Hydraulics basic

Textbook TP 501
Festo Didactic
093281 en
Hydraulics
Basic Level
093281_cover_textbook_tp501_en.indd 1 11.04.2005 17:00:09

Order No.: 093281
Description: HYDRAUL.LEHRB
Designation: D:LB-TP501-1-GB
Edition: 11/2003
Author: D. Merkle, B.Schrader, M. Thomes
Graphics: D. Schwarzenberger
Layout: 25.11.2003, M. Göttfert, G. Heigl, W. Schreiner
© Festo Didactic GmbH & Co. KG, 73770 Denkendorf/Germany, 2003
Internet: www.festo.com/didactic
e-mail: did@festo.com
The copying, distribution and utilization of this document as well as the
communication of its contents to others without expressed authorization is
prohibited. Offenders will be held liable for the payment of damages. All rights
reserved, in particular the right to carry out patent, utility model or ornamental
design registration.

Table of contents
1 Tasks of a hydraulic installation __________________________________ 7
1.1 Stationary hydraulics____________________________________________ 8
1.2 Mobile hydraulics _____________________________________________ 10
1.3 Comparison of hydraulics with other control media __________________ 11
2 Fundamental physical principles of hydraulics _____________________ 13
2.1 Pressure _____________________________________________________ 13
2.2 Pressure transmission__________________________________________ 18
2.3 Power transmission ____________________________________________ 19
2.4 Displacement transmission _____________________________________ 21
2.5 Pressure transfer ______________________________________________ 23
2.6 Flow rate_____________________________________________________ 25
2.7 Continuity equation____________________________________________ 26
2.8 Pressure measurement _________________________________________ 30
2.9 Temperature measurement______________________________________ 31
2.10 Measurement of flow rate _______________________________________ 31
2.11 Types of flow _________________________________________________ 31
2.12 Friction, heat, pressure drop ____________________________________ 35
2.13 Energy and power _____________________________________________ 41
2.14 Cavitation____________________________________________________ 51
2.15 Throttle points ________________________________________________ 53
3 Hydraulic fluid________________________________________________ 57
3.1 Tasks for hydraulic fluids _______________________________________ 57
3.2 Types of hydraulic fluid _________________________________________ 58
3.3 Characteristics and requirements_________________________________ 59
3.4 Viscosity_____________________________________________________ 60
4 Components of a hydraulic system _______________________________ 67
4.1 Power supply section __________________________________________ 67
4.2 Hydraulic fluid ________________________________________________ 67
4.3 Valves_______________________________________________________ 68
4.4 Cylinders (linear actuators)______________________________________ 70
4.5 Motors (rotary actuators) _______________________________________ 71
© Festo Didactic GmbH & Co. KG • TP 501 3

Table of contents
5 Graphic and circuit symbols_____________________________________ 73
5.1 Pumps and motors_____________________________________________ 73
5.2 Directional control valves _______________________________________ 74
5.3 Methods of actuation __________________________________________ 75
5.4 Pressure valves _______________________________________________ 76
5.5 Flow control valves ____________________________________________ 78
5.6 Non-return valves _____________________________________________ 79
5.7 Cylinders ____________________________________________________ 80
5.8 Transfer of energy and conditioning of the pressure medium __________ 82
5.9 Measuring devices_____________________________________________ 83
5.10 Combination of devices_________________________________________ 83
6 Design and representation of a hydraulic system ___________________ 85
6.1 Signal control section __________________________________________ 86
6.2 Hydraulic power section ________________________________________ 87
6.3 Positional sketch ______________________________________________ 90
6.4 Circuit diagram________________________________________________ 91
6.5 Components plus technical data__________________________________ 92
6.6 Function diagram______________________________________________ 94
6.7 Function chart ________________________________________________ 95
7 Components of the power supply section__________________________ 97
7.1 Drive ________________________________________________________ 97
7.2 Pump _______________________________________________________ 99
7.3 Coupling____________________________________________________ 107
7.4 Reservoir ___________________________________________________ 107
7.5 Filters ______________________________________________________ 109
7.6 Coolers _____________________________________________________ 120
7.7 Heaters_____________________________________________________ 122
8 Valves______________________________________________________ 123
8.1 Nominal sizes________________________________________________ 123
8.2 Design _____________________________________________________ 125
8.3 Poppet valves________________________________________________ 126
8.4 Spool valves_________________________________________________ 127
8.5 Piston overlap _______________________________________________ 129
8.6 Control edges________________________________________________ 134
4 © Festo Didactic GmbH & Co. KG • TP 501

Table of contents
9 Pressure valves ______________________________________________ 137
9.1 Pressure relief valves _________________________________________ 137
9.2 Pressure regulators ___________________________________________ 144
10 Directional control valves______________________________________ 149
10.1 2/2-way valve _______________________________________________ 153
10.2 3/2-way valve _______________________________________________ 157
10.3 4/2-way valve _______________________________________________ 159
10.4 4/3-way valve _______________________________________________ 162
11 Non-return valves ____________________________________________ 167
11.1 Non-return valve _____________________________________________ 168
11.2 Piloted non-return valve _______________________________________ 172
11.3 Piloted double non-return valve _________________________________ 175
12 Flow control valves ___________________________________________ 179
12.1 Restrictors and orifice valves ___________________________________ 180
12.2 One-way flow control valve_____________________________________ 184
12.3 Two-way flow control valve_____________________________________ 185
13 Hydraulic cylinders___________________________________________ 193
13.1 Single-acting cylinder _________________________________________ 194
13.2 Double-acting cylinder ________________________________________ 196
13.3 End position cushioning _______________________________________ 199
13.4 Seals_______________________________________________________ 200
13.5 Types of mounting____________________________________________ 202
13.6 Venting_____________________________________________________ 202
13.7 Characteristics _______________________________________________ 203
13.8 Buckling resistance ___________________________________________ 205
13.9 Selecting a cylinder ___________________________________________ 207
14 Hydraulic motors_____________________________________________ 211

Table of contents
15 Accessories _________________________________________________ 215
15.1 Flexible hoses _______________________________________________ 217
15.2 Pipelines____________________________________________________ 223
15.3 Sub-bases __________________________________________________ 226
15.4 Bleed valves_________________________________________________ 228
15.5 Pressure gauges _____________________________________________ 229
15.6 Pressure sensors _____________________________________________ 230
15.7 Flow measuring instruments____________________________________ 231
16 Appendix ___________________________________________________ 233

1. Tasks of a hydraulic installation
Hydraulic systems are used in modern production plants and manufacturing
installations.
By hydraulics, we mean the generation of forces and motion using hydraulic fluids.
The hydraulic fluids represent the medium for power transmission.
The object of this book is to teach you more about hydraulics and its areas of
application. We will begin with the latter by listing the main areas for the application
of hydraulics.
The place held by hydraulics in (modern) automation technology illustrates the wide
range of applications for which it can be used. A basic distinction is made between:
• stationary hydraulics
• and mobile hydraulics
Mobile hydraulic systems move on wheels or tracks, for example, unlike stationary
hydraulic systems which remain firmly fixed in one position. A characteristic feature
of mobile hydraulics is that the valves are frequently manually operated. In the case
of stationary hydraulics, however, mainly solenoid valves are used.
Other areas include marine, mining and aircraft hydraulics. Aircraft hydraulics
assumes a special position because safety measures are of such critical importance
here. In the next few pages, some typical examples of applications are given to
clarify the tasks which can be carried out using hydraulic systems.
What do we mean
by hydraulics?

The following application areas are important for stationary hydraulics:
• Production and assembly machines of all types
• Transfer lines
• Lifting and conveying devices
• Presses
• Injection moulding machines
• Rolling lines
• Lifts
Machine tool construction is a typical application area.
Lathe
In modern CNC controlled machine tools, tools and work pieces are clamped by
means of hydraulics. Feed and spindle drives may also be effected using hydraulics.
1.1
Stationary hydraulics

Press with elevated reservoir

Typical application fields for mobile hydraulics include:
• Construction machinery
• Tippers, excavators, elevating platforms
• Lifting and conveying devices
• Agricultural machinery
There is a wide variety of applications for hydraulics in the construction machinery
industry. On an excavator, for example, not only are all working movements (such as
lifting, gripping and swivelling movements) generated hydraulically, but the drive
mechanism is also controlled by hydraulics. The straight working movements are
generated by linear actuators (cylinders) and the rotary movements by rotary
actuators (motors, rotary drives).
Mobile hydraulics
1.2
Mobile hydraulics

There are other technologies besides hydraulics which can be used in the context of
control technology for generating forces, movements and signals:
• Mechanics
• Electricity
• Pneumatics
It is important to remember here that each technology has its own preferred
application areas. To illustrate this, a table has been drawn up on the next page
which compares typical data for the three most commonly used technologies –
electricity, pneumatics and hydraulics.
This comparison reveals some important advantages of hydraulics:
• Transmission of large forces using small components, i.e. great
power intensity
• Precise positioning
• Start-up under heavy load
• Even movements independent of load, since liquids are scarcely
compressible and flow control valves can be used
• Smooth operation and reversal
• Good control and regulation
• Favourable heat dissipation
Compared to other technologies, hydraulics has the following disadvantages:
• Pollution of the environment by waste oil (danger of fire or accidents)
• Sensitivity to dirt
• Danger resulting from excessive pressures (severed lines)
• Temperature dependence (change in viscosity)
• Unfavourable efficiency factor
1.3
Comparison of hydraulics
with other control media

Electricity Hydraulics Pneumatics
Leakage Contamination No disadvantages apart from
energy loss
Environmental
influences
Risk of explosion in certain areas,
insensitive to temperature.
Sensitive in case of temperature
fluctuation, risk of fire in case of
leakage.
Explosion-proof,
insensitive to temperature.
Energy storage Difficult, only in small quantities
using batteries.
Limited, with the help of gases. Easy
Energy transmission Unlimited with power loss. Up to 100 m,
flow rate v = 2 – 6 m/s,
signal speed up to 1000 m/s.
Up to 1000 m,
flow rate v = 20 – 40 m/s,
signal speed 20 – 40 m/s.
Operating speed v = 0.5 m/s v = 1.5 m/s
Low High Very highPower supply costs
0.25 : 1 : 2.5
Linear motion Difficult and expensive, small
forces, speed regulation only
possible at great cost
Simple using cylinders, good speed
control, very large forces.
Simple using cylinders, limited
forces, speed extremely, load-
dependent.
Rotary motion Simple and powerful. Simple, high turning moment, low
speed.
Simple, inefficient, high speed.
Positioning accuracy Precision to ±1 µm and easier to
achieve
Precision of up to ±1 µm can be
achieved depending on
expenditure.
Without load change precision of
1/10 mm possible.
Stability Very good values can be achieved
using mechanical links.
High, since oil is almost
incompressible, in addition, the
pressure level is considerably
higher than for pneumatics.
Low, air is compressible.
Forces Not overloadable.
Poor efficiency due to downstream
mechanical elements.
Very high forces can be realized.
Protected against overload, with
high system pressure of up to 600
bar, very large forces can be
generated F < 3000 kN.
Protected against overload,
forces limited by pneumatic
pressure and cylinder diameter
F < 30 kN at 6 bar.

Hydraulics is the science of forces and movements transmitted by means of liquids.
It belongs alongside hydro-mechanics. A distinction is made between hydrostatics –
dynamic effect through pressure times area – and hydrodynamics – dynamic effect
through mass times acceleration.
Hydro-mechanics
Hydrostatic pressure is the pressure which rises above a certain level in a liquid
owing to the weight of the liquid mass:
pT
sT
= h ⋅ ρ ⋅ g
pT
sT
= hydrostatic pressure (gravitational pressure) [Pa]
h = level of the column of liquid [m]
ρ = density of the liquid [kg/mT
3T
]
g = acceleration due to gravity [m/sT
2T
]
In accordance with the SI international system of units, hydrostatic pressure is given
in both Pascal and bar. The level of the column of liquid is given the unit “metre”,
the density of the liquid “kilograms per cubic metre” and the acceleration due to
gravity “metres per second squared”.
2. Fundamental physical principles of hydraulics
2.1
Pressure
Hydrostatic pressure

The hydrostatic pressure, or simply “pressure” as it is known for short, does not
depend on the type of vessel used. It is purely dependent on the height and density
of the column of liquid.
Hydrostatic pressure
Column: h = 300 m
ρ = 1000 kg/mT
3T
g = 9.81 m/sT
2T
= 10 m/sT
2T
pT
ST
= h ⋅ ρ ⋅ g = 300 m ⋅ 1000 3
m
kg
⋅ 10 2
s
m
= 3 000 000 23
sm
mkgm
⋅
⋅⋅
= 3 000 000 2
m
N
pT
ST
= 3 000 000 Pa = 30 bar
Reservoir: h = 15 m
ρ = 1000 kg/mT
3T
g = 9.81 m/sT
2T
= 10 m/sT
2T
pT
ST
= h ⋅ ρ ⋅ g = 15 m ⋅ 1000 3
m
kg
⋅ 10 2
s
m
= 150 000 23
sm
mkgm
⋅
⋅⋅
= 150 000 2
m
N
pT
ST
= 150 000 Pa = 1,5 bar
Elevated tank: h = 5 m
ρ = 1000 kg/mT
3T
g = 9.81 m/sT
2T
= 10 m/sT
2T
pT
ST
= h ⋅ ρ ⋅ g = 5 m ⋅ 1000 3
m
kg
⋅ 10 2
s
m
= 50 000 23
sm
mkgm
⋅
⋅⋅
= 50 000 2
m
N
pT
ST
= 50 000 Pa = 0,5 bar

Every body exerts a specific pressure p on its base. The value of this pressure is
dependent on the force due to weight F of the body and on the size of the area A on
which the force due to weight acts.
A1 A2
F
F
Force, area
The diagram shows two bodies with different bases (AT
1T
and AT
2T
). Where the bodies
have identical mass, the same force due to weight (F) acts on the base. However, the
pressure is different owing to the different sizes of base. Where the force due to
weight is identical, a higher pressure is produced in the case of a small base than in
the case of a larger base (“pencil” or “concentrated” effect).
This is expressed by the following formula:
A
F
p =
Unit: 1 Pa = 1 2
m
N
2
m
N
1 bar = 100 000 = 10T
5T
Pa
p = Pressure Pascal [Pa]
F = Force Newton [N] 1 N = 1 2
s
mkg ⋅
A = Area Square metre [mT
2T
]
Rearrangement of the formula produces the formulae for calculating force and area:

A cylinder is supplied with 100 bar pressure, its effective piston surface is equal to
7.85 cmT
2T
. Find the maximum force which can be attained.
Given that: p = 100 bar = 1000 N/cmT
2T
A = 7.85 cmT
2T
F = p ⋅ A = 2
2
cm
cm85.7N1000 ⋅
= 7850 N
A lifting platform is to lift a load of 15 000 N and is to have a system pressure of
75 bar.
How large does the piston surface A need to be?
Given that: F = 15 000 N
P = 75 bar = 75 ⋅ 10T
5T
Pa
N
mN
002.0
Pa1075
N00015
p
F
A
2
5
⋅
=
⋅
== = 0.002 mT
2T
= 20 cmT
2T
Instead of making calculations it is possible to work with a diagram. The stiction in
the cylinder is not taken into consideration.
Given that: Force F = 100 kN
Operating pressure p = 350 bar.
What is the piston diameter?
Reading: d = 60 mm
Example
Example
Example

2.5
3
4
5
6
7
8
9
10
15
20
30
40
50
60
70
80
90
100
150
200
300
400
500
600
700
800
900
1000
1500
2000
3000
kN
Force
10
15
20
25
30
40
50
60
70
80
90
100
150
200
250
mm
400
Piston diameter
350 bar
300 bar
200 bar
160 bar
125 bar
100 bar
80 bar
50 bar
(5000 kPa)
Piston diameter, force and pressure

If a force FT
1T
acts via an area AT
1T
on an enclosed liquid, a pressure p is produced which
extends throughout the whole of the liquid (Pascal’s Law). The same pressure
applies at every point of the closed system (see diagram).
Pressure transmission
Owing to the fact that hydraulic systems operate at very high pressures, it is
possible to neglect the hydrostatic pressure (see example). Thus, when calculating
the pressure in liquids, the calculations are based purely on pressure caused by
external forces. Thus, the same pressure acts on the surfaces AT
2T
, AT
3T
as on AT
1T
. For
solid bodies, this is expressed by means of the following formula:
A
F
p =
Given that: AT
1T
= 10 cmT
2T
= 0.001 mT
2T
F = 10 000 N
22
m
N
00000010
m001.0
N00010
A
F
p === = 100 ⋅ 10T
5T
Pa (100 bar)
2.2
Pressure transmission
Example

Given that: P = 100 ⋅ 10T
5T
Pa
AT
2T
= 1 cmT
2T
= 0.0001 mT
2T
N1000
m
mN
1000m0001.0Pa10100ApF 2
2
25
=
⋅
=⋅⋅=⋅=
The same pressure applies at every point in a closed system. For this reason, the
shape of the container has no significance.
Power transmission
Where a container is formed as shown in the diagram, it is possible to transmit
forces. The fluid pressure can be described by means of the following equations:
1
1
1
A
F
p = and
2
2
2
A
F
p =
The following equation applies when the system is in equilibrium:
pT
1T
= pT
2T
When the two equations are balanced, the following formula is produced:
2
2
1
1
A
F
A
F
=
The values FT
1T
and FT
2 T
and AT
1T
and AT
2T
can be calculated using this formula.
Example
2.3
Power transmission

For example, FT
1T
and AT
2T
are calculated as shown here:
2
21
1
A
FA
F
⋅
= and
1
21
2
F
FA
A
⋅
=
Small forces from the pressure piston can produce larger forces by enlarging the
working piston surface. This is the fundamental principle which is applied in every
hydraulic system from the jack to the lifting platform. The force FT
1T
must be sufficient
for the fluid pressure to overcome the load resistance (see example).
A vehicle is to be lifted by a hydraulic jack. The mass m amounts to 1500 kg.
What force FT
1T
is required at the piston?
Power transmission
Given that: Load m = 1500 kg
Force due to weight FT
2T
= m · g = N00015
s
m
10kg1500 2
=⋅
Given that: AT
1T
= 40 cmT
2T
= 0.004 mT
2T
AT
2T
= 1200 cmT
2T
= 0.12 mT
2T
N500
m12.0
N00015m004.0
A
FA
F 2
2
2
21
1 =
⋅
=
⋅
=
Example

It has been proved that the force FT
1T
of 100 N is too great for actuation by hand lever.
What must the size of the piston surface AT
2T
be when only a piston force of FT
1T
= 100 N
is available?
2
2
1
21
2
2
21
1
m6.0
N100
N00015m004.0
F
FA
A
A
FA
F
=
⋅
=
⋅
=
⋅
=
If a load FT
2T
is to be lifted a distance sT
2T
in line with the principle described above, the
piston PT
1T
must displace a specific quantity of liquid which lifts the piston PT
2T
by a
distance sT
2T
.
Displacement transmission
The necessary displacement volume is calculated as follows:
VT
1TT
= TsT
1T
· AT
1T
and VT
2 T
= sT
2T
· A T
2
Since the displacement volumes are identical (V1 = V2), the following equation is
valid:
s1 · A1 = s2 · A2
From this it can be seen that the distance s1 must be greater than the distance s2
since the area A1 is smaller than the area A2.
Example
2.4
Displacement
transmission

The displacement of the piston is in inverse ratio to its area. This law can be used to
calculate the values s1 and s2. For example, for s2 and A1.
2
11
2
A
As
s
⋅
= and
1
22
1
s
As
A
⋅
=
Displacement transmission – example
Given that: A1 = 40 cmT
2T
A2 = 1200 cmT
2T
s1 = 15 cm
2
2
2
2
11
2 cm5.0
cm
cmcm
1200
4015
A
As
s =
⋅⋅
=
⋅
=
Given that: A2 = 1200 cmT
2T
s1 = 30 cm
s2 = 0.3 cm
2
2
2
22
1 cm12
cm
cmcm
30
12003.0
A
As
A =
⋅⋅
=
⋅
=

Pressure transfer
The hydrostatic pressure p1 exerts a force F1 on the area A1 which is transferred via
the piston rod onto the small piston. Thus, the force F1 acts on the area A2 and
produces the hydrostatic pressure p2. Since piston area A2 is smaller than piston
area A1, the pressure p2 is greater than the pressure p1. Here too, the following law
applies:
A
F
p =
From this, the following equations can be formulated for the forces F1 and F2:
F1 = p1 ⋅ A1 and F2 = p2 ⋅ A2
Since the two forces are equal (F1 = F2), the equations can be balanced:
P1 ⋅ A1 = p2 ⋅ A2
The values p1, A1 and A2 can be derived from this formula for calculations.
For example, the following equations result for p2 and A2:
2
11
2
A
Ap
p
⋅
= and
2
11
2
p
Ap
A
⋅
=
2.5
Pressure transfer

In the case of the double-acting cylinder, excessively high pressures may be
produced when the flow from the piston rod area is blocked:
Pressure transfer by double-acting cylinder
Given that: P1 = 10 ⋅ 10T
5T
Pa
A1 = 8 cmT
2
T= 0.0008 mT
2T
A2 = 4.2 cmT
2
T= 0.00042 mT
2T
)bar19(Pa1019
mm
mN
00042.0
0008.01010
A
Ap
p 5
22
25
2
11
2 ⋅=
⋅
⋅⋅⋅
=
⋅
=
Given that: p1 = 20 ⋅ 10T
5T
Pa
p2 = 100 ⋅ 10T
5T
Pa
A1 = 8 cmT
2
T= 0.0008 mT
2T
22
2
5
5
2
11
2 cm6.1m00016.0
Pa
mPa
10100
0008.01020
p
Ap
A ==
⋅
⋅
⋅⋅
=
⋅
=

Flow rate is the term used to describe the volume of liquid flowing through a pipe in
a specific period of time. For example, approximately one minute is required to fill a
10 litre bucket from a tap. Thus, the flow rate amounts to 10 l/min.
Flow rate
In hydraulics, the flow rate is designated as Q. The following equation applies:
t
V
Q =
Q = Flow rate [mT
3T
/s]
V = Volume [mT
3T
]
t = time [s]
The equations for the volume (V) and the time (t) can be derived from the formula for
the flow rate. The following equation is produced:
V = Q · t
2.6
Flow rate

Given that: Q = 4.5 l/s
t = 10 s
V = Q ⋅ t =
smin
minsl
60
102.4
⋅
⋅⋅⋅
= 0.7 l
A flow rate of 4.2 litres per minute produces a volume of 0.7 litres in 10 seconds.
Given that: V = 105 l
Q = 4.2 l/min
t =
l
minl
2.4
105
Q
V ⋅
= = 25 min
25 minutes are required to transport a volume of 105 litres at a flow rate of 4.2 litres
per minute.
If the time t is replaced by s/v (v = s/t) in the formula for the flow rate (Q = V/t) and
it is taken into account that the volume V can be replaced by A⋅s, the following
equation is produced:
Q = A · v
Q = Flow rate [mT
3T
/s]
v = Flow velocity [m/s]
A = Pipe cross-section [mT
2T
]
From the formula for the flow rate, it is possible to derive the formula for calculating
the pipe cross-section and flow velocity. The following equation applies for A or v.
v
Q
A = results in
A
Q
v =
Example
Result
Example
Result
2.7
Continuity equation

