Synchronized Hydraulic System for Controlled Structure
Displacements
Hugo Daniel Almeida Rodrigues
Dissertação para obtenção do Grau de Mestre em
Engenharia Mecânica
Júri
Presidente:
Prof. Luís Manuel Varejão de Oliveira Faria
Orientador:
Prof. Eduardo Joaquim Anjos de Matos Almas
Co-Orientador:
Prof. Carlos Baptista Cardeira
Vogal:
Prof. Alberto José Antunes Marques Martinho
Vogal:
Prof. José Alberto de Jesus Borges
Maio 2010
i
Abstract
Nowadays heavy structural movements are mostly made using hydraulic systems
with manual operation control. This paper discusses the development of an automatic and
synchronized system, suitable for engineering operations, versatile in raising and thrusting of
large buildings. The system allows controlling pressures and movements synchronously and
accompanied by position readings measured by position sensors. The system works
simultaneously, with a maximum ensemble of units working together, typically in a number
that fits the operations needs. The system provides information and acts in real time in order
to perform predefined movements, using the measurements of position and compensating
for flow variations and / or pressure.
iii
Resumo
Movimentações de grandes estruturas são actualmente feitas, em grande parte dos
casos, usando sistemas hidráulicos de actuação manual. Este trabalho aborda o
desenvolvimento de um sistema automático e sincronizado, adequado a trabalhos de
engenharia, que seja versátil em levantamentos e empurres de construções de grande
porte. O sistema pressupõe uma manipulação controlada das pressões e controlo dos
movimentos, acompanhada por leitura de posicionamento medida por sensores de
deslocamento. O sistema prevê o controlo, em simultâneo e sincronizadamente, de um
número máximo de unidades de actuação que deverá ser adequado às necessidades dos
trabalhos onde será aplicado, trabalhando modularmente com qualquer quantidade de
actuadores até ao número máximo definido. Este sistema fornece informações e actua em
tempo real de modo a efectuar os movimentos predefinidos, recorrendo às leituras de
posição, compensando com variações de caudal e/ou pressão.
iv
Keywords:

Heavy lifting

Hydraulic Synchronism

Discrete valves control

On/off valves

Hydraulic simulation
Palavras-chave:

Movimentação de estruturas

Sincronismo hidráulico

Controlo discreto de válvulas

Válvulas tudo ou nada

Simulação hidráulica
v
Acknowledgements
Estou grato acima de tudo aos meus Pais pelo que já aturaram e por todo o
apoio que sempre providenciam. À Lucy, pela excelente vontade e pela força que
me deu na realização deste trabalho. Agradeço especialmente ao Eng. Carlos Moniz
por ter tornado possível esta tese ao acreditar no meu trabalho e em me dar esta
Oportunidade. Ao Prof. José Borges por me ter dado o empurrão para avançar com
este tema e pelo apoio ao longo do seu desenvolvimento. Aos meus orientadores
Prof. Eduardo Matos Almas e Prof. Carlos Cardeira por terem aceite trabalhar
comigo com o tema que lhes propus e por acompanharem o meu trabalho tornandoo possível.
vi
List of symbols
A - actuator area
- Reynolds number inside pipes
Cf - coefficient of viscous friction
- Reynolds number inside pipes
Cx - valve flow coefficient
s – complex variable Laplace Transform
Cp - actuator flow coefficient
v - piston velocity
- diameter
V - flow velocity
- pipes internal diameter
- flow velocity inside pipes
- hose diameter
- flow velocity inside hoses
- friction factor
- pump volumetric displacement
- pipe friction factor
Vi - valve input signal
- hose friction factor
– platform parallel axis
fE - external load force
X - valve position
g - gravity force
z - lifting points position
h - fluid column height
β - bulk modulus
- virtual rotational line position
- fluid density
k – local hydraulic loss
- fluid viscosity
KT - pressure transducer gain
- measured fluid viscosity
KV - valve gain
- selected fluid viscosity
KA - amplifier gain
- pump volumetric efficiency
- fluid piping length
- pump total efficiency
- hydraulic loss
- pipes length
- hydraulic loss across pipes
- hoses length
- piston mass
- hydraulic loss across hoses
- radial velocity
- valve hydraulic loss
- elbow hydraulic loss
p - system pressure
P - electric motor power
- vertical oil column pressure
Q - flow rate
- measured fluid press. loss
- selected fluid press. loss
- Reynolds number
- rotation platform length
vii
Abbreviations
PV – proportional valve
SV – servo valve
EH – Electro-hydraulic
MIMO – Multiple input multiple output
SISO – Single input single output
P – Pressure port in a valve
A – Directional valve port A
B – Directional valve port B
T – To reservoir connection
PI – Proportional integrative
PD – Proportional derivative
SP – Pressure line solenoid signal
SR – Return line solenoid signal
viii
Contents
ABSTRACT .......................................................................................................................................... III
RESUMO ............................................................................................................................................ IV
KEYWORDS: ........................................................................................................................................ V
PALAVRAS-CHAVE: .............................................................................................................................. V
ACKNOWLEDGEMENTS ...................................................................................................................... VI
LIST OF SYMBOLS .............................................................................................................................. VII
ABBREVIATIONS .............................................................................................................................. VIII
CONTENTS ......................................................................................................................................... IX
LIST OF TABLES .................................................................................................................................. XI
LIST OF FIGURES................................................................................................................................ XII
1
INTRODUCTION ........................................................................................................................... 1
1.1 BACKGROUND ..................................................................................................................................... 1
1.2 UNDERSTANDING THE PROBLEM ............................................................................................................. 1
1.3 ELECTRO-HYDRAULIC COMPONENTS......................................................................................................... 3
1.4 SYNCHRONIZATION OF MULTIPLE ACTUATORS IN HEAVY LIFTING .................................................................... 8
1.5 STATE OF THE ART ................................................................................................................................ 8
2
DESIGN AND SELECTION OF HYDRAULIC SYSTEM ...................................................................... 10
2.1 COMPONENT SELECTION ..................................................................................................................... 11
2.2 CIRCUIT SELECTION ............................................................................................................................ 13
2.3 CIRCUITS DISCUSSION ......................................................................................................................... 16
2.4 WORKING OVER THE ELEMENTARY CIRCUIT.............................................................................................. 24
2.5 PUMP OPTIONS ................................................................................................................................. 32
2.5.1
Single pumping system with auxiliary gear pump ............................................................. 32
2.5.2
Multi-outlet pump .............................................................................................................. 33
2.5.3
Separated units .................................................................................................................. 34
ix
3
HYDRAULIC CONTROL ............................................................................................................... 35
3.1 ADVISED RULES FOR HYDRAULIC CONTROL .............................................................................................. 35
3.2 ABOUT VALVES ................................................................................................................................. 39
3.3 TYPES OF CONTROL............................................................................................................................. 39
3.4 PRESSURE CONTROL ........................................................................................................................... 41
3.4.1 Components ......................................................................................................................... 41
3.4.2 Control diagram ................................................................................................................... 44
3.4.3 Scheme ................................................................................................................................ 47
3.4.4 Controller parameters ........................................................................................................ 47
3.5 DIRECTIONAL CONTROL...................................................................................................................... 49
4
CIRCUIT DIMENSIONING ............................................................................................................ 56
4.1 PARAMETERS .................................................................................................................................... 56
4.2 FLUID .............................................................................................................................................. 57
4.3 VALVES ............................................................................................................................................ 58
4.4 PIPES AND HOSES............................................................................................................................... 60
4.5 HYDRAULIC LOSSES ............................................................................................................................ 61
4.6 PUMP AND MOTOR ........................................................................................................................... 64
5
SIMULATION RESULTS ............................................................................................................... 67
5.1 MODEL ........................................................................................................................................... 67
5.2 SIMULATION FEATURES ...................................................................................................................... 69
5.3 RESULTS .......................................................................................................................................... 72
6
ECONOMICAL VIABILITY ............................................................................................................ 76
7
CONCLUSIONS AND FUTURE WORK........................................................................................... 81
7.1 SIMULATIONS .................................................................................................................................... 82
7.2 PROSPECT IMPROVEMENTS AND ADAPTATION POSSIBILITIES TO INNOVATION ................................................. 82
8
BIBLIOGRAPHY .......................................................................................................................... 83
APPENDIX ......................................................................................................................................... 86
A - HYDRAULIC CIRCUIT DIAGRAM .............................................................................................................. 87
B - ELECTRICAL DIAGRAM .......................................................................................................................... 88
C - OIL TEMPERATURE DEPENDENCE OF VISCOSITY .......................................................................................... 89
D – VALVES TECHNICAL DATA ..................................................................................................................... 90
E – SIMULINK MODEL .............................................................................................................................. 91
x
List of tables
TABLE 1: DEFINITION OF SYSTEM REQUIREMENTS ............................................................................................................. 12
TABLE 2: COMBINATION OF SOLENOID SIGNALS FOR THE DIRECTIONAL CONTROL .................................................................... 22
TABLE 3: ASSIGNED COMPONENTS OF THE SYSTEM ........................................................................................................... 25
TABLE 4: FEEDBACK AND CONTROL COMBINATIONS (NATCHWEY P.) .................................................................................... 40
TABLE 5: CIRCUIT FRACTIONS FOR HYDRAULIC LOSSES CALCULATION..................................................................................... 61
