limiting
factors are the conductor’s pressure rating and the fluid’s maximum operating
temperature.
When flushing a system, the valving and actuators must be “jumpered”
for safety reasons so that the only resistance to fluid flow is the pressure drop
in the conductors and fittings. When
flow becomes turbulent, the pressure
drop is proportional to the square of the
velocity. Extrapolating this relationship
to its maximum, the highest possible
velocity occurs when the pressure drop
in the conductor generated by fluid flow
is equal to the maximum test pressure
of the conductor. Flushing a system at
these high flows and pressures has the
added advantage of expanding and contracting the conductors and fittings as
the pressure fluctuates while inducing
highly turbulent flow. This optimizes
the flushing action.
By equating the pressure drop in a
conductor to the maximum pressure
rating of that conductor, the maximum
fluid velocity possible, along with the
corresponding Reynolds number, can
be calculated. The temperature of the
fluid directly affects its viscosity and is
the other variable that can control the
Reynolds number. Flushing pressure
also affects viscosity, but this is hard to
quantify because pressure in the pipe
being flushed will vary from maximum
at the pumping source to atmospheric at
the conductor outlet.
The equation used to calculate head
loss in the turbulent zone is:
HL = f × L × V2/(2D)
where HL is head loss,
f is the friction factor found in the
Moody diagram, and
L is the conductor length in ft.
This equation will calculate the maximum velocities and Reynolds numbers
that can be achieved for any given maximum flushing pressure.
Determining friction for pipe flow
requires iterative calculations using the
Moody diagram. Given the pressure rating, ID, length, and relative roughness
of the conductor, assume a friction factor and then calculate the fluid velocity.
Next calculate Reynolds number and
determine a new friction factor from the
Moody diagram. Repeat the calculation
until the friction factor converges.
The flushing table shows velocities
and Reynolds numbers that have been
calculated for 200 ft of Schedule-80 pipe
using the maximum test pressure for the
pipe and a surface roughness of 0.00015
ft for wrought-iron pipe. These calculations did not take into account the pressure drop produced by the various fittings normally used, so the values for the
attainable fluid velocities and Reynolds
numbers are optimistically high. Also,
fluids with lower viscosities or flushing at higher temperatures to reduce the
fluid viscosity can increase the Reynolds
number.
The values determined for maximum
flushing velocity and flow rate indicate
that some of these conditions—mainly
for lines with inside diameters smaller
than ¾ in.—can be satisfied using conventional high-pressure pumps of appropriate flow capacity. However, it may
be difficult to induce the pressure fluctuations needed to dislodge contaminants. For systems with larger conductors, special methods must be used to
achieve the necessary pressures, fluid
velocities, and Reynolds numbers to
properly flush the lines.
THIS ARTICLE was originally authored
by Patrick Jones of Consolidated Fluid
Power Ltd., Dartmouth, Nova Scotia.
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FLUID POWER FUNDAMENTALS
Hydraulic Power Units
A hydraulic power unit driven by an electric motor
must be sized differently from one driven by an
internal combustion engine—due to differences in
their torque-speed curves.
W
hen specifying components
for a hydraulic power unit,
the prime mover is sized
based on torque, speed, and power
requirements of the hydraulic pump.
This is fairly straightforward for electric
motors because they generally have a
starting torque that far exceeds running
torque. Often, though, designers specify
motors sized larger than necessary. This
results in wasted energy because the
motor operates at less-than-maximum
efficiency.
Diesel and gasoline engines are another matter. They have a much flatter torque-speed curve, so they deliver
roughly the same torque at high speed as
they do at low speed. This means an internal combustion engine may develop
high enough torque to drive a loaded
pump, but not enough to accelerate it to
operating speed. Consequently, with all
other factors being equal, a power unit
requiring an electric motor of a given
power rating usually requires a gasoline or diesel engine with a power rating more than double that of the electric
motor.
Selecting the Optimum Motor Size
The cost of electricity to operate an
electric motor over its entire lifespan
generally is many times that of the cost
of the motor itself. Therefore, sizing the
motor correctly for a hydraulic power
unit can save a sizable amount of money
over the life of the machine. If system
pressure and flow are constant, motor sizing simply involves the standard
equation:
hp = (Q ×P) ÷ (1,714×EM),
1. Calculation for root mean square power.
where:
hp is horsepower,
Q is flow in gpm,
P is pressure in psi, and
E M is the pump’s mechanical efficiency.
However, if the application requires
different pressures during different
parts of the operating cycle, you often
can calculate root mean square (RMS)
power and select a smaller, less-expensive motor. Along with the calculation of
rms power (Fig. 1), the maximum torque
required at the highest pressure level
of the application also must be found.
Actually, the two calculations are quite
simple.
For example, such an application
might use a 6-gpm, 3,450-rpm gear
pump to power a cylinder linkage that
operates for an 85-sec cycle (Fig. 2). The
system requires 3,000 psi for the first 10
sec, 2,200 psi for the next 30 sec, 1,500
psi for the next 10 sec, and 2,400 psi for
the next 10 sec. The pump then coasts
at 500 psi for 20 sec, followed by 15 sec
with the motor off.
It’s tempting to use the standard formula, plug in the highest-pressure segment of the cycle, and then calculate:
hp = (6 × 3000) ÷ (1714 × 0.9)
= 11.7 hp for 10 sec.
To provide this power, some designers would choose a 10-hp motor; others would be ultra-conservative and
use a 15-hp motor; a few might take a
chance with 7½ hp. These motors in
open drip-proof C-face models with
feet would carry a relative price of about
$900, $1,200, and $600, respectively, so
you could save hundreds of dollars per
power unit by choosing the 7½-hp motor—if it will do the job.
To determine this, first calculate the
power required for each pressure segment of the cycle:
hp1 = (6 × 2200) ÷ (1714 × 0.9)
= 8.5 hp for 30 sec.
hp2 = (6 × 1500) ÷ (1714 × 0.9)
= 5.8 hp for 10 sec.
hp3 = (6 × 500) ÷ (1714 × 0.9)
= 1.9 hp for 30 sec.
The RMS horsepower is calculated
by taking the square root of the sum of
these power values squared, multiplied
by the time interval at that power, and
divided by the sum of the times plus the
term (toff ÷ F), as indicated in Fig. 1.
Substituting the example values into
the boxed equation and solving reveals
that hp rms = 7.2. Thus, a 7½-hp motor can be used from the standpoint
of power alone. However, the second
item, maximum torque, still must be
checked before reaching a final decision.
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FLUID POWER FUNDAMENTALS
The maximum torque required to drive
this particular pump will be found at
the highest pressure—because the gear
pump’s output flow is constant. Use this
equation:
T = DP ÷ (12 × 6.28 × EM), where
T is torque in ft-lb, and
D is displacement in in. 3
For this example,
D = (6 × 231) ÷ (3,450)
= 0.402 in. 3
Then
T = (0.402 × 3,000) ÷ (12 × 6.28 × 0.9)
= 17.8 ft-lb.
Because electric motors running at
3450 rpm generate 1.5 ft-lb/hp, the 17.8
ft-lb of torque requires 11.9 hp (17.8÷1.5)
at 3000 psi. This matches closely enough
for the example application. (At other
standard motor speeds: 1725 rpm generates 3 ft-lb per hp; 1,150 rpm, 4.5 ft-lb per
hp; 850 rpm, 6 ft-lb per hp.)
Now the second criteria can be
checked against what the suggested motor can deliver in torque. What is the
pull-up torque of the 7½-hp motor selected? Because the torque is least as the
motor accelerates from 0 to 3450 rpm, it
must be above 11.9 ft-lb with an acceptable safety margin. Note that a motor
running 10% low on voltage will produce only 81% of rated pull-up torque:
in other words, (208÷230)2 = 0.81. Reviewing motor manufacturers’ performance curves will show several available 7½-hp models with higher pull-up
torque. Any of these motors could be a
good choice for this application.
Both motor criteria now have been
verified. The RMS power is equal to or
less than the rated motor’s power. The
motor’s pull-up torque is greater than
the maximum required.
Gas and Diesel Engine Power
Correctly sizing an electric motor
for a hydraulic power unit is a straightforward procedure. And if load pressure and flow remain fairly constant,
determining the power requirement is
relatively simple by using the familiar
equation:
2. Multiple-pressure duty cycle for 6-gpm gear pump from example with calculated horsepower values.
hp = (q × p) ÷ (1714 × EM) where:
q is flow, gpm (and accounts for the
pump’s volumetric efficiency),
p is system pressure at full load, psi,
and
E M is the pump’s mechanical efficiency
For example, assume an application
requires a flow of 13.7 gpm at a maximum pressure of 2,000 psi, and with a
pump efficiency of 0.80. From the equation above:
hp = (13.7 × 2,000) ÷ (1,714 × 0.80)
= 20 hp.
It may seem that a gas or diesel engine as the prime mover would have the
same power rating as an electric motor.
However, the general rule of thumb is to
specify an internal-combustion engine
with a power rating 2½ times that of an
equivalent electric motor (Fig. 2). This
is due primarily to the fact that internal combustion engines have different
torque-speed relationships than electric motors do. Examining the different
torque characteristics will provide the
understanding to make a choice based
on solid reasoning, rather than putting
faith in a rule-of-thumb.
a pump runs slowly, it will still pump
fluid. However, if the prime mover does
not develop enough torque to drive the
pump, the pump will not produce any
output flow.
To determine the torque required by a
hydraulic pump, use the following equation:
T = (p × D ) ÷ (6.28 × 12 × EM)
where:
T is torque, lb-ft, and
D is displacement, in. 3/revolution
Pump displacement is provided in
manufacturer’s literature. Continuing
with the example introduced at left, if
the pump has a displacement of 1.75
in.3/rev., required torque is calculated
as follows:
T = (2,000 × 1.75) ÷ (75.36 × 0.80)
T = 58 lb-ft
Torque can also be calculated using
the familiar horsepower equation:
hp = (T × n) ÷ 5,250
where:
n is shaft speed, rpm.
Substituting values from the example:
20 = (T × 1,800) ÷ 5250
T. = 58 lb-ft.
Pump Torque Requirements
To understand the differences in
power characteristics between an electric motor and internal-combustion engine, we’ll first examine characteristics
of a standard 3-phase electric motor.