Given that: Q = 4.21 l/min =
s
m
1007.0
s60
dm2.4 3
3
3
−
⋅=
v = 4 m/s
ms
sm
4
1007.0
v
Q
A
33
⋅
⋅⋅
==
−
= 0.00002 mT
2T
= 0.2 cmT
2T
To achieve a flow velocity of 4 m/s with a flow rate of 4.2 l/min, a pipe cross-section
of 0.2 cmT
2T
is required.
Given that: Q = 4.2 l/min = 0.07 ⋅ 10T
-3T
mT
3T
/s
A = 0.28 cmT
2T
= 0.28 ⋅ 10T
-4T
mT
2T
s/m5.2
s
m
10
28.0
7.0
ms
m
1028.0
1007.0
A
Q
v 1
2
3
4
3
=⋅=
⋅⋅
⋅
== −
−
In a pipe with a cross-section of 0.28 cmT
2T
, a flow rate of 4.2 l/min brings about a flow
velocity of 2.5 m/s.
A
s
Cylinder
If in the formula for the flow rate
t
V
Q =
the volume replaced by the displacement volume V
V = A ⋅ s results in
t
sA
Q
⋅
=
Example
Result
Example
Result

Given that: A = 8 cmT
2T
s = 10 cm
t = 1 min
min
cm
08.0
min
cm
80
min
cmcm
1
108
t
sA
Q
332
==
⋅⋅
=
⋅
=
If a cylinder with a piston surface of 8 cmT
2T
and a stroke of 10 cm is to extend in one
minute, the power supply must generate a flow rate of 0.08 l/min.
The flow rate of a liquid in terms of volume per unit of time which flows through a
pipe with several changes in cross-section is the same at all points in the pipe (see
diagram). This means that the liquid flows through small cross-sections faster than
through large cross-sections. The following equation applies:
Q1 = A1 ⋅ v1 Q2 = A2 ⋅ v2 Q3 = A3 ⋅ v3 etc.…
As within one line the value for Q is always the same, the following equation of
continuity applies:
A1 ⋅ v1 = A2 ⋅ v2 = A3 ⋅ v3 = etc...
Time (t)
A1
A3
A2
QQ
s1 s3
s2
Flow rate
Example
Result

Given that: v1 = 4 m/s
v2 = 100 m/s
A1 = 0.2 cmT
2 T
= 0.2 ⋅ 10T
-4T
mT
2T
A2 = 0.008 cmT
2 T
= 0.008 ⋅ 10T
-4T
mT
2T
Q = A ⋅ v
Q1 = 0.2 ⋅ 10T
-4T
mT
2T
⋅ 4 m/s
Q2 = 0.008 ⋅ 10T
-4T
mT
2T
⋅ 100 m/s
Q = 0.8 ⋅ 10T
-4T
mT
3T
/s
A2
A1
V1
V2
Cylinder
Given that:
Pump delivery
s
cm
60
1010
min
cm
1010
min
dm
10
min
l
10Q
333
3
3
⋅
=⋅===
Inlet internal diameter d1 = 6 mm
Piston diameter d2 = 32 mm
To be found: Flow velocity v1 in the inlet pipe
Extension speed v2 of the piston
Q = v1 ⋅ A1 = v2 ⋅ A2
s
m
95.5
s
cm
595
scm
cm
28.060
1010
cm28.0
s60
cm1010
A
Q
v
cm0.8
4
cm2.3
4
d
A
cm28.0
4
cm6.0
4
d
A
2
33
2
33
1
1
2
222
2
2
222
1
==
⋅
=
⋅
⋅
=
⋅
==
=
π⋅⋅
=
π⋅
=
=
π⋅⋅
=
π⋅
=
s
m
21.0
s
cm
8.20
scm
cm
860
1010
cm8
s60
cm1010
A
Q
v 2
33
2
33
2
2 ==
⋅
=
⋅
⋅
=
⋅
==
Example
Example

To measure pressures in the lines or at the inputs and outputs of components, a
pressure gauge is installed in the line at the appropriate point.
A distinction is made between absolute pressure measurement where the zero point
on the scale corresponds to absolute vacuum and relative pressure measurement
where the zero point on the scale refers to atmospheric pressure. In the absolute
system of measurement, vacuums assume values lower than 1, in the relative
system of measurement, they assume values lower than 0.
4 3
0 -1
1 0
2
Atmospheric pressure
Vacuum
Absolute
pressure measurement
Relative
pressure measurement
Pressure above
atmospheric pressure
p in barabs
p = general pressure
p = absolute pressure
p = relative pressure
abs
e
Measure-
ment scale
p in bare
Measurement
scale
1
3 2
Absolute pressure, relative pressure
p
bar
7
p = 4 bare
p = 5 barabs
p = -0.3 bare
± 5% atmospheric approx.
p = 0.7 barabs
3
0
4
1
5
2
Example
2.8
Pressure measurement

The temperature of hydraulic fluid in hydraulic installations can either be measured
using simple measuring devices (thermometers) or else by means of a measuring
device which sends signals to the control section. Temperature measurement is of
special significance since high temperatures (> 60 degrees) lead to premature
ageing of the hydraulic fluid. In addition, the viscosity changes in accordance with
the temperature.
The measuring devices may be installed in the hydraulic fluid reservoir. o keep the
temperature constant, a pilotherm or thermostat is used which switches the cooling
or heating system on as required.
The simplest method of measuring flow rate is with a measuring container and a
stop watch. However, turbine meters are recommended for continuous
measurements. The speed indicated provides information about the value of the
flow rate. Speed and flow rate behave proportionally.
Another alternative is to use an orifice. The fall in pressure recorded at the orifice is
an indication of the flow rate (pressure drop and flow rate behave proportionally),
measurement by orifice is scarcely influenced by the viscosity of the hydraulic fluid.
A distinction is made between laminar and turbulent flow.
vm
laminar turbulent
vmax
Laminar and turbulent flow
2.9
Temperature measurement
2.10
Measurement of flow rate
2.11
Types of flow

In the case of laminar flow, the hydraulic fluid moves through the pipe in ordered
cylindrical layers. The inner layers of liquid move at higher speeds than the outer
layers. If the flow velocity of the hydraulic fluid rises above a certain point (known as
the critical speed), the fluid particles cease to move in ordered layers. The fluid
particles at the centre of the pipe swing out to the side. As a result, the fluid
particles affect and hinder one another, causing an eddy to be formed; flow becomes
turbulent. As a consequence of this, power is withdrawn from the main flow.
A method of calculating the type of flow in a smooth pipe is enabled by the
Reynolds’ number (Re). This is dependent on
• the flow velocity of the liquid v (m/s)
• the pipe diameter d (m)
• and the kinetic viscosity ν (m2/s)
ν
⋅
=
dv
Re
The physical variable “kinematic viscosity” is also referred to simply as “viscosity”.
A value for Re calculated with this formula can be interpreted as follows:
• laminar flow: Re < 2300
• turbulent flow: Re > 2300
The value 2300 is termed the critical Reynolds’ number (Recrit) for smooth round
pipes.
Turbulent flow does not immediately become laminar on falling below (Recrit).
The laminar range is not reached until 1/2 (Recrit).

80
1
40
4
5
30
10
5
6
7
8
10
15
20
30
40
50
60
70
80
100
20
20
4
30
3
50
10
100
2
5
200
300
1
50
100
200
500
1000
2000
5000
2 • 10
4
3 • 10
4
10
4
Pipe
diameter
d
Flow
velocity
of the
liquid
ν
Reynolds'
number
Re
Flow
rate
Q
[mm] [cSt = 10 m /s]
-6 2
[dm /min]
3
[-]
3
1
50
3
60
2
70
Determining of the Reynolds’ number (Prof. Charchut)
Q = 50 dmT
3T
/min
d = 25 mm
ν = 36 cSt
Re = 1165
The critical velocity mentioned above is the velocity at which the flow changes from
laminar to turbulent.
d
2300
d
Re
v crit
krit
ν
=
ν⋅
=
Example

To prevent turbulent flow causing considerable friction losses in hydraulic systems,
(Recrit ) should not be exceeded.
The critical speed is not a fixed value since it is dependent on the viscosity of the
hydraulic fluid and the diameter of the pipe. Therefore, empirically determined
values are generally used in practice. The following standard values for vcrit are valid
for the flow velocity in lines.
• Pressure line: to 50 bar operating pressure: 4.0 m/s
to 100 bar operating pressure: 4.5 m/s
• Suction line: 1.5 m/s
• Return line: 2.0 m/s
Types of flow

Given that: v1 = 1 m/s v3 = 4 m/s v4 = 100 m/s
ν = 40 mmT
2T
/s
d1 = 10 mm d3 = 5 mm d4 = 1 mm
The type of flow at cross-sections A1, A3, A4 is to be found.
2500
mm40s
smm1mm000100
Re
500
mm40s
smm5mm4000
Re
250
mm40s
smm10mm1000
Re
dv
Re
24
23
21
1
=
⋅
⋅⋅
=
=
⋅
⋅⋅
=
=
⋅
⋅⋅
=
ν
⋅
=
The flow is only turbulent at cross-section A4 since 2500 > 2300. The flow becomes
laminar again at cross-section A3 after the throttling point as 500 < 1150. However,
this is only after a steadying period.
Friction occurs in all devices and lines in a hydraulic system through which liquid
passes.
This friction is mainly at the line walls (external friction). There is also friction
between the layers of liquid (internal friction).
The friction causes the hydraulic fluid, and consequently also the components, to be
heated. As a result of this heat generation, the pressure in the system drops and,
thus, reduces the actual pressure at the drive section.
The size of the pressure drop is based on the internal resistances in a hydraulic
system. These are dependent on:
• Flow velocity (cross-sectional area, flow rate),
• Type of flow (laminar, turbulent),
• Type and number of cross-sectional reductions in the system of lines (throttles,
orifices),
• Viscosity of the oil (temperature, pressure),
• Line length and flow diversion,
• Surface finish,
• Line arrangement.
Example
Result
2.12
Friction, heat,
pressure drop

The flow velocity has the greatest effect on the internal resistances since the
resistance rises in proportion to the square of the velocity.
1
0
10 5m/s
v
432
2
3
4
5
6
7
8
9
10
11
12
13
14
bar
16
p
Influence of flow velocity on pressure loss

The friction between the flowing layers of liquid and the adhesion of the liquid to the
pipe wall form a resistance which can be measured or calculated as a drop in
pressure.
Since the flow velocity has an influence on the resistance to the power of two, the
standard values should not be exceeded.
Flow resistance in pipelines per 1 m length
For hydraulic fluid with ρ=850 kg/mT
3T
(K) at approx. 15 °C (ν = 100 mmT
2T
/s); (W) at approx. 60 °C (ν = 20 mmT
2T
/s)
v (m/s) 0.5 1 2 4 6
d (mm) K W K W K W K W K W
Re 30 150 60 300 120 600 240 1200 360 1800
λ 2.5 0.5 2.25 0.25 0.625 0.125 0.312 0.0625 0.21 0.04
6
∆p
bar/m
0.44 0.09 0.88 0.177 1.77 0.35 3.54 0.70 5.3 1.02
Re 50 250 100 500 200 1000 400 2000 600 3000
λ 1.5 0.3 0.75 0.15 0.375 0.075 0.187 0.037 0.125 0.043
10
∆p
bar/m
0.16 0.03 0.32 0.064 0.64 0.13 1.27 0.25 1.9 0.65
Re 100 500 200 1000 400 2000 800 4000 1200 6000
λ 0.75 0.15 0.375 0.075 0.187 0.037 0.093 0.04 0.062 0.036
20
∆p
bar/m
0.04 0.008 0.08 0.016 0.16 0.03 0.32 0.136 0.47 0.275
Re 150 750 300 1500 600 3000 1200 6000 1800 9000
λ 0.5 0.1 0.25 0.05 0.125 0.043 0.062 0.036 0.042 0.032
30
∆p
bar/m
0.017 0.003 0.035 0.007 0.07 0.024 0.14 0.082 0.214 0.163
Flow resistance
in pipelines

Flow resistance in pipelines per 1 m length (Continuation)
For hydraulic fluid with ρ=850 kg/mT
3T
(K) at approx. 15 °C (ν=100 mmT
2T
/s); (W) at approx. 60 °C (ν=20 mmT
2T
/s)
v (m/s) 0.5 1 2 4 6
d (mm) K W K W K W K W K W
Re 200 1000 400 2000 800 4000 1600 8000 2400 12000
λ 0.375 0.075 0.187 0.037 0.093 0.04 0.047 0.033 0.045 0.03
40
∆p
bar/m
0.01 0.002 0.02 0.004 0.04 0.017 0.08 0.056 0.172 0.114
Re 250 1250 500 2500 1000 5000 2000 10000 3000 15000
λ 0.3 0.06 0.15 0.045 0.075 0.037 0.037 0.031 0.043 0.028
50
∆p
bar/m
0.006 0.001 0.013 0.004 0.025 0.012 0.05 0.042 0.13 0.085
Re 300 1500 600 3000 1200 6000 2400 12000 3600 18000
λ 0.25 0.05 0.125 0.043 0.062 0.036 0.045 0.03 0.04 0.027
60
∆p
bar/m
0.004 0.0008 0.009 0.003 0.017 0.01 0.05 0.034 0.1 0.007
A flow with a velocity of v = 0.5 m/s flows through a pipeline with a nominal width of
6 mm.
The kinematic velocity amounts to = 100 mmT
2T
/s at 15 °C.
The density ρ= 850 kg/mT
3T
.
Calculate the pressure loss ∆p for 1 m length.
2
v
2d
l
p ⋅
ρ
⋅⋅λ=∆
Figure for resistance of pipes
Re
75
=λ (resistance value)
In order to calculate the friction value λ, it is first necessary to calculate the
Reynolds’ number Re:
ν
⋅
=
dv
Re
Example for calculating
the values in the table

Given that: ν = 100 mmT
2T
/s = 1 ⋅ 10T
-4T
mT
2T
/s
d = 6 mm = 0.006 m
v = 0.5 m/s
30
101
006.05.0
Re 4
=
⋅
⋅
= −
(comp. with table)
Figure for resistance of pipes 5.2
30
75
Re
75
===λ (comp. with table)
table)with(comp.bar4427.0m/N44270p
sm
mkg
44270)s/m5.0(
m2
kg850
mm6
mm1000
5.2v
2d
l
p
2
22
2
3
2
==∆
⋅
⋅
=⋅⋅⋅=⋅
ρ
⋅⋅λ=∆
bar1bar10
m/N1
sm
mkg
1
N1
s
mkg
1
5
2
22
2
=
=
⋅
⋅
=
⋅
Flow reversal causes a considerable drop in pressure in curved pipes, T-pieces,
branches and angle connections. The resistances which arise are chiefly dependent
on the geometry of the formed parts and the flow value.
These pressure losses are calculated using the form coefficient ξ for which the most
common shapes are set as a result of experimental tests.
2
v
p
2
⋅ρ
⋅ξ=∆
Since the form coefficient is heavily dependent on the Reynolds’ number, a
correction factor b corresponding to the Re number is taken into consideration.
Thus, the following applies for the laminar range:
2
v
bp
2
⋅ρ
⋅⋅ξ=∆
Table for correction factor b
Re 25 50 100 250 500 1000 1500 2300
b 30 15 7.5 3 1.5 1.25 1.15 1.0
Pressure losses
through formed parts

5 ... 1521.3 0.5 - 1ξ
T-piece 90° bend Double angle
1.2
90° angle Valve
Table for the form coefficient
Calculate the pressure drop ∆p in an elbow with the nominal size 10 mm.
Given that: Flow speed v = 5 m/s
Density of the oil ρ = 850 kg/mT
3
Viscosity ν = 100 mm2
/s at 150 °C
First Re is calculated:
500
m0001.0s
sm01.0m5dv
Re 2
=
⋅
⋅⋅
=
ν
⋅
=
Factor from the table b = 1.5
Form coefficient from the table ξ = 1.2
bar19.0m/N19125
2sm
m25kg850
5.112
2
v
bp 2
23
22
==
⋅⋅
⋅
⋅⋅=
⋅ρ
⋅⋅ξ=∆
The pressure loss in the valves can be derived from the ∆p-Q-characteristics of the
manufacturer.
Example
Pressure losses in
the valves

The energy content of a hydraulic system is made up of several forms of energy. As
stated in the law of conservation of energy, the total energy of a flowing liquid is
constant. It only changes when energy in the form of work is externally supplied or
carried away. The total energy is the sum of the various forms of energy:
• Static – Potential energy
– Pressure energy
• Dynamic – Motion energy
– Thermal energy
Potential energy is the energy which a body (or a liquid) has when it is lifted by a
height h. Here, work is carried out against the force of gravity. In presses with large
cylinders, this potential energy is used for fast filling of the piston area and for pilot
pressure for the pump. The amount of energy stored is calculated on the basis of an
example.
Diagram – press with elevated reservoir
2.13
Energy and power
Potential energy

W = m ⋅ g ⋅ h
W = Work [J]
m = mass of the liquid [kg]
g = acceleration due to gravity [m/s2
]
h = height of the liquid [m]
from: W = F⋅s F = m⋅g
is produced: W = m⋅g⋅h s = h
unit: 1 kg⋅m/s2
⋅m = 1 Nm = 1 J = 1 W/s [1 J = 1 Joule, 1 W = 1 Watt]
Given that: m = 100 kg
g = 9.81 m/s2
≈ 10 m/s2
h = 2 m
J2000Nm2000
s
mmkg
2000m2s/m10kg100hgmW 2
2
==
⋅⋅
=⋅⋅=⋅⋅=
If a liquid is pressurized, its volume is reduced, the amount by which it is reduced
being dependent on the gases released. The compressible area amounts to 1-3 % of
the output volume. Owing to the limited compressibility of the hydraulic fluid,
i.e. the relatively small ∆V, the pressure energy is low. At a pressure of 100 bar ∆V
amounts to approx. 1 % of the output volume. A calculation based on these values is
shown overleaf.
Pressure energy
Pressure energy

W = p ⋅ ∆V
p = Liquid pressure [Pa]
∆V = Liquid volume [m3
]
from: W=F⋅s and F=p⋅A
is produced: W = p ⋅ A ⋅ s
A⋅s is replaced by ∆V, producing: W = p⋅∆V
Unit: 1 N/m2
⋅m3
= 1 Nm = 1 J
Given that: p = 100 ⋅ 105
Pa
∆V = 0.001 m3
J00010
m
mN
101.0m001.0Pa10100VpW 2
3
535
=
⋅
⋅=⋅⋅=∆⋅=
Pressure energy is obtained from the resistance with which the fluid volume meets
the compression.
All matter is compressible, i.e., if the initial pressure p0 is increased by the value ∆p,
the initial volume V0 is reduced by the value ∆V. This compressibility is increased
even further by the gases dissolved in the oil (to 9%) and by the rising temperature.
In the case of precision drives, the compressibility of the oil must not be neglected.
The characteristic value for this is the compression modulus K which is also often
referred to as the modulus of elasticity for oil = Eoil. This modulus can be calculated
in the usual pressure range using the following approximate formula.
V
p
VK 0
∆
∆
⋅≈ [ ]22
cm/Norm/N
V0 = output volume
∆V = volume reduction
The value K represents air-free oil at 50 °C ≈ 1.56 · 105
N/cm2
. Since the oil generally
contains air, the K value of 1.0 to 1.2 · 105
N/cm2
is used in practice.
Example

200 bar counter pressure is applied to the oil volume for a cylinder with a diameter
of 100 mm and a length of 400 mm (l0). By how many mm is the piston rod pushed
back?
Compression modulus
The area ratio piston side to piston rod side amounts to 2:1 and the compression
modulus K = 1.2 · 105
N/cm2
(the elasticity of the material and the expansion of the
cylinder barrel are not taken into consideration).
The area ratio 2:1 produces an additional 100 bar of pressure on the constrained oil
volume.
From:
V
p
VK 0
∆
∆
⋅=
is produced:
K
p
VV 0
∆
⋅=∆
00 lAV
lAV
⋅=
∆⋅=∆
mm33.3
cm/N102.1
cm/N1000
mm400
K
p
ll
K
p
lAlA
25
2
0
0
=
⋅
⋅=
∆
⋅=∆
∆
⋅⋅=∆⋅
Therefore, the piston rod is pushed back by 3.33 mm. For this calculation, the
increase in volume caused by changes in temperature was not taken into
consideration. This is because the changes in pressure are generally so fast that an
adiabatic change in status (i. e. one proceeding without heat exchange) may be
assumed.
Example
Solution

This example shows that compressibility can be neglected in many cases (e. g. in
presses). However, it is advisable to keep pipe lines and cylinders as short as
possible.
Thus, instead of long cylinders, spindle drives or similar devices which are driven by
hydraulic motors are used for linear movements on machine tools.
Motion energy (also known as kinetic energy) is the energy a body (or fluid particle)
has when it moves at a certain speed. The energy is supplied through acceleration
work, a force F acting on the body (or fluid particle).
The motion energy is dependent on the flow velocity and the mass.
Motion energy
Motion energy

2
vm
2
1
W ⋅=
v = velocity[m/s]
a = acceleration [m/s2
]
W = F ⋅ s = m ⋅ a ⋅ s
F = m ⋅ a s = 2
ta
2
1
⋅ v = a ⋅ t
2222
vm
2
1
tam
2
1
ta
2
1
amW ⋅=⋅⋅=⋅⋅⋅=
Unit: 1 kg⋅(m/s)2
= 1 kg ⋅ m2
/s2
= 1 Nm = 1 J
Given that: m = 100 kg
v1 = 4 m/s
J800
s
mkg
800)s/m4(kg100
2
1
vm
2
1
W 2
2
22
=
⋅
=⋅⋅=⋅=
J000500
s
mkg
000500)s/m100(kg100
2
1
vm
2
1
W 2
2
22
=
⋅
=⋅⋅=⋅=
Every change in the flow velocity (in the case of a constant flow rate) automatically
results in a change in the motion energy. Its share of the total energy increases
when the hydraulic fluid flows faster and decreases when the speed of the hydraulic
fluid is reduced.
Owing to varying sizes of line cross-section, the hydraulic fluid flows in a hydraulic
system at various speeds as shown in the diagram since the flow rate, the product of
the flow velocity and the cross-section are constant.
Example

Thermal energy is the energy required to heat a body (or a liquid) to a specific
temperature.
In hydraulic installations, part of the energy is converted into thermal energy as a
result of friction. This leads to heating of the hydraulic fluid and of the components.
Part of the heat is emitted from the system, i.e. the remaining energy is reduced. The
consequence of this is a decrease in pressure energy.
The thermal energy can be calculated from the pressure drop and the volume.
Thermal energy
W = ∆p ⋅ V
∆p = Pressure loss through friction [Pa]
Unit: J1Nm1
m
m
N1mPa1 2
3
3
===⋅
Given that: ∆p = 5 ⋅ 105
Pa
V = 0.1 m3
J00050m
m
N
105.0m1.0Pa105VpW 3
2
535
=⋅=⋅⋅=⋅=
Thermal energy
Example

Power is usually defined as work or a change in energy per unit of time. In hydraulic
installations, a distinction is made between mechanical and hydraulic power.
Mechanical power is converted into hydraulic power, transported, controlled and
then converted back to mechanical
power.
Hydraulic power is calculated from the pressure and the flow rate.
The following equation applies:
P = p ⋅ Q
P = Power (W) [Nm/s]
P = Pressure [Pa]
Q = Flow rate [m3
/s]
P = F • v
P = p • Q
P = 2 n • Mπ
P T
BA
P T
Ts
M
P
T
Mechanical
power
Hydraulic
power
Mechanical
power
Electrical
power
M = Turning
moment (Nm)
in watts
Load
Power
Power