TABLE 6: MULTI-OUTLET PUMP SYSTEM QUOTATION ........................................................................................................ 77
TABLE 7: 8 UNITS SINGLE PUMP EQUIPPED QUOTATION ..................................................................................................... 78
TABLE 8: INVESTMENT CHARACTERISTICS ........................................................................................................................ 79
TABLE 9: MULTI-OUTLET PUMP SYSTEM CASH FLOW ......................................................................................................... 79
TABLE 10: 8 UNITS SINGLE PUMP EQUIPPED CASH FLOW .................................................................................................... 80
xi
List of figures
FIGURE 1: ACTUATORS ON PARALLEL AND SERIES CONNECTION ............................................................................................. 2
FIGURE 2: FLOW CONTROL VALVE WITH REGULATION [PARKER HANNIFIN] .............................................................................. 3
FIGURE 3: DIGITAL SELECTION OF FORWARD SPEEDS (BOSCH, 1989) ..................................................................................... 4
FIGURE 4: COMPONENTS OF A POSITION CONTROL HYDRAULIC SYSTEM (DE NEGRI 1999) ......................................................... 5
FIGURE 5: PROPORTIONAL VALVE (JONES J.C. 1997) ......................................................................................................... 5
FIGURE 6: TWO STAGE SERVO VALVE (JAMES E. JOHNSON PENTON - MEDIA, INC) ................................................................... 6
FIGURE 7: PROPORTIONAL VALVE APPLICATION MATRIX (EATON CORP.) ............................................................................... 7
FIGURE 8: SYSTEM PRACTICAL LAYOUT ........................................................................................................................... 11
FIGURE 9: A) B) C) ACTUATOR SPEED CONTROL CIRCUITS: METER IN, OUT AND BLEED OFF RESPECTIVELY (ERIKSSON, B. - A AND B )
(PARKER TRAINING - C ) ...................................................................................................................................... 13
FIGURE 10: ACTUATOR SPEED CONTROL CIRCUIT USING A SERVO VALVE (CUNDIFF J.) .............................................................. 13
FIGURE 11: DIFFERENT MOVEMENTS TO BE ACHIEVED: A) PARALLEL LIFTING AND B) ROTATION. ............................................... 15
FIGURE 12: HYDRAULIC SCHEME ON A FIRST APPROACH..................................................................................................... 16
FIGURE 13: CIRCUIT 1 ANALYSIS ................................................................................................................................... 17
ND
FIGURE 14: HYDRAULIC SCHEME 2 VERSION ................................................................................................................. 17
FIGURE 15: CIRCUIT 2 ANALYSIS REGARDING PUMP PERMANENT LOADING ............................................................................ 18
FIGURE 16: CIRCUIT 2 ANALYSIS FOR PRESSURE CONTROL .................................................................................................. 18
RD
FIGURE 17: HYDRAULIC SCHEME 3 VERSION.................................................................................................................. 19
TH
FIGURE 18: HYDRAULIC SCHEME 4 VERSION .................................................................................................................. 20
FIGURE 19: CIRCUIT 4 ANALYSIS WITH SINGLE-ACTING CYLINDERS (LEFT) AND WITH DOUBLE-ACTING CYLINDERS (RIGHT) ............... 20
FIGURE 20: HYDRAULIC SCHEME 5TH VERSION ................................................................................................................ 21
FIGURE 21: SET OF 3 DIRECTIONAL VALVES ..................................................................................................................... 21
FIGURE 22: ELEMENTARY CIRCUIT. ................................................................................................................................ 23
FIGURE 23: ELEMENTARY CIRCUIT ANALYSIS .................................................................................................................... 24
FIGURE 24: ELEMENTARY CIRCUIT WITH DETAILS .............................................................................................................. 25
FIGURE 25: QUICK COUPLERS C604 (ENERPAC) .............................................................................................................. 26
FIGURE 26: PRESSURE RELIEF VALVE (BIERI HYDRAULIC AG) .............................................................................................. 27
FIGURE 27: PRESSURE TRANSDUCER S-10 (WIKA INSTRUMENT CORPORATION)..................................................................... 27
FIGURE 28: SHUT-OFF VALVE (BIERI HYDRAULIC AG) ....................................................................................................... 28
FIGURE 29: SOLENOID OPERATED DIRECTIONAL VALVES ..................................................................................................... 29
xii
FIGURE 30: PROPORTIONAL PRESSURE RELIEF VALVE (HAWE HYDRAULIK SE)......................................................................... 30
FIGURE 31: RADIAL PISTON PUMP (HAWE HYDRAULIC SE)................................................................................................. 30
FIGURE 32: WIRE DISPLACEMENT SENSOR (MICRO-EPSILON GMBH) ................................................................................... 31
FIGURE 33: HIGH PRESSURE-LOW PRESSURE PUMPING CONFIGURATION ............................................................................... 32
FIGURE 34: MULTI-OUTLET RADIAL PISTON PUMP WITH FOUR EXITS (DYNEX/RIVETT INC) ........................................................ 33
FIGURE 35: CONFIGURATION WITH MULTI-OUTLET PUMP .................................................................................................. 33
FIGURE 36: SYSTEM CONFIGURATION WITH VARIOUS SEPARATED PUMPS .............................................................................. 34
FIGURE 37: FUNCTIONAL PRINCIPLE OF A FORCE-CONTROLLED PROPORTIONAL VALVE [5]......................................................... 41
FIGURE 38: FUNCTION OF FORCE ALONG THE STROKE IN A SPOOL OF A PROPORTIONAL FORCE-CONTROLLED VALVE FOR DIFFERENT
CURRENT VALUES............................................................................................................................................... 42
FIGURE 39: CROSS SECTIONAL VIEW OF PROPORTIONAL PRESSURE RELIEF VALVE (CHAPPLE P.) .................................................. 43
FIGURE 40: DIAGRAM OF THE CORE OF A PROPORTIONAL PILOT-OPERATED PRESSURE RELIEF VALVE (CHAPPLE P.) ........................ 43
FIGURE 41: SIMPLIFIED SYSTEM BLOCK DIAGRAM (CHAPPLE P.) .......................................................................................... 45
FIGURE 42: PRESSURE CONTROL BLOCK DIAGRAM (CHAPPLE P.) ......................................................................................... 45
FIGURE 43: PRESSURE CONTROL SCHEME ....................................................................................................................... 47
FIGURE 44:PRESSURE CONTROL IN PLASTIC FORMING MACHINE EXAMPLE ............................................................................. 48
FIGURE 45: MANUFACTURER GRAPH FOR PDV700 (BIERI HYDRAULIC) ............................................................................... 49
FIGURE 46: DIRECTIONAL CONTROL PROCESSES DIAGRAM .................................................................................................. 50
FIGURE 47: DIAGRAM OF A SOLENOID ACTUATED DIRECTIONAL VALVE (BOSCH-REXROTH) ....................................................... 51
FIGURE 48: VSL STRAND JACK LIFTING SYSTEM WITH DOUBLE-ACTING CYLINDERS ................................................................... 52
FIGURE 49: LIFTING AND PUSHING SCHEMATIZATION......................................................................................................... 53
FIGURE 50: SCHEMATIZATION OF A ROTATION ................................................................................................................. 54
FIGURE 51: ROTATION EXAMPLE ................................................................................................................................... 54
FIGURE 52: ROTATION EXAMPLE WITH DIFFERENT ROTATIONAL CENTRE................................................................................ 55
FIGURE 53: SHELL TELLUS OILS T CHARACTERISTICS .......................................................................................................... 58
FIGURE 54: VALVE CHARACTERISTIC DIRECTIONAL VALVE BIERI WV700 4/3-U NG6 ............................................................ 59
FIGURE 55: TYPICAL CONFIGURATION FOR AN ASSEMBLY ................................................................................................... 64
FIGURE 56: CHARACTERISTIC CURVES FOR FIXED DISPLACEMENT RADIAL PUMP TYPE PR4 (BOSCH-REXROTH) .............................. 65
FIGURE 57: HYDRAULIC COMPONENTS SIMULATION DETAIL ................................................................................................ 68
FIGURE 58: PRESSURE CONTROL SUBSYSTEM ................................................................................................................... 68
FIGURE 59: EXAMPLE OF THE USE OF LOCAL SOLVERS ON PORTIONS OF THE MODEL WITH OTHER EXPLICIT SOLVER. (MILLER ET
AL.,2009) ....................................................................................................................................................... 70
FIGURE 60: FLOWCHART DEPICTING THE PROCESS OF MOVE FROM DESKTOP SIMULATION TO REAL-TIME SIMULATION (MILLER ET AL.,
2009) ............................................................................................................................................................ 71
FIGURE 61: POSITION-TIME FOR A PARALLEL LIFTING WITH 8 CYLINDERS POSITION-TIME FOR A GIVEN TIME RANGE ....................... 72
FIGURE 62: ZOOM FOR THE ASSIGNED INTERVAL .............................................................................................................. 72
FIGURE 63: CORRESPONDING VALVE SWITCHING FOR THE 8 DIRECTIONAL VALVES................................................................... 73
FIGURE 64: POSITION-TIME FOR 8 CYLINDER ROTATION..................................................................................................... 74
xiii
FIGURE 65: PRESSURE LEVEL MAINTAINED AT 100 BAR ..................................................................................................... 74
xiv
1 Introduction
1.1 Background
Nowadays there are specialized operations in civil engineering or industrial plants that are
related with the displacement of large and heavy structures. The modern engineering scenario
imposes for economical or technical reasons, lifting and displacements in situations such as in
bridges, roof structures or industrial assemblies.
The increasing use of the so called heavy lifting has to do not only with situations where it is
absolutely necessary, but also with the viability of construction and installation of prefabricated
structures. On the other hand it is also very useful in the reconditioning of aged structures where a
lifting or a lowering is sometimes needed.
In the first case the situation is ruled by economical reasons or by the best ratio between lifting
power and the possibility of building something assembled off-site before mounting it, which is
sometimes the only practicable way.
The second is true in an age where roads, buildings, bridges and other types of infrastructures
have increased in numbers like never before. All these infrastructures need maintenance in order to
preserve them for extended life time, or to implement new construction technologies in them.
The complexity of some works, because of their movement types, forces involved and
available space sometimes make unusable the conventional systems like cranes, or electrical based
systems. Hydraulic technology is the one that normally goes further when massive force is required for
some applications. It also has interesting characteristics like the possibility of placing relatively small
actuators (comparing to their force capacity) distant from their power source, which is very important,
mainly for the operation types discussed on this paper.
1.2 Understanding the problem
Not being new, computer controlled systems for hydraulic applications of this type are an
emergent and effective technology, although most of the operations are done manually with hand
operated hydraulic circuits. Some cases require higher precision than those resulting from the simple
control of a worker. A normal manual operated system should be monitored to control the
displacements in every lifting point, but the accuracy is not always sufficient to some applications
1
where internal stresses can result from the structure dislocation. The tolerances achieved with digital
controlled systems could be in an order of magnitude smaller than a millimeter, nevertheless the
conditioning lower velocities for tight values of tolerance. However in an operation, time is quite
dispensable in comparison with tolerances importance according to the experts.
There are minimum requirements in order to provide a well functional power lift system for a
multiple jack system. A summary of these conditions was described by [S. Price, 1997]:

All jacks in a system operating at the same speed,

Capability to preset and adjust the pumping pressure,

Capability to open or close the flow to any of the jacks in a system,

Monitoring of the actual pumping pressure,

Simple operating controls.
It is feasible to assume theoretically that to obtain the first, second and fourth characteristics it
is simply required an equal flow distribution in each cylinder, a pressure valve and a pressure gauge.
In practice the simplest and logical ways of getting synchronism without control are not sufficiently
precise and have some issues, that will be discussed further. The next figure shows actuators
connected in parallel and in series, which represents the first and easiest method to achieve hydraulic
synchronism.
Figure 1: Actuators on parallel and series
connection
2
In the parallel assembly even if the cylinders were equal sized the effect of the common flow
would be that the one that found less opposite force would be the first rising. So there would be a lag
from the cylinder having the minimum opposite force till the one that had the maximum, which would
result in an impossible synchronism. With the second, in series, the problem seems to be solved by
neglecting oil leaks between the piston and the cylinder; but to have always equal relations (of flow
and area) and double acting cylinders would be a requirement and a drawback.
The individual flow control on each cylinder should be the only way to assure the use of
different cylinder types as well as their synchronization. However a basic solution just on a manual
flow control basis (as a flow control valve, figure 2) is not ideal because the piston movement occurs in
function of the instantaneous reactions it finds.
Figure 2: Flow control valve with regulation [Parker Hannifin]
The reason is that a flow control valve works with the pressure control valve that diverts the
excess of flow according to the pressure differential imposed by the regulated orifice. In the circuit, the
pressure-relief valve is always closed till the defined pressure level is achieved. When it gets opened
there are valve oscillations, as well as pressure variations, and consequently the movement of the
actuator piston could be irregular [Speich Bucciarelli, 1978]. Anyway, both to control all flows of the
different cylinders, and to regulate the pressure and flow oscillations, an electronic based control
system is the way to achieve the best precisions.
1.3 Electro-hydraulic components
Advances in technology, after the Second World War, showed that separate elements of
control drive and hydraulic components could be embedded into a single system concept. An
hydraulic drive system, was after this period (post-war till 1960), able to receive fluid from a power
source and transmit it with more precise control in terms of force, speed or position, due to
3
improvements. These improvements were related with better reliability, linearity and stability with
outside disturbances such as contamination, pressure change and acceleration force [Jones 1997].
The control of the hydraulic circuit with directional, flow or pressure valves and actuators can be split
in discrete or continuous actuation systems. This specifies the type of signal that the electro-hydraulic
(EH) components receive to control the variables of oil direction, pressure and flow. An example of
how to control the velocity with discrete valves can be seen on figure 3:
Figure 3: Digital selection of forward speeds (Bosch, 1989)
By using on/off valves configured with different orifice diameters, it is possible to control seven
different velocities combining the different actuation possibilities of the directional valves between
each other. On the other side, by using continuous actuation systems it is possible to answer
mechanically to an activation signal whether it is mechanical, electric or hydraulic. In Figure 4 is shown
schematically an hydraulic position control system, which can be built with a servo or a proportional
valve.
4
Figure 4: Components of a position control hydraulic system (De Negri 1999)
The main difference between servo (SV) and proportional valves (PV) is that the proportional
ones do not have feedback but both of them are programmable orifices. PVs cannot compensate
external disturbances alone like the servo but they are more tolerant to contaminant and cheaper. A
PV can be found in two types: force-controlled and stroke-controlled. A stroke-controlled solenoid has
feedback of the spool position, not feedback of the load position.
Proportional valve circuits do not compensate for external disturbances; so they are
recommended for applications in which the load is approximately constant for each cycle [Cundiff,
2002].
A lot of applications, where the same part is manipulated in the same way during each cycle,
have an approximately constant load cycle. A series of settings can be programmed into the
proportional valve controller, and it will cycle the cylinder with good repeatability.
Figure 5: Proportional valve (Jones J.C. 1997)
5
As described by [Cundiff, 2002] the spool in a PV can have different shapes, and better
controllability is achieved if a proportional valve with 2:1 ratio spool is used with a 2:1 area ratio
cylinder. A 2:1 area ratio spool is one that has notches with twice the area on the side of the land as
the notches on the other side.
The pressure drops on both sides of the PV directional, between ports P-A and B-T are an
inevitable loss in efficiency, but some pressure drop is required for the metering which is
indispensable for a good control. Besides that, the loss of efficiency can be exchanged for a better
control provided by the PV directionally controlled. The control current to a proportional valve can be
programmed to position the spool and create the orifice area, and thus the pressure drop needed to
accelerate a load, maintains a constant velocity, and then decelerate the load to a stop.
The interaction of these characteristics determines the final control settings that are
programmed into the proportional valve amplifier. Once a PV circuit has been programmed the circuit
behavior is very satisfactory. Its initial cost is less than a SV circuit, and it costs less to operate. The
PV circuit has no feedback. When load increases beyond the level at which the programmed settings
were defined, pressure rises. The consequences can be reflected by: pump leakage increases with
pressure, flow across the relief valve increases with pressure, and flow to the cylinder changes. The
controller does not compensate for these changes. It positions the spool of the PV in accordance with
the programmed settings.
SV circuits must be used to maintain a precise output when load fluctuations, or extreme tight
tolerances are required. A SV schematization is shown on Figure 6:
Figure 6: Two stage servo valve (James E. Johnson Penton - Media, Inc)
The hydraulic valves technology behind the systems is still an important choice factor. The next
diagram schematize the performance needs and applications of valves used in hydraulics.
6
Figure 7: Proportional Valve Application Matrix (Eaton Corp.)
As shown on this diagram, the performance of the SVs is not a requirement for the type of
application discussed in this subject, and for pressures used on this applications (up 700bar / 10153
Psi) it is difficult to find components on most of the manufacturer’s catalogs. Here starts the path
where options should be taken in order to choose the control type used for this application and valve
kind selection. Analyzing to the level of standard technology developed for electronic controlled
products of this pressure range and size available on the manufacturers, it is observable how different
these systems can be.
Prior to discussing configurations, the systems can be split in two categories, the ones
achieving levels around 300-350 bar (a level defined by most of the manufacturers) for which much
more components are available but for which larger actuators are required. The second category, up
to 600-700 bar, is a pressure value typically used with jacking tools in civil engineering applications,
especially across Europe. This doubling of the value makes unfeasible the development of a system
respecting the first category condition (till 300-350 bar).
7
1.4 Synchronization of multiple actuators in heavy lifting
The synchronization of multiple actuators is difficult in heavy lifting applications where the
irregular loadings and differences in actuators have much impact on the system performance. The lift
distance among the actuators can be different and manual operated systems can be unstable in
complex movements due to their tolerances, in the sense that differences in lift distance could
increase and could result on the toppling of a load. There are three approaches proposed by [Sun et
al., 1999] to solve the synchronization of multiple actuators:

Flow divider circuit that will keep the same cylinder velocity by maintaining the same
flow rate on the cylinders. This is the simplest one but will depend on the flow divider
performance.

Physically connect the hydraulic actuators (as detailed for example in section 1.2).

Using a EH system, controlling each cylinder individually.
For the control, [Xiong et al. 1992] proposed a model-reference adaptive control algorithm
together with a cross-coupled controller to improve synchronization performance and to attempt to
handle the parameter variation associated with the hydraulic systems. [Chiu, 1994] formulated the
synchronization of multiple motion axes in a geometrical framework and proposed three different
approaches to explicitly address the motion synchronization issue.
[Sun, 2001] and [Chiu et al., 1999] proposed both linear and nonlinear approaches to achieve
the synchronization. The research of EH synchronization control has focused on the linear approach
instead of the nonlinear.
Finally [Sun et al., 2000] proposed a 2-step controller design approach that employs a linear
time-invariant MIMO robust motion synchronization control to address the load uncertainly and motion
synchronization aspects of the problem in the outer loop. A SISO perturbation observer-based
nonlinear EH force/pressure control is used for each individual cylinder to address the nonlinearity and
uncertainties associated with EH systems in the inner loop.
1.5 State of the art
The present market shows some types of synchronized lifting systems originally developed by
hydraulic manufacturers like Enerpac (700 bar), Larzep (350 bar) or by enterprises of the construction
sector like VSL France (700 bar) or Hydrospex (210 bar). Two different systems were taken for
analysis, the first: named Verso from VSL France and the second SLS from Enerpac. They are very
8
different from the design point of view and there are reasons to explain these differences: construction
period, types of control used and versatility versus size.
Taking a look at the constitution of Verso and SLS, different approaches on the hydraulic
scheme types are identified.
The first uses an individual pump for each actuator. This allows ruling the oil flow with an
electronic controlled motor speed. Thus the flow is controlled in a manner, perhaps, completely
different that with a valve. The advantage of this different control system, due to the fact that every
pressure line has its own pump and circuit, is reflected by a more portable solution that can be placed
easily in the application field, with reduction in hose lengths.
The control of both pressure and flow should be closer between each other with Enerpac
system, because having various actuators connected to the same unit, they are controlled
independently only with the electro valves. The SLS represents a different approach that results in a
larger unit with additional hoses length but with an inferior complexity level.
There are several ways to build a system with these functions as well as opinions about their
proper characteristic. It is important to refer that these opinions do not respect a consensus at all. This
work is organized in seven parts that validate and lead to a system design based on requirements and
features defined by professionals. Starting from this point several designs are evaluated in order to
elect one.
Chapter 2 describes the system requirements and develops an elementary hydraulic circuit
following an evolutional process. It is also made a component selection regarding the processes
involved and the components available on the market. Chapters 3 and 5 explain more in detail the
hydraulic control processes and the corresponding simulation in Matlab. In chapter 4 the system is
dimensioned and its parameters are calculated. An economical analysis has been elaborated in
chapter 6 with quotations from several manufacturers of hydraulic, electrical and other components,
with the aim of analyzing the most acceptable investment from two mainly different possibilities.
9
2 Design and selection of hydraulic system
Hydraulic technology as a substitution of mechanical systems has brought different
possibilities to machinery manufacturers. However a drawback of the hydraulics is that the results
during essays are not always as desired because there are factors and variables which are probably
more unexpected that in other mechanical systems with no fluid power required. The numerous
catalogs and manuals of the hydraulic products manufacturers with a lot of explanations,
considerations and applications can give an image like if it was Lego the combination of various
standard products, resulting in easy solutions.
If designed machinery in general should admit demanding tests and proof requirements
before going out to the market, this in general is even more valid for hydraulic systems. As it was
referred by [Speich Bucciareli,1978], one of the reasons is that a designer, other than planning the
system itself, should be aware and familiar with hydraulic features. This is important because the
objective could be a specific characteristic (control of an object, movement, force or other) for which
designers have the knowledge, but it is common that they do not have the knowhow in terms of
hydraulic system analysis.
Typically the solution is the cooperation between the designers and experts on the theme.
Five important considerations can be schematized as follows:

The hydraulic control in a machine requires crucial changes on it, and it is advisable to explore
the latest developments in order to know the latest advances. The final solution should justify the
efforts taken on the searches.

The project development and its accomplishment should be the result of a strict cooperation from
the designer and from the hydraulic experts since the very beginning.

According to the service conditions and the elected hydraulic circuit type, it is necessary to
respect the demands of the mechanical parts with respect to the hydraulics and vice-versa.

The standard components can experience some difficulties when working in some applications
even if they are very successful in most of the machinery.
10

In general an hydraulic controlled machine has nothing in common with the mechanical or
electro-mechanical versions of the same equipment.
However it is not sufficient to have a correct hydraulic circuit design. Equilibrium should also
exist in the equipments construction. For this paper purpose, other than the mechanical and hydraulic
combination, it is also important that the system meets the third factor: controllability. For that a correct
component selection is crucial, as well as the definition of all system requirements, as described below
on Table 1. On Figure 8 are schematized the system principles.
Figure 8: System practical layout
2.1 Component selection
Even when the system functionalities are well defined, the number of different possible circuits
is assorted. Additionally, there are different designs of the hydraulic components available and brands
that influence in different ways the performance of the circuit and consequently the system. The
selection of the components is rarely a trivial assignment especially because it requires knowledge of
some points as detailed bellow:

Available hydraulic components range;

Functional and operating characteristics;

Available possibilities of hydraulic circuits;

Analytical methods for determining systems performances to meet the machine
requirements.
All of these items contribute in increasing the correct choice complexity, of all the system
components and their interaction.
11
REQUIREMENTS
Actuators
Cylinder Types
Pressure
Tolerance
Average speed
Reservoir capacity
8 Independent. Simple or double-acting.
With different areas and displacements capacity. Double and simple actuated
700 Bar
1 mm
2 - 10 mm/s
(20L / cylinder)
FUNCTIONS AND COMMAND TYPE
Move type under tolerance
Complex movements
Stop
Command
Connection +8
Quick approach
Reset to zero
During lifting and lowering. Control in position terms.
Rotations: Different speeds controlled at different points.
Possible anytime and possibility of changing to manual control.
Automatic and Manual.
Possibility of connect to a twin unit.
In the beginning of a work.
After quick approach.
HYDRAULIC CIRCUIT
Nº independent circuits
Pump number
8, one per cylinder.
1 or 8
WORKING CONDITIONS
External
Works duration
Distance between actuators
Rain, dust, transportations.
Up to 48h.
Variable and different between cylinders.
MEASURING DEVICES
Displacement
Encoders external to hydraulic circuit.
OTHER FEATURES
Hydraulic pressure blocking
Pressure discharge
Manual operated valve close to the cylinders before hose connection.
Shut off valve after directional valve.
Oil heat changer
Possible to install for longtime operations.
Table 1: Definition of system requirements
12
2.2 Circuit selection
Once identified the system requirements, it is appropriate to introduce the component types to
install and their performance prior to plan the hydraulic circuit. To distinguish the influences of each
component in the hydraulic parameters (such as pressure or flow) is essential, not only for control
reasons, but also to determine the hydraulic circuit type. As referred, an hydraulic circuit could be
designed with different configurations. The more the system is complex the more this is true.
On figures 9 and 10 are shown different possibilities to control an actuator speed:
Figure 9: a) b) c) Actuator speed control circuits: Meter in, out and bleed off respectively (Eriksson, B. - a and b ) (Parker
Training - c )
Figure 10: Actuator speed control circuit using a servo valve (Cundiff J.)
13
On the images above is possible to recognize four different basic circuits for controlling the
velocity of an actuator:

Meter in;

Meter out;

Bleed Off;