Figure 3 shows the torque-speed relationship of a 20 hp, 1,800 rpm, NEMA
Power, of course, is the combination of torque and rotational speed. A
pump’s torque requirement is the main
factor that determines whether a motor or engine is suitable for an application. Speed is less critical, because if
Electric Motor Torque Signature
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FLUID POWER FUNDAMENTALS
Design B motor. Upon receiving power,
the motor develops an initial, lockedrotor torque, and the rotor turns. As
the rotor accelerates, torque decreases
slightly, then begins to increase as the
rotor accelerates beyond about 400 rpm.
This dip in the torque curve generally is
referred to as the pull-up torque. Torque
eventually reaches a maximum value
at around 1,500 rpm, which is the motor’s break-down torque. As rotor speed
increases beyond this point, torque applied to the rotor decreases sharply. This
is known as the running torque, which
becomes the full-load torque when the
motor is running at its rated full-load
speed—usually 1,725 or 1,750 rpm.
The torque-speed curve for a 3,600rpm motor would look almost identical to that of the 1,800-rpm motor. The
difference would be that speed values
would be doubled, and torque values
would be halved.
Common practice is to ensure that
torque required from the motor will
always be less than breakdown torque.
Applying torque equal to or greater
than breakdown torque will cause the
motor’s speed to drop suddenly and severely, which will tend to stall the motor and most likely burn it out. If the
motor is already running, it is possible
to momentarily load a motor to near its
breakdown torque. But for simplicity of
discussion, assume the electric motor is
selected based on full-load torque.
Note that Fig. 3 shows a temporary
large torque excess that can provide additional muscle to drive the hydraulic
pump through momentary load increases. These types of electric motors
also can be run indefinitely at their rated
hp plus an additional percentage based
on their service factor—generally 1.15
to 1.25 (at altitudes to 3,300 ft).
Catalog ratings for electric motors
list their usable power at a rated speed.
If the load increases, motor speed will
decrease, and torque will increase to a
value higher than full-load torque (but
less than breakdown torque). So when
operating the pump at 1,800 rpm, the
electric motor has more than enough
torque in reserve to drive the pump.
Torque Signature of Engines
A gasoline engine has a dramatically
different torque-speed curve (Fig. 4)
than an electric motor does. This means
a gasoline engine exhibits a much less
variable torque output throughout its
speed range. Depending on their design, diesel engines with the same power
ratings may generate slightly higher or
lower torque at lower speeds than gasoline engines do, but diesels exhibit a
similar torque curve throughout their
operating speed range.
Calculations above determined that
58 lb-ft of torque is required to drive
the pump at any speed. Referring to Fig.
4, the 20-hp gasoline engine develops
a maximum torque of only 31 lb-ft—
clearly not enough to drive the pump.
This is because its 20-hp rating is based
on performance at 3,600 rpm. Maximum torque occurs at speeds near 3,000
rpm but is still well below the 58 lb-ft
required by the pump. Even if the engine
produced enough torque at this speed,
power would still be inadequate due to
the lower speed.
This is where the 2½ sizing rule
comes from. An HPU requiring a 20-hp
electric motor to drive the pump at 1,800
rpm would require a gas or diesel engine
rated at about 50 hp. Moreover, these
values are based on an engine operating
at its maximum torque and power ratings. However, manufacturers recommend that gasoline and diesel engines
only operate continuously at about 85%
of their maximum rated values to prevent seriously shortening of their service
lives. So referring again to Fig. 4, a 20-hp
gasoline engine would develop just over
26 lb-ft of maximum torque, and only 24
lb-ft at 3,600 rpm.
It is also interesting to compare
this performance with fuel consumption. The fuel consumption chart (Fig.
5) shows that a 20-hp gasoline engine
achieves greatest fuel efficiency at about
2,400 rpm, where it consumes just over
8.2 lb/hr (0.41 lb/hp × 20 hp). At 3,600
rpm, the engine would be considerably
less fuel-efficient.
Actions to Take
By now it should be clear that specifying a gasoline or diesel engine to drive a
hydraulic power unit follows a different
procedure than that for specifying an
electric motor. If you are accustomed to
specifying electric motors for hydraulic
power units, you may be tempted to size
a pump to be driven at 1,800 rpm, then
specify an oversized motor that can develop enough torque to drive the pump
at this speed. This technique will pro-
3. The torque-speed curve of an ac electric motor reveals that much higher torque can be
generated at low speed than is needed to drive a hydraulic pump at full-load speed.
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FLUID POWER FUNDAMENTALS
duce a reliable power unit, but one that
is relatively heavy, bulky, inefficient, and
noisy.
Instead of following this procedure,
any of several options should be considered. One would be to drive the pump
at a speed higher than 1,800 rpm. Pump
literature for mobile equipment should
list ratings at a variety of speeds. If it
doesn’t, consult the pump manufacturer.
Driving the pump at a higher speed
decreases its required displacement,
thereby reducing its size, weight, and
torque requirement. So operating the
power unit at higher speed more closely
matches engine performance to the application by increasing torque produced
by the engine and reducing the torque
required by the pump.
More specifically, operating the pump
in our example at 2,800 rpm would increase engine torque to more than 30
ft-lb and reduce torque required by the
pump to perhaps 38 ft-lb. Although the
engine torque still would fall short of
that required, it obviously comes much
closer to matching pump torque than
when operating at 1,800 rpm.
Designers may be tempted to run a
gas or diesel engine at or near the speed
at which it exhibits optimum fuel efficiency. However, an operating speed
where the engine produces maximum
torque generally takes priority. This is
because if the engine doesn’t generate
enough torque at its optimum fuel efficiency speed, a larger engine would be
required. But a larger engine consumes
more fuel, which would defeat the purpose of trying to conserve fuel by operating at a specific speed.
In addition, pumps generally have
a speed range at which they are most
efficient. So even if an engine operates
a few hundred rpm above or below its
optimum fuel efficiency speed, torque
produced and pump dynamics generally
have a more pronounced effect on overall efficiency of the power unit. Therefore, the speed at which the gas or diesel
engine operates should take all of these
considerations into account.
As far as pump performance, many
designs exhibit higher mechanical and
volumetric efficiencies when operated
at speeds greater than 1,800 rpm. On
the other hand, operating a pump at a
speed higher than what it was designed
for would reduce its service life. Therefore, it is important to choose a pump
speed that offers the best combination of
pump and engine performance.
Perhaps an even better alternative
would be to provide a gearbox or other
type of speed reducer between the engine and pump. Although this would add
components to the power unit, it would
increase torque and reduce speed while
allowing both the engine and the pump
to operate at their optimum speeds. The
additional cost of the speed reducer may
be offset by the lower cost of a smaller,
lighter, and less-expensive engine.
Other Considerations
Because gas and diesel engines do
not exhibit the torque reserve of electric
motors—especially when accelerating
from rest—it is especially important that
the pump be unloaded whenever the
HPU is started. This can be done hydraulically, or mechanically through a
centrifugal clutch or other type of drive
element.
Finally, as with HPUs driven by electric motors, pump size—and, therefore,
size of the prime mover—often can be
reduced by incorporating accumulators
into the hydraulic system. If the hydraulic system operates in cycles where full
flow is needed only for brief periods, an
accumulator can store hydraulic power
during periods of low flow demand
and release this energy when full flow
is needed.
4. The torque-speed
curve for an internal
combustion engine
is much more linear
than that for an electric motor. This illustrates that to provide
the torque to drive a
hydraulic pump at low
speeds, gas and diesel
engines must have a
higher power capacity
than an electric motor
for driving the same
pump.
5. Depending on its design, a gas or diesel engine’s optimum fuel efficiency often occurs at a
speed other than where it produces maximum torque.
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☞ Back to
Table of Contents
FLUID POWER FUNDAMENTALS
Reducing Noise from
Hydraulic Systems
Quieter systems don’t just have a perception of
higher quality, but can also improve the health,
safety, and productivity of machine operators.
T
he National Institutes of Health
estimates that 15% of Americans between the ages of 20 and
69 have suffered hearing loss—mostly
permanent—due to exposure to noise
at work or in leisure activities. At the
workplace, the combination of a quiet
pump, well-engineered vibration and
pulsation controls, and good, economical installation practices will result in a
product with a distinct advantage in the
marketplace.
Sound is formed by vibrations that
create an audible mechanical wave of
pressure through a medium, usually
air or water. In hydraulics, noise can be
grouped into three categories: airborne
noise, which travels from the air to the
ear; fluid-borne noise, which is transmitted through the hydraulic system;
and structure-borne noise, which is created when one component of a system
propagates vibration through another
component.
The factors that influence noise generation are summarized in Fig. 1. Unfortunately, people often reference only input
excitation and sound pressure or sound
power. They tend to avoid the other factors that make up the physics of noise
generation. Sometimes one part is dominant while others are not. Therefore, one
must consider all of these factors when
designing for low noise. Furthermore,
the process applies separately to airborne
and fluid- and structure-borne noise.
Each application is unique, so you can’t
assume that what works in one system or
assembly will work in another.
A Closer Look at Noise Sources
Simply put, noise is any unwanted
sound. More technically, it is the unwanted byproduct of fluctuating forces
in a component or system. As mentioned, noise can be transmitted in three
ways: through the air, through the fluid,
and/or through the system’s physical
structure.
Airborne—We generally think of
noise as traveling only through air, going directly from its source to some receiver—our ears. This is airborne noise.
Airborne noise, however, must have a
source within some component of the
system or application. That component
can be—but is not always—the pump.
All noise heard by the operator is
technically airborne noise. From the
perspective of the noise, vibration, and
harshness (NVH) engineer, airborne
noise refers to noise that came directly
from the surface of the source.
Fluid-borne—Whether it’s a piston,
vane, or gear pump, these positive-dis-
x
Force
=
Transfer
function
placement pumps all have some level
of pressure ripple (Fig. 2). As a result,
uneven flow characteristics and pressure
pulsations are created and transmitted through the fluid. This is known as
fluid-borne excitation. The fluid-borne
excitation generates vibration at the
surface of the hose, which can be transferred into adjacent structures via the
hose clamps/supports, or due to direct
contact of the hose to the structure when
under pressure.
The pressure pulsations of fluidborne excitation, in turn, create corresponding force fluctuations. The vibrations in the hydraulic hoses are known
as fluid-borne excitation. These result in
vibrations that create fluid-borne noise.