Given that: p = 60 ⋅ 105
Pa
s/m1007,0s/m10
60
2,4
min/m102,4min/l2,4Q 333333 −−−
⋅=⋅=⋅== T
W420
sm
Nm
102,4s/m1007,0Pa1060QpP 2
3
2335
=⋅=⋅⋅⋅=⋅= −
The following applies if the equation is changed around to express the pressure:
Q
P
p =
Given that: P = 315 W
s/m1007.0s/dm
60
2.4
min/l2.4Q 333 −
⋅===
)bar45(Pa1045)Pa(m/N104500
ms
sNm
1007.0
315
p 523
33
⋅=⋅=
⋅
⋅
⋅
= −
p
P
Q =
Given that: P = 150 W
p = 45 ⋅ 105
Pa
min/l2s/dm033.0s/m103.3
Ns
mNm
103.3
Pa1045
W150
Q 335
2
5
5
==⋅=
⋅
⋅
⋅=
⋅
= −−
The input power in a hydraulic system does not correspond to the output power
since line losses occur. The ratio of the output power to the input power is
designated as efficiency (h).
powerinput
poweroutput
Efficiency =
In practice, distinction is made between volumetric power loss caused by leakage
losses and hydro-mechanical power loss caused by friction.
Example
Example
Example
Efficiency

In the same way, efficiency is divided into:
• Volumetric efficiency (ηvol):
This covers the losses resulting from internal and external leakage losses in the
pumps, motors, and valves.
• Hydro-mechanical efficiency (ηhm):
This covers the losses resulting from friction in pumps, motors, and cylinders.
The total losses occurring in pumps, motors, and cylinders during power conversion
are given as the total efficiency (ηtot) and calculated as follows:
ηtot = ηvol ⋅ ηhm
The following example illustrates how the different types of efficiency need to be
taken into consideration when calculating the input and output power in a hydraulic
system. The values given are experimental values which need to be replaced by
manufacturers’ values for practical application.
70% / 75% 25% / 30%
hydr. power
loss
P = p • Q
P T
Ts
M
P
T
F
v
P T
BA
P = 2 n • MO O Oπ
P = 2 n • MI I Iπ
Output power of the motor:
( 330 W at P = 467 W)~ I
Output power of the cylinder:
( 350 W at P = 467 W)~ I
5% cylinder or
10% motor
10% valves and
lines
10% pump
5% electric motor
Output power
PO
Electrical power
Input power PI
Hydraulic
power
Input power which
the motor delivers
to the pump
P = F • vO
M
n
O
O
Calculation of input and output power

Cavitation (Lat. cavitare = to hollow out) refers to the releasing of the smallest
particles from the surface of the material. Cavitation occurs on the control edges of
hydraulic devices (pumps and valves). This eroding of the material is caused by local
pressure peaks and high temperatures. Flash temperatures are sudden, high
increases in temperature.
What causes the pressure drop and the flash temperatures?
Motion energy is required for an increase in flow velocity of the oil at a narrowing.
This motion energy is derived from the pressure energy. Because of this, pressure
drops at narrow points may move into the vacuum range. From a vacuum of
pe ≤ - 0.3 bar onwards, dissolved air is precipitated. Gas bubbles are formed. If the
pressure now rises again as a result of a reduction in speed, the oil causes the gas
bubbles to collapse.
3
Pressure
Pressure drop
Pressure collapse
Relative
vacuum
2
1
0
0.7
bar
Pressure drop at the narrow point
2.14
Cavitation

-0.3 bar
v4
v < v3 4
v3
Cavitation
After the narrowing, the pressure rises again, the bubbles burst and the following
cavitation effects might occur:
• Pressure peaks
Small particles are eroded from the pipe wall at the point where the cross-section
is enlarged. This leads to material fatigue and often to fractures. This cavitation
effect is accompanied by considerable noise.
• Spontaneous ignition of the oil/air mixture
When the air bubbles burst, the oil displaces the bubbles. Owing to the high
pressure after the narrowing, very high temperatures are produced as a result of
compression of the air on the bubbles bursting. As with a diesel engine, this may
lead to spontaneous ignition of the oil/air mixture in the bubbles (diesel effect).
There are various explanations for the presence of air in a hydraulic system:
Liquids always contain a certain quantity of air. Under normal atmospheric
conditions, hydraulic oils contain approx. 9 % air vol. in soluble form. However, this
proportion varies according to the pressure, temperature, and type of oil. Air can
also get into the hydraulic system from outside, particularly at leaky throttle points.
In addition, it is possible that hydraulic oil taken in by the pump already contains air
bubbles. This may be caused by the return line running incorrectly into the oil
reservoir, by the hydraulic oil being kept in the oil reservoir for too short a time, or
by insufficient air releasing properties in the hydraulic oil.

The subjects covered in this chapter – types of flow, friction, heat, pressure drop,
energy, power, and cavitation – are all illustrated by examples based on a throttle
point:
Throttle point
At throttle points, the value of the Reynolds’ figure is far above 2300. The reason for
this is the cross-sectional narrowing which, owing to the constant flow rate, results
in an increase in flow velocity. Thus, the critical speed at which the flow changes
from laminar to turbulent is achieved very quickly.
The Law of Conservation of Energy states that the total energy in a system always
remains constant. Therefore, if the motion energy increases as a result of a higher
flow velocity, one of the other types of energy must be reduced. Energy conversion
takes place from pressure energy into motion energy and thermal energy. The
increase in the flow velocity causes the friction to rise; this leads to heating of the
hydraulic fluid and an increase in thermal energy. Part of the heat is emitted from
the system. Consequently, the flow rate after the throttle point has the same flow
velocity as before the throttle point. However, the pressure energy has been
reduced by the amount of the thermal energy resulting in a fall in pressure after the
throttle point.
2.15
Throttle points

The decrease in energy at throttle points leads to power losses. These can be
determined by measuring the pressure loss and the increase in temperature.
Pressure losses are dependent on:
• viscosity
• flow velocity
• type and length of throttle
• type of flow (laminar, turbulent).
Poiseuille’s formula:
ρ
∆⋅
⋅⋅α=
p2
AQ D
α = Flow reference number
AD = Throttle cross-section [m2
]
∆p = Pressure drop [Pa]
ρ = Density of the oil [kg/m3
]
Q = Volumetric flow rate [m3
/s]
can be expressed more simply by leaving out the constants:
pQ ∆≈
Flow through a throttle is dependent on the pressure difference.
3 Pressure
Pressure drop
Pressure collapse
Relative
vacuum
2
1
0
0.7
bar
Pressure drop

If the pressure at the throttle point drops into the vacuum range, the air exits from
the oil and bubbles are formed which are filled with oil gas and air (cavitation).
If the pressure increases again after the throttle point at the transition to the
enlarged cross-section, the bubbles burst. This leads to cavitation effects – eroding
of the material in the area of the cross-sectional enlargement and, potentially, to
spontaneous ignition of the hydraulic oil.

3. Hydraulic fluid
In principle, any liquid can be used to transfer pressure energy. However, as in
hydraulic installations, other characteristics are also required of hydraulic fluids, the
number of suitable fluids is considerably restricted.
As a hydraulic fluid, water causes problems related to corrosion, boiling point,
freezing point and low viscosity.
Hydraulic fluids with a mineral oil base – also known as hydraulic oils – fulfil most
normal requirements (e.g. for machine tools). They are used very widely.
Hydraulic fluids with low inflammability are required for hydraulic systems with high
risk of fire such as, for example:
• hard coal mining
• die-casting machines
• forging presses
• control units for power station turbines
• and steel works and rolling mills.
In all these applications, there is a risk that hydraulic fluids based on mineral oils
will catch fire on intensively heated metal parts. Oil mixtures containing water or
synthetic oils are used here in place of standard oils.
The hydraulic fluids used in hydraulic installations must fulfil very varied tasks:
• pressure transfer,
• lubrication of the moving parts of devices,
• cooling, i.e. diversion of the heat produced by energy conversion
(pressure losses),
• cushioning of oscillations caused by pressure jerks,
• corrosion protection,
• scuff removal,
• signal transmission.
3.1
Tasks for hydraulic fluids

3. Hydraulic fluid
Within these two groups – hydraulic oils and hydraulic fluids with low inflammability
– there are various types of fluid with different characteristics. These characteristics
are determined by a basic fluid and small quantities of additives.
In DIN 51524 and 51525 hydraulic oils are divided according to their characteristics
and composition into three classes:
• Hydraulic oil HL
• Hydraulic oil HLP
• Hydraulic oil HV.
The designations for these oils are composed of the letter H for hydraulic oil and an
additional letter for the additives. The code letter is supplemented by a viscosity
code defined in DIN 51517 (ISO viscosity classes).
Designation Special characteristics Areas of application
HL Increased corrosion
protection and ageing
stability
Systems in which high thermal
demands are made or corrosion
through immersion in water is possible.
HLP Increased wearing protection Like HL oil, also for use in systems
where variable high friction occurs
owing to design or operating factors.
HV Improved
viscosity-temperature
characteristics
Like HLP oil, for use in widely
fluctuating and low ambient
temperatures.
Hydraulic oil for hydraulic systems
H hydraulic oil
L with additives to increase corrosion protection and/ or ageing stability
P with additives to reduce and/or increase load carrying, ability
68 Viscosity code as defined in DIN 51517
3.2
Types of hydraulic fluid
Hydraulic oils
Hydraulic oil HLP 68

3. Hydraulic fluid
Where these hydraulic fluids are concerned, a distinction is made between hydrous
and anhydrous synthetic hydraulic fluids. The synthetic hydraulic fluids are
chemically composed so that their vapour is not flammable.
The table shown here provides an overview of hydraulic fluids with low flammability
(HF liquids). They are also described in VDMA standard sheets 24317 and 24320.
Abbreviated code VDMA standard
sheet no.
Composition Water content in %
HFA 24 320 Oil-water emulsions 80 – 98
HFB 24 317 Water-oil emulsions 40
HFC 24 317 Hydrous solutions,
e.g. water-glycol
35 – 55
HFD 24 317 Anhydrous liquid,
e.g. phosphate ether
0 – 0.1
Hydraulic fluids with low flammability
For hydraulic oils to be able to fulfil the requirements listed above, they must exhibit
certain qualities under the relevant operating conditions. Some of these qualities
are listed here:
• lowest possible density;
• minimal compressibility;
• viscosity not too low (lubricating film);
• good viscosity-temperature characteristics;
• good viscosity-pressure characteristics;
• good ageing stability;
• low flammability;
• good material compatibility;
In addition, hydraulic oils should fulfil the following requirements:
• air release;
• non-frothing;
• resistance to cold;
• wear and corrosion protection;
• water separable.
The most important distinguishing feature of hydraulics is viscosity.
Hydraulic fluids with
low inflammability
3.3
Characteristics and
requirements

3. Hydraulic fluid
The word “viscosity” can be defined as “resistance to flow”. The viscosity of a liquid
indicates its internal friction, i.e. the resistance which must be overcome to move
two adjacent layers of liquid against each another. Thus, viscosity is a measure of
how easily a liquid can be poured.
The international system of standards defines viscosity as “kinematic viscosity”
(unit: mmT
2T
/s).
It is determined by a standardised procedure, e.g.:
• DIN 51562: Ubbelohde viscometer;
• DIN 51561: Vogel-Ossag viscometer.
The ball viscometer can also be used to determine kinematic viscosity. It can be
used to measure viscosity values precisely across a broad area. Measurements are
made to determine the speed with which a body sinks in a liquid under the influence
of gravity. To find the kinematic viscosity, it is necessary to divide the value
determined using the ball viscometer by the density of the fluid.
Ball viscometer
3.4
Viscosity

3. Hydraulic fluid
One important method of identifying hydraulic oils is the specification of viscosity
class. The ISO standard and the new draft of DIN 51524 explain that the viscosity
classes lay down the minimum and maximum viscosity of hydraulic oils at 40 °C.
kinematic viscosity (mm²/s) at 40 °CISO
viscosity classes
max. min.
ISO VG 10 9.0 11.0
ISO VG 22 19.8 24.2
ISO VG 32 28.8 35.2
ISO VG 46 41.4 50.6
ISO VG 68 61.2 74.8
ISO VG 100 90.0 110.0
Viscosity classes (DIN 51502)
Thus, six different viscosity classes are available for the various types of hydraulic
oil HL, HLP and HV. The table below specifies areas of application for the different
viscosity classes; it is necessary here to match the viscosity class to the ambient
temperatures.
For storage reasons, high-grade motor or gear lubricating oil is also used in
hydraulic installations. For this reason, the SAE viscosity classification is also listed
here. However, this classification allows fairly large tolerance zones as can be seen
from a comparison between the two methods of classification.

3. Hydraulic fluid
SAE classes ISO-VG Areas of application
30
100
Stationary installations in closed areas
at high temperatures
68
20, 20 W
4610 W
32
At normal temperatures
5 W
22 For open air applications – mobile hydraulics
(15)
10
In colder areas
SAE viscosity classification
In practice viscosity margins play an important role:
Where viscosity is too low (very fluid), more leakages occur. The lubricating film is
thin and, thus, able to break away more easily resulting in reduced protection
against wear. Despite this fact, fluid oil is preferred to viscous oil since pressure and
power losses are small owing to the lower friction. As viscosity increases, the
internal friction of the liquid increases and, with that, the pressure and power loss
caused by the heat also increases.
High viscosity results in increased friction leading to excessive pressure losses and
heating particularly at throttle points. This makes cold start and the separation of air
bubbles more difficult and, thus, leads to cavitation.

3. Hydraulic fluid
Kinematic viscosity
Lower limit
10
s
mm2
Ideal viscosity range
15 – 100
s
mm2
Upper limit
750
s
mm2
Viscosity limits
When using hydraulic fluids, it is important to consider their viscosity-temperature
characteristics, since the viscosity of a hydraulic fluid changes with changes in
temperature. These characteristics are shown in the Ubbelohde’s viscosity-
temperature diagram. If the values are entered on double logarithmic paper, a
straight line is produced.
0
0
10
20
50
100
500
1000
5000
mm /s
2
10000
20 40 60 80 °C 100 Temperature
ν
1400
1200
1000
800
600
400
200
0
over-pressure
(bar)
Ubbelohde’s viscosity temperature diagram
63 © Festo Didactic GmbH & Co. • TP501

3. Hydraulic fluid
The viscosity index (VI) is used as a reference value for viscosity-temperature
characteristics.
It is calculated in accordance with DIN ISO 2909. The higher the viscosity index of a
hydraulic oil, the less the viscosity changes or the greater the temperature range in
which this hydraulic oil can be used. In the viscosity-temperature diagram, a high
viscosity index is shown as a level characteristic line.
Mineral oils with a high viscosity index are also referred to as multigrade oils. They
can be used anywhere where very changeable operating temperatures arise; such as
for mobile hydraulics, for example. Where oils with a low viscosity index are
concerned, a distinction must be made between summer oils and winter oils:
• Summer oils:
with higher viscosity so that the oil does not become too fluid causing the
lubricating film to break up.
• Winter oils:
with lower viscosity so that the oil does not become too thick and a cold start is
possible.
The viscosity-pressure characteristics of hydraulic oils are also important since the
viscosity of hydraulic oils increases with increasing pressure. These characteristics
are to be noted particularly in the case of pressures from a ∆p of 200 bar.
At approx. 350 to 400 bar the viscosity is generally already double that at 0 bar.
0 2000 4000 8000 10000bar Pressure6000
0.1
1
10
100
1000
10000
0°C
mm /s
2
100000
1000000
Kinem.
viscosity
40°C
100°C
200°C
Viscosity-pressure characteristics

3. Hydraulic fluid
If the characteristics of hydraulic fluids described in this chapter are summarized,
the following advantages and disadvantages of hydraulic fluids with low
flammability result when compared to hydraulic oils on a mineral oil base:
Advantages Disadvantages
Greater density Difficult intake conditions for
pumps.
Low compressibility Hydraulic oil less fluid Higher pressure peaks possible.
Unfavourable air venting
properties
Increase dwell time in reservoir by
using larger reservoirs.
Limited operating
temperatures
50 °C may not be exceeded as
otherwise too much water
vaporises.
Favourable viscosity
temperature characteristics
In the case of HFC liquids, the
viscosity changes less sharply
in case of temperature
fluctuations.
In the case of HFD liquids, the
viscosity changes with
temperature fluctuations.
Wearing properties HFD liquids erode conventional
bunan seals, accumulator
diaphragms and hoses.
Price Characteristics of HFD liquids
correspond to those of
hydraulic oil when
appropriate cooling and
heating equipment is in use.
HFD liquids are more expensive
than hydraulic oils.
Advantages and disadvantages of hydraulic fluids with low flammability

4. Components of a hydraulic system
The modules and devices used in hydraulic systems are explained in some detail in
this chapter.
The power supply unit provides the necessary hydraulic power – by converting the
mechanical power from the drive motor.
The most important component in the power supply unit is the hydraulic pump. This
draws in the hydraulic fluid from a reservoir (tank) and delivers it via a system of
lines in the hydraulic installation against the opposing resistances. Pressure does
not build up until the flowing liquids encounter a resistance.
The oil filtration unit is also often contained in the power supply section. Impurities
can be introduced into a system as a result of mechanical wear, oil which is hot or
cold, or external environmental influences. For this reason, filters are installed in the
hydraulic circuit to remove dirt particles from the hydraulic fluid. Water and gases in
the oil are also disruptive factors and special measures must be taken to remove
them.
Heaters and coolers are also installed for conditioning the hydraulic fluid. The
extent to which this is necessary depends on the requirements of the particular
exercise for which the hydraulic system is being used.
The reservoir itself also plays a part in conditioning the hydraulic fluid:
• Filtering and gas separation by built-in baffle plates,
• Cooling through its surface.
This is the working medium which transfers the prepared energy from the power
supply unit to the drive section (cylinders or motors). Hydraulic fluids have a wide
variety of characteristics. Therefore, they must be selected to suit the application in
question. Requirements vary from problem to problem. Hydraulic fluids on a mineral
oil base are frequently used; these are referred to as hydraulic oils.
4.1
Power supply section
4.2
Hydraulic fluid

Valves are devices for controlling the energy flow. They can control and regulate the
flow direction of the hydraulic fluid, the pressure, the flow rate and, consequently,
the flow velocity.
There are four valve types selected in accordance with the problem in question.
These valves control the direction of flow of the hydraulic fluid and, thus, the
direction of motion and the positioning of the working components. Directional
control valves may be actuated manually, mechanically, electrically, pneumatically
or hydraulically. They convert and amplify signals (manual, electric or pneumatic)
forming an interface between the power control section and the signal control
section.
Directional control valve
4.3
Valves
Directional control valves

These have the job of influencing the pressure in a complete hydraulic system or in a
part of the system. The method of operation of these valves is based on the fact that
the effective pressure from the system acts on a surface in the valve. The resultant
force is balanced out by a counteracting spring.
Pressure relief valve
These interact with pressure valves to affect the flow rate. They make it possible to
control or regulate the speed of motion of the power components. Where the flow
rate is constant, division of flow must take place. This is generally effected through
the interaction of the flow control valve with a pressure valve.
Flow control valve
Pressure valves
Flow control valves

In the case of this type of valve, a distinction is made between ordinary non-return
valves and piloted non-return valves. In the case of the piloted non-return valves,
flow in the blocked direction can be released by a signal.
Non-return valve
Cylinders are drive components which convert hydraulic power into mechanical
power. They generate linear movements through the pressure on the surface of the
movable piston. Distinction is made between the following types of cylinder:
The fluid pressure can only be applied to one side of the piston with the result that
the drive movement is only produced in one direction. The return stroke of the
piston is effected by an external force or by a return spring.
Examples: – Hydraulic ram
– Telescopic cylinder
The fluid pressure can be applied to either side of the piston meaning that drive
movements are produced in two directions.
Examples: – Telescopic cylinder
– Differential cylinder
– Synchronous cylinder
Non-return valves
4.4
Cylinders
(linear actuators)
Single-acting cylinders
Double-acting cylinders

Double-acting cylinder
Like cylinders, hydraulic motors are drive components controlled by valves. They too
convert hydraulic power into mechanical power with the difference that they
generate rotary or swivel movements instead of linear movements.
Hydraulic motor (gear motor)
4.5
Motors
(rotary actuators)

Simple graphic and circuit symbols are used for individual components to enable
clear representation of hydraulic systems in diagrams. A symbol identifies a
component and its function, but it does not provide any information about its
design. The symbols to be used are laid down in DIN ISO 1219. The most important
symbols are dealt with in this chapter.
An arrow drawn at an angle through the symbol indicates that setting possibilities
exist.
Hydraulic pumps and motors are represented by means of a circle which shows
where the drive or output shaft is located. Triangles within the circle give
information about the direction of flow. These triangles are filled in, since hydraulic
fluids are used for hydraulics. If a gaseous pressure medium were being used, as is
the case in pneumatics, the triangles would not be filled in. The symbols for
hydraulic motors and hydraulic pumps can only be distinguished from one another
by the fact that the arrows indicating the direction of flow are drawn pointing one
way for the pumps and the other for the motors.
Hydraulic pumps with fixed displacement
Hydraulic motors with fixed displacement
– with one flow direction
Fluids
Gases
– with single direction of rotation
– with two flow directions
– with two directions of rotation
Fixed displacement hydraulic pumps and motors
5. Graphic and circuit symbols
Note
5.1
Pumps and motors

Directional control valves are shown by means of several connected squares.
• The number of squares indicates the number of switching positions possible for a
valve.
• Arrows within the squares indicate the flow direction.
• Lines indicate how the ports are interconnected in the various switching
positions.
There are two possible methods of port designation. One method is to use the
letters P, T, A, B and L, the other is to label ports alphabetically A, B, C, D, etc. The
former method is generally preferred. Ports should always be labelled with the valve
in the rest position. Where there is no rest position, they are allocated to the
switching position assumed by the valve when the system is in its initial position.
The rest position is defined as the position automatically assumed by the valve on
removal of the actuating force.
When labelling directional control valves, it is first necessary to specify the number
of ports followed by the number of switching positions. Directional control valves
have at least two switching positions and at least two ports. In such an instance, the
valve would be designated a 2/2-way valve. The following diagrams show other
directional control valves and their circuit symbols.
2/2 – way valve
4/2 – way valve
3/2 – way valve
4/3 – way valve
Port designations
Number of switching positions
Number of ports
A
A
A
A
B
B
P
T
T
T
P
P
P
P
A
or:
T
B
A
B
C
D
L
L
pressure port
pressure port
return port
return port
power ports
power ports
leakage oil
leakage oil
5.2

The switching position of a directional control valve can be changed by various
actuation methods. The symbol for the valve is elaborated by the addition of the
symbol indicating the actuation method. In the case of some of the actuation
methods shown, such as push button, pedal, lever with detent, a spring is always
necessary for resetting. Resetting may also be achieved by switching the valve a
second time, e.g. in the case of a valve with hand lever and detent setting.
Listed below are the symbols for the most important actuation methods. Refer to
DIN ISO 1219 for other methods of actuation.
– general symbol with spring return and bleed port
– by lever
– by pedal and spring return
– by push button with spring return
– by lever with detent setting
Mechanical actuation
– by spring
– by stem or push button
– by roller stem
Mechanical actuation (continuation)
5.3
Methods of actuation

* Type of actuation to be specified
where no standard symbol exists
General symbol
Pressure valves are represented using squares. The flow direction is indicated by an
arrow. The valve ports can be labelled P (pressure port) and T (tank connection) or A
and B.
The position of the valve within the square indicates whether the valve is normally
open or normally closed.
open flow from
P to A
T closed
closed
A A
B P T
P
T
Pressure valves
A further distinction is made between set and adjustable pressure valves. The latter
are indicated by a diagonal arrow through the spring.
set adjustable
P P
T T
Pressure valves
5.4
Pressure valves