With servo valve.
These four examples describe different possibilities to satisfy a single requirement using
hydraulics. On this specific and simple case: the control of the speed of an actuator. Basically there
are two major different groups on these examples which are: flow control using a flow control valve
and flow control using a servo valve.
The first three examples use a flow control valve and the fourth uses a SV. The meter-in type
is an accurate method used on applications where a load resists in an opposite direction of the
actuator movement, like pushing or lifting a weight with a vertical cylinder. Meter-out on the other
hand, controls the stream on the return line, forcing the fluid to flow from the actuator’s exit through an
orifice. This method is quite used on applications where the load is not constantly in contact with the
actuator contact surface or tends to escape.
In the bleed-off circuit, flow control is obtained with fluid deviation to the tank, by conducting
the excessive flow amount directly to the reservoir. A disadvantage of this method mentioned by
[Parker,1999], consists on a lower control precision. When the flow is routed to the tank and not to the
actuator, makes the cylinder liable to pump oscillations. An advantage is that less capacity in flow rate
is required for valves in this method when comparing with the first two. However this is not a common
application for flow control circuits.
The fourth situation in Figure 10 represents the possibility of controlling the actuator speed
with a SV, which acts like a programmed orifice in the circuit.
However there are other possibilities of controlling the speed of an actuator, such as:
Variable displacement pumps: allow controlling the amount of oil per time unit on the circuit
by varying their volumetric capacity while the hydraulic process is running. Thus varying their
volume, it’s possible to control the flow and consequently an actuator speed.
Control of pump rotational speed: a typical hydraulic pump delivers a certain amount of oil
per revolution. Therefore the flow is expressed by pump volumetric capacity and rotational
speed of the electric motor attached to the pump. Another possibility to control the speed of an
actuator is to vary the motor speed.
14
All these examples represent an overview of hydraulic circuit selection mission, which is
assorted. Nevertheless there are other important factors.
It is important to calculate the fluids performance as well as it is very important to dimension
the components. The components cannot cover the complete range of results that are possible to
achieve for every specific system. This renders true that circuits design should be as flexible as
possible in order to use standard components.
As the desirable was obviously to have the ideal components for optimized results and this is
not always possible, it is reasonable trying to know at the best, the components available on the
market in advance, to avoid harder work. The choice of using standard components is normally
justified by economical reasons and it is commonly mandatory to adapt designs to standard
components [Roca, 1998]. The price of a non-standard component for a specific request could be
unsustainable.
System requirements established before, allow reporting the desirable functions for the system
but not only. From this information, it matters to convert all the dynamics into hydraulic functionalities:
velocities, forces, cycle times, pressures, flow directions, control or others.
Example: Imagining that an actuator is being loaded with greater force that the fluid can do
inside the piston chamber, and it gets overloaded. Damages will occur in the weakest link, which can
be a hose, a cylinder, a tube or other circuit component. To prevent this, safety valves can be installed
in the circuit upside, between directional valves and jacks couplers, and this represents an important
change on the hydraulic circuit.
The successful design of the hydraulic circuit depends on including the functions and define
limits the best as possible. Also a recreation of working cycles is important to validate the options. A
key characteristic to carry out the simulations is represented in Figure 11:
Figure 11: Different movements to be achieved: a) parallel lifting and b) rotation.
The motion represented on b) imposes that some classical methods for synchronization fail
here. This reveals that all the features should be well taken into account for a successful design.
15
2.3 Circuits discussion
To sketch the correct circuit for the desired application is definitely an important factor for the
success. In order to respect the system requirements defined earlier, various options can be
discussed. The sketches are made for a four cylinder single-acting type (instead of 8 due to graphical
reasons) model. The aim is to arrive to an elementary circuit focusing the main system requirements.
Once achieved an efficient circuit, other details will be added in order to get the final solution.
Circuit 1. Four hydraulic lines
Figure 12: Hydraulic scheme on a first approach
An initial illustration of 4 independent oil tracks on the system is based in post-tensioning
hydraulic pumps [Belchior et al., 2006]. In a first analysis, this suits the actuators independence
purpose.
Assuming that 3 directional valves are closed during a lifting and the fourth one is opened, the
pressure will increase in the fourth branch, while the pressure remains unchanged on the other three.
As the branches share a common point on the circuit, when the other directional valves are opened, a
balance on the pressures occurs, unsetting the specific pressures of each actuator hydraulic line.
Figure 13, illustrates such situation:
16
Figure 13: Circuit 1 analysis
Simulations in Automation Studio 5.0 allow schematizing the system performance under many
cases. It was taken the situation where the second hydraulic branch (in red) is pressurized, while in
the others the pressure is zero. When opening the valves of the three other branches, they all will
have the same pressure, because the fluid comes from a shared point. Pressure in the red area
remains equal while the valve is closed. If opened, the pressure will drop and there is no independent
pressure between branches.
Circuit 2. Splitting the pressures
nd
Figure 14: Hydraulic scheme 2 version
17
In a second analysis, the pressure should be individually controlled by adding four pressure
relief valves, each one to be regulated according to each branch needs. Regarding the directional
valves, the center is now closed to prevent flow being deviated directly to the tank, leaving the
possibility to analyze this detail’s influence.
However if pressure is controlled, in the other hand it is not possible to path the oil to the tank
directly because the pump is always loaded. The oil is dumped over at 100% by the pressure relief
valves, which would heat up incredibly the oil temperature.
Figure 15: Circuit 2 analysis regarding pump permanent loading
Figure 15 illustrates occurrences when the pump is loaded prior to the actuator’s
displacement. Another disadvantage is that to control pressure, an inefficient method is necessary:
control the pressure on a branch independently requires that each pressure valve should be regulated;
in the meantime all other branches have the directional valves on center positions. This is the same of
having just one pressure relief valve, as exemplified with an example:
Figure 16: Circuit 2 analysis for pressure control
18
Simulations revealed that to achieve the situation of Figure 16 it is necessary to set each
th
pressure valve to maintain the pressures independently. The 4 directional valve (on the right side) is
opened, and the pressure value for this branch is 200 bar. To make this possible, the other branches
must have for now their directional valves closed and the respective pressure valves should be
regulated not to open before 200 bar.
The conclusion therefore is that with this circuit, it is possible to have independent pressures,
but working with each cylinder at different time intervals, and for this, one pressure relief valve is
sufficient. This method would result in very slow operations since each cylinder would be moved one
by one if different pressures were required.
Circuit 3. Multi pump solution
rd
Figure 17: Hydraulic scheme 3 version
By adding a pump per branch, it’s now possible to have an independent situation in the
branches. As there is no flow control valves available on the market for pressures up to 600-700bar, to
control the flow in this condition, the solution will be the valve switching frequency as used before by
[Liu et al., 2000]. It is possible to individualize now flows and pressures.
19
Circuit 4. Controlled lowering
Circuit 3 does not take into account that operations done by jacks are heavy weight. A normal
lifting situation involves the vertical ascension of several tons, which should be controlled with the
same precision both lowering and rising. Figure 18 shows an option to control the decreasing
movements, constantly powered by gravity, in a smoother manner. Denoting that valve switching
control is still used.
th
Figure 18: Hydraulic scheme 4 version
This scheme shows a line with a throttle and a 2/2 valve, which reduces the area of the pipe
and allows decreasing the velocity of the lowering operation, which is ruled by the structure’s weigh.
Without this, a lowering is done by the return line, which should be normally sized (as the pressure
line) because it should accomplish operations not only on the vertical direction (but also parallel) and
so the velocities for return should be capable to have the same magnitude in both flow directions.
This circuit was tested and the results are illustrated in Figure 19:
Figure 19: Circuit 4 analysis with single-acting cylinders (left) and with double-acting cylinders (right)
20
The system provides reduced lowering speeds with single-acting cylinders. However with
double-acting cylinders (which is a common case in lifting works) there is an impracticality in working
with this 2/2 valve: if the 4/3 valve is closed, there is no connection from the cylinder to the reservoir,
and the return line goes in vacuum (pink line). A double-acting cylinder remains blocked at this point.
Circuit 5. 3 Directional Valves
The vacuum problem is solved with a set of 3 directional valves: One for pressure, one for
return and a third one (as in circuit 4) for lowering. It is now possible to use single or double-acting
cylinders. A fourth valve was included to do the oil by-pass, since with the three directional valves
closed the oil should flow directly to the tank.
Figure 20: Hydraulic scheme 5th version
Synchronized movements are a consequence of all the control orders at the ensemble of the
directional on/off valves. Each valve has its function and their positional configurations allow the
different operations. In Figure 21 are represented the directional valves:
Figure 21: Set of 3 directional valves
21
Valve 1: Pressure valve - allows/blocks oil flow towards the cylinder first chamber.
Valve 2: Lowering valve - controls the passage of the flow in case of a lowering when the jacks are
loaded.
Valve 3: Return valve - Allows the normal return of a cylinder.
The switching combination between these directional valves 2/2 and 3/2 allow to define
whether the oil is flowing towards the pressure chamber, return chamber or simply lowering. Not only
the order to one valve is necessary, but the combination of orders to all of them is essential. The
orders for the different functions are:
Movement
Valve
Lifting
Lowering
Return
One
On
Off
Off
Two
Off
On
On
Three
Off
Off
On
Table 2: Combination of solenoid signals for the directional control
This circuit provides the complete solution for the movement types. And it is easier to find its
components since there are more manufacturers producing 2/3 and 3/2 valves for high pressure, like
for example the worldwide famous Bosch-Rexroth. Material for such high-pressures is sometimes not
available, and few manufacturers worldwide produce directional 4/3 solenoid-controlled valves for this
industrial application type. It is quite important to refer that solutions provided by hydraulic sellers and
designers mostly depend on the hydraulic brands they represent which directly interferes on the
solutions they point up.
It was mentioned before that the contact with professionals and their practical knowledge has
great importance for a successful project, nonetheless is not less true, that for specific equipment like
this it’s also elementary to know what is available.
22
th
The 5 circuit controls all the desired functions; however its disadvantages are:

The fact of having more manufactures/selling points and being easier to find 2/2 and 3/2
valves that 4/3 is not rewarding in the system’s overall price. A solution with four valves in
place of one is much more expensive.

A controller with much more outputs would be necessary. Standard controllers for hydraulic
pressure and positioning could not be implemented here because they do not have such
number of variables per independent axis.

The system is more intricate and complicated.