Proper hydraulic-line configuration
can be used to maintain vibration isolation when pumps and electric motors are
mounted on isolators. A proper combination of rigid and flexible conduit can
provide a more stable configuration,
providing reduced vibration and noise.
x
Vibration
=
Radiation
efficiency
Sound
pressure
or power
1. Factors that influence the generation of noise range from input excitation (far left) to sound
pressure or sound power (far right).
☞ LEARN MORE @ hydraulicspneumatics.com | 22
FLUID POWER FUNDAMENTALS
Result from nine pistons
Piston #1
Piston #2
Piston #3
Piston #4
Flow
Piston #9
0
20
40
60
80
Angle of rotation, deg.
180
2. Characteristics of the noise generated by hydraulic pumps are determined by many factors,
including its design and number of pumping chambers. This illustration shows individual and
combined flow pulsations in a nine-piston pump.
Sound-pressure level amplitude, dBA
Isolation of hydraulic lines and hoses
from the application structure (i.e.,
frame, supports, or panels) offers another opportunity to reduce noise in
the design of the machine. Panels and
shields can often act as speakers and amplify relatively low vibration levels into
high noise sources.
Hydraulic hoses and tubing can be
transmitters of fluid-borne vibration
in hydraulic hoses and tubing, turning
structural components into “speakers”.
It’s important to address the position
of hoses or tubes when designing quiet
hydraulic equipment in order to achieve
maximum noise reduction.
Structure-borne—Structure-borne
noise is the result of vibration transmitted only through the structure of the ap2nd harmonic
shaft frequency
plication. The vibration, as shown in Fig.
1, is the combination of the force and
the response of the component, and the
radiation efficiency of the component.
These structures then emit an audible
sound, or airborne noise, which is what
hydraulic equipment operators physically notice.
Structure-borne noise starts with vibration from an external source or component and is transferred directly into
the electric motor, structure, or frame of
an application. Once the vibration enters
the structure, it propagates through the
structure at the speed of sound of the
structure (most likely steel), which can
excite other components and cause them
to become radiators of noise—i.e., speakers. Components on the machine—such
Pumping frequency
as panels, shields, supports, and reservoirs—can radiate noise at pumping frequencies, and multiples of pumping frequencies, very effectively (Fig. 3). That’s
because these types of components have
many resonant frequencies. Components such as these are known as high
modal density components.
Vibration control can be used to minimize transmission of vibration from
pumps and drives to machine structures
and equipment. This can be achieved by
isolating the pump and/or motor from
rigid foundations by using subplates or
other base isolators.
Large areas of thin metal in systems
can also radiate noise effectively. This
noise can be reduced by strategically
placing engineered stiffening ribs or
damping treatment to the metal surfaces.
Understanding Noise Parameters
Evaluating noise can become confusing, because multiple vibration paths
can exist at the same time. One must
understand the source ranking of the
noise to properly evaluate the system
transmission paths and effectiveness of
each in any and all operating conditions.
A noise source often is surrounded by
a box-like enclosure to provide a physical barrier between the noise sources,
which can be caused by hydraulic-power
units, valves, hydraulic manifolds, motors, cylinders, hoses/tubing, and additional machine equipment. These barriers are designed to reduce the sound
3rd 4th
5th
6th
3rd harmonic
shaft frequency
Shaft
frequency
4th harmonic
shaft frequency
30
60
120
300
7th
8th
9th
10th
11th
2nd harmonic
pumping
frequency
900
Frequency (Hz)
3,000
5,000
3. This spectrum of structure- or fluid-borne output identifies shaft and pumping frequencies and their harmonics.
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10,000
FLUID
POWER
FUNDAMENTALS
FLUID
POWER
FUNDAMENTALS
FLUID
POWER
FUNDAMENTALS
Sound-pressure level amplitude, dBA
Flow
as panels, shields, supports, and resergenerated
hydraulic
equipment at the
Enclosure
Resultbyfrom
nine pistons
voirs—can radiate noise at pumping freoperator or bystander locations.
Free field
quencies, and multiples of pumping freAcoustic leakage around door seals,
58 dB
50
dB
quencies, very effectively (Fig.
3). That’s
etc., can also greatly affect the ability
detected
detected
because
these
types
of
components
have
of an enclosure
to
reduce
generated
Piston #9
Piston #1
Piston #2
Piston #3
Piston #4
50
dB
50
dB
many
resonant
frequencies.
Composound. As a rule of thumb, a 1% “hole”
noisesuch
source
noise source
nents
as these are known as high
in an acoustic enclosure will permit 50%
modal density components.
of the noise measured in it. When en1 can
meter
1 meter
Vibration control
be used to minclosed, amplitude of the noise within the
imize transmission of vibration from
enclosure actually increases; the noise
pumps and drives to machine structures
reflects within the enclosure, rather than
and equipment.
Thissource
can beinachieved
by
projecting
out.
4. Enclosures are often used to isolate noise. However,
placing a noise
an enclosure
0
20
40
60
80
180
isolating
the
pump
and/or
motor
from
Noise amplitude within the enclosure
can increase noise inside the enclosure by 5 to 8 dBA, which translates to a 45 to 60% increase
Angle of rotation,
deg.
rigid foundations by using subplates or
depends on the distance away from the over that without an enclosure.
other base isolators.
2.
Characteristics
the noise
generated
by hydraulic pumps are determined by many factors,
dominant
sourceof that
the noise
is meaLargewhen
areas evaluating
of thin metal
in systems
including
its
design
and
number
of
pumping
chambers.
This
illustration
shows
individual
and
sured. As a general rule, the amplitude points of effectiveness, cost, and prac- greater
noise
with a
also radiate
noise
effectively.
This
combined
pulsations
a nine-piston
of a noiseflow
source
when inplaced
inside pump.
of ticality. At the onset of developing a can
systematic
approach
rather
than simply
can individual
be reducedcomponents.
by strategically
an enclosure can increase noise inside noise-control program, it’s best to start noise
selecting
An
placing
engineered
stiffening
ribs
or
plication.
The
vibration,
as
shown
in
Fig.
Isolation
of
hydraulic
lines
and
hoses
the enclosure by five to eight decibels, at the source: the pump. Of course, the informed team, cognizant of the various
treatment
to the
metal
surfaces.
is the
combinationisofresponsible
the force and
from
structure
(i.e., 1,
or
45%the
to application
60% greater than
the source
pump
manufacturer
for damping
components
and roles
in the
overall
sysresponse
of thepump.
component,
and the tem, can help identify noise sources and
frame, supports,
or panels)
without
an enclosure
(Fig. 4). offers an- the
delivering
a quiet
Subsequently,
efficiencystrategy
of the component.
other
opportunity
tofactor
reduce
Noise Parameters
Another
important
in noise
terms in
of radiation
the most common
is to use a Understanding
design for low noise.
These
structures
then
emit
an
audible
the design is
ofabsorption
the machine.
Panels and
Evaluating noise can become confusenclosures
coefficient.
All porting design to minimize the presor airborneatnoise,
which israted
what Sound
shields can often
as speakers
and am- sound,
ing, because
multiple
vibration
paths
enclosures
have act
some
level of internal
sure pulsations
the pump’s
Quality
in Hydraulic
Systems
plify relatively
vibration
levels into
exist at the
time. the
Onesource
must
absorption,
butlow
adding
additional
ab- hydraulic
speed and equipment
pressure. operators physi- canHydraulics
is same
not always
cally
notice.
high
noise
sources.
understand
the
source
ranking
of
the
sorption material will help reduce noise.
At the component level, designers of a noise problem, but hydraulics freStructure-borne
startsvariablewith vi- noise
Hydraulic
hoses and
be may
togets
properly
evaluate
system
Larger
enclosures
will tubing
have a can
lower
want to start noise
off with
quently
the blame.
The the
reason
has
from anInexternal
source or drive
com- transmission
transmitters offactor
fluid-borne
vibration
paths
effectiveness
of
amplification
than smaller
en- bration
speed pumps.
variable-speed
more to do with
the and
quality
of the sound
ponent
and
is
transferred
directly
into
in
hydraulic
hoses
and
tubing,
turning
each
in
any
and
all
operating
conditions.
closures. Gaps or holes in the enclosure (VSD) systems, the pump speed varies produced than with its volume or preselectricthe
motor,
structure,
or frame of sure.
structural
into of
“speakers”.
A noise
is surrounded
by
will
reducecomponents
the effectiveness
noise re- the
to match
duty-cycle
requirement.
Mostsource
readersoften
are familiar
with the
application.
vibration
enters
It’s important
the position
box-like quality
enclosure
provide awhine.
physiduction
outsidetoofaddress
the enclosure.
Even an
This
will lower Once
noise,the
because
speeds
are aannoying
of atohydraulic
structure,
propagates
the cal
hoses
or tubes
barrierobjectively,
between the
noise
sources,
aoftiny
hole
or gapwhen
in andesigning
enclosurequiet
can the
reduced
when itnot
needed bythrough
the system.
Measured
that
whine
typistructure
at
the
speed
of
sound
of
the
hydraulic
equipment
in
order
to
achieve
can be have
caused
by hydraulic-power
significantly reduce its effectiveness in
Although quieter individual compo- which
cally doesn’t
a lot
of sound power.
structure
likely steel),
which
can units,
maximum
noise reduction.
valves,
curbing
sound.
nents may(most
contribute
greatly
to noise
However,
it is hydraulic
unpleasantmanifolds,
and tonal, moand
other components
cause
them
Structure-borne—Structure-borne excite
cylinders,
hoses/tubing,
andeven
adreduction,
additionaland
gains
can
be tors,
that makes
the actual
sound seem
to
become
radiators
of
noise—i.e.,
speaknoise
is
the
result
of
vibration
transmitditional
machine
equipment.
These
barachieved by reviewing the overall system louder.
Quieter Products and Systems by
Components
on the to
machine—such
ted only through the structure of the ap- ers.
are designed
to reduce
sound
design
for opportunities
reduce noise. riers
Therefore,
in addition
to thethe
objective
Design
A successful noise-control program Vibration control works to minimize issue of how much the hydraulic system
of vibration
from pumps contributes to overall sound levels, marequires a team effort by individuals in transmission
Pumping
frequency
2nd harmonic
3rd 4th
and
electric
motors
to
machine
strucchine builders also have to address the
several areas
of
expertise.