Pressure valves are divided into pressure relief valves and pressure regulators:
pressure relief valve 3-way pressure regulator
pressure valves
P(A) P(A)
T(B) A(B)
Pressure valves
In the normally closed position the control pressure is detected at the input. This
pressure acts on a valve via the control passage coming from the input on a piston
surface which is held against the control pressure by a spring. If the force resulting
from the pressure and the effective piston surface exceeds the spring force, the
valve opens. In this way, it is possible to set the limiting pressure to a fixed value.
In the case of a normally open pressure regulator, the control pressure is detected at
the output. This pressure is effective in the valve via the control passage on a piston
surface and generates a force. This force works against a spring. The valve begins to
close when the output pressure is greater than the spring force. This closing process
causes a pressure drop from the input to the output of the valve (caused by the flow
control). When the output pressure reaches a specified value, the valve closes
completely. The specified maximum system pressure is set at the input of the valve,
the reduced system pressure at the output. Thus, the pressure regulator can only be
set to a smaller setting value than that set at the pressure relief valve.
Pressure regulator

In the case of flow control valves, a distinction is made between those affected by
viscosity and those unaffected. Flow control valves unaffected by viscosity are
termed orifices. Throttles constitute resistances in a hydraulic system.
The 2-way flow control valve consists of two restrictors, one setting restrictor
unaffected by viscosity (orifice) and one adjustable throttle. The adjustable throttle
gap is modified by changes in pressure. This adjustable throttle is also known as a
pressure balance. These valves are depicted as a rectangle into which are drawn the
symbol for the variable throttle and an arrow to represent the pressure balance. The
diagonal arrow running through the rectangle indicates that the valve is adjustable.
There is a special symbol to represent the 2-way flow control valve.
adjustable
set
A
A
B
B
adjustable
set
A
A
B
B
Throttle Orifice
Throttle and orifice
adjustable
A B
adjustable
A B
with throttle with orifice in detail
2-way flow control valve
5.5
Flow control valves

The symbol for non-return valves is a ball which is pressed against a sealing seat.
This seat is drawn as an open triangle in which the ball rests. The point of the
triangle indicates the blocked direction and not the flow direction. Pilot controlled
non-return valves are shown as a square into which the symbol for the non-return
valve is drawn. The pilot control for the valve is indicated by a control connection
shown in the form of a broken line. The pilot port is labelled with the letter X.
Shut-off valves are shown in circuit diagrams as two triangles facing one another.
They are used to depressurise the systems manually or to relieve accumulators. In
principle, wherever lines have to be opened or closed manually.
A A
B B
spring loaded unloaded
Non-return valve
A B
B
A X
shut-off valve pilot-controlled non-returned valve
Shut-off valve and pilot-controlled non-return valve
5.6
Non-return valves

Cylinders are classified as either single-acting or double-acting.
Single acting cylinders just have one port, i.e. only the full piston surface can be
pressurised with hydraulic fluid. These cylinders are returned either by the effect of
external forces – indicated by the symbol with the open bearing cap – or by a spring.
The spring is then also drawn into the symbol.
single acting cylinder,
with spring return
single acting cylinder,
return by external force
single acting telescopic cylinder
Single acting cylinder
Double acting cylinders have two ports for supplying either side of the piston with
hydraulic fluid.
It can be seen from the symbol for a double acting cylinder with single piston rod
that the piston area is greater than the annular piston surface.
Conversely, the symbol for the cylinder with a through piston rod shows that these
areas are of the same size (synchronous cylinder).
The symbol for the differential cylinder can be distinguished from that for the
double-acting cylinder by the two lines added to the end of the piston rod. The area
ratio is 2:1.
Like single-acting telescopic cylinders, double-acting ones are symbolized by
pistons located one inside the other.
In the case of the double-acting cylinder with end position cushioning, the
cushioning piston is indicated in the symbol by a rectangle.
5.7
Cylinders
Double acting cylinder

double-acting cylinder
with end position cushioning at both ends
with through piston rod
double-acting telescopic cylinder
with single piston rod
differential cylinder
with single end position cushioning
double acting cylinder
with adjustable end position cushioning
at both ends
Double-acting cylinders

The following symbols are used in circuit diagrams for energy transfer and
conditioning of the pressure medium.
– electric motor
– hydraulic pressure source
– non-electric drive unit
– pressure, power, return line
– control (pilot) line
– exhaust, continuous
– lines crossing
– reservoir
– filter
– cooler
– heater
– quick-acting coupling connected
with mechanically opening non-return valves
– flexible line
M
M
Energy transfer and conditioning of the pressure medium
5.8
Transfer of energy and
conditioning of the
pressure medium

Measuring devices are shown in the circuit diagrams by the following symbols:
– pressure gauge
– thermometer
– flow meter
– filling level indicator
If several devices are brought together in a single housing, the symbols for the
individual devices are placed into a box made up of broken lines from which the
connections are led away.
P T
Ts
M
Hydraulic power pack
B2
A2
B1
A1
Pilot-operated double non-return valve
5.9
Measuring devices
5.10
Combination of devices

A hydraulic system can be divided into the following sections:
• The signal control section
• The power section
Drive section
Powerflow
Power
control
section
Power supply
section
Energy conversion
Pressure medium
preparation
Hydr. power section
Signal control section
Signal
input
Control energy supply
Signal
processing
Diagrammatic representation of the structure of a hydraulic system
6. Design and representation of a hydraulic system

The signal control section is divided into signal input (sensing) and signal
processing (processing).
Signal input may be carried out:
• manually
• mechanically
• contactlessly
Signals can be processed by the following means:
• by the operator
• by electricity
• by electronics
• by pneumatics
• by mechanics
• by hydraulics
P T
A B
M
P T
T
A
P
P
Drive section
Powerflow
Power
control
section
Power supply
section
Energy conversion
Pressure medium
preparation
Hydr. power section
Signal
input
Signal
processing
Signal
output
Interface
Hydraulic system (Design)
6.1

The hydraulic power can be divided up into the power supply section, the power
control section and the drive section (working section). The power supply section is
made up of the energy conversion part and the pressure medium conditioning part.
In this part of the hydraulic system, the hydraulic power is generated and the
pressure medium conditioned. The following components are used for energy
conversion – converting electrical energy into mechanical and then into hydraulic
energy:
• Electric motor
• Internal combustion engine
• Coupling
• Pump
• Pressure indicator
• Protective circuitry
The following components are used to condition the hydraulic fluid:
• Filter
• Cooler
• Heater
• Thermometer
• Pressure gauge
• Hydraulic fluid
• Reservoir
• Filling level indicator
6.2
Hydraulic power section

filter
P T
A B
M
P T
T
A
P
P
pressure gauge
filling level
indicator
pump
motor
pressure
relief
valve
Drive section
Powerflow
Power
control
section
Hydr. power section
Signal
input
Signal
processing
Power supply
section
Energy conversion
Pressure medium
preparation
Hydraulic system (Design)
The power is supplied to the drive section by the power control section in
accordance with the control problem. The following components perform this task:
• directional control valves
• flow control valves
• pressure valves
• non-return valves.
The drive section of a hydraulic system is the part of the system which executes the
various working movements of a machine or manufacturing system. The energy
contained in the hydraulic fluid is used for the execution of movements or the
generation of forces (e. g. clamping processes). This is achieved using the following
components:
• cylinders
• motors

P T
A B
M
P T
T
A
P
pressure gauge
directional control
valve
non-return valve
filling level
indicator
pump
pressure valve
motor filter
flow control valveDrive section
Powerflow
Power
control
section
Power supply
section
Energy conversion
Pressure medium
preparation
Hydr. power section
Signal
input
Signal
processing
Hydraulic system ( Design)
A suitable type of representation is required in order to reproduce movement
sequences and operating statuses of working elements and control elements clearly.
The following types of representation are of importance:
• positional sketch
• circuit diagram
• displacement-step diagram
• displacement-time diagram
• function diagram
• function chart.

The positional sketch is a drawing or schematic diagram of a production installation
or machine etc. It should be easily understandable and should include only the most
important information. It shows the spatial arrangement of the components.
The positional sketch in the Figure shows the position of cylinder Z1 and its function:
Z1 is intended to lift the hood of the tempering furnace.
Z1
Positional sketch
6.3
Positional sketch

The circuit diagram describes the functional structure of the hydraulic system.
P
A
T
T
1Z1
0M10P1
50l
1V3
1V2
0Z2
0Z1
1A
T
P
P
m
1V1
M
Signal
input
Drive
section
Power
control
section
Power
supply
section
Designation of the components
The power supply section of the system with filter (0Z1), pressure-relief valve (0Z2),
pump (0P1) and electric motor (0M1) is depicted in the lower part of the circuit
diagram shown for the hydraulic device of the tempering furnace.
The power control section with the non-return valve (1V1), the 3/2-way valve (1V3)
and the pressure-relief valve (1V2) is located in the centre of the circuit diagram. The
3/2-way valve (1V3) with the hand lever for signal input forms the “system-person”
interface.
Like the drive section, the power control section is assigned to the power section. In
this hydraulic device, the drive section consists of the single-acting cylinder 1A.
6.4
Circuit diagram

In the circuit diagram, the technical data are often additionally specified with the
devices in accordance with DIN 24347.
P
A
T
T
1Z1
0M10P1
50 l
1V3
1V2
0Z2
0Z1
32/22 x 200
1.1 kW2.8 cm
3
6000 kPa
(60 bar)
5000 kPa
(50 bar)
100 kPa
(1 bar)
NG6
T
P
P
1V1
M
1A
m
Circuit diagram with technical data
6.5
Components plus
technical data

Furthermore, the circuit diagram can be supplemented by tables:
Equipment Specifications Example values
Volume in litres to the highest
permissible oil level
Max. 50 lReservoirs
Type of hydraulic fluid ISO VG 22 type Al or HLP
Rated capacity in kW 1.1 kWElectric motors
Rated speed in rpm 1420 rpm
Fixed displacement
pumps and variable-
displacement pumps
Geometric delivery rate in cm³ Gear pump 2.8 cm³/revolution
Pressure valves Set pressure in bar or permissible
pressure range for the system
Operating pressure 50 bar
Non-return valve Opening pressure 1 bar
Cylinder Cylinder inner diameter/piston rod
diameter ⋅ stroke in mm.
The function (e. g. clamping,
lifting, flat turning etc.) must be
entered above every cylinder
32/22 ⋅ 200
1A lifting
Filter Nominal flow rate in l/min
ß...at ∆p...bar
Flexible hose Nominal diameter (inner diameter)
in mm
6 mm
Hydraulic motor Capacity in cm³
Speed in rpm
v = 12.9 cm³
n = 1162.8 rpm
at
Q = 15 cm³/min
M = 1 Nm
Directional control valve Nominal size NG 6

Function diagrams of working machines and production installations can be
represented graphically in the form of diagrams. These diagrams are called function
diagrams. They represent statuses and changes in status of individual components
of a working machine or production installation in an easily understood and clear
manner.
The following example shows a lifting device controlled by electromagnetic
directional control valves.
Time
Designation
Pump 0P1
Directional control
valve
1V1
Cylinder 1A
Directional control
valve
2V1
Cylinder 2A
On
Off
Y2
Y1
1
0
Y4
Y3
1
B1
B0
S0
S1
0
Signal
Identi-
fication
Components
Step
1 2 3 4 5 6 7 8 9 10
p
Function diagram
6.6
Function diagram

A function chart is a flow chart in which the control sequence is strictly divided into
steps. Each step is executed only after the previous step has been completed and all
step enabling conditions have been fulfilled.
S
S
S
S
Step
Transmission
condition
Action Acknowledgement
signal
3S2
1S2
3S1
1S1
0
1
2
3
4
&
4.1: 1S1
1.1: 3S2
2.1: 1S2
3.1: 3S1
A1
Start 1S3
A 4
Close gripper 3A+
Swivel 1A+
Open gripper 3A-
Swivel back 1A-
Function chart
6.7
Function chart

The power supply section provides the energy required by the hydraulic system. The
most important components in this section are:
• drive
• pump
• pressure relief valve
• coupling
• reservoir
• filter
• cooler
• heater
In addition, every hydraulic system contains service, monitoring and safety devices
and lines for the connection of hydraulic components.
Hydraulic power unit
Hydraulic systems (with the exception of hand pumps) are driven by motors (electric
motors, combustion engines). Electrical motors generally provide the mechanical
power for the pump in stationary hydraulics, whilst in mobile hydraulics combustion
engines are normally used.
In larger machines and systems, the central hydraulics are of importance. All
consuming devices in a system with one or several hydraulic power supply units and
with the help of one or more reservoirs are supplied via a common pressure line. The
hydraulic reservoir stores hydraulic power which is released as required. The
reservoir is dealt with in greater detail in the TP502 Advanced Course.
Pressure, return and waste oil lines are all ring lines. Space and power requirements
are reduced by employing this type of design.
7. Components of the power supply section
7.1
Drive

This diagram shows a processing station from a transfer line.
S3
S3
P T
A
P
Pressure line
Return line
Waste oil line
Circuit diagram

The pump in a hydraulic system, also known as a hydraulic pump, converts the
mechanical energy in a drive unit into hydraulic energy (pressure energy).
The pump draws in the hydraulic fluid and drives it out into a system of lines. The
resistances encountered by the flowing hydraulic fluid cause a pressure to build up
in the hydraulic system. The level of the pressure corresponds to the total resistance
which results from the internal and external resistances and the flow rate.
• External resistances:
come about as a result of maximum loads and mechanical friction and static load
and acceleration forces.
• Internal resistances:
come about as a result of the total friction in the lines and components, the
viscous friction and the flow losses (throttle points).
Thus, the fluid pressure in a hydraulic system is not predetermined by the pump. It
builds up in accordance with the resistances – in extreme cases until a component is
damaged. In practice, however, this is prevented by installing a pressure relief valve
directly after the pump or in the pump housing at which the maximum operating
pressure recommended for the pump is set.
The following characteristic values are of importance for the pump:
The displacement volume V (also known as the volumetric displacement or working
volume) is a measure of the size of the pump. It indicates the volume of liquid
supplied by the pump per rotation (or per stroke).
The volume of liquid supplied per minute is designated as volumetric flow rate Q
(delivery). This is calculated from the displacement volume V and the number of
rotations n:
Q = n ⋅ V
7.2
Pump
Displacement volume

Calculation of the delivery of a gear pump.
Given that: Number of rotations n = 1450 minT
-1T
Displacement volume V = 2.8 cmT
3T
(per rev.)
To be found: Delivery Q
min/l06.4
min
dm
06.4
min
cm
4060cm8.2.m.p.r1450VnQ
33
3
===⋅=⋅=
The operating pressure is of significance for the area of application of pumps. Peak
pressure is specified. However, this should arise only briefly (see diagram) as
otherwise the pump will wear out prematurely.
Peak pressure
p3
Maximum
pressure p2
Continuous
pressure p1
Pressure
p
Time t
Duty cycle
Operating pressure
A pressure relief valve is installed in some pumps for safety reasons.
The drive speed is an important criterion for pump selection since the delivery Q of a
pump is dependent on the number of rotations n. Many pumps are only effective at a
specific r.p.m. range and may not be loaded from a standstill. The most usual
number of rotations for a pump is n = 1500 r.p.m. since pumps are mainly driven by
three-phase asynchronous motors whose number of rotations is not dependent on
the supply frequency.
Example
Operating pressure
Speeds

Mechanical power is converted by pumps into hydraulic power resulting in power
losses expressed as efficiency.
When calculating the total efficiency ηT
totT
of pumps, it is necessary to take into
consideration the volumetric (ηT
volT
) and the hydro-mechanical (ηT
hmT
) efficiency.
ηT
tot = TηT
vol ⋅ TηT
hmT
In practice, characteristic lines are made use of in the evaluation of pumps. VDI
recommendation 3279 provides a number of characteristic lines, for example for:
• delivery Q
• power P
• and efficiency η as a function of the pressure at a constant speed.
The characteristic line for the delivery as a function of the pressure is designated the
pump characteristic. The pump characteristic shows that the effective pump delivery
(QT
effT
) is reduced according to pressure build-up. The actual delivery (QT
wT
) can be
determined when the waste oil from the pump (QT
LT
) is taken into consideration. A
minimum leakage in the pump is necessary to maintain lubrication.
The following information can be derived from the pump characteristic:
• where p = 0, the pump supplies the complete delivery Q.
• where p > 0, Q is reduced owing to the leakage oil.
• The course of the characteristic line provides information about the volumetric
efficiency (ηT
volT
) of the pump.
Efficiency

In the diagram, the pump characteristic for a pump in working order and for a worn
(defective) pump.
Volumetric flow rate
Q
dm /min
3
9.2
8.6
0 50 150
Defective pump
Pump in working order
100 200 bar 250 Pressure
p
0
8.8
9.0
9.4
9.6
10.0
<7%
13%
Pump characteristic
Characteristic for the new pump: The leakage oil from the pump amounts to 6.0 % at
230 bar. This results in:
QT
(p = 0)T
= 10.0 dmT
3T
/min
QT
(p = 230)T
= 9.4 dmT
3T
/min
QT
LT
= 0.6 dmT
3T
/min
min/dm0.10
min/dm4.9
3
3
vol =η = 0.94
Characteristic for the defective pump: The leakage oil from the pump amounts to
14.3 % at 230 bar. This results in:
QT
(p = 0)T
= 10.0 dmT
3T
/min
QT
(p = 230)T
= 8.7 dmT
3T
/min
QT
LT
= 1.3 dmT
3T
/min
min/dm0.10
min/dm7.8
3
3
vol =η = 0.87

Therefore, on the basis of the pump characteristic, there is a possibility of
calculating the volumetric efficiency (ηT
volT
) of a pump.
In order to be able to use pumps correctly, the characteristic values and curves
which have been described must be known. Using this information, it is easier to
compare devices and select the most suitable pump.
Other design features of a pump may also be of significance:
• type of mounting
• operating temperatures
• noise rating
• hydraulic fluid recommendations
• pump type.
Three basic types of hydraulic pump can be distinguished on the basis of the
displacement volume:
• constant pumps: fixed displacement volume
• adjustable pumps: adjustable displacement volume
• variable capacity pumps: regulation of pressure, flow rate or power, regulated
displacement volume.
Hydraulic pump designs vary considerably; however, they all operate according to
the displacement principle. The displacement of hydraulic fluid into the connected
system is effected, for example, by piston, rotary vane, screw spindle or gear.
Gear pump
External gear pump
Internal gear pump
Annular gear pump
Screw pump
Constant pump
Rotary vane pump
Hydraulic pumps
Internally pressurized
Externally pressurized
Constant, adjustable and variable capacity pumps
Piston pump
Radial piston pump
Axial piston pump
Hydraulic pump

Hydraulic pump: gear pump
Gear pumps are fixed displacement pumps since the displaced volume which is
determined by the tooth gap is not adjustable.
Operation principle of the gear pump
The gear pump shown in the diagram is in section. The suction area S is connected
to the reservoir. The gear pump operates according to the following principle:
One gear is connected to the drive, the other is turned by the meshing teeth. The
increase in volume which is produced when a tooth moves out of a mesh causes a
vacuum to be generated in the suction area. The hydraulic fluid fills the tooth gaps
and is conveyed externally around the housing into pressure area P. The hydraulic
fluid is then forced out of the tooth gaps by the meshing of teeth and displaced into
the lines.
Fluid is trapped in the gaps between the teeth between suction and pressure area.
This liquid is fed to the pressure area via a groove since pressure peaks may arise
owing to compression of the trapped oil, resulting in noise and damage.
Example

The leakage oil from the pump is determined by the size of the gap (between
housing, tips of the teeth and lateral faces of the teeth), the overlapping of the
gears, the viscosity and the speed.
These losses can be calculated from the volumetric efficiency since this indicates the
relationship between the effective and the theoretically possible delivery.
Owing to the minimal permissible flow velocity, the suction area in the suction lines
is greater than the pressure area. The result of an undersize suction pipe cross-
section would be a higher flow velocity since the following is valid for v:
A
Q
v =
Where there is a constant flow rate and a smaller cross section, an increase in the
flow velocity results. Consequently, pressure energy would be converted into motion
energy and thermal energy and there would be a pressure drop in the suction area.
Since, whilst hydraulic fluid is being drawn into the suction area, there is a vacuum
in the suction area, this would increase resulting in cavitation. In time, the pump
would be damaged by the effects of cavitation.
The characteristic values and pump characteristics are of importance for the correct
selection and application of pumps.
The table below lists the characteristic values for the most common constant pumps.
Characteristic values for other hydraulic pumps are contained in VDI recommen-
dation 3279.