The maintenance is harder as well as potential problems are harder to identify. For example
pressure loss in a system is often related with problems in the pump or in directional valves
and this system have three directional valves and another one for by-pass.
Circuit 6. 4/3 Directional valve with regulation: elementary circuit
Economical reasons among higher simplicity were the main reasons for purposing the next
circuit. The concession taken here was the selection of a circuit at the cost of more difficulties in
assistance, installation or expert advertisement, since it is sparse to find official licensed dealers of all
components at national level.
Figure 22: Elementary circuit.
23
Implementing a manual variable non return valve, allows decreasing the flow area only in one
direction, this answers to the same problems as circuit 5 did. A major advantage is noticed when using
in the same operation, cylinders that work in different directions as shown in Figure 23:
Figure 23: Elementary circuit analysis
The non return throttle valves of the branches on the right side could be regulated to be opened
at 100% since they will not realize a lowering, (further, Figure 49 shows a situation where this occurs)
while the branches on the left can be pre configured to full speed or to provide low speed depending of
the requirements.
This circuit called from now on, elementary circuit, will be completed with necessary implements
to its functionality.
2.4 Working over the elementary circuit
The so called elementary circuit aim is to congregate and fulfill the basic hydraulic functions
for the desired result. It is compatible with most of the features and requirements expressed before on
table 1. However with more intensive analysis, other components must complete the circuit.
From now on, add-ons will be made over the elementary circuit for economical, control, safety,
metering, or even constructional reasons.
24
Figure 24: Elementary circuit with details
On Table 3 are exposed the components that have been added to the elementary circuit:
1
2
3
4
5
6
7
8
9
10
11
Quick couplers for hose connection
Safety valve
Pressure gauge
Pressure transducer
Shut-off valve
Variable non return throttle valve
Directional 4/3 valve solenoid operated
Check valve
Proportional pressure-relief valve
Radial piston pump
Filter
Table 3: Assigned components of the system
25
Quick couplers
The coupling of hoses to the system should be easy, quick and safe. The best way is to equip
the system with high quality quick couplers, leakage free that can connect two lines with reliability on
time. They represent the border that separates the system from the actuators and their hoses.
Figure 25: Quick couplers C604 (Enerpac)
Safety valve
There was added a second pressure relief valve, for which there are two reasons that justify
this option.
The first reason is that system maximum pressure could be generated, in a circuit fraction
after the directional valve, if it occurs when directional valve is at centre position. This can happen due
to a non predicted structure weight. For this condition the first pressure relief valve only controls the
pressure below the directional valve.
The second reason is to protect the jacks with an independent safety from the control system.
This device should be able to work with different jacks. This difference supposes diverse models, and
so there are unequal limits on actuators maximum pressures. Jacks with maximum pressure bellow
the system nominal value are extra protected, once not all jacks have safety internal valves.
26
Pressure-relief valve
Features: Up to 700 bar; up to 12l/min
Manufactures: BIERI, HAWE, Rexroth-Bosch
Quantity: 8
Figure 26: Pressure relief valve (Bieri Hydraulic AG)
Pressure metering
The pressure metering is very important to know always the pressure/force involved on an
operation and should be traced continuously by a sensor (transducer). The transducer can be used
also to make calculations for structure weight, or else to notice unpredictable forces in a certain lifting
point.
More than provide digital information for the equipments reference, a pressure gauge installed
on the pressure line is a guarantee to always know the maximum pressure achieved directly without
interferences which is also important in case of some electronic failure. Hence both transducer and the
pressure gauge are essential for the system.
Pressure transducer
Features: 4-20 mA 2-wire output signal
Model: Wika S-10
Quantity: 8
Figure 27: Pressure transducer S-10 (Wika Instrument Corporation)
27
Shut-off valve
The shut-off valve included is related with the need of releasing the pressure inside the circuit
anytime in a safe and directly manual operated way. The valve is located on the fraction that connects
the directional valve to hose coupler. This is the line fraction that keeps the pressure when the
actuator final position is achieved and which will conserve it till the beginning of the opposite
movement (a lowering in opposition to a lifting). It can be used to relief pressure contained inside the
system after or before a work.
Shut-off valve
Features: Up to 1000bar, up to 30l/min
Model: BIERI type SPV2
Quantity: 8
Figure 28: Shut-off valve (Bieri Hydraulic AG)
Variable non return throttle valve
Manual regulated variable non return throttle valve, allows unconstrained flow only in one
direction. In the final product assembly, there should be indications of throttle opening percentage to
ensure the equipment users about its correct setting. This component has introduced a greater
simplicity in the overall system as explained with circuit 6, page 23.
Directional 4/3 valve solenoid operated
The 4/3 directional valve controlled with on/off solenoids is one of the most important
components and allows to control the stream across pressure and return lines. It is actuated with two
solenoids and in case of a power failure the springs maintain the valve on the centre position. The
28
open centre provides the possibility of by-passing the oil directly to the tank, from the pump, which is a
very important aspect as referred earlier on circuit’s discussion.
Directional 4/3 valves
Features: Up to 700bar, porting NG6; direct
operated, different solenoids available.
Manufactures: BIERI, HAWE
Quantity: 8
Figure 29: Solenoid operated directional valves
Check valve
The check valve allows running the flow on a single direction, and is a logical and safe part
simultaneously. It assures that the oil flow does not return towards the tank during the lifting, if
something breakdowns in the system, or in power failure situation.
Proportional pressure-relief valve
The proportional pressure relief valve is essential along with the directional valve. It allows to
control the pressure continuously by using a proportional solenoid that creates a pressure differential
from the pressure line to the tank, maintaining the desired pressure value. Pressure must be
controlled by levels in order not to introduce excessive forces on structures.
29
Proportional pressure relief valve.
Features: NG6; up to 700bar; proportional solenoid
force
controlled;
with
electrical
loop
control:
automatic adjustment of system pressure to a preselected pressure; pressure setting with PLC,
manual operated potentiometer or via BUS.
Models: BIERI PDV700, HAWE PMV
Quantity: 8
Figure 30: Proportional pressure relief valve (Hawe Hydraulik SE)
Pump
Pressures up to 700 bar require radial piston-type pumps, which have a compact design and
provide pressures up to 1000 bar. The pistons expand against the cylinder wall and with positive
acting check valves the huge pressure values are achieved. The positive acting check valves also
maintain the oil flow to the correct direction. It’s doubtless the possible choice for this application.
Radial piston pump
Features: High pressures (up to 1000bar),
self-priming,
self-venting,
various
outlets
possibility.
Manufactures: Rexroth, Bieri, Hawe, Dynex
Quantity: 1 or 8
Figure 31: Radial piston pump (Hawe Hydraulic SE)
30
Filter
Always important on an hydraulic system, is the filter which allows keeping the oil clean. It also
blocks the passage of undesired particles across the hydraulic circuit which can damage valves and
other components. It is a vital component since the outside working conditions are tough and dusty.
The works performed by the equipment, include practicalities, like for example, the need to
close the circuit after a lifting (or other movement). Between a lifting and the corresponding lowering,
the displaced structure should be motionless to perform other operations. These tasks should be done
in maximum safety, so the oil should be absolutely contained in the jacks and in the minimum length of
hydraulic circuit possible. The normal and correct solution is to install a manually operated valve in the
end of the hose connections right before the actuator. The directional valves are located some meters
forward, and so they are not reliable for maintaining the structure position for long time.
Another important aspect not represented on the hydraulic scheme is the monitoring of the
position and velocity which is outputted by a position sensor. These sensors are a very important
component for the rotations because a rotation is a performed with different velocities performed at
different lifting points.
Figure 32: Wire displacement sensor (Micro-Epsilon GmbH)
31
2.5 Pump options
High pressures lead unavoidably to choose a radial piston pump for this system. However some
configurations can be made from this presupposition. They are essentially: the number of pumps,
number of outlets and the existence of auxiliary pumps. The next lines show the hypothesis studied.
2.5.1 Single pumping system with auxiliary gear pump
This first option combines a radial piston pump for high pressure and low flow, with a secondary
gear pump which has a greater flow but only supports low pressures. The advantage of having this
second pump is that at an operation beginning, the approaching position of the cylinders takes less
time. Also the filling of all pipes and hoses from the pump to the jack is quicker. This solution implies
an integrated pump, or in the other hand: two separated pumps powered by the same shaft, one more
safety valve for the gear pump and one check valve to obstruct fluid (at nominal pressures) entering
the gear pump line.
Figure 33: High pressure-low pressure pumping configuration
This option still being admissible is only justified when installing one pump group per cylinder
(referring to circuit 3 analysis), consequently subdividing in subsystem units. Although the unit
subdivision, has other advantages, as it will be shown on section 2.5.3.
32
2.5.2 Multi-outlet pump
The multi-outlet pumps are modular designed pumps fitted up with various flow outlets and
can be installed in devices with diverse hydraulic circuits. With only one pump and one motor, is
possible to have various oil lines with individual pressures and flows, similar to having various pumps
connected in parallel but in a much more compact solution. These pumps (represented in figure 34)
make possible to build a single body unit.
Figure 34: Multi-outlet radial piston pump with four exits (Dynex/Rivett Inc)
The pump design with check balls along with isolated chambers provide the output of each piston to
be used separately.
Figure 35: Configuration with multi-outlet pump
33
This solution is an interesting choice and it can be used if a single body construction is desired for the
system. It will be taken into account.
2.5.3 Separated units
This configuration provides a single pump per circuit and consequently the system separation
in eight units which are interconnected by a central command.
A larger volume pump is here
substituted by smaller pumps, one per each lifting point.
Figure 36: System configuration with various separated pumps
Being a convincing choice, it can be still equipped with a two speed electric motor, avoiding
the use of two pumps per circuit like in 2.5.1 which would increase the unit size. Also the fact of using
an electronic command to input different velocities could have some positive impact for future
applications as it will be referred further on this study.
In one hand the multi-axis control of 2.5.3 is also the most energy efficient design because the
different drives can be adapted independently to control speed and torque. In the other hand the multioutlet pump design of 2.5.2 is less expensive and easier to build.
Hence both systems provide valid solutions and offer two different design approaches. These
designs represent differences in the operations field, especially in hose length, transportation and
working position.
34
3 Hydraulic control
Synchronizing the motion of various cylinders is never an easy task. Stroke length, depends
on the volume of oil delivered to a cylinder, and moving them together requires equal oil flow to each
cylinder as well as diameters. Oil leakages, pump slip, changing workloads, and varying friction loads
could also affect oil delivery among other factors. On a system like this one, uneven loadings are
expected and can be a regular situation during operations. To ensure synchronization of all the
cylinders, [Natchwey, 2009] advises that each one must be powered and controlled from a separate
circuit that allows independent, variable and reversible oil delivery.
Nowadays, however, fluid power systems are capable of high performance motion control.
Well designed closed loops systems can execute works involving heavy forces with accuracies of less
than a millimeter. To achieve this, there are several factors that are not only dependent of basic
hydraulic components as valves, pumps and cylinders. The combination, of the metering quality,
(sensors and transducers for pressure, position or flow), controllers and programming techniques and
special fluids selection is extremely important, in order to achieve the best results.
3.1 Advised rules for hydraulic control
There are some rules normally applied to the design and control of hydraulic systems. It is
relevant to mention that for this specific application, some of the standard and basic advices should be
changed under the particular working conditions. These features described by [Natchwey, 2009] are:
35
RULES TO FOLLOW:
Use servo-quality valves
As referred in advance, the high hydraulic pressures involved, are not suitable to use servo
valves or even proportional, which are not available for such high hydraulic pressures. The solution is
with on/off valves.
Place valves on cylinders
Recommendations to mount valves, both on the end and on the top of cylinders keeping a
minimum amount of oil trapped between the valve and the piston, as small as possible, is not possible
here. From the valves till the cylinders, it goes a hose connection with some meters typically because
the pumping unit is far from the actuator. In this case, trapping oil between cylinders and valves is a
mandatory feature, for which this recommendation feature does not apply here.
Use accumulators and pre-charge them
For this pressure range an accumulator is not advised. In case of energy failure, the check
valve maintains the position at the given moment.
Size the pump correctly
Size the pump of the average oil flow plus 10%.
Size the cylinders correctly
The suggestion of using cylinders with diameters large enough to get the desired
accelerations or decelerations rates doesn’t fit on this system. The cylinders to be used with the
2
2
equipment could have differences from 80 cm area to 1000 cm or more. And the control of system
should be flexible enough to support these differences.
36
Size hydraulic lines correctly
Should be used adequate pressure and return line with a minimum number of 45º and 90º
joints.
Size valves correctly
Valves should be correctly sized: large enough to allow the system to accelerate at the desired