A
quiet
hy3rd harmonic
shaft frequency
5th
tures. This can be achieved by isolating subjective
issue of how the quality of
draulic pump does not guaranteeshaft
a quiet
frequency
6th
application’s sound affects overall
system. Choosing a quiet pump should the pump and/or electric motor from their7th
8th
its quality. The rumbling
be only oneShaft
part of a multifaceted pro- rigid supports via sub-plates or other perception of9th
frequency
10th
of an engine is typically
much louder
gram that
calls upon the talents of the base isolators.
11th
System testing and evaluation can than hydraulic whine, but the percepsystem designer, fabricator, installer, and
reduction of engine noise is one of power and
maintenance technicians. A breakdown provide further insight into noise2nd
harmonic
4th harmonic
pumpingstrength.
In properly
designed
testing
areas,
in any of these areas can unravel the en- tion.shaft
frequency
frequency
isolating components from background
tire noise control program.
30
60 a key role
120 noise makes
300
900
5,000 was 10,000
it possible to focus
on noise THIS 3,000
System designers
play
INFORMATION
provided by
Frequency (Hz)
in achieving successful noise control. sources, transmission paths, and op- Mike Beyer, senior specialist—noise and
3.
Thismust
spectrum
of structureor fluid-borne output
identifies shaft
and pumping
frequencies
their harmonics.
portunities
for reduction.
The
potentialandvibration
They
evaluate
every noise-control
at Eaton’s Hydraulics Div., Eden
technique available from the stand- for successfully reducing noise becomes Prairie, Minn. www.eaton.com.
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Table of Contents
FLUID POWER FUNDAMENTALS
Pneumatic Quick-Acting
Couplings
Servo and servo proportional valves control
pressure or flow—and ultimately, force or velocity.
Unlike simple directional valves, they can maintain
any position between fully open in one direction or
the other.
I
f a hose or tubing in a pneumatic
system will be connected and disconnected more frequently than
once a week, then chances are a quickacting coupling will pay for itself rapidly
by improving productivity. Although
simple in concept, many quick-acting
couplings are precisely engineered for
specific applications. Their widespread
use over many years has produced a
wide variety of standard designs.
Regardless of the manufacturer, all
quick-acting couplings have some elements in common. All have two parts: a
plug and a socket. The plug is the male
half and the socket is the female half.
When connected properly, these parts
seal and lock the joint effectively to contain internal pressures and resist any
tensile forces that tend to pull the joint
apart. The two parts are easily disconnected without tools by disengaging a
locking mechanism and separating the
parts.
One common application is in assembly workstations, where a worker
may have to rapidly switch from impact wrench to drill to riveter. With
one quick-acting coupling half on every
tool and the mating half on the air line,
tool changing is accomplished in seconds. Without the couplings, separate
air lines would be needed for each tool;
the mass of tools and lines would clutter
the workstation and could slow down
production.
Basic Components
A plug may be one- or two-piece
construction. The one-piece is machined to accept the mating locking
mechanism of the female or coupler
half. The two-piece is similar, but two
machined parts are used to provide
flexibility for a variety of end terminations. It may also be used as the retainer
for a valve assembly.
The socket is constructed to provide
a leak-tight interface with the plug. This
requires a sealing surface between the
plug and socket. A socket may also be
a one- or two-piece construction. The
one-piece is single part machined to
accept the configuration of the mating
plug and to provide a leak-tight seal. The
two-piece is similar but has a second
part to retain an internal valve assembly,
or to provide flexibility for a variety of
end configurations.
Pneumatic systems generally use a
single-shutoff valve coupling. In this
design, the valved coupling half prevents air loss from the system while the
joint is disconnected, and the unvalved
coupling half allows downstream air
to bleed off. (In hydraulic applications,
both coupling halves often are valved,
to minimize fluid leakage and limit the
amount of air, dirt, and water that can
enter the system.)
Coupling Designs
Ball-lock is a common design and
has a wide range of applications. A group
of balls is positioned in holes located
around the ID of the socket body. These
holes normally are tapered or stepped to
reduce their diameter at the socket body
inner diameter (ID), so the balls do not
fall into the cavity vacated by the plug
when the coupling is disconnected.
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FLUID POWER FUNDAMENTALS
Pin-lock couplings
use pins arranged in a
Ring-lock couplings
Ball-lock couplings are the
truncated-cone formation to grip
secure by pushing plug
most popular quick-acting
and hold the plug in the socket.
into socket; they disconnect
by rotating the socket’s outer sleeve.
coupling in use today and are offered
by many manufacturers.
A spring-loaded sleeve around the
socket body’s outer diameter (OD)
forces the balls toward the socket body
ID. To connect the plug, the sleeve is
pushed back, which opens clearance
so the balls are free to move outward.
Once the plug is in place, releasing the
sleeve forces the balls inward against a
locking groove on the OD of the plug.
To disconnect, pushing the sleeve back
provides the balls with clearance to
move outward and allow the plug to be
removed.
Pin-lock couplings allow push-toconnect joining using only one hand
because the outer sleeve does not need
to be retracted to make a connection.
In this design, pins are mounted around
the socket body ID in a truncated-coneshaped formation. Pushing the plug
into the socket moves the pins back and
outward, due to a ramp on the plug.
Shear across pins locks the plug into the
socket. Retracting the spring-loaded
sleeve, which forces the pins back out
of the locking groove, releases the plug
from the socket.
Ring-lock couplings use a split ring
seated in a groove and slot in the socket.
Pushing the plug into position causes
a ramp on the plug to spread the ring
apart at the split until the ring snaps
closed behind a retention shoulder on
the plug. Rotating an external sleeve expands the ring, thus releasing it from
the retention shoulder so the halves can
be pulled apart. This design provides
maximum flow in a small envelope for
normal shop air applications. A variation of this design uses jaws instead of a
split ring to lock the parts together.
Coupling Selection
Roller-lock coupling
A twist of the sleeve
design positions rollers
secures the plug once it
circumferentially around the
has been inserted into the
ID of the socket to grip the plug.
socket of the bayonet-type coupling.
Roller-lock couplings use locking rollers or pins spaced end-to-end
in grooves or slots around the socket’s
ID. As the plug is inserted, a ramp on
the plug OD pushes the rollers outward.
Once the plug is inserted the prescribed
distance, the rollers slip into a retention
groove on the plug’s OD. Retracting the
locking sleeve, which allows the ramp
on the plug OD to move the rollers outward, releases the plug.
Bayonet couplings rely on the familiar twist locking arrangement and are
widely used in a variety of applications,
especially in plastic couplings for lighterduty pneumatic equipment. To join the
coupling halves, lugs on the OD of the
plug engage slots in the socket sleeve
as the plug is pushed into the socket. A
quick turn locks the lugs into position.
Turning the plug in the opposite direction allows the halves to be pulled apart.
Before selecting a coupling, questions
must be answered regarding its expected
performance. How often will the coupling be connected and disconnected?
What type and diameter of hose or tubing will be used? Will the coupling or
hose be subjected to abuse such as impact from falling objects, severe vibration, or contamination from the work
environment?
A wide variety of O-ring and seal
materials—elastomers, PTFE, etc.—is
available. Material chosen for the plug
and socket also is important. Steel, stainless steel, brass, and aluminum are common. Lighter and inexpensive couplings
made from plastics are often used in
pneumatic applications. Most industrial
pneumatic couplers are made of brass
or steel.
Pressure rating relates to values that
provide optimum service life and maxi-
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FLUID POWER FUNDAMENTALS
mum pressure that can be tolerated
without failure. Manufacturer literature
should provide data for both of these
ratings of pressure. Information should
also help in determining safety factors if
service rating may be exceeded.
Literature should include data for
determining pressure drop through the
coupling at expected flows and pressures.
Many of these calculations are based on
flow of water at 60°F. Calculations for air
are more complex because a gas’ density
varies widely with its pressure and
temperature. A general rule to estimate
maximum air flow at 100 psig inlet and
5-psi pressure drop is to multiply flow
coefficient of the coupling by 25. Often,
literature contains more detailed data
on maximum air flow at prescribed inlet
pressures and pressure drop. Therefore,
precise values for pressure drop for
specific couplings should be obtained
from the manufacturer.
Also be aware that couplings may be
subjected to pressures well above the
maximum operating pressure. Sudden
shifting of valves or abrupt application
of heavy loads can cause system pressure to quickly rise and fall within milliseconds. These pressure spikes often
go undetected in a system, but still can
damage seals and locking elements of
the coupling. Ultimately, then, the coupling would develop leaks, become difficult to disconnect or reconnect, or any
combination of these. To prevent these
problems from occurring, select a coupling with a pressure rating substantially
higher than the anticipated maximum
operating pressure.
Multi-tube Connectors
As pneumatic systems become more
complex and the trend toward modular
automation increases, the need to connect and disconnect the growing number of pressure and control lines grows
almost exponentially.
Multi-tube connectors are the fluid
equivalent to electrical Cannon-style
connectors. They quickly and easily
connect or disconnect several tubing
lines, while maintaining a correct line
orientation and discrete flow paths during reconnection.
Radial seal multi-connectors use a
pin-shaped passage which is inserted
into a cavity containing a retained Oring.
Axial seal style multi-connectors
merely mate, instead of sliding into each
other. This allows for reduced insertion
force, decreased wear, and a smaller
overall profile.
Multi-tube connectors quickly connect
multiple lines of tubing in a specific
orientation.