Types of design Speed range
r.p.m.
Displacement
volume
(cmT
3T
)
Nominal
pressure
(bar)
Total
efficiency
Gear pump,
externally toothed
500 – 3500 1.2 – 250 63 – 160 0.8 – 0.91
Gear pump,
internally toothed
500 – 3500 4 – 250 160 – 250 0.8 – 0.91
Screw pump 500 – 4000 4 – 630 25 – 160 0.7 – 0.84
Rotary vane pump 960 – 3000 5 – 160 100 – 160 0.8 – 0.93
Axial piston pump ……. – 3000
750 – 3000
750 – 3000
100
25 – 800
25 – 800
200
160 – 250
160 – 320
0.8 – 0.92
0.82 – 0.92
0.8 – 0.92
Radial piston pump 960 – 3000 5 – 160 160 – 320 0.90

Couplings are located in the power supply section between the motor and the pump.
They transfer the turning moment generated by the motor to the pump.
In addition, they cushion the two devices against one another. This prevents
fluctuations in the operation of the motor being transferred to the pump and
pressure peaks at the pump being transferred to the motor. In addition, couplings
enable the balancing out of errors of alignment for the motor and pump shaft.
Examples: – rubber couplings
– spiral bevel gear couplings
– square tooth clutch with plastic inserts.
The tank in a hydraulic system fulfils several tasks. It:
• acts as intake and storage reservoir for the hydraulic fluid required for operation
of the system;
• dissipates heat;
• separates air, water and solid materials;
• supports a built-in or built-on pump and drive motor and other hydraulic
components, such as valves, accumulators, etc.
Oil reservoir (tank)
7.3
Coupling
7.4
Reservoir

From these functions, the following guidelines can be drawn up for the
design of the reservoir.
Reservoir size, dependent on:
• pump delivery
• the heat resulting from operation in connection with the maximum permissible
liquid temperature
• the maximum possible difference in the volume of liquid which is produced when
supplying and relieving consuming devices (e.g. cylinders, hydraulic fluid
reservoirs)
• the place of application
• the circulation time.
The volume of liquid supplied by the pump in 3 to 5 minutes can be used as a
reference value for deciding the size of reservoir required for stationary systems. In
addition, a volume of approx. 15% must be provided to balance out fluctuations
in level.
Since mobile hydraulic reservoirs are smaller for reasons of space and weight, they
alone are not able to perform the cooling operations (other cooling equipment is
necessary).
High reservoirs are good for heat dissipation, wide ones for air separation.
These should be as far away from one another as possible and should be located as
far beneath the lowest oil level as possible.
This is used to separate the intake and return areas. In addition, it allows a longer
settling time for the oil and, therefore, makes possible more effective separation of
dirt, water and air.
The base of the tank should slope down to the drain screw so that the deposited
sediment and water can be flushed out.
Reservoir size
Reservoir shape
Intake and return lines
Baffle and separating plate
Base plate

To balance the pressure in case of a fluctuating oil level, the reservoir must be
ventilated and exhausted. For this purpose, a ventilation filter is generally integrated
into the filler cap of the feed opening.
Ventilation and exhaust are not necessary in the case of closed reservoirs as used
for mobile hydraulics. There, a flexible bladder which is prestressed by a gas cushion
(nitrogen) is built into the air-tight container. Because of this, there are fewer
problems with pollution through contact with air and water and premature ageing of
the hydraulic fluid with these containers. At the same time, prestressing prevents
cavitation in the intake line since there is a higher pressure in the reservoir.
Filters are of great significance in hydraulic systems for the reliable functioning and
long service life of the components.
Piston clearance
Z
Valve seat Dirt particles
Detail Z
HIGH PRESSURE
LOW PRESSURE
Effects of polluted oil
Contamination of the hydraulic fluid is caused by:
• Initial contamination during commissioning by metal chips, foundry sand, dust,
welding beads, scale, paint, dirt, sealing materials, contaminated hydraulic fluid
(supplied condition).
• Dirt contamination during operation owing to wear, ingress via seals and tank
ventilation, filling up or changing the hydraulic fluid, exchanging components,
replacing hoses.
Ventilation and exhaust
(air filter)
7.5
Filters

It is the task of the filter to reduce this contamination to an acceptable level in order
to protect the various components from excessive wear. It is necessary to use the
correct grade of filter and a contamination indicator is required in order to check the
efficiency of the filter. Systems are often flushed using economical filters before
commissioning.
Selection and positioning of the filter is largely based on the sensitivity to dirt of the
hydraulic components in use.
Dirt particles are measured in µm, the grade of filtration is indicated accordingly.
Distinction is made between:
• Absolute filter fineness
indicates the largest particle able to pass through a filter
• Nominal filter fineness
particles of nominal pore size are arrested on passing through everal times
• Average pore size
measurement of the average pore size for a filter medium as defined in the
Gaussian process
• β-value
indicates how many times more particles above a specific size are located in the
filter intake than in the filter return
βT
50T
= 10 means that 10 x as many particles larger than 50 µm are located in the filter
intake than in the filter outlet.
Proposed grade of
filtration x in µm,
where β x = 100
Type of hydraulic system
1 – 2 To prevent the most fine degree of contamination in highly sensitive systems
with an exceptionally high level of reliability; mainly used for aeronautics or
laboratory conditions.
2 – 5 Sensitive, powerful control and regulating systems in the high pressure range;
frequently used for aeronautics, robots and machine tools.
5 – 10 Expensive industrial hydraulic systems offering considerable operational
reliability and a planned service life for individual components.
10 – 20 General hydraulic and mobile hydraulic systems, average pressure and size.
15 – 25 Systems for heavy industry or those with a limited service life.
20 – 40 Low pressure systems with considerable play. Grades of filtration and areas of
application
Grades of filtration and areas of application
Grade of filtration
Example

Return filters are built straight onto the oil reservoir, return power filters are
installed in the return line. The housing and filter insert must be designed in such a
way as to stand up to pressure peaks which may occur as a result of large valves
opening suddenly or oil being diverted directly to the reservoir via a by-pass valve
with fast response. The complete return flow is to flow back through the filter. If the
return flow is not concentrated in a common line, the filter may also be used for he
partial flow (in the by-pass flow). Return filtering is cheaper than high pressure
filtering.
Important characteristic values
Operating pressure depending on design, up to max. 30 bar
Flow rate up to 1300 l/min (in the case of filters for reservoir installation)
up to 3900 l/min (large, upright filters for pipeline installation)
Grade of filtration 10 – 25 µm
Perm. Differential
pressure ∆p
Up to approx. 70 bar, dependent on the design of the filter element.
Double filters are used to avoid down times for filter maintenance. In this type of
design, two filters are arranged parallel to one another. If the system is switched
over to the second filter, the contaminated one can be removed without the system
having to shut down.
A
B
Filter unit, reversible
Return filtering

These filters are located in the suction line of the pump; as a result, the hydraulic
fluid is drawn from the reservoir through the filter. Only filtered oil reaches the
system.
Grade of filtration: 60 – 100 µm
These filters are mainly used in systems where the required cleanliness of the
hydraulic fluid cannot be guaranteed. They are purely to protect the pump, and
exhibit a low degree of filtration as particles of 0.06 -0.1 mm are still able to pass
through the filter. In addition, they aggravate pump intake as a result of a
considerable fall in pressure or an increased degree of filter contamination.
Consequently, these filters must not be equipped with fine elements as a vacuum
would be built up by the pump leading to cavitation. In order to ensure that these
intake problems do not occur, suction filters are equipped with by-pass valves.
Suction filter with by-pass
These filters are installed in the pressure line of a hydraulic system ahead of devices
which are sensitive to dirt, e.g. at the pressure port of the pump, ahead of valves or
flow control valves. Since this filter is subjected to the maximum operating pressure,
it must be of robust design. It should not have a by-pass but should have a
contamination indicator.
Important characteristic values
Operating pressure Up to 420 bar
Flow up to 300 l/min
Grade of filtration 3 – 5 µm
Perm. Differential
pressure ∆p
Up to 200 bar, depending on the design of the filter element.
Suction filters
Important characteristic
values
Pressure filters

Hydraulic filters can be arranged in various different positions within a system.
A distinction is made between
• filtering of the main flow: return, inlet and pressure filtering
• filtering of the by-pass flow: only one part of the delivery is filtered.
Circuit diagram
M M M
Advantages
Return flow filter Pump inlet filter Pressure line filter
Filtering of the main flow
economical
simple maintenance
protects pump from
contamination
smaller pore size
possible for valves
sensitive to dirt
Disadvantages contamination can only
be checked having
passed through the
hydraulic components
difficult access,
inlet problems with
fine pored filters.
Result: cavitation
expensive
Remarks frequently used can also be used
ahead of the pump
as a coarse filter
requires a pressure-tight
housing and
contamination indicator
M
By-pass flow filtering
smaller filter possible as an
additional filter
lower dirt-filtering capacity
only part of the delivery is filtered
Filtering of the main flow and By-pass flow filtering
The various possible filter arrangements are listed in the diagram above. The most
favourable filter arrangement is decided by considering the sensitivity to dirt of the
components to be protected, the degree of contamination of the hydraulic fluid and
the costs involved
Filter arrangement

Hydraulic devices Filtration
principle
Arrangement of the
filter in the circuit
Nominal filter
in µm
Axial piston machine Full flow filter Return line and/or
pressure line
≤ 25
Low pressure line ≤ 25 (10)
Gear pumps, radial piston
pumps.
Full flow filter Return line ≤ 63
directional control valves,
pressure valves, flow valves,
non-return valves
Partial flow filter
(additional)
Inlet line ≤ 63
working cylinders
Average speed hydraulic
motors
Full flow filter Return line ≤ 25
Recommended grades of filtration
These filters consist of a thin layer of woven fabric, e.g. metal gauze, cellulose or
plastic fabric. These are disposable filters which are suitable for flushing processes
and for commissioning a system.
These may be made of compressed or multi-layered fabric, cellulose, plastic, glass
or metal fibres or may contain a sintered metal insert. These filters have a high dirt
retention capacity across the same filter area.
Surface filters
Deep-bed filters

Surface filter
Deep-bed filter
Filter design
Filters generally have star-shaped folds in the filter material. In this way, a very large
filter area is achieved with a very small volume.

Specific characteristics are determined by the filter material, the grade of filtration
and the application possibilities. These are shown in the table below.
Element type Grade of filtration
(µm)
Application characteristics
Absolute filter
βT
xT
= 75
3, 5, 10, 20 Safeguards operation and service life of sensitive
components, e.g. servo and proportional valves.
Nominal filter
Polyester
Paper Mat/web
Metal Web
1, 5, 10, 20 Safeguards operation and service life of less
sensitive components; low flow resistance; good
dirt retention capacity.
Wire gauze
Braid weave
25
25, 50, 100
Water and liquids which are difficult to ignite,
employing special steel filter material; high
differential pressure resistance; high dirt retention
capacity.
Operating temperature of 120 °C possible in
special design.
Selection criteria for filter components (HYDAC Co.)
Every filter causes a pressure drop. The following reference values apply here:
Pressure filter ∆p ~ 1 to 1.5 bar at operating temperature
Return line filter ∆p ~ 0.5 bar at operating temperature
Intake filter 1 ∆p ~ 0.05 to 0.1 bar at operating temperature
Main stream filtering

The by-pass pump delivery should be approx. 10% of the tank content. To keep
pressure losses low, the filter should be made sufficiently large. Viscosity also has
an effect on total pressure loss as does the grade of filtration and flow rate.
The viscosity factor f and the pressure loss ∆p from the housing and filter element
are specified by the manufacturer.
The total differential pressure of the complete filter is calculated as
follows:
∆pT
totalT
= ∆pT
housingT
+ f · ∆pT
elementT
Determining the differential pressure for a pressure filter
A pressure loss ∆pT
totalT
is to be calculated for a flow rate of 15 l/min. Filter fineness is
to be 10 µm, kinematic viscosity ν = 30 mmT
2T
/s. The following diagrams are shown as
examples of company specifications.
0 5 10 15 20 25 l/min 30
0
0.4
0.8
1.2
1.6
2.0
bar
Q
∆p
Housing characteristic
By-pass flow filtering
Example

0 5 10 15 20 25 l/min 30
0
0.4
0.8
1.2
1.6
2.0
3 mµ
5 mµ
10 mµ
20 mµ
bar
Q
∆p
Pressure filter-element characteristic
10 30 50 70100 200 300 mm /s
2
1000
0.1
1
5
15
10
20
30
3
Operating viscosity
Factorf
Viscosity factor f

Using these tables, the following values are read off:
∆pT
housingT
= 0.25 bar
∆pT
elementT
= 0.8 bar
f = 1.2
This results in a total pressure difference (pressure loss) of
∆pT
totalT
= 0.25 + 1.2 · 0.8 bar = 1.21 bar
If the reference value for pressure filters amounts to a ∆p of ~ 1 to 1.5 bar, the filter
has been correctly selected.
It is important that the effectiveness of the filter can be checked by a contamination
indicator. The contamination of a filter is measured by the drop in pressure. As the
contamination increases, the pressure ahead of the filter rises. This pressure acts on
a spring-loaded piston. As the pressure increases, the piston is pushed against the
spring.
There are a number of different display methods. Either the piston movement is
visible or else it is converted into an electrical or optical indicator by electrical
contacts.
A
B
Flowdirection
Contamination indicator
Contamination indicators

In hydraulic systems, friction causes energy losses when the hydraulic fluid flows
through the lines and components. This causes the hydraulic fluid to heat up. To a
certain extent, this heat is given off to the environment via the oil reservoir, the lines
and other components.
Operating temperature should not exceed 50 – 60 °C. Where there is a high
temperature, the viscosity of the oil falls by an unacceptable amount, leading to
premature ageing. It also shortens the service life of seals.
If the cooling system of the installation is not powerful enough, the cooler is
generally switched on by thermostat keeping the temperature within specified
limits.
The following cooling devices are available:
• Air cooler: difference in temperature of up to 25 °C possible;
• Water cooler: difference in temperature of up to 35 °C possible;
• Oil cooling by means of air fan cooler: when large quantities of heat must be
dissipated.
Coolers are almost always necessary for mobile hydraulics since the reservoirs are
too small to ensure adequate removal of the heat emitted from the system.
Air cooler (Längerer & Reich)
7.6
Coolers

Water cooler (Längerer & Reich)
Air cooler Water cooler
Description The hydraulic fluid flows from
the return through a pipe which
is cooled by a fan.
Pipes conveying oil are
by-passed by coolant.
Advantages Low running costs.
Easy installation.
Larger heat losses can be
diverted.
No disturbing noises.
Disadvantages Disturbing noise. Higher operating costs.
Susceptible to contamination
and corrosion (coolant).

Heaters are often required to ensure that the optimum operating temperature is
quickly attained. The aim of this is to ensure that when the system is started up, the
hydraulic fluid quickly reaches the optimum viscosity. Where the viscosity is too
high, the increased friction and cavitation lead to greater wear.
Heating elements or flow preheaters are used for heating and preheating hydraulic
fluid.
Heating element (Längerer & Reich)
Stationary systems: 35 – 55 °C in the oil reservoir
Mobile systems: 45 – 65 °C in the oil reservoir
7.7
Heaters
Estimated hydraulic
fluid temperatures

8. Valves
In hydraulic systems, energy is transferred between the pump and consuming device
along appropriate lines. In order to attain the required values – force or torque,
velocity or r.p.m. – and to maintain the prescribed operating conditions for the
system, valves are installed in the lines as energy control components. These valves
control or regulate the pressure and the flow rate. In addition, each valve represents
a resistance.
The nominal sizes of valves are determined by the following characteristic
values:
Nominal diameter in mm
4; 6; 10; 16; 20; 22; 25; 30; 32; 40; 50; 52; 63; 82; 100; 102
Pressure in bar (Pascal) at which hydraulic devices and systems are designed to
work under defined operating conditions; Pressure stages as defined in VDMA
24312: 25; 40; 63; 100; 160; 200; 250; 315; 400; 500; 630
Quantity of oil (l/min) that flows through the valve at a pressure loss of ∆p = 1 bar
(oil viscosity 35 mmT
2T
/s at 40 °C)
The largest quantity of oil (l/min) which can flow through the valve with
correspondingly large pressure losses.
e.g. 20 – 230 mmT
2T
/s (cSt);
e.g. 10 – 80 °C;
8.1
Nominal sizes
Nominal size NW
Nominal pressure NP:
(operating pressure)
Nominal flow QT
nT
Maximum flow QT
maxT
Viscosity range
Hydraulic fluid
temperature range

8. Valves
1 2 3 4 5 6 7 8 9 10 11 12 bar 140
0
8
16
26
6
14
24
22
32
2
10
18
28
4
12
20
l/min
∆p
Q
ϑ
ν
:
:
25°C
20mm /s (cST)
2
P
A
B
A; B
P
T
T
∆p-Q characteristic curve for a 4/2-way valve NW 6
Actuating force
In the case of some types of poppet valve, the actuating force, which is dependent
on pressure and area, may be very great. To avoid this, there must be pressure
compensation at the valves (right-hand diagram).
However, in most cases, it is not possible to design poppet valves to incorporate
pressure compensation. For this reason, high switching forces are required for
actuation which must be overcome by lever transmission or pilot control.
Example

8. Valves
The control edges of the valve are by-passed by oil causing dirt particles to be
washed away (self-cleaning effect). As a result, poppet valves are relatively
insensitive to dirt. However, if dirt particles are deposited on the valve seat, the
valve only partially closes resulting in cavitation.
Various aspects are taken into consideration when classifying valves:
• Function
• Design
• Method of actuation.
A selection is made between the following types of valve based on the tasks they
perform in the hydraulic system:
• Pressure valves
• Directional control valves
• Non-return valves
• Flow control valves.
Poppet valves and piston slide valves are distinguished from one another by the
difference in their design. Overlapping and the geometry of the control edges are
also of significance for the switching characteristics of the valve.
Poppet principle and Slide principle
8.2
Design

8. Valves
In poppet valves a ball, cone, or occasionally a disk, is pressed against the seat area
as a closing element. Valves of this design form a seal when they are closed.
Ball poppet valves
Cone poppet valves
Disk poppet valves
Sectional diagramValve type Advantages and disadvantages/use
simple manufacture; tendency for
ball to vibrate when flow is passing
through producing noise;
Non-return valves
considerable precision is required
to manufacture the cones,
good sealing properties;
only small stroke area;
Shut-off valves
Poppet valves
According to the poppet principle, a maximum of three paths can be opened to a
device by a control element. Overlapping is negative. This means that a valve which
has more than three paths must be constructed from a number of control elements.
A 4/2-way valve on the poppet principle may consist internally of two 3/2-way
valves.
8.3
Poppet valves
Example

8. Valves
A distinction is made between longitudinal and rotary slide valves. A rotary slide
valve is made up of one or more pistons which are turned in a cylindrical bore.
as a rule, shorter than
longitudinal slide valves,
when used as directional
control valves.
Rotary slide valve
The elongated spool valve consists of one or more connected pistons which are
axially displaced in a cylindrical drilled hole. Moving these pistons within the spool
valves can open up, connect together or close any number of connection channels.
Both a 3-way pressure regulator and a 6/4-way directional control valve can be
realised by this principle.
A
P
Elongated spool valve
8.4
Spool valves
Example

8. Valves
To actuate elongated spool valves, it is only necessary to overcome the frictional
and spring forces. Forces resulting from the existing pressure are balanced out by
the opposing surfaces.
Actuating force
A spool must be installed with a certain amount of clearance. This clearance results
in continuous leakage which causes losses in the volumetric flow rate at the valve.
The spring chamber therefore must be connected with a leakage oil line. To prevent
the piston being pressed against the side, the piston skirt area is provided with
circular grooves. When the piston is shifted, only fluid friction arises.
If the hydraulic oil is contaminated, dirt particles appear between the spool and
bore. They act as abrasives and cause the bore to be enlarged. This results in
increased leakage.
Spool principle Poppet principle
flow leakage good sealing
sensitive to dirt non-sensitive to dirt
simple construction even in the case of multi-
position valves
complicated design as multi-position valves
pressure-compensated pressure compensation must be achieved
long actuation stroke short actuation stroke
Comparison of valve constructions

8. Valves
The switching characteristics of a valve are decided by the piston overlap. A
distinction is made between positive, negative and zero overlap. The type of overlap
for the piston control edges can also be varied.
zero
= 0
negative
< 0> 0
positive
Piston overlap
In addition to determining piston clearance, the piston overlap also determines
the oil leakage rate.
Overlapping is significant for all types of valve. The most favourable overlap is
selected in accordance with the application:
• Positive switching overlap
During the reversing procedure, all ports are briefly closed against one another;
no pressure collapse (important in the case of systems with reservoirs);
switching impacts resulting from pressure peaks; hard advance;
• Negative switching overlap
During the reversing procedure, all ports are briefly interconnected; pressure
collapses briefly (load drops down);
• Pressure advanced opening
The pump is first of all connected to the power component, then the power
component is discharged to the reservoir;
• Outlet advanced opening
The outlet of the power component is first discharged to the reservoir before the
inlet is connected to the pump;
• Zero overlap
Edges meet. Important for fast switching, short switching paths.
8.5
Piston overlap

8. Valves
In the case of multi-position valves, piston overlapping within a valve may vary with
the application. Even switching overlaps are adapted to requirements. When repairs
are necessary, it is important to ensure that the new piston has the same overlaps.
The effect of positive and negative overlap is shown below based on the example of
a single-acting cylinder, triggered by a 3/2-way valve.

8. Valves
P
A
T
T
P
50 bar
50 bar
P
A
T
T
P
50 bar
50 bar
P T
A
m
m
Port P A is opened
only after A T is closed.
Positive switching overlap
System pressure affects the cylinder immediately, hard advance.

8. Valves
P
A
T
T
P
50 bar
~0 bar
P
A
T
T
P
50 bar
50 bar
P T
A
m
m
Port P A is opened
although port A T is
not closed yet.
Thus, all ports are
briefly interconnected.
Negative switching overlap
Pressure is reduced during the reversing procedure, gentle build-up of pressure for
approach.

8. Valves
As with spool valves, any switching overlap can be achieved with 2/2-way poppet
valves.
R P
BA
T
x1 x3x2 x4
Switching overlap with poppet valves
In the case of spool valves, the switching overlap is decided by the geometry of the
control edge and the inflexible connection of the control piston.
Where poppet valves are concerned, the desired switching overlap is achieved by
varying response times of the various valves and can be changed, if required, by
altering the switching times.

8. Valves
The control edges of the piston are often either sharp, chamfered or notched. This
profiling of the control edge has the effect that there is gradual rather than sudden
throttling of the flow on switching.
control edge with axial notches
chamfered control edge
sharp control edge
Control edges
The pressure in the valve causes the piston to be pressed against the bore in the
housing. This results in considerable frictional forces and, consequently, high
actuating forces being produced. The pressure is balanced out by annular grooves
on the piston circumference. The piston is then supported on a film of oil. On
actuation, only the fluid friction needs to be overcome.
Annular grooves
There are various methods of actuation for valves. In addition, valves may also be
electrically, pneumatically or hydraulically actuated.
8.6
Control edges
Actuating force

8. Valves
There are two methods of port designation. The ports can be labelled either with the
letters P, T, A, B and L or they can be labelled alphabetically.
Valves have several switching positions. The following rule is applied to determine
which ports are interconnected and which ones are closed against each other:
• A horizontal line between the letters for the ports (e.g. P-A) means that the ports
are connected together;
• An individual letter separated by a comma (e.g. P-A, T) signifies that this port
(here: T) is blocked.
P-A-B-T: all ports are interconnected.
B
T
A
P
P-A-B, T: P, A and B are connected, T is blocked.
B
T
A
P
Port designations
Examples

Pressure valves have the task of controlling and regulating the pressure in a
hydraulic system and in parts of the system.
• Pressure relief valves
The pressure in a system is set and restricted by these valves. The control
pressure is sensed at the input (P) of the valve.
• Pressure regulators
These valves reduce the output pressure where there is a varying higher input
pressure. The control pressure is sensed at the output of the valve.
The symbols for the different pressure valves are shown below.
P(A)
T(B)
A(B)
P(A)
P(A) T
A(B)
2-way pressure regulator
Pressure valves
Pressure relief valves are designed in the form of poppet or slide valves. In the
normal position,
• a compression spring presses a sealing element onto the input port
or
• a slide is pushed over the opening to the tank connection.
9. Pressure valves
9.1
Pressure relief valves

9. Pressure valves
P T
BA
P T
Ts
M
P
T
Pressure relief valves (circuit diagram)
Pressure relief valves (sectional diagram)

9. Pressure valves
Pressure relief valves operate according to the following principle: The input
pressure (p) acts on the surface of the sealing element and generates the force
F = pT
1T
⋅ AT
1T
.
The spring force with which the sealing element is pressed onto the seat is
adjustable.
If the force generated by the input pressure exceeds the spring force, the valve
starts to open. This causes a partial flow of fluid to the tank. If the input pressure
continues to increase, the valve opens until the complete pump delivery flows to the
tank.
Resistances at the output (tank line, return line filter, or similar) act on the surface
AT
2T
. The resultant force must be added to the spring force. The output side of the
valve may also be pressure-compensated (see pressure relief valve with cushioning
and pressure compensation).
Cushioning pistons and throttles are often installed in pressure relief valves to
eliminate fluctuations in pressure. The cushioning device shown here causes:
• fast opening
• slow closing of the valve.
By these means, damage resulting from pressure surges is avoided (smooth valve
operation). Pressure knocks arise when the pump supplies the hydraulic oil to the
circuit in an almost unpressurised condition and the supply port is suddenly closed
by a directional control valve.
In the circuit diagram shown here, the total pump delivery flows at maximum
pressure via the pressure relief valve to the tank. When the directional control valve
is switched, the pressure in the direction of the cylinder decreases and the
cushioned pressure relief valve closes slowly. An uncushioned valve would close
suddenly and pressure peaks might occur.