rate and the maximum desired speed. This point should be consistent also with the fact that at high
flow across the valve the pressure drop across the piston will drop, reducing the ability to accelerate.
Use a suitable motion controller
The hydraulic control requires parameters, transducers interfaces, and algorithms not
available on motion controllers intended for motor controls. Specific algorithms or even specific
controllers applied to hydraulic applications are advised.
Use a precision position feedback transducer
For the equipment, the sensors that are planned to be used are wire sensors with current
output. This displacement sensors were already tested on these working conditions and have high
precision.
TO AVOID:
Flexible hoses between valves and cylinders
This feature is noted as very important, and once again cannot be followed. The advisement is
to avoid the use of flexible hoses between the valves and the cylinders because this leads to difficult
tuning and poor control. However there is absolutely no chance of not doing this, because this is the
way of connecting the circuit units to the jacks on the field during the construction works.
37
Overlapped Spools
Overlapped spools are those with closed-center. When an overlapped spool shifts from one
side to another to change direction of the motion, the system tend to stop with a bump which leads to
a more difficult tuning around critical zero point and results on a poorer control. With open center
spools there are conditions for more smooth transitions between opposite directions
Separate Valve Amplifier
It is desirable to avoid valves with separate amplifiers, unless it is possible to set up the
amplifier properly, since many valves have the electronic built in the own valves and are set up by the
factory.
Run wires in same raceways
It’s not recommended to run the high current Pulse Width Modulated signal, from the amplifier
to the valve, in the same conduit or raceway as the wires from the position feedback transducer.
Go half way on closed-loop control
Closed loop control is simple and effective. In some cases it is normal to find two different
valves on the system. One for open-loop control and one for close loop control, along with plumbing to
switch between the two. This renders the system more complex requiring more programming and
maintenance than a simple closed-loop control. On this systems case, the control tries to follow this
advisement.
Counter-balance valves
The use of counter-balance valves is unadvised, because they counteract the motion
controller.
38
3.2 About valves
The choice of the valve components to develop heavy lifting equipment for this range of
pressures will also interfere in the control type used on it. In Figure 7 highest performances required
are essentially for uses that need a good speed characteristic (vibration exciters), or a good precision
(flight simulators) or even a combination of both (machine tools).
Trying to include heavy lifting equipment in a category based on a similar matrix, means to
insert it on a low/standard proportional valves group. At least when comparing qualitatively with some
examples given, such as: material handling, cranes or construction vehicles, due to the similar values
of velocity and process class size.
However these referred examples despite being more similar, are still far from the application
in study. Taking a look on the system requirements defined before on Table 1, pressure values
achieved are certainly higher and time for cylinder extension and retraction are definitely slower. This
means that the hydraulics involved in heavy lifting works are particular. Thus the control used should
be chosen having in mind that the reduced velocities determine the use of directional on/off control
valves. This renders the system simpler and cheaper.
The solution for pressure control is chosen based on different considerations especially for this
equipment’s mission. The continuous monitoring of the pressure is an essential feature, more than in
other applications discussed, which are concerned specifically with the displacement control.
Generally forces are not unknown on some works like they could be on a bridge or roof lifting.
In other applications pressure should be available in order to respect the system requirement
(Hydraulic Tools, Rolling Mills, Utility Construction Vehicles) while in heavy lifting operations the
system oil pressure should be controlled within varying limits.
3.3 Types of control
On this system the control types are respectively on/off, also known as bang-bang control
(directional), and closed-loop control (pressure). The selection of these two types of control was based
on different processes to be controlled such as pressure and position, but also with the valve types
available as seen previously on this text.
Options taken are related with the inexistence of SV’s for this application pressures. There is
consequently the impossibility of using closed loop control to manage the oil flow and direction with a
precise tuning using the spool of a valve to control position and velocity. In the other hand the
existence of proportional pressure relief valves allows to control pressure continuously, with closedloop.
The discrete control is often called bang-bang control due to the noise that an actuator makes
when it stops abruptly. It cannot accomplish certain tasks and ratios of precision and velocity that a
closed-loop control with continuous signal with continuous feedback and output control does. However
as demonstrated further in chapter 5, the conditions involved are satisfied when this control type is
39
used. Table 4 shows a correlation between the combinations of control and feedback possibilities,
illustrating therefore the discussion:
Control type
Feedback
On-off
Open-loop
Closed-loop
Least Precise
Limited speed control
-
Discrete
Stop at position
Stop at position with ramping
-
Continuous
Flexible stop at position
Ramp to position
Precise control
None
Table 4: Feedback and control combinations (Natchwey P.)
The first thing to notice is that it is possible to use on/off valves with a combination of different
feedback signals and there are three ways to combine bang-bang control. Thus the quality in the
feedback does difference.
Assuming that an on/off valve controls a cylinder which is position feed backed with a discrete
signal: every time a motion tolerance is adjusted, it is necessary to change the valve switching
frequency and consequently the quality of metering information must be well assured in order to
respect tolerances.
For directional purposes in this case, the feedback is the reference to the valves and must be
continuous. It allows a more precise response, especially because a continuous comparison between
eight cylinders is being made.
The signal comes from an encoder that measures position on each point across the travel
range of every cylinder. By providing this information to the controller there is a continuous time
measuring. Thus if some cylinder escapes out the tolerance defined, an order to the corresponding
valve is sent.
However the discrete feedback is not as flexible in velocity information as the continuous one
because with an encoder or a linear displacement transducer is easier to detect how fast the position
changes, which is critical information when cylinders should execute a rotation motion. Obviously the
no-feed control is not an option here because it is very limited on its flexibility. The only position that
can be achieved with certainty is when a jack is either fully extended or retracted.
The slow travels of the cylinders on this system justify controlling the directional valves with
discrete actuated valves.
40
3.4 Pressure control
3.4.1 Components
An hydraulic system only produces pressure because a fluid inside a circuit gets confined.
This condition results from the circuit components or from control orders. Otherwise, the only pressure
resulting would be the pressure of the flow similarly to a fluid transport system. Thus the pressure on
an hydraulic system appears only when the oil flow gets restricted or is being limited against a motor
or an actuator (with their respective resistances). As a result, the pressure control is a different
process from the directional control.
In another pressure range (till 300-350 bar) to control a pressure on a system, it is common to
use an electronic controlled variable displacement pump in a closed loop control system. To control
the pressure up to 700 bar in this case, a normal fixed displacement pump for high pressures will be
used and also an external proportional pressure control valve. These valves are solenoid force
controlled. The functional and essential part of a force-controlled proportional valve is composed by
the spool that stands in position by a spring:
Figure 37: Functional principle of a force-controlled proportional valve [5]
Current passes in the solenoid coil and the plunger pushes the spool, which allows opening
the valve. The higher the current applied to the solenoid, the higher is the deflection of the spring and
consequently the force. It is important to refer that the force contained on a solenoid is nearly a linear
function of the input current.
41
Figure 38: Function of force along the stroke in a spool of a proportional force-controlled valve for different current values
Comparing with a conventional pressure relief valve, a solenoid force-controlled replaces a
spring and a screw. These are the main components in several types of pressure control valves. This
substitution allows adjusting continuously a desired value instead of having a fixed limit manually
defined.
For high pressure or flows, these valves could be pilot-operated. As described by [Chapple,
2003], in these valves, the pilot section drains across port Y (figure 39) to drain a line directly
connected to the reservoir. The precision of the solenoid requires that the pilot be drained directly
back to the reservoir. If back pressures in the T line causes instability and drain lines outside the valve
should be provided, the back pressures of internal drains could cause wrong measurements and valve
operation.
42
Figure 39: Cross sectional view of proportional pressure relief valve (Chapple P.)
Figure 40: Diagram of the core of a proportional pilot-operated pressure relief valve (Chapple P.)
As detailed by [Cundiff, 2002], the response time for these valves is good and normally the
time for a power amplifier to change from a pressure level to another in response to the milliamp
signal from the amplifier, ranges from 50 to 150ms, still depending these values from the valve size.
43
Regarding electronic problems, if there is a power failure, the signal goes to zero and the pilot
poppet will open for very low pressures. In the other hand if the failure inputs a high current and a
high pressure a mechanically adjusted pilot can be set up to relieve a pressure just over the solenoid
setting, ensuring the circuit protection.
3.4.2 Control diagram
To control the pressure the valve will operate under a closed loop system. Closed-loop
systems work by comparing an output response with an input demand signal. In this EH control
system, the input signal is a voltage and the output is fed back as a voltage signal, which exits from a
pressure transducer.
Both signals are evaluated, and the difference measured between them is an error, which will
be amplified as a current signal and supplied to the valve. The ratio obtained by [Chapple, 2003]
between the input and the output is obtained from the following equations, where valve position is
given by:
–
(3.1)
X - valve position [m]
p - pressure [bar]
KT - pressure transducer gain [V/m]
KV - valve gain [m/A]
KA - amplifier gain [A/V]
Vi - input signal, [V]
44
Figure 41: Simplified system block diagram (Chapple P.)
This diagram is used when it is necessary to control the force of a cylinder. This represents
also the closed-loop system for the pressure control using a proportional pressure relief valve and a
pressure transducer. As this type of control allows controlling a force of an actuator, the whole process
is affected by the load on the opposite side of the cylinder and its characteristics. Also the mechanical
stiffness of the system contributes with its influence for the process. Regarding this aspects, the block
diagram could be approached on a more precise manner by:
Figure 42: Pressure control block diagram (Chapple P.)
A - Area of actuator
Cf - Coefficient of viscous friction
Cx - Flow coefficient valve
Cp - Flow coefficient actuator
fE - External load force
m - Piston mass
s - Complex variable Laplace Transform
β - Bulk modulus
45
The purpose is to control system force being applied during lifting within certain limits, or
setting a lifting step by step, for example from 50 on 50 bar. Analyzing the diagram for a constant load
force, there is:
(3.2)
β
In the closed loop system, the transfer function is given by:
(3.3)
β
Thus, the steady state gain is:
(3.4)
This expresses the situation for which the load force over a jack is constant, the piston is
moving at a velocity such that the viscous friction, C f, absorbs the actuator force from the pressure
increase.
46
3.4.3 Scheme
Connected to the pressure line is the proportional pressure relief valve. The valve is
commanded by the controller that takes information from the pressure transducer. The transducer
directly connected to the hydraulic line, measures the pressure and sends the value to the controller,
which receives the pressure signal and compares with the desired pressure inputted by the worker.
The controller defines the solenoid current to open the valve and keep a pressure differential
that allows maintaining the inputted value inside the pressure line. In Figure 43: Pressure control
scheme follows the representation of the pressure control interaction:
Figure 43: Pressure control scheme
3.4.4
Controller parameters
In many cases the controller’s design aim is to supply high gains at low frequencies in order to
minimize steady state errors.
Since the pressure must be regulated by levels as in a saturator, the pressure control inside
the system only requires, a proportional gain acting on the proportional valve solenoid. The control is
assumed to be closed-loop. Saturation is defined by the fact that the controller should simply avoid
that the pressure should not surpass a defined value, previously inputted by the equipment user. This
also defines the control type as linear, and it will be adequate for this purpose, avoiding complex
47
parameter analysis as in equation 3.4. To explain this, an antagonism will be used with an opposite
situation:
Figure 44:Pressure control in plastic forming machine example
Considering a plastic forming machine and assuming that the process should keep specific
pressure value for the injection operation (admitting that force variations during the injection process in
excess or in lack could both origin defects on the final product). If this force is controlled hydraulically
by a proportional pressure-relief valve, the control should be closed-loop and substantially more
complex, once every pressure fluctuation should be suppressed and stability should be maintained. In
this case a PI or PD controller should be required.
Once for the system in study the objective is only to maintain pressure levels, a proportional
gain is used on the controller. The gain value in simulations is K = 10 [mA/ Bar]. In Figure 45:
Manufacturer graph for PDV700 (Bieri Hydraulic) is shown a graph illustrating the magnitude order of
the gains for proportional pressure relief valves.
48
Figure 45: Manufacturer graph for PDV700 (Bieri Hydraulic)
3.5 Directional control
On/off valves solenoid actuated are generally used in hydraulic control systems of engineering
machines because have interesting advantages such as:

Simple constructional structure;

Low mass;

High resistance to pollution;

High reliability and easy application

Easy maintenance;

Low cost;
Directional control is responsible for directly act over the cylinders, commanding whether a
cylinder is rising, lowering, or stops.
Control is achieved when for each individual movement the cylinders respect the given
tolerances to produce a displacement, being the tolerances measured during each programmed
movement.
49
All cylinders are constantly compared in order to see if there is one or more passing out the
comparative tolerance. If this is verifiable, the valve order should be annulled till the delayed cylinders
(according to the specified movement direction), rise the sufficient to maintain the given tolerance. The
next diagram shows the scheme of the directional control:
Figure 46: Directional control processes diagram
50
The directional valves are commanded with two solenoids, each one is activated regarding the
cylinder movement of raising or lowering. When the movement is initiated, if it is a lifting, the first
solenoid is activated and the valve allows flow from P to A.
The corresponding order to stop if a tolerance is surpassed is not allowing flow from P to B
(which would return after energizing the second solenoid and power off the first one), but switching off
the first one, returning the valve to the centre position. Fluid will flow from P directly to T till the
movement is again under geometrical acceptance.
This situation could be repeated as many times as necessary, being the switching frequency
proportional to the tolerance decreasing values. Exactly the opposite operations occur for a lowering in
contrast with the lifting. Figure 47 represents a on/off directional valve:
Figure 47: Diagram of a solenoid actuated directional valve
(Bosch-Rexroth)
The directional control allows working both with single or double-acting cylinders. Typically in a
lifting applied to these civil engineering operations, single-acting cylinders are used, once the pushing
is made in vertical direction. Thus the down movement does not require other help that the own weight
of the heavy structure by itself.
Special movements, such as the parallel pushing of a structure, require returning stroke
motion to be made with pump direct pressure. These operations are driven by special double-acting
cylinders:
51
Figure 48: VSL strand jack lifting system with double-acting cylinders
Figure 48: VSL strand jack lifting system, illustrates a parallel pushing or lifting with strand jacks.
This system works with a double grip system. A strand is fixed by 2 grips, one on the jack that pulls
the strand and consequently the structure, and the second that keeps the structure at final stroke
position.
The grip system allows returning the actuator’s piston to the initial position, while the structure
is fixed. Thus the use of the fluid return line is so important like in the pressure.
Another condition that requires a total independent control on each circuit is in case of
complex movements where two or more cylinders should be synchronized in a first movement of
pushing, and a second group also synchronized to accomplish a lifting. Figure 49 identifies an
example:
52
Figure 49: lifting and pushing schematization
Another interesting movement type is the rotation, which is a particular ascending movement.
It requires that all the cylinders involved, respect a ratio between the variations of the structure
declination in each point and not a tolerance numerical value, as in translations. In every defined point
is a jack, which working arrangement should be well known prior to define stroke limits. Distribution
arrangements of the cylinders should be defined during the engineering project that anticipates the
works. The project always defines the number of cylinders and positions to allow a successful and
safe work. Some requirements normally presented on the projects are related with:

Necessary number of jacks to assure the necessary force according to the predicted structure
weight;

Distribution of applied forces by all the cylinders across the structure and the momentums
supported;

Forces values applied by the jacks;

In case of a concrete structure: the maximum surface tension of the concrete. Actuators must
have contact areas large enough to avoid excessive surface pressures applied. Mistakes
could lead to a penetration on the structure caused by the actuators.
These characteristics must be defined a priori and are very important in all displacement types
to ensure a safe and triumphant work. But in a rotation (schematized on figure 50) they are
fundamental because the position of the cylinders is previously necessary to calculate how much a
cylinder should rise:
53
Figure 50: Schematization of a rotation
In a rotation, each cylinder stroke value is formerly calculated and defined before operations.
When this value is introduced into the system, the electronic controller will compare positions between
all dynamic points. System will evaluate if each point conserves the inclination reference during the
operation time. To schematize this, Figure 51 illustrates an example:
Figure 51: Rotation example
The structure is rotating due to the lifting points 1 to 3 represented on the left side of Figure 51
by z’s positions. The geometric ratio between assigned points on a virtual straight line is represented
on the left side by h’s. The h’s relation is the following:
54
(3.5)
System will compute the data on a rotation movement following the inequation:
(3.6)
The assigned values for the tolerances in rotations are not defined in millimeters, or length
units in opposition with parallel displacements. Tolerance is defined as an inclination percentage,
being thus dimensionless.
However the system should predict other arrangements and variations that can occur for the
movements. Figure 52, shows a rotation where one cylinder rise down and others are rising up. In this
case other control orders should be given:
Figure 52: Rotation example with different rotational centre
For specific motions like this, a manual configuration interface when programming the
actuators travel is the best option. For this specific case the cylinders b and c should follow the
algorithm of a rotation with 2 cylinders while the cylinder a (manually assigned by the user) will
perform a lowering with the same inclination tolerance of the motion during the interval time of the
working act.
55
4 Circuit dimensioning
Once known the application of the hydraulic system and its components functions, as well as
its technical requirements, the next step is to dimension the whole circuit. In this task it is important to
consider the standard components that better fit to the needs.
In this system, parameters have a vast range of possible values: for example the hydraulic
cylinder areas. However for circuit dimensioning, velocities, forces, times or cycles should be
admitted. The circuit diagrams along with specifications defined before and the use of hydraulic
formulas allow setting off the dimensioning.
4.1 Parameters
Identifying nominal values for system operations like suggested by [Roca, 1998]:
One of the largest jacks that could be used on this works for reference is:

Model: SLU 120/ZPE200

Max. Force per lifting point = 615 10
Area: 325.7 cm
5
2
Maximum pressure 615 bar
-4
325.7 10 = 2003 kN
Ideal flow rate for an average cylinder, with regular pressure area on lifting applications

Model: HFP 80
Area: 201cm
2
Maximum pressure 400 bar
Acceptable lifting speed: 3 mm/s
56
Corresponding flow rate will be:
Flow rate: Q
=
=
=
s
3
Q - Flow rate [m /s]
- Piston velocity [m/s]
A - Actuator area [m]
Desired flow rate: 3.6 l/min
4.2 Fluid
Operational reliability, long service life and well functioning are characteristics that depend
from the selection of the fluid. As defined by [Götz W., 1998] the demands imposed on the hydraulic
fluid are diverse, as can be seen in the following list:

Transfer of hydraulic power from the pump to the actuators

Lubrication of moving parts, such as sliding surfaces of pistons and spools, bearings and
switching elements etc.

Protection of the metal surfaces actually wetted by the hydraulic fluid

The removal of contamination and dirt, abrasion, water, air etc.