Pneumatic Coupler Design Guide
Most standards are build around the nipple—therefore, the nipples may look the same,
but the coupler design will vary by manufacturer
Interchange
Feature
Application
Identifying feature
Tru-flate
Most popular in
automotive plants in
the midwest, usually
1/4- to 1/2-in. sizes
Drop lines and hoses
for pneumatic
hand tools and
equipment; original
quick disconnect
Short leading nose,
rounded features
Industrial
Interchange
Most common
interchange, used in
most general
industrial plants,
usually
1/4- to 3/4-in. sizes
Most plant pneumatic
drop lines for air tools
and equipment
Longer leading edge,
sharper corners
ARO-210
Most popular in
southeast furniture
and textile plants,
usually 1/4-in. sizes
Textile equipment and
pneumatic lines
Similar to Tru-flate
but with a
longer lead-in
Lincoln
Longneck
Interchange
Most popular in plant
lubrication systems,
usually
1/4- to 1/2-in. sizes
Grease and
lubrication systems
Very long lead-in
Feature
Application
Benefit
High-flow
1/4-in. body typically
provides as much flow
as standard 3/8-in.
body coupler,
3/8-in. provides
1/2-in. flow
Can use existing
air lines with new
style pneumatic tools
that require higher
air flows without
changing air lines,
very common in
European equipment
Increase air flow in
existing system
without replacing
compressor
Exhaust
Exhaust air in coupler
and connect at
zero pressure
Easy to connect large
air tools and reduces
dangerous hose whip
Used to create a safer
work environment
and meets ISO 4414
Twist-lock
Twist coupler to
lock in place
Lab equipment and
breathing
air applications
Locks coupler to
protect against
accidental
disconnection
Style
Universal
Plastic
Image
Image
Connects Tru-flate,
Used in equipment
Industrial and ARO-210
that travels to plants
nipples to one
using different couplers
coupler style
or unknown styles
Plastic body or
rubber boot on
coupler
☞ LEARN MORE @ hydraulicspneumatics.com | 27
Used to protect end
product from being
damaged or scratched
by coupler
☞ Back to
One coupler for
many applications
Popular in automotive
body shops and other
applications where
the finish can not be
marked
Table of Contents
FLUID POWER FUNDAMENTALS
Seals for
Hydraulic Cylinders
Pressure has been mounting to design and produce
cylinders that offer superior sealing performance
for enhanced reliability. This article provides an
overview of hydraulic cylinders along with the
various types of hydraulic seals for specification.
T
he type of cylinder and the application which it is used for are
two of the main criteria when
selecting the appropriate seals and
guides. Applications are referred to as
light-duty, medium-duty, or heavyduty applications. The duty levels are
typically characterized by the following criteria:
• Light-duty cylinders are used for
stationary equipment in a factory
environment and are characterized
by system pressures up to 160 bar
(2,300 psi) and temperatures up to
70°C (160°F).
• Medium-duty cylinders often are
used in agriculture off-highway
equipment with system pressures up
to 250 bar (3,625 psi) and temperatures up to 90°C (195°F).
• Heavy-duty cylinders are found in
off-highway earthmoving, mining,
and forestry equipment and are
characterized by system pressures
to 400 bar (5,800 psi) or greater and
with temperatures exceeding 90°C
(195°F), and perhaps intermittently
to 110°C (230°F).
Hydraulic Cylinder Seals
Hydraulic cylinder seals are used to
seal the opening between various components in the hydraulic cylinder. They
are designed to retain hydraulic fluids,
exclude solid or liquid contaminants,
and maintain hydraulic pressure. These
tasks require a variety of different seal
Different types are used in hydraulic cylinders to prevent fluid from leaking out of the
cylinder, from allowing contaminants in, and to provide low friction for long and reliable
cylinder life.
designs and performance-enhancing
features. Seal material must conform to
irregularities in metal surfaces to block
fluid passage. To adjust to clearance
gap size changes, the seal must expand
or compress rapidly to follow dimensional variations. Finally, to resist being extruded into gaps, the seal must
have sufficient modulus and hardness
to withstand shear stress produced by
system pressure.
Successful sealing involves containment of fluid within fluid power systems
and components while excluding contaminants. The surfaces in contact with
a seal determine what type to use. The
surface can either be static or dynamic—
in motion or without movement. Static
seals are typically used when there is no
relative motion between mating surfaces. Dynamic seals are the opposite.
They are used when there is motion be-
tween surfaces. This can be either reciprocating or oscillating motions.
Rod and Buffer Seals
Rod and buffer seals maintain sealing contact with a sliding motion between the cylinder head and the piston
rod. Depending on the application, a
rod sealing system can consist of a rod
seal and a buffer seal or a rod seal only.
Rod sealing systems for heavy-duty applications typically consist of a combination of both seal types. The buffer seal is
arranged between the rod seal and the
piston in the cylinder head.
Rod seals act as a pressure barrier to
keep the operating fluid inside the cylinder. They also provide a thin lubrication
film on the piston rod that lubricates the
rod seals and wiper seals. The lubricant
also inhibits corrosion of the piston rod
surface. However, the lubrication film
☞ LEARN MORE @ hydraulicspneumatics.com | 28
FLUID POWER FUNDAMENTALS
must be thin enough so it returns internal to the cylinder during the return
stroke. Selecting profiles and materials for a rod sealing system is a complex
task, considering all possible cylinder
designs and application criteria. Rod
and buffer seals come in many different
profiles and in a wide range of materials, series, and sizes to perform under
a variety of operating conditions and
applications.
Buffer seals protect rod seals by reducing the magnitude of pressure peaks.
Abrupt pressure peaks can occur by external forces acting on the piston rod,
initiated by the fluid inside the cylinder
and create higher fluid pressures in the
cylinder. These pressure peaks can be
in excess of the system operating pressure. Buffer seals—in combination with
rod seals—provide an effective rod sealing system for cylinders in heavy-duty
applications at high temperatures and
pressures.
Rod seals may have the most difficult task
in sealing hydraulic cylinders because
they must keep hydraulic oil from leaking
into the surrounding environment while
maintaining a thin coating of oil on the rod
surface.
Piston Seals
Piston seals maintain sealing contact between the piston and the cylinder bore. Differential pressures acting
on the piston to extend or retract the
piston rod can exceed 400 bar (5,800
psi). The pressure acting on the piston
Piston seals maintain sealing contact
between the piston and the cylinder
bore. Pressure acting on the piston
side is open to atmosphere. Therefore,
the piston seal must leave minimal oil
film when passing along the cylinder
bore since the transportation of oil otherwise would result in a leakage to the
exterior.
In single-acting cylinders, the open
end may push air out and draw air in
as the piston reciprocates. This air may
carry moisture and contaminants into
the cylinder, which can lead to seal
damage. Vent filters can be fitted to
the open side of the cylinder to reduce
contaminants entering the inside of the
cylinder. The cylinder bore may also
be hard chromium plated to prevent
corrosion.
seal increases contact forces between
the piston seal and cylinder surface.
Wiper Seals
Therefore, the surface properties of the
Hydraulic cylinders operate in a variety of applications and environmental
conditions, including exposure to dust,
debris, or outside weather conditions.
To prevent these contaminants from
entering the cylinder assembly and hydraulic system, wiper seals (also known
as scrapers, excluders, or dust seals) are
fitted on the external side of the cylinder
head.
Wiper seals maintain sealing contact
to the piston rod when the equipment is
stationary (static, no reciprocating motion of rod) and in use (dynamic, reciprocating rod). Without a wiper seal, the
retracting piston rod could transport
contaminants into the cylinder.
sealing surfaces are critical to proper seal
performance.
seal increases contact forces between the
piston seal and cylinder surface. Therefore, the surface properties of the sealing surfaces are critical to proper seal
performance.
Piston seals serve as a pressure barrier and prevent fluid from passing the
piston, which is important for controlling the cylinder motion or maintaining
the position when at rest. Piston seals
are typically classified into single-acting
(pressure acting on one side only) and
double-acting (pressure acting on both
sides) seals.
Double-acting piston seals have a
symmetrical profile and identical sealing functions in both directions. Typically, double-acting piston seals consist
of a slide ring and an energizer. Because
double-acting cylinders contain fluid
on both sides of the piston, a relatively
thick lubrication film can be permitted
between the piston seal and the cylinder
bore to minimize friction and wear.
A single-acting piston seal is designed
for cylinders where pressure is applied
from one side only. The piston in single-acting cylinders may have oil on the
pressure side only, while the opposite
Guide Lubrication & Guide Rings
Rod guides are typically placed inward of both the rod and buffer seal
and should be lubricated on assembly
with the same medium as used in the
system. The guide must receive ample
lubrication at all times and should not
be outside the rod seal. However, in certain conditions, guides with polytetrafluoroethylene (PTFE) added may be
used outside the rod seal due to their
self-lubricating properties.
Guide rings provide effective guidance of components that are in relative
motion to each other and accommodate
☞ LEARN MORE @ hydraulicspneumatics.com | 29
FLUID
POWER
FUNDAMENTALS
FLUID
POWER
FUNDAMENTALS
FLUID
POWER
FUNDAMENTALS
must
thinacting
enough
returns inradialbe
loads
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theitcylinder
asternal
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sembly.toThe
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reals
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is a complex
quires
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task,
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Rod and
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prevent
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ponents,
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cylinder
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als,
and
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and series,
keep the
piston
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radiaally
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and
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applications.
within acceptable limits for the seals.
Buffer
seals protect
rod sealsforbyperreThese
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are important
ducing
theofmagnitude
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formance
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and
Abrupt
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ternal forces acting on the piston rod,
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Sealing Material
Selection
Key
andIndustrial
create higher
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the
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peaks condican be
range of challenging
operating
in
excess
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temperature,
sure.
Buffer
seals—in
combination
pressure,
and
aggressive
chemicals.with
To
rod
seals—provide
an effective
rod sealhandle
these and other
harsh conditions,
ing
fortocylinders
heavy-duty
it is system
essential
select theinmost
suitable
applications
at
high
temperatures
and
sealing materials. Several factors impact
pressures.
material selection, including exposure
to media, pressure, temperature, and
potentially stringent regulatory requirements common in food and beverage or
oil and gas applications. Types of sealing
materials include:
Rubbers—NBR, FKM, and HNBR
are commonly used rubber materials
in hydraulic applications. They are extremely flexible and can be stretched
and deflected by exerting relatively little
force. Many of them deliver excellent resistance to mineral oils, greases, or other
media.
Thermoplastic elastomers—These
Rod
seals
may havetypical
the most
task
offers
advantages
ofdifficult
both rubber
in
sealing
hydraulic
cylinders
because
and
plastic
materials.
SKF’s
high-perthey
must keep
hydraulic oil polyurethanes
from leaking
formance
thermoplastic
into
the surrounding
environment
while and
(TPUs)
combine excellent
abrasion
maintaining
a thin coating
of oil on the rod
wear resistance,
low compression
set,
surface.
and high tear strength and outstanding
pressure resistance.