9. Pressure valves
P T
BA
P T
Ts
M
P
T
Pressure relief valve (circuit diagram)
Pressure relief valve with cushioning (sectional diagram)

9. Pressure valves
Pressure relief valves are used as:
• Safety valves
A pressure relief valve is termed a safety valve when it is attached to the pump,
for example, to protect it from overload. The valve setting is fixed at the
maximum pump pressure. It only opens in case of emergency.
• Counter-pressure valves
These counteract mass moments of inertia with tractive loads. The valve must be
pressure-compensated and the tank connection must be loadable.
• Brake valves
These prevent pressure peaks, which may arise as a result of mass moments of
inertia on sudden closing of the directional control valve.
• Sequence valves (sequence valves, pressure sequence valves)
These open the connection to other consuming devices when the set pressure
is exceeded.
• There are both internally and externally controlled pressure relief valves.
Pressure relief valves of poppet or slide design may only be used as sequence
valves when the pressure is compensated and loading at the tank connection has
no effect on the opening characteristics.
P
BA
P T
Ts
M
T
T
P
160 bar
(16 MPa)
100 bar
m
Break valve
Application example: brake valve

9. Pressure valves
The diagram below shows a cushioned pressure valve of poppet design.
Pressure relief valve, internally controlled, cushioned
Pressure relief valve, externally controlled

9. Pressure valves
P T
Ts
M
m
T
P
P T
BA
P
T
Counter-balance valve
20 bar
System pressure limit
100 bar
Safety valve
160 bar
Application example: counter-balance valve

9. Pressure valves
Pressure regulators reduce the input pressure to a specified output pressure.
They are only used to good effect in systems where a number of different pressures
are required. To clarify this, the method of operation is explained here with the help
of an example with two control circuits:
• The first control circuit operates on a hydraulic motor via a flow control valve in
order to drive a roller. This roller is used to stick together multi-layer printed
wiring boards.
• The second control circuit operates on a hydraulic cylinder which draws a roller
towards the boards at a reduced, adjustable pressure. The roller can be lifted
with a cylinder to allow the boards to be inserted (piston rod extends).
P T
Ts
M
P
A
P PT T
T
T
A A
P
P
P
A
FPulling
Example: 2-way pressure regulator
9.2
Pressure regulators

9. Pressure valves
The pressure regulator in the circuit diagram operates according to the following
principle:
The valve is opened in the normal position. The output pressure at (A) is transmitted
to the piston surface (1) via a control line (3). The resultant force is compared to the
set spring force. If the force of the piston surface exceeds the set value, the valve
starts to close as the valve slide moves against the spring until an equilibrium of
forces exists. This causes the throttle gap to be reduced and there is a fall in
pressure. If the pressure at output (A) increases once again, the piston closes
completely. The pressure present in the first control circuit prevails at output (A).
Pressure regulators of poppet design open and close very quickly in the case of
short strokes and may as a result flutter with fast changes in pressure; this is
prevented by adding cushioning.
In the case of slide valves, it is also possible to influence opening characteristics by
having control edges shaped in such a way that the opening gap increases slowly.
This will result in greater control precision and lead to improvements in the
oscillation characteristics of the valve.

9. Pressure valves
The 2-way pressure regulator dealt with earlier might be used, for example, when a
constant low pressure is required for a clamping device in a by-pass circuit of the
hydraulic installation.
In the example shown here, however, problems may arise with the 2-way pressure
regulator.
P T
Ts
M
P
A
P PT T
T
T
A A
P
P
P
A
Circuit with 2-way pressure regulator
If the 2-way pressure regulator closes, thickening of the workpiece material causes a
further pressure increase at output (A) of the pressure regulator. This increase in
pressure above the set value is not desired. One method of rectifying this would be
to install a pressure relief valve at the output.
Example

9. Pressure valves
The 2-way pressure regulator is rarely used in practice. Its design does not permit a
reduction from a high set pressure to a low pressure.
AA(B)
P(A)
L
T
Pressure relief valve to prevent increases in pressure
This pressure relief valve can be set in various ways:
• PRV setting greater than that for pressure regulator;
• PRV setting equal to that of pressure regulator;
• PRV setting lower than that of pressure regulator.
These settings produce various characteristics in the pressure regulator.
Another method of reducing these increases in pressure is to use a 3-way pressure
regulator.
The method of operation of a 3-way pressure regulator is identical to that of a 2-way
pressure regulator with respect to flow from P to A.
However, an increase in pressure above that which has been set at output (A)
causes a further shift of the piston. The built-in pressure relief function comes into
force and opens a passage from A to T.

9. Pressure valves
P T
Ts
M
P PT T
T
T
A A
P
P
P
A
P T
A
Circuit diagram for a 3-way pressure regulator
In the case of the 3-way pressure regulator, the overlap forms part of the design.
However, where a 2-way pressure regulator is combined with a pressure relief valve,
the overlap is adjustable.
As external forces act on the cylinder in this pressure roller, a 3-way pressure
regulator or a 2-way pressure regulator combined with a pressure-relief valve should
be installed.
It is a good idea to use the 3-way pressure regulator with negative overlap (T opens
before P closes). Where a 2-way pressure regulator is combined with a pressure
relief valve, the pressure relief valve should be set to a lower pressure than the
pressure regulator.
Note

Directional control valves are components which change, open or close flow paths
in hydraulic systems. They are used to control the direction of motion of power
components and the manner in which these stop. Directional control valves are
shown as defined in DIN ISO 1219.
A P LA P L
A
P L
2/2-way valve
10. Directional control valves

The following rules apply to the representation of directional control valves:
• Each different switching position is shown by a square.
• Flow directions are indicated by arrows.
• Blocked ports are shown by horizontal lines.
• Ports are shown in the appropriate flow direction with line arrows.
• Drain ports are drawn as a broken line and labelled (L) to distinguish them from
control ports.
Each individual switching position is shown in a square
Two flow paths
Flow paths are indicated by means of arrows within the square
Closed position
Two ports are connected, two are closed
Three ports are connected, one is closed
All ports are connected
Switching positions
P P
P P
B B
B B
A A
A A
T T
T T
Examples: switching positions
Symbols for directional
control valves

There are two types of directional control valve: continually operating and binary*
directional control valves.
(* two values possible (0 or 1): 1 = output present, 2 = output not present)
In addition to two end positions, these valves can have any number of intermediate
switching positions with varying throttle effect. Proportional and servo valves which
are discussed in the TP 700 training books are examples of this type of valve.
These always have a fixed number (2, 3, 4, ...) of switching positions. In practice,
they are known simply as directional control valves. They are central to hydraulics
and form an important part of the subject matter of this book.
Directional control valves are classified as follows according to the number of ports:
• 2/2-way valve
• 3/2-way valve
• 4/2-way valve
• 5/2-way valve
• 4/3-way valve.
The diagram on the following page shows the symbols used for directional control
valves. For the sake of simplicity, the actuation methods have been omitted.
Many other designs are available for use in particular fields of application.
Continuously operating
directional control valves
Digitally operating
directional control valves

4/3-WV
4/3-WV
4/3-WV
4/2-WV
4/3-WV
4/3-WV
5/2-WV
2/2-WV
3/2-WV
Directional control valve
P
P
P
P
P
B
B
B
B
B
A
A
A
A
A
T
T
T
T
T
P
P
P
P
R
P
A
A
A
A
A
B
B
A
T
T
T
P
T
Normal position
"closed" (P, A)
Normal position
"closed" (P, T A)
Normal position
"flow" (P B, A T)
Mid position
"closed" (P, A, B, T)
Mid position "Pump
re-circulating" (P T, A, B)
Mid position
"By-pass" (P A B, T)
"H" mid position
(P A B T)
Mid position "working lines
de-pressurised"
(P, A B T)
Normal position
"flow" (A R, P B, T)
Normal position
"flow" (P A)
Normal position
"flow" (P A, T)

The 2/2-way valve has a working port (A) and a pressure port (P) (see diagram). It
controls the delivery by closing or opening the passage. The valve shown here has
the following switching positions:
A P L A P L
AA
PP L
2/2 way valve, spool design
• Normal position: P to A closed
• Actuated position: Flow from P to A
2/2-way valve, poppet design
10.1
2/2-way valve

Symbols for poppet valves are often drawn to include the symbol for the valve seat.
This representation is not standard. This valve is also available with “flow from P to
A” in the rest position.
A
P
Symbol, poppet valve
P T
Ts
M
TP
PA
m
Triggering a single acting cylinder (circuit diagram)

P T
Ts
M
T
A LP
P
m
Triggering a single acting cylinder (sectional diagram)
Other possible applications:
• As a by-pass, e.g. rapid traverse feed circuit, pressurizes pump by-pass;
• Switching on or off various flow or pressure valves;(pressure stage circuit)
• Triggering a motor in a single direction.

M
Further application possibilities

The 3/2-way valve has a working port (A), a pressure port (P) and a tank connection
(T). It controls the flow rate via the following switching positions:
• Normal position: P is closed and A to T is open;
• Actuated position: Outlet T is closed, flow from P to A.
3/2-way valve can be normally open, i.e. there may be a flow from P to A.
3/2-way valve
P T
A
P T
Ts
M
L
Triggering a single acting cylinder
10.2
3/2-way valve

Triggering a single acting cylinder, sectional diagram
2 l/min 4 l/min
Heizer Kühler
In use as shunt

The 4/2-way valve has two working ports (A, B), a pressure port (P) and a tank
connection (T).
• Normal position: flow from P to B and from A to T;
• Actuated position: flow from P to A and from B to T.
4/2-way valve with three control pistons
T
P T
Ts
M
LP
BA
Triggering a double acting cylinder – circuit diagram
10.3
4/2-way valve

Triggering a double acting cylinder – sectional diagram
4/2-way valves are also constructed with just two control pistons. These valves do
not require any drain ports. It should be borne in mind that tank connection T and
working ports A and B are routed via the end cap of the valve in this design.
For this reason, in data sheets about these valves, a smaller maximum pressure is
specified from the tank connection than for the pressure side because the pressure
at this port is effective at the cover cap.
4/2-way valve with two control pistons
The simplest type of design for 4/2-way valves is that of the spool valve. 4/2-way
valves of poppet design, on the other hand, are complicated as they are put together
from two 3/2-way or four 2/2-way valves.

Overlapping positions are an important consideration in the selection of valves. For
this reason, they are often indicated in detailed representations of the symbol. As no
actual switching positions are shown, the relevant box in the diagram is drawn with
thinner, broken lines.
Symbol: positive switching overlap
Symbol: negative switching overlap
Overlapping position 4/2-way valve
Possible applications of the 4/2-way valve:
• Triggering of double-acting cylinders;
• Triggering of motors with either clockwise or anti-clockwise rotation;
• Triggering of two circuits.
A 5/2-way valve may also be used in place of the 4/2-way valve.
T
P
R
A B
T
P
5/2-way valve
Overlapping positions

4/3-way valves constructed as spool valves are of simple construction, whilst those
constructed as poppet valves are complex in design. 4/3-way valves of poppet valve
design may be composed, for example, of four individual two-way valves.
Mid position – closed
Mid position – pump by-pass
Mid position – by-pass
H – mid position
Mid position – working lines de-pressurised
4/3-way valves
The overlapping positions are specified for 4/3-way valves:
Overlap positions –example
The 4/3-way valve shown here has positive overlap in the mid position. Left-hand
and right-hand overlap positions are a combination of positive and negative overlap.
10.4
4/3-way valve

The mid position is decided by the control problem. Multi-position valves are also
constructed as 5-way valves.
5/3-way valve
4/3-way valve with pump by-pass (re-circulating)

Only one control loop system can be driven by this valve.
T
P T
Ts
M
BA
P L
Pump by-pass
Pump by-pass, sectional diagram

4/3-way valve, mid position closed

If a number of control circuits are to be powered, 4/3-way valves with mid position
closed can be used to trigger individual control circuits. When an operational system
is to be switched to pump by-pass, a 2/2-way valve is used.
Application examples
One of the main applications of 4/3-way valves consists in triggering double acting
cylinders and motors (stop, clockwise rotation, anticlockwise rotation).

Non-return valves block the flow in one direction and permit free flow in the other.
As there must be no leaks in the closed direction, these valves are always of poppet
design and are constructed according to the following basic principle:
The sealing element (generally a ball or cone) is pressed against an appropriately
shaped seat. The valve is opened by volumetric flow in the flow direction, the sealing
element being lifted from the seat.
Non-return valves are distinguished as follows:
• Non-return valves (unloaded, spring-loaded)
• Lockable and unlockable non-return valves.
B
A
X
B
A X
B2
A2
B1
A1
Non-return valve, unloaded
Non-return valve, spring-loaded
Lockable non-return valve,
opening of the valve is prevented by
a pilot air supply or hydraulic supply
De-lockable non-return valve,
closing of the valve is prevented by
a pilot air supply or hydraulic supply
Shuttle valve
De-lockable (piloted) double non-return valve
Non-return valves
11. Non-return valves

p1p2
pF
Symbol:
ACone
Flow openFlow blocked
Sealing conePressure spring
Spring loaded non-return valve
If a pressure (pT
1T
) operates on the sealing cone, this is lifted from its seat releasing
the flow when the valve is not spring-loaded. Counter pressure pT
2T
must be overcome
here. As the non-return valve shown here is spring-loaded, the spring force operates
on the sealing cone in addition to the counter pressure pT
2T
and flow is produced
when:
PT
1T
> pT
2T
+ pT
FT
The following equation is valid for the pressure exercised by the spring:
cone
spring
F
A
F
p =
11.1
Non-return valve
168 © Festo Didactic GmbH & Co. • TP501

The diagrams show possible applications of non-return valves.
P T
Ts
M
P T
T
BA
P
m
Pump protection
Possible applications

P T
Ts
M
P T
T
BA
P
m
Pump protection
When the electric motor is switched off, the load pressure cannot drive the pump
backwards. Pressure peaks which occur in the system do not affect the pump but are
diverted by the pressure relief valve.

By pass
PRV as brake valve
Flow valve only effective
in one direction
Suction retaining valve
for a press
By-passing contaminated filter
(opening pressure 0.5 – 3 bar)
Graetz-rectifer circuit
Suction retaining valve for
a rotating mass
By pass
flow regulator
Applications

In piloted non-return valves, flow can be released in the closed position by pilot
control of the valve poppet. This takes place according to the following principle:
Flow is possible from A to B, flow is blocked from B to A.
Flow blocked from B to A
Flow from A to B
Flow from B to A
If the hydraulic fluid is to flow from B to A, the valve poppet with the de-locking
piston must be lifted away from its seat. The de-locking piston is pressurised via
control port X.
11.2
Piloted non-return valve

For reliable de-locking of the valve, the effective surface on the pilot piston must
always be greater than the effective surface on the sealing element. The area ratio is
generally 5 : 1. Piloted non-return valves are also made with pre-discharge.
The method of operation of a piloted non-return valve in a hydraulic system is
explained below using circuit diagrams:
P T
Ts
M
P PT T
BA A
B
A X
m
De-lockable non-return valve
The 3/2-way valve blocks the hydraulic flow in the normal position. Oil flow is
released at the 4/2-way valve on the piston rod side. The piston rod cannot retract
as the non-return valve is blocked. Once the 3/2-way valve is actuated, the pilot
piston is pressurised and the sealing element of the non-return valve opens. This
allows the hydraulic fluid to flow away from the piston side via the 4/2-way valve to
the reservoir.
When the 4/2-way valve is actuated, the hydraulic fluid flows via the non-return
valve to the cylinder – the piston rod extends.
Method of operation

A piloted non-return valve which is raised only closes when the control oil can be
discharged from the pilot port to the reservoir. For this reason, using a piloted non-
return valve calls for a special mid-position of the 4/3-way valve.
P T
Ts
M
m
1000kg
T
B
A X
BA
P
The piloted non-return valve cannot close immediately as pressure can only escape
from the closed control port X via the leakage from the directional control valve.
Mid-position “closed”

P T
Ts
M
T
B
A X
BA
P
m
1000kg
Since in this mid-position ports A and B are connected to T, and P is closed, both
control port X and port B are exhausted at the non-return valve. This causes the non-
return valve to close immediately.
With the piloted double non-return valve, a load can be reliably positioned above
the cylinder piston even where there is internal leakage. However, this reliable
positioning is only possible with supporting cylinders. Positioning by a piloted
double non-return valve is not possible in the case of hanging cylinders or cylinders
with through-rods.
The diagram below shows both the simplified and complete symbols for a piloted
double non-return valve and its assembly.
Mid-position “Working lines
de-pressurised”
11.3
Piloted double non-return
valve

B2
A2
B1
complete simplified
(not standardised)
A1
B2B1
A2A1
Piloted double non-return valve, symbol
P T
Ts
M
T
BA
P
B2
A2
B1
A1
m
Application example

Piloted double non-return valve, closed
Piloted double non-return valve, open
The piloted double non-return valve operates according to the following principle:
Free flow is possible either in the flow direction from AT
1T
to BT
1T
or from AT
2T
to BT
2T
, flow is
blocked either from BT
1T
to AT
1T
or from BT
2T
to AT
2T
.
If flow passes through the valve from AT
1T
to BT
1T
, the control piston is shifted to the
right and the valve poppet is lifted from its seat. By these means, flow is opened
from BT
2T
to AT
2T
(the valve operates in a corresponding manner where there is flow from
AT
2T
to BT
2T
).

12. Flow control valves
Flow control valves are used to reduce the speed of a cylinder or the r.p.m. of a
motor. Since both values are dependent on the flow rate, this must be reduced.
However, fixed displacement pumps supply a uniform flow rate. Reduction in the
rate of flow supplied to the drive element is achieved according to the following
principle:
A reduction in the flow cross-section in the flow control valve causes an increase in
pressure ahead of this. This pressure causes the pressure relief valve to open and,
consequently, results in a division of the flow rate. This division of the flow rate
causes the flow volume required for the r.p.m. or speed to flow to the power
component and the excess delivery to be discharged via the pressure relief valve.
This results in a considerable energy loss.
In order to save energy, adjustable pumps can be used. In this case, the increase in
pressure acts on the adjustable pump device.
On the basis of their controlling or regulating function, flow control valves are
classified as either:
• flow control valves or
• flow regulating valves.
Examples of flow control valves as restrictors and orifice valves:
A B A B A B
Q = variablepartial Q = constantpartial
Flow control valves
Control valves
Restrictor type
dependent on load independent of load
Orifice type
Regulating valves
Restrictors and orifice valves

Restrictors and orifice valves represent a flow resistance. This resistance is
dependent on the flow cross-section and its geometric form and on the viscosity of
the liquid. When hydraulic fluid is passed through the flow resistor, there is a fall in
pressure as a result of friction and of an increase in the flow velocity. The part of the
pressure drop caused by friction can be considerably reduced by changing the
orifice shape. In order to obtain the required resistance using an orifice, turbulence
must be achieved by increasing the flow velocity (smaller cross-section than that of
a corresponding restrictor). In this way, the resistance of the orifice is determined by
the turbulence and becomes independent of viscosity. For this reason, orifice valves
are used in cases where independence from temperature and, therefore, from
viscosity is required, e.g. in flow gauges.
Restrictor Orifice
Restrictor and orifice
In many control systems, on the other hand, a specified high fall in pressure is a
requirement. In such cases, restrictors are used.
Restrictors and orifice valves control the flow rate together with a pressure relief
valve. The valve resistance causes pressure to build up ahead of these valves.
The pressure relief valve opens when the resistance of the restrictor is greater than
the set opening pressure at the pressure relief valve. As a result, the flow is divided.
Part of the pump delivery flows to the consuming device, the other part is
discharged under maximum pressure via the pressure relief valve (high power loss).
The partial flow passing through the throttling point is dependent on the pressure
difference ∆p. The interrelationship between ∆p and the flow QT
consuming deviceT
corresponds to:
12.1
Restrictors and
orifice valves
2
devicegminconsuQ~p∆

The inlet pressure to the valve is kept at a constant level by the pressure relief valve.
The pressure difference ∆p is changed by altering the load coming from the
consuming device. The result of this is that there is a change in the rate of flow to
the consuming device, i.e.:
The operation of restrictors is flow-dependent.
Consequently, they are not suitable for adjusting a constant flow rate in the case of a
changeable load.
0
At a pressure of 100 bar, the max. volumetric flow
exits via the pressure-relief valve
Opening point of the pressure-relief valve
Opening characteristic of the pressure-relief valve
Total resistance value set with restrictor
Division point
2.5
Q proportion, pressure-relief valve Q proportion, cylinder
5 7.5 l/min
80
90
1
10
00
bar
Qmax.
Settingvalue,
pressure-reliefvalve
Characteristic

Q
M
v
P T
Ts
T
BA
P
P
T
Qconsuming device
Q
QPRV
Restrictor
p
(variable)
2
p
(constant)
1
∆p
variable
Flow division point
Restrictor – Flow division
The requirements for adjustable restrictors are as follows:
• build-up of a resistance;
• constant resistance in the face of changing hydraulic fluid temperatures, i. e.
independent of viscosity;
• sensitive adjustment – the sensitivity of adjustment of a restrictor is dependent
amongst other things, on the ratio of the orifice cross-sectional area to the
control surface area;
• economical design.
Adjustable restrictors

The various designs of adjustable restrictor fulfil these requirements with varying
degrees of success:
Type Resistance Dependence
on viscosity
Ease of adjustment Design
Needle
restrictor
Increase in
velocity, high
friction owing to
long throttling
path
Considerable
owing to high
friction
Excessive cross-
sectional enlargement
with a short
adjustment travel,
unfavourable ratio
area to control surface
Economical,
simple design
Circum-
ferential
restrictor
As above As above, but
lower than for
the needle
restrictor
Steadier cross-
sectional
enlargement, even
ratio area to control
surface, total
adjustment travel only
90°.
Economical,
simple design,
more
complicated
than the needle
restrictor
Longitudinal
restrictor
As above As above As above, however
sensitive adjustment
owing to long
adjustment travel
As for
circumferential
restrictor
Gap restrictor Main part:
increase in
velocity, low
friction, short
throttling path
Low Unfavourable, even
cross-sectional
enlargement,
adjustment travel of
180°
Economical
Gap restrictor
with helix
Increase in
velocity,
maximum
friction
Independent Sensitive, even cross-
sectional enlarge-
ment, adjustment
travel of 360°
Expensive to
produce helix
Adjustable restrictors

The one-way flow control valve where the restrictor is only effective in one direction
is a combination of a restrictor and a non-return valve. The restrictor controls the
flow rate in a single direction dependent on flow. In the opposite direction, the full
cross-sectional flow is released and the return flow is at full pump delivery. This
enables the one-way flow control valve to operate as follows:
The hydraulic flow is throttled in the flow direction from A to B. This results in flow
division as with the restrictor. Flow to the power component is reduced, the speed
being reduced correspondingly.
Flow is not restricted in the opposite direction (B to A) as the sealing cone of the
non-return valve is lifted from its valve seat and the full cross-sectional flow is
released.
With adjustable one-way flow control valves, the throttling point can either be
enlarged or reduced.
One-way flow control valve
12.2
One-way flow control valve

As has already been described in the section on restrictors, there is an
interrelationship between pressure drop ∆p and volumetric flow Q:
∆p ~ Q2.
If, in the case of a changing load, an even flow rate to the consuming device is
required, the pressure drop ∆p via the throttle point must be kept constant.
Therefore, a restrictor (2) (adjustable restrictor) and a second restrictor (1)
(regulating restrictor or pressure balance) are built-in for the desired flow rate.
These restrictors change their resistance according to the pressures present at the
input and output of the valve. The total resistance of the two restrictors combined
with the pressure relief valve causes the flow division.
The regulating restrictor (1) can be installed either ahead of or behind the adjustable
restrictor.
The valve is open in the normal position. When flow passes through the valve, input
pressure pT
1T
is produced ahead of the adjustable restrictor. A pressure drop ∆p is
produced at the adjustable restrictor, i.e. pT
2T
< pT
1T
. A spring must be installed on the
side FT
2T
so that the regulating restrictor retains its equilibrium. This spring causes the
constant pressure difference across the adjustable throttle. If a load passes from the
consuming device to the valve output, the regulating restrictor reduces the
resistance by the amount by which the load has increased.
12.3
Two-way flow control valve

During idling, the spring helps to keep the regulating restrictor in equilibrium and
the valve provides a certain resistance which is adjusted in line with the desired flow
rate using the adjustable restrictor.
If the pressure at the output of the valve increases, the pressure pT
3T
also increases.
As a result, the pressure difference is modified via the adjustable restrictor. At the
same time, pT
2T
operates on the piston surface AT
P2T
. The force which arises combines
with the spring force to act on the regulating restrictor. The regulating restrictor
remains open until there is once more a state of equilibrium between the forces FT
1T
and FT
2T
and, therefore, the pressure drop at the adjustable restrictor regains its
original value. As with the restrictor, the residual flow not required at the 2-way flow
control valve is discharged via the pressure relief valve to the tank.
Q
M
p1
p2
∆pkonstant
P T
Ts
P
T
Pressure balance
Pressure balance
Adjusting restrictor
Adjusting
restrictor
If the pressure pT
3T
at the output of the valve falls, the pressure difference ∆p
increases. As a result, the pressure acting on the piston surface AT
P2T
is also reduced
with the consequence that the force FT
1T
becomes greater than FT
2T
. The regulating
restrictor now recloses until an equilibrium is established between FT
1T
and FT
2T
.