Piston
Seals
PTFE—Engineered
to handle exPiston
seals maintain
sealing
contreme
conditions,
PTFE and
its comtact between
piston and
the cylinpounds
can the
withstand
aggressive
der bore. Differential
pressures acting
chemicals
plus high temperatures
and
on the piston
or retract low
the
pressures.
Duetotoextend
their extremely
piston rod can
exceedthey
400can
baralso
(5,800
coefficients
of friction,
tolpsi). The
pressureconditions.
acting on the piston
erate
dry running
Plastics—Plastic materials can meet
higher temperature, chemical, and mechanical property requirements and can
range from engineering plastics to highperformance plastics. Backup rings are
typically made of plastics and used to
enhance the pressure carrying capability
of a rod or piston seal.
Criteria for Seal Specification
Designing sealing and guide systems
in hydraulic cylinders requires careful
attention to the interaction between all
cylinder components and the operatPiston
seals maintain
sealing
contact
ing conditions
as well
as the
application
between
the
piston
and
the
cylinder
requirements. The selection of the right
bore.
Pressure
acting
on thefor
piston
seal profile
and
material
a given apseal
increases
contact
forces between
plication
requires
consideration
of many
the
piston
seal
and
cylinder
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factors. For any application factors outTherefore,
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properties
of the
side of thethe
ordinary
to specify
sealing
sealing
critical to cylinder
proper sealdesystemssurfaces
in neware
hydraulic
performance.
signs, a certain amount of expertise may
be required.
sealBefore
increases
between
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can beforces
selected,
certain
piston
seal and
cylinder surface.
Thereapplications,
parameters,
and informafore,
the surface
properties
the sealtion should
be collected.
Theoffollowing
ing
surfaces
are application
critical to proper
seal
most
common
considerperformance.
ations are almost always required when
Pistonhydraulic
seals serve
as a pressure barselecting
seals:
rierFluid
and prevent
from passing
pressurefluid
range—the
rangethe
of
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important
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ling
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as frequency
and
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the
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Piston
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areTemperature
typically classified
into single-acting
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and
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double-acting
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acting
on
both
when operating and at rest
sides)
seals. stroking speed of the reSpeed—the
Double-acting
piston seals have a
ciprocating
piston rod
symmetrical
profile
sealFluid media—theand
typeidentical
and viscosity
ing
functions
both
directions. Typiof fluid
used ininthe
system
cally,
double-acting
piston sealsrod
consist
Hardware
dimensions—the
and
of
a slide
ring and
angroove
energizer.
Because
bore
diameters,
seal
dimensions
double-acting
cylinders
containcylinfluid
and gaps (if already
specified),
on
sides
of the
a relatively
derboth
overall
length
andpiston,
stroke length,
and
thick
lubrication
film can be (if
permitted
surface
finish specifications
already
between
specified)the piston seal and the cylinder
bore
to minimizeof
friction
and wear.
Application
the cylinder—the
A single-acting
seal is designed
type
of equipmentpiston
the cylinder
will be
for
cylinders
where
pressure
is
applied
used on and how the cylinder will operate
from
side only.
The as
piston
in sinin theone
equipment
as well
installation,
gle-acting
cylinders
may have oilfactors
on the
duty cycles,
and environmental
pressure
side only, while
the opposite
(external temperature
or contaminants).
side
is open to
atmosphere.
Therefore,
Customized
Solutions
for Unique
the
piston seal must leave minimal oil
Applications
film
when passing
along
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Performance
issues
in specific
applibore
since
of oil
othcations
are the
not transportation
always solved with
a stanerwise
result
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to For
the
dard orwould
catalog
range
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exterior.
difficult and constantly evolving fluid
In single-acting
the open
sealing
applications,cylinders,
seal engineers
can
end
mayapush
air out sealing
and draw
air in
develop
customized
solution.
as
the piston reciprocates.
Thissolution
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Development
of this custom
carry
and contaminants
shouldmoisture
include failure
analysis and into
systhe
which caninvestigations,
lead to seal
tem cylinder,
operating conditions
damage.
Vent filters
can be specificafitted to
testing according
to customer
the
open
side
of
the
cylinder
to
tions and performance standards reduce
as well
contaminants
entering the inside of the
as technical training.
cylinder.
Theseals
cylinder
may
also
Hydraulic
have abore
crucial
impact
be
hard
chromium
plated
to
prevent
on system performance in many applicorrosion.
cations. Factors such as temperatures,
speeds, pressures, lubricants, and other
Wiper
Sealsoperating conditions can
application
Hydraulic
cylinders
in a the
vagreatly
impact
seal life.operate
Specifying
riety
applications
and
environmental
right of
seal
helps boost
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mance, optimize
operations,
andto
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machine’s
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of ownership.
To prevent these contaminants from
entering
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and hyTHIS
MATERIAL
wasassembly
contributed
by
draulic
system,
wiper
seals
(also
known
Tadd McBride, Customized Molded
as scrapers,
excluders,
or dust
seals)Visit
are
Seals
engineering
manager
at SKF.
fitted on
the external
side ofseals
the cylinder
SKF’s
website
on hydraulic
at bit.ly/
head.
HP_SKF.
Wiper seals maintain sealing contact
to the piston
theofequipment
Table
Contentsis
☞ Backrodtowhen
stationary (static, no reciprocating motion of rod) and in use (dynamic, reciprocating rod). Without a wiper seal, the
retracting piston rod could transport
contaminants into the cylinder.
Guide Lubrication & Guide Rings
Rod guides are typically placed inward of both the rod and buffer seal
and should be lubricated on assembly
with the same medium as used in the
system. The guide must receive ample
lubrication at all times and should not
be outside the rod seal. However, in certain conditions, guides with polytetrafluoroethylene (PTFE) added may be
used outside the rod seal due to their
self-lubricating properties.
Guide rings provide effective guidance of components that are in relative
motion to each other and accommodate
☞ LEARN MORE @ hydraulicspneumatics.com | 30
FLUID POWER FUNDAMENTALS
Rotary Unions & Swivels
Even the best hose assemblies can fail prematurely
if adequate allowance for movement of machine
elements is not provided. An effective solution
uses a rotary union fitting to allow hoses to pivot,
which prevents or reduces stress from bending,
twisting, stretching, and kinking.
A
rotary union is the fitting used
to transfer the pressurized
fluid from a fixed inlet to rotating outlet, without obstructing the flow
of the fluid or air. This fitting is sometimes referred to as a swivel joint, rotating manifolds, or rotary coupling.
Figure 1 shows a swivel fitting containing a single circuit to transmit
fluid. These fittings pay for themselves
many times over by reducing stress on
a hose, thereby extending its life. This
describes one end of the swivel spectrum.
At the other end are rotary unions
that transmit fluid for multiple circuit
lines through a single manifold that
rotates continuously. In general, fluid
enters one or more ports in the stationary portion of the manifold and
exits through one or more ports on
Rotary unions often are used in applications such as this excavator fitted with a log grapple. A rotary union mounted between the turret and
track dive transmits hydraulic fluid between the rotating and non-rotating assemblies, respectively. The grapple also uses a swivel to allow
continuous 360-deg. rotation for hydraulic fluid and electrical power.
☞ LEARN MORE @ hydraulicspneumatics.com | 31
FLUID POWER FUNDAMENTALS
the other portion, which rotates with
the machine. A rotary seal between
the two halves contains the pressurized fluid, yet allows relative rotation
between the halves. For simplicity of
discussion, the term rotary union will
be used here as an all-inclusive term
to describe swivel fittings and rotating
manifolds.
The rotary seal is probably the most
critical part of the device, whether a
swivel fitting or rotary manifold. This
is because the seal between the rotating and stationary halves must be tight
enough to prevent leakage of pressurized fluid, while introducing as little
torque drag as possible. Torque drag is
a measure of the swivel joint’s resistance
to rotation.
These seals vary in complexity depending on the application. For simple
swivel fittings undergoing less than
360 deg. of rotation, the seal may be
little more than two machined surfaces
loaded against each other. Rotating
manifolds, however, may require ball
bearings and spring-loaded seals with
auxiliary loading by fluid pressure. If
the seal is not pressure balanced (fluid
pressure acting on opposite sides of the
seal), torque drag may increase with
fluid pressure.
As with any custom-engineered
product, manufacturers can supply a
swivel joint to meet virtually any specification. However, a variety of standard
swivels is available to keep costs reasonable.
Configurations
Most swivel fittings are standard
catalog items considered specialty fittings. However, depending on their
complexity and manufacturer, rotating
manifolds often are engineered items
that must be special-ordered, especially
if more than four independent flow
paths are required. Standard configurations of swivel joints include straight
through (where flow paths are coaxial)
and right-angle (where outlet ports are
perpendicular to inlet ports). A less
common design is the offset configuration, which is essentially a straightthrough design with a 90 deg. elbow at
each end.
Available space and fluid line routing generally determine which configuration should be used. Keep in
mind that axial length of a rotating
manifold increases with the number
of independent flow paths. In some applications, directional control valves
can be mounted on the rotating end of
the machine to allow routing only two
common flow paths (pressure and return) through the rotating manifold. In
this case, all valves connect to the common flow paths through a conventional
manifold or line fittings.
In some instances, a valve is built
into the rotating manifold to allow or
block fluid flow as the rotating member
advances through a revolution. Internal passageways open and close as the
Fluid ports
Rotating
member
Seals
Fluid
port
Fluid port
Rotating member
Seals
Wear ring
(bearing)
Stationary
member
Wear ring
(bearing)
Fluid
ports
Stationary member
Cutaway view of single port swivel fitting.
Cutaway view of the multi-port rotating manifold. Ball-bearing construction allows
device to accommodate side loading. Drain port prevents any fluid that leaks past
seals from pressurizing, thereby reducing cross-channel mixing of fluids.
☞ LEARN MORE @ hydraulicspneumatics.com | 32
FLUID POWER FUNDAMENTALS
manifold turns, allowing fluid to flow
only when the rotating member is in
certain positions—a setup that operates
much like a camshaft and cam follower.
As with a cam, this arrangement is not
as easy to reconfigure as using electrically actuated valves. However, it can be
very practical for applications that have
a repetitive, fixed operation— such as an
indexing table.
O t h e r c o n s i d e r at i o n s i n c l u d e
through holes and integral valves. A
hole through the center of the rotating
manifold may be necessary to provide
access for electrical lines, a shaft, or
other machine elements that must be
routed from the stationary member to
the rotating one.
Because improper mounting can
cause vibration, how the rotary union
is secured to the equipment is also
important. Proper support for the rotary union based its weight and center
of gravity must be provided. The location of a torque arm with respect to
the mounting flange also plays an important role because an inappropriate
location can transmit side load, which
can increase the torque drag or damage the seals of the rotary union. The
mounting arm and the torque arm can
be welded or bolted to the rotary union
depending upon the application. The
swivel joint should also be painted or
provided some other means to prevent
metal corrosion.
Type of Motion
Just as swivel joints and rotating
manifolds should exhibit minimal friction to allow free rotation, hoses and
piping should transmit as little external
load to the swivel joint or manifold as
possible unless the swivel joint is designed with adequate bearings to support external loads. Otherwise, seals
can wear prematurely and leak. In extreme cases, the rotating joint itself may
fracture.
Just the weight of components—
hoses, tube assemblies, and fittings—
may be substantial enough to transmit
an external load to the swivel. For example, the weight of a 10-ft section of
spiral-wound hose (and the weight of
the fluid in it) can easily be underestimated or overlooked altogether. However, it can transmit a substantial side
load or bending moment on a swivel
joint.
Size and Mounting
Obviously, the swivel joint must
have ports of the correct size and geometry to accommodate hose or tubing assemblies mating to it. Ensure
that enough room is available on the
equipment structure to accommodate the swivel joint. For swivel fittings—as with any fitting—the higher
the flow rating, the larger the ID and
external envelope of the fitting. For
rotating manifolds, enough clearance
must be provided between ports to allow threading and unthreading hoseand tube-end fittings to the manifold.
Also keep in mind the physical size of
the rotating manifold. The more fluid
lines routed through the manifold, the
longer its axial length. The greater the
flow through the manifold, the larger
its required OD.
A means must exist to either mount
the swivel to the structure or to mount
the connecting hose and/or tubing to
the structure adjacent to the swivel joint.
This practice helps prevent misalignment from long runs of unsupported
hose or tubing. Misalignment can transmit side loads to the swivel, causing the
detrimental effects outlined above. Side
loading can also be introduced by forcing misaligned rigid tubing into position for mounting. The assembly may fit
together, but life and performance of the
swivel joint may suffer.
cause fluid leakage by pushing fluid past
the joint’s rotary seals.
Excessive pressure can also increase
friction, leading to premature wear and
higher torque drag. Excessive torque
drag can damage hoses because motion
is transmitted to the hose instead of the
swivel. The seals also play an important
role in controlling friction, as harder
material seals increase the friction but
are more stable at higher pressure and
temperature. However, seals made out
of a softer material may reduce friction,
but they are not stable at high pressure
and temperatures.
Also ensure that the swivel joint is
compatible with the application environment—the chemical composition
of the fluid being used, its temperature,
and the external environment. Swivel
joints are readily available in steel, stainless steel, brass, and other popular materials to match the chemistry and temperature of the fluid and surroundings.
Perhaps more importantly, a variety of
seal materials is available to accommodate virtually any hydraulic fluid at virtually any temperature.
Whenever p ossible, mount the
swivel joint where it will have minimal exposure to abrasive or corrosive
particles. In some applications, an elastomeric boot, bellows, or cover may be
necessary to help isolate the seal area of
the swivel joint from an extremely dirty
environment.
VINAY PATIL is a mechanical design
engineer at United Equipment Accessories.
For more information, call or visit www.ueainc.com.
Selection Considerations
When selecting swivel joints, not adhering to manufacturers’ specifications
can result in leakage, premature failure
of the joint, premature failure of hose, or
all of these conditions. Exceeding manufacturers’ published pressure ratings can
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FLUID POWER FUNDAMENTALS
Water Hydraulics
Water hydraulics combines the high-power density
of hydraulics with the clean and fire-proof operation
of water. But water’s inherent physical properties
pose some design challenges.
W
ater-based hydraulic systems traditionally have been
used in longwall mining
applications and in hot-metal areas of
steel mills. The obvious advantage of
water systems in these industries is their
fire resistance. Water-based hydraulic
systems also have not-so-obvious cost
advantages over oil-based fluid. First,
non-toxic, biodegradable synthetic additives cost much less per gallon of fluid
than oil-based fluids do.
Considering the costs associated
with preventing and cleaning up environmental contamination, water-based
hydraulic systems hold the potential for
tremendous cost savings at the plant
level. Oil that has leaked or been drained
from a system can’t just be dumped
down the drain. It must be collected,
properly contained, and hauled away by
a certified carrier—an expensive propo-
sition. Water containing synthetic additives, however, usually can be dumped
into plant effluent systems.
Cost savings at the plant level don’t
stop at the lower cost of the fluid and its
disposal. Because water-based hydraulic
fluid consists of 10 parts water and one
part synthetic additive, 5 gal of additive
mixes with water to make 100 gallons of
water-based fluid. A 50-gal container is
certainly easier to handle than two 55gal drums, so warehousing is simpler,
cleaner, and less cluttered. Transportation costs also are lower.
Other potential plant-wide savings
include improved safety for workers because the water-based fluid is non-toxic
and non-flammable. These attributes
can reduce plant insurance rates. Spills
cost less to clean up because granular
absorbents or absorbent socks are unnecessary.
Fighting Freeze
Water-based hydraulic systems do, of
course, have limits to their applications.
One is the potential of freezing. This
possibility is probably the most significant blockade to more widespread application of water-based systems for mobile equipment. By far, longwall mining
is the largest sector of mobile equipment
that has been able to take advantage of
water-based systems. Temperatures underground do not approach the freezing
point of water, and fire resistance is essential. Mobile and even marine equipment used in temperate climates could
cash in one the advantages of waterbased systems, but there is no guarantee
that such equipment always will be used
in above-freezing temperatures.
Nevertheless, adding an antifreeze
to a water-based fluid can depress its
freezing temperature to well below 32°F.
Most components used in water hydraulic systems, such as these cylinders, make extensive use of stainless steel for its strength and high
corrosion resistance. (Courtesy: The Water Hydraulics Co. Ltd.)
☞ LEARN MORE @ hydraulicspneumatics.com | 34
FLUID POWER FUNDAMENTALS
Ethylene glycol—used in automotive
antifreeze—is toxic and is not biodegradable, so its use for antifreeze in
water-based hydraulic fluid would defeat the environmental advantage water-based fluid has. Propylene glycol is
an alternative, which is not toxic and is
biodegradable. It costs more than ethylene glycol and is not quite as effective an
antifreeze, so it must be used in slightly
higher concentrations.
Two additional techniques to reduce
freezing potential are to keep fluid circulating continuously and use hose
where practical. Hose insulates fluid
from exterior temperatures better than
metal tubing does.
Sealing the System
Two more perceived problems with
water hydraulic systems are bacterial
infestation and difficulty in maintain
proper concentrations. Sealing the system from atmosphere can hold bacterial
growth in check. Addition of an antibacterial agent to the fluid can have a
lasting effect on preventing bacterial
buildup if air is excluded from the system. Sealing the system from the atmosphere also keeps out most airborne
contaminants—a common cause of
component failure.
A sealed reservoir eliminates another
problem suffered by many hydraulic systems: water ingression. Dissolved suspended water contaminates hydraulic
oil. The only detriment water ingression
has in a water-based system, though, is
that is alters the concentration of additive. Water ingression is still undesirable, but its occurrence is far less detrimental in a water-based system than in
one using oil.
This addresses another perceived
issue with water-based systems: water-based systems must be closely
monitored to ensure that the additive
concentration stays within tolerance.
That is because water evaporates from
the reservoir more readily than the additive does. Consequently, water evaporation causes the additive concentration
Water-hydraulic systems are widely used in rolling mills and other hot-metal
applications where the fireproof nature of water provides the highest level of safety.
to increase. When new fluid is added to
a system, samples of the existing fluid
must be taken to determine the concentration of additive in solution. These results then reveal the ratio of additive to
fluid that must be added so that fluid
concentration is correct.
With a system sealed from the atmosphere, the evaporation problem is virtually eliminated. Any fluid that leaks
out is a solution containing water and
additive. Therefore, the quantity of fluid
in the system changes, but concentration
does not. System fluid is replenished
simply by adding a pre-mixed solution
of water and additive to the reservoir.
Special Considerations
An important consideration for water-based systems is that major components should be designed specifically for
use with water fluid, not just modified
from versions originally intended for oil
service. An oil valve retrofitted for water
service may work, but its compromise
in performance will be obvious when
compared to a valve designed for water
service.
Tubing, hose, and fittings usually
can be identical to those for oil systems.
Pumps, valves, and actuators for water
service, however, exhibit some significant differences from components for
oil systems. Pump gears, for example,
should be made of super-hard alloys to
resist wear. A pump’s gear face should be
wider than that of an oil pump because
water’s low viscosity requires a larger
area to form an adequate lubricant film.
Cylinders used in water systems should
have bronze-clad pistons to minimize
wear between pistons and cylinder
walls. Spring- or O-ring-energized seals
should be used to minimized leakage
across the piston.
Valves for Water
Valves for water-based fluid usually
are packed with seals separating metal
parts to prevent metal-to-metal contact.
This is because water—even with lubricant additives—does not provide the
full-film lubrication of oil. In valves for
oil service, lapped spools can be used because oil forms a film on metal components to keep surfaces separated. Metal
surfaces in relative motion in valves for
water-based fluid are separated by bearing-type materials. Moreover, because of
its much lower viscosity, water can readily leak through the clearances found in
non-packed valves for oil service.
Valves for water service also are
slightly larger than those for oil. This may
be another reason why water-based systems have not gained wide acceptance.
Originally, the larger size of components
for water-based fluid created a handicap
when designing systems, and more costly
construction inflated prices of valves
for water-based fluid to three times or
more that of valves for oil. Now, however,
☞ LEARN MORE @ hydraulicspneumatics.com | 35
FLUID POWER FUNDAMENTALS
FLUID POWER FUNDAMENTALS
valve sizes are comparable to those for
valve
sizesvalves
are comparable
those
for
oil. Many
are availabletowith
stanoil.