The same regulating function operates with fluctuating input pressures. With
changed input conditions, ∆p via the adjustable restrictor and, thus, also the
delivery to the consuming device remain constant.
As previously discussed, the function of the regulating restrictor is to balance out
changes in load at the input or output through modification of its flow resistance,
and, by these means, to maintain a constant pressure difference via the adjustable
restrictor. For this reason, there must be an equilibrium of forces at the regulating
piston so that it can adjust to changing loads; i.e. FT
1T
= FT
2T
.
FT
1T
is produced from the area AT
P1T
and the pressure pT
1T
. FT
2T
results from the area AT
P2T
,
which is equal to AT
P1T
and the pressure pT
2T
. Since the pressure pT
2T
is reduced by the
resistance of the adjustable restrictor, a spring must be installed for the purposes of
balance.
FT
1T
= FT
2T
AT
K1T
= AT
K2T
FT
1T
= AT
K1T
⋅ pT
1T
FT
2T
= AT
K2T
⋅ pT
2T
+ FT
FT
AT
K1T
⋅ pT
1T
= AT
K1T
⋅ pT
2T
+ FT
FT
AT
K1T
(pT
1 T- pT
2T
) = FT
FT
(pT
1 T- pT
2T
) =
1K
F
A
F
This means: The constant spring force FT
FT
divided by the piston area AT
P1T
equals the
pressure difference ∆p. This difference across the adjustable restrictor is always
kept constant as shown by the following examples.
In practice, adjustable restrictors are generally designed in the form of adjustable
orifices so that the flow control valve remains to a large degree unaffected by
viscosity.
Tasks of the
regulating restrictor
Note

Q =
3 l/min
CD
p = 5 bar3
p = 144 bar2
p = 148 bar1
p = 150 bar
Q = 10 l/minp
Q = 7 l/minPRV
p = 150 bar
∆p = 4 bar
∆p = 139 bar
T
BA
A
P
P
P
T
2-way flow control valve, loading of the consuming device (idling)

Q =
3 l/min
CD
p = 40 bar3
p = 144 bar2
p = 148 bar1
p = 150 bar
Q = 10 l/minp
Q = 7 l/minPRV
p = 150 bar
∆p = 4 bar
∆p = 104 bar
T
BA
A
P
P
P
T
F
2-way flow control valve, loading of the consuming device (under load)

Q =
3 l/min
CD
p = 30 bar3
p = 104 bar2
p = 108 bar1
p = 110 bar
Q = 10 l/minp
Q = 0 l/minPRV
p = 150 bar
∆p = 4 bar
∆p = 74 bar
Q = 7 l/min
T
T
BA
A
A A
P P
P
P
P
P
T
T
F
In connection with other consuming devices

There is both a detailed and a simplified symbol for the 2-way flow control valve.
T T
B BA A
A
P
P P
A
P
AP
M M
P PT T
Ts Ts
AP

2-way flow control valves may be used either in the inlet and/or outlet and for by-
pass flow control.
Disadvantage of by-pass flow control: The uneven pump delivery caused by
fluctuations in speed has an effect on the flow quantity to be regulated.
2-way flow control valves provide a constant flow rate in the face of changing loads
meaning that they are suitable for the following application examples:
• Workpiece slides which operate at a constant feed speed with varying working
loads;
• Lifting gear where the lowering speeds need to be carefully restricted.
The flow control valve is opened when the system is at a standstill. Once the system
has been switched on, there is a higher flow rate until the pressure balance has been
set to the desired position; this procedure is referred to as the initial jump. There are
several ways to reduce the initial jump.
• A by-pass valve opens before the main valve opens.
• Or the measuring restrictor is closed by a spring in unpressurised status.
Note

The hydraulic cylinder converts hydraulic energy into mechanical energy. It
generates linear movements. For this reason, it is also referred to as a “linear
motor”.
There are two basic types of hydraulic cylinder
• single-acting and
• double-acting cylinders.
Sectional views of these two basic types are shown in the diagrams below.
21 6 743 5
1 Mounting screw
2 Vent screw
3 Piston rod
4 Cylinder barrel
5 Piston rod bearing
6 Piston rod seal
7 Wiper
1 2
345
1 Piston
2 Piston rod
3 Piston rod bearing
4 Annular piston surface
5 Piston surface
13. Hydraulic cylinders

In single-acting cylinders, only the piston side is supplied with hydraulic fluid.
Consequently, the cylinder is only able to carry out work in one direction. These
cylinders operate according to the following principle:
The hydraulic fluid flows into the piston area. Owing to the counter force
(weight/load), pressure builds up at the piston. Once this counter force has been
overcome, the piston travels into the forward end position.
During the return stroke, the piston area is connected to the reservoir via the line
and the directional control valve whilst the pressure line is closed off by the
directional control valve. The return stroke is effected either by intrinsic load, by a
spring or by the weight load. In the process, these forces (load forces) overcome the
frictional forces in the cylinder and in the lines and valves and displace the hydraulic
fluid into the return line.
P PT T
Ts Ts
M M
T T
A A
P P
m
Single acting cylinder – hydraulic ram
13.1
Single-acting cylinder

Single-acting cylinders are used wherever hydraulic power is required for only one
direction of motion.
For lifting, clamping and lowering workpieces, in hydraulic lifts, scissor lifting tables
and lifting platforms.
Telescopic cylinder
Designation Description
Hydraulic ram piston and rod form
one unit
longer strokes
Single-acting cylinders can be mounted as follows:
• vertical mounting: when the return movement of the piston is brought about by
external forces (special instance: scissor lifting table);
• horizontal mounting: for single-acting cylinders with spring-return.
In large hydraulic presses, the return stroke is brought about by pullback cylinders.
Scissor lifting table
Possible applications
Examples

In the case of double-acting cylinders, both piston surfaces can be pressurized.
Therefore, it is possible to perform a working movement in both directions. These
cylinders operate according to the following principle:
The hydraulic fluid flows into the piston area and pressurises the piston surface.
Internal and external resistances cause the pressure to rise. As laid down in the law
F = p ⋅ A, a force F is produced from the pressure p and the piston surface area A.
Consequently, the resistances can be overcome and the piston rod extends. This is
possible owing to the conversion of hydraulic energy into mechanical energy which
is made available to a consuming device.
It should be borne in mind that when the piston extends the oil on the piston rod
side must be displaced via the lines into the reservoir. During the return stroke, the
hydraulic fluid flows into the (annular) piston rod area. The piston retracts and the
oil quantity is displaced from the piston area by the piston.
13.2

P PT T
Ts Ts
M M
T T
B BA A
P P
In double acting cylinders with a single-sided piston rod, different forces (F= p ⋅ A)
and speeds are produced for the same flow rate on extension and retraction owing
to the differing surfaces (piston surface and annular piston surface).
The return speed is higher since, although the flow rate is identical, the effective
surface is smaller than for the advance stroke. The following equation of continuity
applies:
A
Q
v =

The following designs of double-acting cylinders exist fulfilling varying
requirements:
Symbol
2 1:
A1 A2=
Differential
cylinder
Area ratio 2:1
(piston surface:
annular piston surface)
piston return stroke twice
as fast as advance stroke.
Synchronous
cylinder
Pressurised area of
equal size.
Advance and return
speeds identical.
Cylinder with
end-position
cushioning
To moderate the speed
in the case of large
masses and prevent a
hard impact.
Telescopic
cylinder
Longer strokes
Pressure
intensifier
Increases pressure
Tandem
cylinder
When large forces are
required and only small
cylinder dimensions are
possible.
Designation Description
Cylinder types

The movements generated by hydraulic cylinders are used for:
• Machine tools
– Feed movements for tools and workpieces
– Clamping devices
– Cutting movements on planing machines; shock-testing machines and
broaching machines
– Movements on presses
– Movements on printing and injection moulding machines, etc.
• Handling devices and hoists
– Tilting, lifting and swivel movements on tippers, fork-lift trucks, etc.
• Mobile equipment
– Excavators
– Power loaders
– Tractors
– Fork-lift trucks
– Tipper vehicles
• Aircraft
– Lifting, tilting and turning movements on landing gear, wing flaps, etc.
• Ships
– Rudder movements, adjustment of propellers
Cylinders with end position cushioning are used to brake high stroke speeds. They
prevent a hard impact at the end of the stroke.
Cushioning is not required for speeds of v < 6 m/min. At speeds of v ≥ 6-20 m/min,
cushioning is achieved by means of restrictors or brake valves. At speeds of
v > 20 m/min, special cushioning or braking procedures are required.
When the piston returns to the retracted end position, the normal discharge of the
hydraulic fluid from the piston area is interrupted by the cushioning piston and flow
is reduced from a certain point until it is finally closed. The hydraulic fluid in the
piston area must then flow away via a restrictor (see diagram).
In this way, the piston speed is reduced and there is no danger of malfunctions
resulting from high speeds. When the cylinder extends, the oil flows unhindered via
the non-return valve, the throttle point being by-passed. To achieve end position
cushioning, the pressure relief valve (flow division) must be used.
13.3
End position cushioning

P PT T
Ts Ts
M
T T
B BA A
P P
CushioningFlow control screw
Non-return
valve
Double-acting cylinder with end position cushioning
In addition to this simple type of end position cushioning, there is also double
cushioning for forward and retracted end positions. With this type of cushioning, a
hard impact is avoided both on advancing and on retracting.
The function of seals is to prevent leakage losses in hydraulic components. Since
pressure losses also occur as a result of leakage losses, seals are of considerable
importance for the efficiency of a hydraulic system.
In general, static seals are inserted between stationary parts and dynamic seals
between movable parts.
• Static seals:
– O-rings for the cylinder housing
– Flat seals for the oil reservoir cover
• Dynamic seals:
– Piston and piston rod seals
– Rotary shaft seals on turning devices
13.4
Seals

The recommended maximum piston speed is approx. 0.2 m/s. and is dependent on
the operating conditions as well as the sealing material and type of seal. Where
extremely low speeds and/or a minimal break-away force are required, special
sealing materials, systems and modified cylinder surfaces must be used.
The seals pictured opposite are used on cylinders according to requirements
(pressure, temperature, velocity, diameter, oil, water):
Cylinder seals

Cylinders are mounted in various ways according to usage. Some types of mounting
are shown in the diagram.
Foot mounting
Flange mounting
Swivel design
Swivel mounting with trunnion
Types of mounting
A hydraulic cylinder must be vented to achieve jolt-free travel of a cylinder piston,
i.e. the air carried along in the lines must be removed. As trapped air always gathers
at the highest point of a system of lines, a vent screw or automatic venting valve
must be positioned at this point.
Hydraulic cylinders are supplied with vent screws at both end positions. These ports
can also be used for connecting pressure gauges.
13.5
Types of mounting
13.6
Venting

The cylinder is selected to suit the load F. The required pressure p is selected in
accordance with the application.
F = p ⋅ A
This can be used for calculating the piston diameter. The hydraulic, mechanical
efficiency ηT
hmT
must be considered here. This efficiency is dependent on the
roughness of the cylinder barrel, the piston rod and the type of sealing system. The
efficiency improves with increases in pressure. It lies between 0.85 and 0.95. Thus,
the piston diameter is derived from:
π⋅η⋅
=
π⋅
=
η⋅⋅=
hm
2
hm
p
F
4
d
A
ApF
π⋅η⋅
=
hmp
F4
d
The volumetric efficiency ηT
v T
takes into consideration the leakage losses at the piston
seal. Where the seal is intact, ηT
vT
= 1.0 and is not, therefore, taken into
consideration.
Cylinder diameter, piston rod diameter and nominal pressures are standardised in
DIN 24334 and DIN ISO 3320/3322. In addition, a preferred ratio ϕ = piston area AT
PT
to annular piston area AT
PRT
is laid down.
Internal diameter of the cylinder
12 16 20 25 32 40 50 63 80
100 125 160 200 220 250 280 320 360 400
Piston rod diameter
8 10 12 14 16 18 20 22 25 28 32 36 40 45 50 63 70 80 90
10
0
11
0
11
2
14
0
16
0
18
0
20
0
22
0
25
0
28
0
32
0
36
0
Nominal pressures
U25U 40 U63U 100 U160U 200 U250U 315 U400U 500 U630U
The values which are underlined are recommended values. The recommended range
of piston strokes is laid down in DIN ISO 4393 and for piston rod threads in DIN ISO
4395.
13.7
Characteristics

In the table below, the area AT
PT
appropriate to the cylinder diameter dT
PT
and the
annular piston area AT
PRT
(not the piston rod area AT
STT
) for the piston rod diameter dT
STT
are assigned to the area ratio ϕ.
KR
K
A
A
=ϕ
Table for the area ratio ϕ
STPKR AAA −=
dT
PT
25 32 40 50 60 63 80 100 125Nominal
value ϕ
AT
PT
cmT
2T
4.91 8.04 12.60 19.60 28.30 31.20 50.30 78.50 123
dT
STT
12 14 18 22 25 28 36 45 56
AT
PRT
cmT
2T
3.78 6.50 10.00 15.80 23.40 25.00 40.10 62.20 98.10
1.25
ϕ Actual value 1.30 1.24 1.25 1.24 1.21 1.25 1.25 1.26 1.25
dT
STT
14 18 22 28 32 36 45 56 70
AT
PRT
cmT
2T
3.37 5.50 8.77 13.50 20.20 21 34.40 54 84.20
1.4
ϕ Actual value 1.46 1.46 1.44 1.45 1.39 1.49 1.46 1.45 1.46
dT
STT
16 20 25 32 36 40 50 63 80
AT
PRT
cmT
2T
2.90 4.90 7.66 11.60 18.20 18.60 30.60 47.70 72.40
1.6
ϕ Actual value 1.69 1.64 1.64 1.69 1.55 1.68 1.64 1.66 1.69
dT
STT
18 22 28 36 40 45 56 70 90
AT
PRT
cmT
2T
2.36 4.24 6.41 9.46 15.70 15.30 25.60 40.00 59.10
2
ϕ Actual value 2.08 1.90 1.96 2.08 1.80 2.04 1.96 1.96 2.08
dT
STT
20 25 32 40 45 50 63 80 100
AT
PRT
cmT
2T
1.77 3.13 4.52 7.07 12.30 11.50 19.10 28.40 44.20
2.5
ϕ Actual value 2.78 2.57 2.78 2.78 2.30 2.70 2.64 2.78 2.78
dT
STT
– – – 45 55 56 70 90 110
AT
PRT
cmT
2T
– – – 3.73 4.54 6.54 11.80 14.90 27.70
5
ϕ Actual value – – – 5.26 6.20 4.77 4.27 5.26 4.43
This table gives the area ratios up to a piston diameter of 125 mm. The complete
table is included in DIN 3320.

Buckling resistance as defined by Euler must be taken into consideration when
deciding on piston rod diameter and stroke length. Manufacturer’s tables are
available for this. When installing the cylinder, it is necessary to insure that no
distortions are possible. In addition, the direction of force must be effective in the
axial direction of the cylinder.
The permissible buckling force FT
permT
for a pressurised load is calculated as follows:
ν⋅
⋅⋅π
= 2
K
2
.perm
l
lE
F
E = Elasticity module ⎥
⎦
⎤
⎢
⎣
⎡
2
cm
daN
(for steel = 2.1 ⋅ 10T
6T
)
I = Area moment [cmT
4T
] (for ∅ =
64
d4
π⋅
= 0.0491 ⋅ dT
4
)T
LT
KT
= Free bucking length [cm]
ν = Safety factor 2.5 - 3.5
The free bucking length IP is dependent on the load in question:
l = lK l = ½K
F F F
l = 2lK
F
l
l
l
l
l = l *K √½
1st method 2nd method
(Basic case)
3rd method 4th method
One end free, one
end firmly clamped
Two ends with
flexible guide
One end with flexible
guide, one end firmly
clamped
Two ends firmly
clamped
Alternative clamping methods as defined by Euler
13.8
Buckling resistance

Cylinders are designed for tensile and pressure forces only. Shearing forces must be
absorbed by guides.
Note:
The type of mounting and installation determines the Euler method on which it
should be based.
mm mm
l
l
l
m
l
l
on method 2 on method 4on method 1 on method 3
Example for determining length l
The following apply in principle:
The length I is calculated from the attachment area of the flange or other fixed
bearing method (pivot pin, etc.). If the flange or pivot pin is at the cylinder head, for
example, the length I is calculated from this point.
Mounting methods three and four should be avoided wherever possible. Distortion
may occur where the load guide is imprecise in these areas.

Lifting device
A differential cylinder with the area ratio ϕ of 2:1 is to lift 40 kN 500 mm in 5 secs.
The maximum system pressure for the pump is to be 160 bar.
Calculate the piston diameter dT
P T
and find the piston rod diameter dT
STT
in the area ratio
table. On the basis of the piston rod diameter dT
STT
, find the maximum possible stroke
length from the buckling resistance diagram (next page). In addition, calculate the
advance and return speeds for the piston and the pump delivery.
The mechanical, hydraulic efficiency of the cylinder amounts to ηT
mhT
= 0.95. Pipe loss
amounts to 5 bar, pressure drop in the directional control valve 3 bar and back
pressure from the return movement 6 bar.
T
BA
P
2 : 1
P
T
M
P T
Ts
m
500mm
Lifting device
13.9
Selecting a cylinder
Example

Buckling resistance diagram

The safety factor ν is already included in the buckling resistance diagram.
Calculate the required piston diameter dT
PT
.
Available system pressure: 160 bar
minus line loss: 5 bar
pressure loss in the directional control valve: 3 bar
pressure from the return movement: when ϕ = 2:1 =
2
bar6
3 bar
Thus, the following pressure force remains at the cylinder
160-11 = 149 bar = 1490 N/cmT
2T
3
2
hm
P
hmP
cm3.28
N95.01490
cmN00040
p
F
A
ApF
=
⋅
⋅
=
η⋅
=
η⋅⋅=
4
d
A
2
P
P
π⋅
=
mm60cm0.6cm36
cm3.284A4
d 2
2
P
P ===
π
⋅
=
π
⋅
=
Chosen piston diameter dT
PT
= 63 mm.
The piston rod diameter dT
STT
= 45 mm is read from the table for the area ratio ϕ = 2:1.
A maximum stroke length of 1440 mm is read from the buckling resistance diagram
for 40 kN and a piston rod diameter dT
STT
= 45 mm. If an area ratio of 2:1 is not
required for the job, a smaller dT
STT
can be selected.
Calculating the advance stroke speed v:
t = 5 sec
Stroke = 500 mm
min/m6s/m1.0
s5
m5.0
t
s
v ====

Required pump delivery QT
PT
:
AT
KT
= 31.2 cmT
2T
= 0.312 dmT
2T
V = 6 m/min = 60 dm/min
min/l7.18min/dm7.18
min
dm60dm312.0
vAQ 3
2
Kp ==
⋅
=⋅=
Calculating the return speed vT
RT
:
PR
PR
A
Q
v
vAQ
=
⋅=
AT
PRT
is read from the table for the area ratio ϕ = 2:1 where dT
STT
= 45 mm:
min/m2.12min/dm122
mindm153.0
dm7.18
v
dm153.0cm3.15A
2
3
22
PR
==
⋅
=
==
When selecting a cylinder, it should be borne in mind that end position cushioning is
necessary for a piston speed of 6 m/min upwards.
Conditional on the area ratio ϕ = 2:1, the return speed of the piston is twice that of
the advance stroke. This also means that the outlet flow of the cylinder is twice that
of the advance stroke. For this reason, you are advised to calculate the speed of the
return flow before a system is sized and, where necessary, to select a larger cross-
section for the return line. The control valve should also be suitable for the
increased return flow, if not, then an additional valve must be installed for the
exhaust.