Many
valves
are
available
with
standard footprints. The price differential has
dard
footprints.
price differential
has
also become
less.The
Components
for wateralso
become
less.
Components
for
waterbased fluid still may cost more than those
based
still may
costmay
more
those
for oil fluid
systems,
but this
bethan
a bargain
for
oil you
systems,
but this
be a bargain
when
consider
themay
cost-saving
powhen
you
consider
the
cost-saving
potential of water-based systems.
tential
of
water-based
systems.
Cartridge valves that fit into cast, ducCartridge
valves
into cast,asductile-iron
bodies
alsothat
arefitavailable,
are
tile-iron
bodies
also
are
available,
as
are
lapped-spool versions of interchangeable
lapped-spool
versions
of
interchangeable
cartridges. Special materials are used incartridges.
Special
are used
instead of seals
whenmaterials
proportional
control
stead
of seals
whenseals
proportional
control
is needed,
because
can promote
unis
needed,
because
seals
can
promote
unacceptable stick-slip operation.
acceptable
stick-slip
operation.
The spool in a valve for oil service
in ainvalve
for oil
service
canThe
ridespool
directly
the valve
body.
Procan
ride
directly
in
the
valve
body.
Proportional valves for water-based fluid,
portional
valves
for
water-based
fluid,
though, often have a spool that rides in
though,
often
have aofspool
rides
in
a cast sleeve
instead
in thethat
valve
body.
aThe
castsleeve
sleeve wears
insteadbecause
of in the it
valve
body.
is softer
The
wears
because
is softer
than sleeve
the spool.
Both
sleeveitand
spool
than
the
spool.
Both
sleeve
and
are hardened to RC 6-72 to reducespool
wear
are
hardened
to
RC
6-72
to
reduce
rates. Valves for water-based fluidwear
also
rates.
Valveslands
for water-based
fluid also
have longer
to reduce leakage.
have longer lands to reduce leakage.
Preventing Leaks
Preventing
Leaks
Leakage continues
to be a nagging
Leakage
continues
to be
a nagging
problem
in many
hydraulic
systems.
New
problem
in
many
hydraulic
systems.
New
seal materials and designs and O-ring
seal
materials
and
designs
and
O-ring
face-seal fittings are powerful weapons in
face-seal
powerful
weapons
the battlefittings
againstare
leakage.
But the
battlein
is
the
battleover
against
leakage.
But the battle is
far from
because
of misapplication,
far
from over
because oformisapplication,
improper
installation,
simple lack of
improper
installation,
or simple
of
understanding. Although
there’s lack
no exunderstanding.
there’s no
excuse for leakageAlthough
in most systems,
it still
cuse
for
leakage
in
most
systems,
it
still
occurs. Assuming that leakage will not
occurs.
Assuming
that
leakage
not
be eliminated
in the
near
future,will
waterbe
eliminated
in
the
near
future,
waterbased fluid can dramatically reduce the
based
fluid can with
dramatically
costs associated
leakage. reduce the
costs
associated
with
leakage.
Internal leakage can
be just as wasteful.
Internal
leakage
can
be just
as wasteful.
For example, lapped-spool
valves
are deFor example,
valves are
designed
to leaklapped-spool
because the leakage
creates
signed
to
leak
because
the
leakage
creates
the oil film necessary to lubricate movthe
oil film
necessary
lubricate
moving parts.
This
leakage to
can
carburize
the
ing by
parts.
This leakage
carburize
the
oil
generating
heat.can
Internal
leakage
oil
by
generating
heat.
Internal
leakage
typically is routed back to tank, so this
typically
routed back
to tank, so
this
techniqueistransforms
mechanical
energy
technique
transforms
mechanical
energy
into heat instead of useful work. Using a
into
heatsteel
instead
of useful
work.seals
Using
stainless
spool
with PTFE
in aa
stainless
steel
spool
with
PTFE
seals
valve for water-based fluid eliminatesin
thea
valve for water-based fluid eliminates the
RESERVOIR DESIGN
RESERVOIR DESIGN
A SEALED RESERVOIR must allow the fluid level to rise and fall without allowing air to
A
SEALED enter
RESERVOIR
allow
the fluidcan
level
to rise and falla without
airbut
to
repeatedly
and exit.must
Several
methods
accommodate
variableallowing
fluid level,
repeatedly
enter
and exit. approach
Several methods
can accommodate
a variable
fluid
level,
a simple and
inexpensive
uses a breather
and two check
valves,
each
withbut
a
a
simple
and
inexpensive
approach
uses
a
breather
and
two
check
valves,
each
with a
different spring rate.
different
With a spring
sealed rate.
system, fluid level is highest at initial startup, before fluid has been
With
a
sealed
system, When
fluid level
highest
initial initially,
startup,airbefore
been
pumped to the system.
the is
system
is at
started
entersfluid
thehas
reservoir
pumped
the system.
When
the the
system
is started
airbeen
enters
the reservoir
through atobreather
as fluid
leaves
reservoir.
Afterinitially,
fluid has
circulated
through
through
a
breather
as
fluid
leaves
the
reservoir.
After
fluid
has
been
circulated
through
the system and returns to the reservoir, air is not allowed to exit through the breather.
the system and returns to the reservoir,
air is not
to exit
through
the breather.
Instead,
theallowed
air pocket
becomes
pressurized.
Instead,
the
air level
pocket
becomes
When the
fluid
rises
further, pressurized.
pressure of the
When
the
fluid
level
rises
further,
pressure
ofthis
the
air pocket eventually will reach
3 to
5 psi. At
air pocketair
eventually
reach through
3 to 5 psi.
At this
pressure,
exits thewill
reservoir
a check
pressure,
air
exits
the
reservoir
through
a
check
valve to avoid overpressurizing the reservoir.
valve
to avoid
overpressurizing
Pressure
in the
reservoir servesthe
thereservoir.
additional
Pressure
in
the
reservoir
serves
the
additional
function of precharging the main pump.
The posi-
function
of precharging
the line
main
pump. The
positive pressure
in the suction
prevents
pump
tive pressure
in the
linedrops,
prevents
pump
cavitation.
When
thesuction
fluid level
instead
of
cavitation.
When
level drops,
instead
of
drawing
in more
air,the
thefluid
air pocket
expands,
which
drawing
more air, pressure.
the air pocket
lowers
the in
precharge
Overexpands,
time, the which
only air in
lowers
theisprecharge
Over time, the only air in
the
system
that whichpressure.
entered initially.
the system is that which entered initially.
Reservoirs constructed of stainless steel are usually the best choice for water hydraulic
Reservoirs
constructed
stainless
steelresistance
are usuallyand
thestrength.
best choice for water hydraulic
systems because
of theirofhigh
corrosion
systems because of their high corrosion resistance and strength.
36
need for clearance between moving comneed
for clearance
moving
components.
Because between
there is no
clearance,
ponents.
Because
there
is
no
clearance,
there is no internal leakage.
there
is no internalvalves
leakage.
Packed-spool
eliminate leakPacked-spool
valves
eliminate leakage and the need for pilot-operated
age
and
the
need
for
pilot-operated
check valves. When the valve centers to
check
valves. When the
valve centers
to
an all-ports-blocked
condition,
pilotan
all-ports-blocked
condition,
pilotoperated checks are not needed to preoperated
checks
areIf not
to prevent cylinder
drift.
thereneeded
is no port-tovent cylinder
drift.
If therewill
is no
port-toport
leakage, the
cylinder
not
drift.
port
leakage,
the
cylinder
will
not
drift.
But beyond the obvious and intanBut
beyond
the
obvious
and
intangible costs of fluid leakage, disposgible
of fluid
leakage,
ing of costs
the fluid
that has
leakeddisposfrom a
ing
of
the
fluid
that
has
leaked
from a
system becomes a concern. Allowing
system
becomes
a
concern.
Allowing
hydraulic oil to enter plant effluent syshydraulic
oil to an
enter
plant effluent
systems becomes
expensive
propositemswhen
becomes
an expensive
proposition
removal
and disposal
costs
tion
when
removal
and
disposal
costs
are considered. Realizing that cleanup
are
considered.
Realizing
that
cleanup
and disposal costs will only go up, and
and disposal
only gosuggests
up, and
that
the price costs
of oil will
is unstable,
that
the
price
of
oil
is
unstable,
suggests
that water-based hydraulics can be an
that
water-based
hydraulics
can be an
economical
solution
to environmental
economical
solution
to
environmental
problems.
problems.
Accepting Water Hydraulics
Accepting
Even theWater
most Hydraulics
expensive water ad-
Evenbecome
the most
expensive
water
additives
attractive
when
designditives
become
attractive
when
designers realize that 1 gal of concentrate can
ers
realize
that
gal ofNo
concentrate
can
make
20 gal
of 1fluid.
wonder, then,
make
20 gal in
of water-based
fluid. No wonder,
that interest
fluids then,
often
that
interest
in
water-based
fluids
often
centers around cost-saving potential.
centers
around
cost-saving
potential.
However, designers must also realize
However,
designers
must also
realize
that they can’t
just change
the fluid
in
that
they
can’t
just
change
the
fluid
in
their systems from oil to water without
their
systems
from
oil
to
water
without
making other substantial changes.
making
substantial
changes. are
Whatother
are viewed
as disadvantages
What
are viewed
disadvantages
are
really
different
rulesas
that
apply to waterreally
different
rules
that
apply
to
waterbased hydraulic systems.
based
hydraulic
systems.resist learning
Designers
probably
Designers
probably resist
learning
more
about water-based
hydraulics
more
about
water-based
hydraulics
because they are intimidated by all the
because
they are
all thea
work required
to intimidated
learn how toby
design
work
required
to
learn
how
to
design
new system or retrofit an older system.a
new
systemtheir
or retrofit
system.
By closing
mindsan
to older
this different
By
closing
their
minds
to
this
different
technology, they may miss the many
technology,
they of
may
miss the many
other advantages
water-based
fluid
other advantages
ofNow
water-based
fluid
beyond
initial cost.
that environbeyond
initial
cost.
Now
that
environmental concerns have added disposal
mental
added disposal
costs toconcerns
the pricehave
of hydraulic
fluids,
costs to the price
of hydraulic
fluids,
water-based
hydraulics
have captured
water-based
hydraulics
have
captured
interest as an environmentally friendly
interest
solution.as an environmentally friendly
solution.
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