Hydraulic motors are components in the working section. They are drive components
(actuators). They convert hydraulic energy into mechanical energy and generate
rotary movements (rotary actuator). If the rotary movement only covers a certain
angular range, the actuator is referred to as a swivel drive.
Hydraulic motors have the same characteristic values as pumps. However, in the
case of hydraulic values we speak of capacity rather than displacement volume.
Capacity is specified by the manufacturer in cm3 per revolution along with the speed
range at which the motor is able to function economically. The following equation
can be used to find the capacity of a hydraulic motor:
V
M
p =
Q = n ⋅ V
p = pressure (Pa)
M = torque (Nm)
V = geometric displacement capacity (cmT
3T
)
Q = flow rate (l/min)
N = speed (r.p.m.)
The flow rate required by the motor is calculated from the capacity and the desired
speed.
A motor with a capacity of V = 10 cmT
³T
is to operate
at a speed of n = 600 revolutions per minute.
What flow rate (Q) is required by the motor?
Q =
min
600cm10 3
⋅
= 6000 cmT
3T
/min = 6 dmT
3T
/min = 6 l/min
Therefore, the pump must supply 6 l/min for the motor to turn at 600 revolutions
per minute.
The mechanical power rating of a hydraulic motor is calculated as follows:
ω = angle velocity in rad/s
ω = 2⋅ π ⋅ n
14. Hydraulic motors
Example

A hydraulic motor with a capacity of V = 12.9 cmT
3T
is driven with a pump delivery of
Q = 15 l/min. At the resultant speed, the turning torque M = 1 Nm. What is this
speed (n) and what is the power rating (P)?
Calculate the torque which arises when the motor brakes suddenly causing a
pressure of 140 bar (140 ⋅ 10T
5T
Pa) to be generated.
Technical Data: Q = 15 dmT
3T
/min
M = 1 Nm
V = 12.9 cmT
3T
Calculation of the r.p.m. n:
.m.p.r1163
minm
m
109.12
1015
minm109.12
m1015
mincm9.12
dm15
V
Q
n
VnQ
3
3
6
3
36
33
3
3
=
⋅⋅
⋅
=
⋅
⋅
===
⋅=
−
−
−
−
Calculation of the power rating p in Watts:
pT
maxT
= 140 · 10T
5T
Pa
W122
s
Nm
60
111632
Nm1.m.p.r1163p2Mn2P =⋅
⋅⋅π⋅
=⋅⋅⋅=⋅⋅π⋅=
Calculation of the torque at the maximum input pressure:
Nm6.180Nm101806M
m
mN
109.1210140m109.12Pa10140VpM
V
M
p
1
2
3
65365
=⋅=
⋅
⋅⋅⋅=⋅⋅⋅=⋅=
=
−
−−
The mechanical-hydraulic and volumetric efficiency were not taken into account for
the purposes of these calculations.
Example

Hydraulic motors are generally designed in the same way as hydraulic pumps. They
are divided up into:
• Constant motors
fixed displacement
• Adjustable motors
adjustable displacement
Both of these basic types includes several different designs.
Geared motor
Externally geared motor
Internally geared motor
Annular gear motor
Constant motor
Vane motor
Hydraulic motor
Internally pressurised
Externally pressurised
Constant, adjustable motors
Piston motor
Radial piston motor
Axial piston motor
Hydraulic motor

15. Accessories
In addition to the hydraulic components described in the previous chapters –
directional control valves, pressure valves, hydraulic cylinders, etc. – the following
accessories are of importance for the functioning of a hydraulic system:
• flexible hoses
• quick-release couplings
• pipes
• screw fittings
• sub-bases
• air bleed valves
• pressure gauges and
• flow gauges
These accessories are mainly used for transporting hydraulic energy (hoses, pipes,
etc.), connecting and mounting components (screw fittings, sub-bases) and for
implementing checking functions (gauges).
The components of a hydraulic system are connected together by means of
hoses or pipes.
Flow cross-sections of hoses and pipes affect the pressure loss within the lines.
To a large extent, they determine the efficiency of a system. To ensure that the
pressure losses in the pipelines, elbows and bends and elbow connectors do not
become too great and, at the same time, that the line dimensions are kept within
certain limits, the system should be designed so that the following flow speeds are
not exceeded:
• Pressure lines: up to 50 bar operating pressure: 4.0 m/s
up to 100 bar operating pressure: 4.5 m/s
• Suction lines: 1.5 m/s
• Return lines: 2.0 m/s

15. Accessories
The required flow cross-section is calculated on the basis of this data with the
following formula:
v
Q
A =
Q = flow rate
V = flow velocity
This equation can be used to determine the required size (diameter) of pipelines
when sizing a hydraulic system.
Calculations to determine the nominal size of lines:
v
Q
A = and
4
d
A
2
⋅
=
π
d = diameter
This results in the following equations for the nominal bore:
v
Q
4
d2
=
⋅π
v
Q4
d2
⋅
⋅
=
π
v
Q4
d
⋅
⋅
=
π
Technical Data: Q = 4.2 dmT
3T
/min = 4.2 l/min
Pressure line to 50 bar
v = 4 m/s
mm7.4mm22m10022.0
s/m
s/m
604
102.44
s/m4
min/dm2.44
d 223
333
==⋅=⋅
⋅⋅π
⋅⋅
=
⋅π
⋅
= −
−
Example

15. Accessories
These are flexible line connections which are used between mobile hydraulic devices
or in places where there is only limited space (particularly in mobile hydraulics).
They are used in cases where it is not possible to assemble pipelines (e. g. on
moving parts). Hoses are also used to suppress noise and vibration. They are made
up of a number of layers:
Structure of the hydraulic hose
The inner tube (1) is made of synthetic rubber, teflon, polyester-elastomer,
perbunan or neoprene. The pressure carrier is a woven intermediate layer of steel
wire and/or polyester or rayon.
This woven section (2) may consist of one or more layers depending on the pressure
range.
The top layer (3) is made of wear-resistant rubber, polyester, polyurethane
elastomer or other materials. The pipelines may be additionally protected against
mechanical damage by external spirals or plaited material.
When deciding on flexible hoses, it is necessary to take into consideration their
function and certain other factors.
In addition to power transmission by fluids, the hoses are subjected to chemical,
thermal and mechanical influences. In particular, it is important to specify the
operating pressure, both dynamic and static.
Pressures arising suddenly as a result of the fast switching of valves may be several
times that of the calculated pressures.
As far as technical data such as nominal size, load, chemical and thermal resistance,
etc. is concerned, only the manufacturer’s specifications are definitive.
The recommendations regarding nominal size and pressure contained in DIN 20021,
20022 and 20023 should be observed. Testing instructions for flexible hoses are laid
down in DIN 20024.
15.1
Flexible hoses
Selecting flexible hoses

15. Accessories
• Maximum permissible operating pressure
is specified by the manufacturer as far as static, and generally also dynamic,
pressure is concerned. Static operating pressure is specified with a fourfold
safety factor, i.e. operating pressure is 1/4 of bursting pressure.
• Bursting pressure
This should be regarded purely as a test value. The hose will not burst or leak
below this pressure.
• Test pressure
Hoses are pressurised to double the operating pressure for at least 30 secs and
at most 60 secs.
• Change in length
Every hose changes in length to a certain extent at operating pressure, the extent
of the change being dependent on the design of the woven intermediate layer.
This change may not amount to more than +2% or less than -4%.
• Bending radius
The specified minimum bending radius is intended for a stationary hose at
maximum operating pressure. For reasons of safety, it is important not to fall
below this minimum value.
• Operating temperature
The specified temperatures refer to the oil passing through the system. High
temperatures considerably reduce the service life of the hose.
The most important thing to ensure when installing flexible hoses is that the correct
length of hose is used. It must be possible to move the parts without the lines being
put under tension. In addition, the bending radii must be sufficiently large.
The following diagram shows some basic rules on the assembly of hoses.
incorrect incorrect
incorrect
correct correct
correct
Installation of hose lines
Definitions of terms

15. Accessories
Hoses are often used as connection components in mobile hydraulics and in many
stationary systems. Therefore, it is necessary that the pressure drop ∆p arising in
the hoses is taken into consideration when sizing these systems.
∆p in bar/m without connection fittings (ρ = 850 kg/mT
3T
; ν = 20 mmT
2T
/s)
NG da
(mm)
10
(l/min)
20 30 50 70 100 125 150 175 200
(l/min)
6 14 0.33 1.13 2.16
18 0.14 0.46 0.88
8 16 0.10 0.31 0.59 1.41 1.2
20 0.045 0.12 0.23 0.55 0.97 0.82 1.2
10 19 0.045 0.12 0.23 0.55 0.97 0.82 1.2
22 0.02 0.04 0.08 0.19 0.37 0.65 0.96 0.68 0.87 1.1
12 20 0.02 0.04 0.08 0.19 0.37 0.65 0.96 0.68 0.87 1.1
26 0.008 0.02 0.03 0.075 0.15 0.27 0.39 0.57 0.73 0.92
16 26 0.01 0.041 0.07 0.14 0.2 0.27 0.35 0.43
30 0.021 0.04 0.073 0.1 0.15 0.186 0.23
20 30 0.012 0.02 0.041 0.06 0.007 0.106 0.136
34 0.013 0.025 0.035 0.05 0.06 0.083
24 36 0.009 0.016 0.023 0.032 0.04 0.051
38.1 0.01 0.015 0.02 0.025 0.033
32 46 0.004 0.006 0.008 0.011 0.014
50.8 0.003 0.004 0.005 0.007 0.009
40 60.3 0.003 0.004
Flow resistance ∆p of hose lines (Prof. Charchut)

15. Accessories
Hose lines may either be connected to the various pieces of equipment or else
connected together by means of screw fittings or quick connection couplings.
Hose support connectors ensure that connections do not affect operation:
Hose – connector
DIN 24950 makes a distinction between the following mounting methods for the
hose side of the support connector:
• Screwed hose support connector
The support required by the hose is made by axial screwing together of
individual parts. This hose fitting can generally be assembled without special
tools and is re-usable.
• Swaged hose support connector
The support required by the hose is achieved by distorting at least one connector
support cone part. This hose fitting can only be assembled using special tools
and is not re-usable.
• Sleeve support
The support required by the hose is created using externally clamped sleeves or
segments. This hose support is re-usable and can be assembled with or without
special tools depending on type.
• Hose binding (hose clamp)
The support required by the hose is achieved through bracing, e.g. using hose
clamps as specified in DIN 3017 or tube straps as specified in DIN 32620. This
hose support can be assembled either with or without special tools, depending
on the design, and is in part re-usable – but is not, however, suitable for high
pressures.
• Push-in hose support
Usually made up of a nipple. The support required by the hose is generally
achieved through the appropriate forming of the nipple. This hose support
connector can be assembled without special tools and is re-usable. However, it is
not suitable for high pressures.

15. Accessories
DIN 24950 distinguishes between the following connections for the connection side
of the hose armature:
• Screw connection
provided with thread
• Pipe connection
provided with pipe, for compression fittings
• Flange connection
provided with flange
• Ring connection
provided with ring
• Coupling connection
provided with a symmetrical or asymmetrical coupling half
• Union connection
provided with union
Connector nut
Pipe end
External thread
Nipple for SAE flange
Hose support connection on connection side

15. Accessories
As shown in the diagram on page 264, the following components also form part of a
hose support connector:
• Connector nut
• Sleeve
The part of a hose support which encircles the hose. Distinction is made between
screwed fixtures, swaged fixtures, clamping fixtures and hose clamps.
• Nipple insert (sleeve, tube support elbow)
Component which is inserted into the hose forming the connection on the
connection side. Even in the case of barbed fittings, DIN 24950 makes a
distinction between a connecting part on the hose side and one on the
connection side:
– On the hose side of the fitting: screw-in, swaged and barbed fittings.
– On the connection side of the fitting: threaded, sealing end, screw-in, pipe,
collar, flanged and ring connections.
Nipple with sealing end connection
Nipple with threaded connection
Nipple with screw-in connection
Nipple with pipe connection
Nipple with collar connection
Nipple with flange connection
Nipple with ring connection
Diagram shows
a sealing cone
with O-ring
Hose support connectors – nipples

15. Accessories
Quick-release couplings can be used for the fast connection and disconnection
of devices.
These couplings are available both with and without a mechanically unlock able
non-return valve. Where there is no pressure, connection is possible via the non-
return valve without bleeding the hydraulic fluid.
Quick coupling socket (1) Sealing cone (3) Spring (5)
Coupling nipple (2) Sealing seat (4) Ring grip (6)
Quick-release coupling
Seamless precision steel tubes are used as pipelines as specified in DIN 2391. The
thickness of the walls of the pipelines is determined by the maximum pressure in the
pipeline and a safety factor for control impacts.
Before installation, pipes can be bent either when cold or by being heated up using
the appropriate bending devices. After being bent when hot, pipes should be
cleaned to remove the scale layer formed during this procedure, for example.
The following components are suitable for pipe to pipe and pipe to device
connection:
• Screwed pipe joints: up to nominal bore 38 (depending on operating pressure)
• Flanged connections: above nominal bore 30.
15.2
Pipelines

15. Accessories
DIN 3850 distinguishes between the following screwed pipe joints:
• Solderless fittings
• Compression fittings
• Double conical ring screwed fittings
• Soldered and welded screwed fittings
• Brazed nipple type fittings
• Ball-type screw fittings
Screwed pipe joint
Owing to ease of use, the compression fitting is the most commonly used type of
screwed fitting. When screwed together, a compression ring (olive) is pushed into
the internal cone of the connector by tightening the connector nut. The olive is
swaged into the pipe as it is pressed against a sealing stop.
Distinction is made in DIN 3850 between the following sealing and connection
components for the specified pipe joints:
Description Defined in DIN
Compression ring 3816
Double conical ring 3862
Spherical-bush 3863
Flanged bushing 3864
Pressure ring 3867
Overview of sealing components

15. Accessories
Description Defined in DIN For sealing component
Compression ring
Double conical ring
Soldered flanged bush
Connector nut A
B
C
3870
Welded flanged bush
Connector nut 3872 Olive with pressure ring
Compression ring
Double conical ring
Spherical bush
Connector screw A
C
3871
Flanged bushing
Overview of connection components
In addition, the following stub-end fittings are used with screwed pipe joints:
• straight connectors
• angle, L-, T- and soldered connectors
• bulkhead fittings, welded hexagon nipples and brazed hexagon nipples
The specified types of connector are available in a number of different designs which
are listed in DIN 3850. Specifications about nominal sizes and pressures for the
standardised screwed pipe joints can also be found in DIN 3850.
Flange connections are also used for larger pipes. The flange may either be screwed
or welded onto the pipe. The diagram shows two flange connections, one for the
pipe and one for the hose. B.S.F thread, metric fine thread and NPT (tapered thread)
are commonly used in hydraulics as connecting threads.
Flange connection

15. Accessories
Direct connection of valves by means of pipes and hoses does not always fulfil
requirements for a compact, economical and safe system. For this reason, sub-bases
are commonly used in hydraulics for connecting equipment. This connection method
allows fast valve exchanges. In addition, it reduces the flow paths of the hydraulic
fluid.
Like the valves, these sub-bases have standardised connection holes defined in DIN
ISO 4401. The valves are screwed onto these bases and then mounted on front
panels or valve supports and connected to hydraulic pipes on the reverse side.
Front panel with tank and pump
15.3
Sub-bases

15. Accessories
To save piping costs, manifold blocks are used for valves switched in parallel (block
hydraulics). Special control blocks of cast steel with the necessary connecting holes
incorporated are manufactured for controls with repeated cycles, e. g. press
controls, meaning that the valves simply need to be screwed on.
These special control blocks can be connected as required to form complex controls
(interlinking of blocks).
Intermediate plate valves are connected together for vertical interlinking and
screwed onto a common sub-base. As a result, only a limited amount of tubing is
required.
A
B
A A
P P
B B
P P
P P
T T
TT
T T
P R X Y
Standardised circuit diagram and vertical linking
Vertical interlinking

15. Accessories
In systems with several control circuits, longitudinal plates are lined up separated by
baffle plates. Either individual valves or a vertical valve arrangement can be screwed
onto the baffle plate.
A further improvement with regard to the realisation of complete controls on a single
block with compact multiple assembly has produced cartridge technology. With this
method, the various control functions are realised by the individual activation of
2/2-way panel-mounted valves. The 2/2-way panel-mounted valves are
standardised in DIN 2432. Panel-mounted valves (control blocks) only become
economical from a nominal diameter of 16 mm upwards and with a larger numbers
of items.
Bleed valves should be fitted at the highest point in a system of lines since this is
where the trapped air collects.
The diagram shows an automatic bleed valve. Figures 1 to 3 illustrate
the following phases:
• Fig. 1
The cylinder has retracted, at the same time the piston of the bleed valve closes.
• Fig. 2
When the piston rod extends, the piston of the bleed valve is lifted. The air is
able to escape via the vent hole until the hydraulic fluid reaches the piston and
pushes it upwards.
• Fig. 3
With the cylinder extended, the piston of the bleed valve is pushed up as far as it
can go by the hydraulic fluid, sealing off the outlet and closing off the air escape
route. If the pressure falls, the spring releases the piston until the vent port is
reopened and the process is repeated.
Longitudinal interlinking
Cartridge technology
15.4
Bleed valves

15. Accessories
Automatic bleed valve
The most commonly used pressure gauge operates on the principle of the Bourdon
tube. The curved Bourdon tube has a flat oval cross-section. When hydraulic fluid
flows into the tube, an identical pressure is produced throughout. Owing to the
difference in area between the outer curved surface and the inner curved surface, a
greater force is produced at the outer area bending the Bourdon tube upwards. This
movement is transferred to the pointer via the lever, rack segment and pinion. The
pressure can then be read off the scale.
This type of gauge is not protected against overpressure. A cushioning throttle must
be installed in the inlet connection to prevent the spring being damaged by pressure
surges. For pressures above 100 bar, a helicoid or screwshaped Bourdon tube is
used in place of the circular one. Pressures of up to 1000 bar can be measured.
These gauges are sensitive with respect to their position and may only be installed
in the position specified.
15.5
Pressure gauges
Bourdon tube gauge

15. Accessories
Bourdon tube gauge
In these gauges, the Bourdon tube is replaced by a pressure-resistant capsule of
corrugated metal or a pressure-resistant diaphragm clamped between two flanges.
When the inside of the capsule or diaphragm is pressurised, it is deflected. This
amount of the deflection determines the pressure being measured and is transferred
to the pointer via a mechanism. The pressure range is dependent on design and may
go up to 25 bar.
In the piston pressure gauge, the hydraulic fluid operates on a piston, the forces of
which work against a pressure spring. The pointer is directly connected to the piston
which displays the relevant pressure at the gauge. Piston pressure gauges are
protected against overloading.
More precise pressure measurements are possible with quartz pressure sensors
which exploit the piezo-electric effect. In these sensors, the pressure operates on a
diaphragm and, consequently, on the quartz crystal which emits an appropriate
voltage or current when under pressure. This electrical signal is electronically
amplified and displayed by an evaluating device in the form of a measurement of
pressure.
Other types of sensor operate with strain gauges which are arranged on a
diaphragm. Under pressure the diaphragm is deformed. The stretching of the gauge
resulting from this is converted into electrical signals. These signals are again
electronically amplified and displayed by a separate piece of equipment. In the case
of these sensors, the electronic section controlling this amplification is integrated
directly into the housing.
Diaphragm pressure gauge
Piston pressure gauge
15.6
Pressure sensors

15. Accessories
Advantages of electronic pressure sensors: The pressure which is displayed can be
evaluated at remote points by connection cables or recorded by operation recorders.
Direct activation of pressure valves via the amplifier is also possible.
If a single measurement is required in order to check the pump delivery or to set a
flow control valve, the simplest method of checking the volumetric flow rate is to use
a measuring container and a stop watch.
If the flow rate in a hydraulic system is to be continually monitored and displayed,
one of the devices on the following pages should be selected to suit requirements
for application and precision.
The hydraulic flow to be measured passes through a measuring tube. A fixed cone is
located in the measuring tube, which can be acted upon by a piston. If the hydraulic
fluid flows through the gauge between the cone and the piston, the piston is
pressed against a spring according to the rate of flow. The piston serves as a mobile
orifice. A flow cross-section is produced corresponding to its position on the cone.
The piston moves until the set pressure difference which moves the piston against
the spring is in equilibrium. As the flow rate is dependent on the pressure difference
at the orifice, the displacement of the piston can be displayed as a measure of the
flow rate. The display error is in the range of 4%.
Flow meter (works diagram UCC)
Measuring turbines, oval disk meters, gear meters, orifice gauges and retarding
disks are used for more precise measurements for the regulation or control of
synchronous cylinders or motors and for positioning control.
Volumetric flow gauges
15.7
Flow measuring
instruments

15. Accessories
The rotor or turbine is set in rotation by the flow rate. The speed is evaluated as a
measurement of flow rate and displayed (diagram).
The gear meter is constructed like a gear motor. Each tooth is inductively sensed by
a measuring device. The speed is shown via a transducer in the form of a flow rate.
The oval disk meter operates by the same principle. Once again, the speed is
measured inductively. Since, as in the case of the gear meter, the chamber volume is
known, the flow rate can be calculated from the speed which is measured.
In the case of the orifice gauge, the ∆p is measured, electronically converted and
displayed as a flow rate.
The baffle plate operational principle is as follows: the flow rate acts on a baffle
plate located in the flow pipe which executes a stroke in accordance with the value
of the flow rate. The stroke length is contactlessly sensed. The electrical output
signal is converted and displayed as a flow rate. Port for determining speed by
inductive means
Port for determining speed
by inductive means
Turbine meter

Values Symbol SI unit Dimension
Displacement s Metre m 1 m = 1000 mm
Force F Newtons N
1 N = 1 2
s
mkg⋅
Time t Seconds s
Velocity v Metre/Seconds m/s
1
s
m
= 60
min
m
Pressure p Newtons/Square metres N/mT
2T
1 2
m
N
T = 1 Pa (Pascal)
1 Pa = 10-5
bar
1 bar = 105
Pa = 10 2
cm
N
Density ρ Kilogram/Cubic metres kg/m
3
1000 3
m
kg
= 1 3
dm
kg
= 1 3
cm
g
Area A Square metres m
2
Volume V Cubic metres m
3
1 m3
= 1000 l (Litre)
1 l = 1 dm3
Volumetric flow rate Q Cubic metres/Seconds m
3
/s
1
s
m3
= 60000
min
l
1
min
l
=
s60000
m1 3
Energy, work W Newton metres Nm 1 Nm = 1 J (Joule)
Power P Watts
Newton metres/Seconds
W
Nm/s
1 kW = 1000 W = 1.36 PS
1 PS = 0.763 kW
1 W = 1
s
Nm
= 1
s
J
Figure for the
friction in pipes
λ
Resistance
coefficient
ξ
Kinematic viscosity ν Square metres/Seconds m
2
/s
Efficiency η
Reynolds’ number Re
16. Appendix

16. Appendix
In hydraulics, the pressure unit bar is generally used owing to the high pressures
which arise. The international system of units SI (Système International) specifies
the use of the pressure units Pascal and, with certain reservations, bar; the units
atm and Torr are to be avoided.
1 Pascal = 1 Pa = 1 2
m
Nm
= 10-5
bar
Pa bar mbar Torr at
1 Pa = 1 N/m
2
1 10
-5
10
-2
7.5 ⋅ 10
-3
1.02 ⋅ 10
-5
1 bar = 10 N/cm
2
10
5
1 10
3
750 1.02
1 mbar = 1 N/dm
2
100 10
-3
1 0.75 1.02 ⋅ 10
-3
1 Torr = 1 mm Hg 1.33 ⋅ 10
2
1.33 ⋅ 10
-3
1.33 1 1.36 ⋅ 10
-3
1 at = 1 kp/cm
2
0.981 ⋅ 10
5
0.981 0.981 ⋅ 10
3
736 1
Conversion of pressure units (Values have been rounded off) DIN 1314 (12.71)
5000 kPa = ? bar
p = 5000 kPa = 5000000 Pa = 5000000 · 10-5
bar =
100000
5000000
bar = 50 bar
Example

16. Appendix
Safety regulations
For hydraulic systems, we advise you to adhere to the technical safety specifications
laid down in DIN 24346. The accident prevention specifications of the Employer’s
Liability Insurance Association (VBG) which are relevant for both individual machines
and complete systems, e.g. “Hydraulic presses” (VBG 7n5.2; UVV 11.064), should
also be taken into consideration.
Some other important safety principles are listed below:
• Never operate a system or press a switch if you are unaware of its function.
• Do not switch on the power supply until all lines are connected up.
Important
check whether all return lines (leakage pipes) lead to the tank.
• Before commissioning, carefully flush the system. Then, change the filter
elements. On initial commissioning of the system, open the system pressure
relief valve almost completely and slowly adjust the system to the operating
pressure. Pressure relief valves must be installed in such a manner that they
cannot become ineffective.
• All setting values must be known.
• Bleed the system and the cylinders.
• Install the EMERGENCY STOP switch in a position where it is easily reached.
• Use only standard parts.
• Incorporate all changes into the circuit diagram without delay.
• Nominal pressure must be clearly indicated.
• Check that the devices installed in the system are permissible for the maximum
operating pressure.
• Suction lines must be designed in such a way as to eliminate the possibility of air
being taken into the system.
• The temperature of the oil in the intake line to the pump must not exceed 60 °C.
• The cylinder piston rods must not be bent; they must not be subjected to lateral
forces.
• Protect piston rods against damage and dirt.
Particular care should be taken in the use of hydraulic reservoirs:
• Before commissioning the reservoir, the manufacturer’s specifications should be
studied.
• The hydraulic lines to the reservoir must be carefully bled. This can usually be
accomplished at the safety and shut-off block of the reservoir.
• Repair work to hydraulic systems can only be carried out after releasing the oil
pressure to the reservoir. Where possible, separate the reservoir from the system
(by means of a valve).
• Never drain off the contents of the reservoir unthrottled!
• For details regarding installation and operation, see “Technical Specifications for
Pressure Reservoirs” (TRB).
• All hydraulic reservoirs are subject to the pressure reservoir standards.

Hydraulics basic

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