The removal of heat loses which have been caused by leakage and friction loses.
The standardized classes as defined by ISO, establish specific properties for the different oil
types, being the appropriate class for this fluid the HV, defined as hydraulic fluids with special highpressure additives which result in increased protection against wear and with extremely low viscositytemperature interdependence. The selected oil market will be a Shell Tellus T 68 for its high
performance.
57
Figure 53: Shell Tellus Oils T characteristics
Temperature dependence of viscosity
Oil manufacturers provide information for the fluid most important characteristic, the viscosity.
The information is presented by its temperature dependence. Considering oil temperature service of
50ºC, recommended for system dimensioning by [Cudell, 1978] and with manufacturer information
(refer to Appendix C) the oil viscosity
is:
= 45 mm2/s
Pressure dependence of viscosity
Although manufacturers do not provide information regarding pressure influence on viscosity,
there are equations to calculate variations of viscosity based on pressure levels. While the increasing
temperature decreases the viscosity value, the pressure has the opposite effect. In theory should be
provided for each oil, a graphical model with the dependence of viscosity with both pressure and
temperature. Challenged by this question, Hidromac, Ld. advised that for the small flows concerning
this application, temperature is the factor that most matters.
4.3 Valves
To assure that valves from different manufacturers can be interchanged without restraint, port
configurations have been standardized at international level. The precise control valves dimensions
are listed in standards DIN 24340 and ISO4401. The abbreviation NG6 figures in both types of
58
electrical valves used in this equipment. It concerns to the connection size and match with the
diameter of the connecting ports in mm. Other sizes available are 10, 16 and 25 being 6 the smallest
available for this purpose and ideal for low flows.
Figure 54: Valve characteristic directional valve BIERI WV700 4/3-U NG6
Hydraulic loss for the valve at the maximum flow, by graphical analysis in figure 54, is p = 3bar
2
This value was measured for a viscosity of 32mm /s. To convert the pressure loss value for the fluid
used on the system, it will be used the equation proposed by [Nachi-Fukikoschi]:
(4.1)
From 4.1 and substituting with Tellus T 68 (at 50º) and manufacturers testing oil viscosities:
3.3 bar
The change isn’t significant, anyway the pressure loss on the valve is:
p = 3.3bar
59
4.4 Pipes and hoses
According to speed limits generally accepted for hydraulic system, and mentioned by [Götz W.,
1998], 6 m/s is a reference value on pressure lines over 100 bar pressures. Thus by the simple
relation:
(4.2)
V - Flow speed at internal pipes [m/s]
Q - Flow rate [l/min]
D - Diameter of the pipes [mm]
For the given velocity, and flow defined above (3.6 l/m) it will result a pipe diameter of:
3.58 mm
For a standard pipe diameter of 3mm, therefore:
D = 3mm
V = 8.49m/s
For the hoses, an important factor referred by [Enerpac, 2007] should be taken into account
and to reduce hydraulic losses across long lengths: a greater internal diameter should be considered
for the hoses. It will be chosen a model commonly used in post-tensioning applications (pressures up
to 700bar with 6.4mm internal diameter) and it will be admitted 20m hose length. The hoses model is
from Enerpac: model HA 6.4.
60
4.5 Hydraulic Losses
For a lifting at nominal pressure, which will be the more severe system condition, calculated
hydraulic losses across the pressure line are:
Fraction
Valve
P-A way
Pipes
1.5 m
Elbows
3 x 90°
Hoses
20 m
Oil Column
15 m
Table 5: Circuit fractions for hydraulic losses calculation
Flowing characteristics are considered to be laminar for Reynolds numbers under 2000 [Matos Almas
et al.,1998]. Re number is given by:
(4.3)
Using equation 4.3 for the pipe flow:
The flow is laminar, and under this condition the friction factor is given by:
(4.4)
61
Thus, from equation 4.4 the friction factor :
Having 20 meters of hose line, the friction factor must also be calculated in this fraction.
From
equation 4.2 flow speed across hoses will be:
With
= 6.4mm from chapter 4.4. Using equation 4.3 the Reynolds number for hose line is:
= 266 < 2000
Being laminar, with equation 4.4 the friction factor on the hoses is:
The hydraulic loses across pipes and hoses will produce pressure differential obtained by:
(4.5)
For pipes:
= 1.5m (admissible value for internal pipes length from pump to quick couplers)
= 8.49 m/s
= 3mm
= 0.11
= 877 Kg/m
3
From equation 4.5:
62
For hoses:
= 20 m
= 1.87 m/s
= 6.4mm
= 0.24
= 877 Kg/m
3
From equation 4.5:
bar
Other hydraulic losses are the PA path on the directional valve calculated with equation 4.1:
= 3.3 Bar
Additional hydraulic losses, are the local ones. Thus, with equation 4.6 :
(4.6)
- Local hydraulic loss factor
Three 90 elbows were considered in the internal pipes, having each one a
to [Esposito, 1980].
For 3 elbows the sum result is:
63
factor of 0.75 according
Figure 55: Typical configuration for an assembly
Considering a more severe case of 15m vertical oil column from the pump till the actuator:
=
=
[bar]
= 1.3 bar
The total hydraulic loss
is:
+
+
4.6 Pump and Motor
The characteristic curves for the pumps are below represented on Figure 56 for different available
sizes. The ideal size should be:
(4.7)
- pump volumetric displacement [l/rot]
Q - flow rate [l/min]
N - rotational velocity [rpm]
- pump volumetric efficiency (approx. 1 for radial piston pumps)
64
Thus for a nominal radial velocity of 1450 rpm and 3.6 l/min of flow rate:
2.48 cm3
3
For this manufacturer, the closest pump volumetric displacement is 2.59 cm . Represented on the
graph by the curve referenced as [2.50-700] .
Figure 56: Characteristic curves for fixed displacement radial pump type PR4 (Bosch-Rexroth)
With a total efficiency
of 0.9 the power required by the motor [Cudell, 2008] is given by:
(4.8)
P - power [kW]
Q - flow rate [l/min]
p - pressure [bar]
For a pressure of 615 bar for the jack plus all hydraulic losses, the total pressure at the most
demanding case will be 648.91 bar and the required power from the motor:
P = 4.32 kW
65
The values for the pump and motor are valid for each independent branch of the circuit having
its own pump. If a multi-outlet pump is considered, the flow of each single pump should be the nominal
of each outlet. However to determine the motor required power, the manufacturers of this pumps
should be consulted to provide an accurate calculation method.
66
5 Simulation results
5.1 Model
To construct virtual prototypes before building a real system, definitely brings other options
and possibilities on an equipment conception. For hydraulic models, interactive tools and specific
software can save hours of fluid flow paths imagination and recreation, as well as the need of experts
consulting, which can be saved for more specific questions and practicalities.
Even if a real prototype should be built, prototype software previous allows streamlining
processes and to identify correct and incorrect options. These tests can however go from small essays
of functionality to much more precise levels.
Optimization and details related with controls, losses, velocities, etc., can be evaluated in an
interesting and inexpensive way. Clearly it is not cheap because is not possible for an enterprise to
acquire licenses for specific software to develop a product that it is not its main project activity.
However the partnership between enterprises and universities is a fine opportunity, having this last
ones experience with a large range of technologies.
The simulation model built using the Simhydraulics toolbox from Simulink is presented in
Appendix E. In the bottom part of the model; are represent eight processes where each one, is an
hydraulic branch with the respective actuator connected.
Represented are also the input signals which simulate the force exerted on each actuator, as
well as valve signals: SP (pressure line signal) and SR (return line signal). SP and SR designate the
directional valve solenoid commands. As an output, the Z positions are sent to a control unit where the
values are constantly compared and according to the movement type defined by the control, the unit
sends the orders to open or close the valves.
Figure 57 presents the design implemented for hydraulic components:
67
Figure 57: Hydraulic components simulation detail
The cylinder is motion powered by the pump, which is connected to a radial velocity source.
The directional valve receives signals from a double-acting solenoid which is inputted by the SP and
SR signals. The cylinder is linked with sensors that measure the position/velocity of the movements,
as well as the force. The cylinder edge is also linked to the loading value that represents the weight of
the structure acting on the cylinder.
The control of the pressure is done in another subsystem, which is represented in Figure 58:
Figure 58: Pressure control subsystem
68
The proportional pressure relief valve was here simulated using a valve orifice controlled by a
proportional solenoid, which applies the force for the exact opening in order to maintain the desired
pressure differential. The considered pressure value is introduced as a signal.
5.2 Simulation features
The simulation of the model follows some rules in order to optimize the hydraulic process.
Parameters should be well selected to detect whether the system is real-time capable or too heavy.
Firstly it matters defining real-time capable for the purpose of this system study.
A system is real-time capable if the simulation runs before the simulated period counted in real
time. This factor in determinant to validate if the system can be controlled directly using MATLAB
connected to real hardware components, consequently saving the cost of real implementation. More
important to this study, is the indication given by the simulation time. Even if it is not supposed to run
an implementation like this in real-time, the simulation time when running the system is an important
factor to measure the computational burden of the simulation system.
Real-time simulation of physical systems with various domains (mechanical, electrical,
hydraulic, etc.) requires a combination of several factors: model complexity, solver choice and solver
settings. What is being outlined here is the arrangement of all this parameters and the comparison
between the desktop simulation and the real-time one.
The sampling period of a simulation can be influenced by several parameters. Primarily it is
influenced by the computer hardware. The compilation of a simulation model is quite affected in case
of a complex system and a difference is obtained when running it from a notebook with Intel Core 2
(1.66GHz) to a Intel Dual Core T4400 (2.2Ghz), however the processor and memory by themselves
are not sufficient to drastically reduce the simulation time, nonetheless they are very important. It is
also possible to save a lot of time in rebuilding and compiling models with a high performance
computer.
Projecting consistently the model, avoiding unnecessary connections, loops or other
unnecessary items is the starting point for having success when running a model. Many occurrences
at the same time on a simulation, such as electric motors, hydraulic pumps, switches, valves, input
signals, physical sensors and other physical events can render challenging the panorama of the
conditions during the simulation. The choice of the correct solver is very important and simulation
execution time per step should be consistent. .
Advances in solver technology were made in the last years according to [Miller et al., 2009].
Features added to simulation tools, like fixed-cost algorithms and local solvers added to Simscape
produce other possibilities to simulate even complex models like hydraulic pipelines in real time. In this
study the local solver is very important and can reduce considerably the computational burden for the
simulation by putting a local solver in each hydraulic subsystem. This option is quite powerful on
physical networks and when executing it, it is possible to use a fixed-step solver only on the stiff
portions of the model. This reduces the computations per time step, making the model faster.
69
Figure 59: Example of the use of local solvers on portions of the model with other explicit solver. (Miller et al.,2009)
The system got slower with the increment of the number of cylinders, and a local solver in
each cylinder proved to be a feasible solution, by delivering superior results. Options between realtime and desktop simulation should start here to be made.
Typically systems like hydro mechanical are simulated with variable-step solvers, which take
smaller steps to accurately capture events occurring during simulation. Although changing the step
size is impossible for real-time simulation and a fixed-step solver must be used. If the challenge here
is to validate the system simulation, the fixed-step solver option can be neglected especially if the
advantages are not crucial.
Efforts to make the real-time simulation can be hard. Both model and fixed-step solvers must
be configured in order to accurately capture system dynamics without changing the step size. The
advantages of using real-time simulation, according to [Miller et al., 2009] are:

Ability to test conditions that would damage equipment or personnel;

Ability to test systems where no prototypes exist;

Reduced costs in the later phases of development;

Ability to test 24 hours a day, 7 days a week;
This simulation type is still used in the final setup or in products that have human interaction
on the simulation loop (ex: flight simulators).
In this equipment’s case, both costs and test conditions can be simulated using a desktop
simulation because there are conditions that can be simulated in the program, such as all the devices,
components and sensors.
Also referred by [Erkkinen et al., 2007] is that all necessary prototype hardware can be
connected and tested before real hardware implementation. For that objective this model-based
design supports a wide variety of C/C++ code generation applications, which include rapid simulation
prototype, or hardware-in-the-loop testing.
70
The compilation of Simscape models for standalone simulations allows accelerating analyses
with parameter studies and Monte Carlo simulations. As referred by [Miller et al., 2009], the possibility
of generating code in a system as this one, which includes several n-subsystems, is very important
since these subsystems can be configured to evaluate the physical networks in parallel.
The simulation of the model is validated accordingly to a variable-step solver which proved to
be a consistent solver in achieving tolerances and features expected from the system.
The procedure to optimize the simulation results is a combination of solver and step size as
referred, and also the number of solver iterations that determine model size and fidelity. To help on the
optimization procedure, the procedure of Figure 60 was followed. It was created to reduce simulation
times in models, specifically containing hydraulic, electrical, mechanical pneumatic and thermal
components including linear and nonlinear components. The following figure illustrates the flowchart
that describes the procedure:
Figure 60: Flowchart depicting the process of move from desktop simulation to real-time simulation (Miller et al., 2009)
71
To conclude, this procedure was important in reducing simulation time and validating the
model. Although to transform it real-time capable was not the aim. Due to this procedure it was also
possible to identify and correct problems related with excessive/defective stiffness, overruns, and
inadequate time steps.
5.3 Results
A parallel lifting movement for eight cylinders was simulated for a .001 m tolerance and for
unbalanced conditions, where eight equal sized actuators, were connected to equal pumps, with
difference in the velocity of 50 rpm from one to other (1500,1450,1400 ...1100). The positions across
time are represented in figure 62:
Figure 61: Position-time for a parallel lifting with 8 cylinders position-time for a given time range
Red area zoom in:
Figure 62: Zoom for the assigned interval
72
Corresponding valve switching:
Figure 63: Corresponding valve switching for the 8 directional valves
The order presented is descending in terms of cylinder lifting difficulty level, where the valve
commanding the last cylinder is always open (= 1). In opposition the first is the valve that remains
closed (= 0) for the most of the time.
These switching frequencies are noticed as high and result from abnormal conditions between
cylinders, to see how further can go the tolerances using only valve switching. However these valve
types, despite having vey high switching times (see Appendix D), should not overpass 1 switching per
second for periods over than one hour, For similar situations it is recommended to reduce the oil flow
throw the valves by reducing the pump rotational speed. This option allows to preserve the valves lifetime and leaves the possibility to increase the tolerances if necessary.
Figure 64 shows a position-time graph for a rotation movement type of 8 cylinders:
73
Figure 64: Position-time for 8 cylinder rotation
The pressure control is ruled by the proportional pressure relief valve, which holds the
pressure at the pre selected level. FFigure 65 represents the variation of pressure on a system branch
controlled by the proportional valve.
Figure 65: Pressure level maintained at 100 bar
74
For the interval (14:16) seconds, the actuator finds a resistive force, which makes the pressure
rises as expected, however the proportional valve set up for 100 bar (1e7 Pa) does not allow that the
value reaches nominal system values, being the level maintained at 100 bar with high precision and
without instability.
The results presented on this chapter are positive and show that the controlling method
chosen is valid. For the directional valves, the valve switching is possible both in translations and in
rotations due to reduced system flow rates. More important is that tolerances defined are possible to
be reached. For the pressure controls, the closed-loop linear control with a proportional gain
accomplishes with success its objective.
The tolerances achieved for the position are quite acceptable for civil engineering words,
despite the possibility of still reducing it by reducing oil flows from the pump if necessary.
75
6 Economical viability
An evaluation of the economic value of the equipment is essential. To decide whether to invest
on new acquisitions will depend of its financial potential. In this specific case, even if the equipment is
tailored to the business core of an enterprise of the construction sector, the figures must justify the
deal. There are interesting advantages in developing and building devices or work tools within a
company, instead of acquiring them outside on the market. For the present case in study, advantages
are:

Acquisition costs reduced around 50% - 75%

No need to rent equipment from the competition and consequently not subjected to
availability.

Possibility of improving self skills and knowhow in solving future challenges as well in
creating new market opportunities.

Modernizes, improves and enhances company’s image on the marketplace.
Despite the referred items denote positive earnings, they should be evidenced from the
financial viewpoint. One of the first important testimonies of this affirmation is that equipments as this
should be budgeted mainly in hydraulic components. A reckless analysis of the global constitution of
the system leads easily to budget overruns. The following Table 6Table 7 and 7 illustrate the power
and control components cost weight, comparing with hydraulics. They also illustrate the price
difference for both selected constructive approaches: Single body structure with multi-outlet pump of
section 2.5.2 and respectively the separated unit’s equipment as illustrated in 2.5.3.
76
MULTI-OUTLET PUMP UNIT
Description
Price/Unit
Multi-outlet radial piston pump Bieri MRK8
Quantity
Total
1.553,00
1
1.553,00
4/3 Directional valve solenoid oper. Bieri WV700
465,00
8
3.720,00
Proportional pressure relief valve Bieri PDV700
334,00
8
2.672,00
Pressure relief valve Bieri DV700
73,00
8
584,00
Sub plate mounting port NG6
131,00
8
1.048,00
Pressure transducer Wika S-10
303,00
8
2.424,00
Pressure gauge Wika 0-1000/63
75,90
8
607,20
Couplings Enerpac C604
Shut-off valve Bieri 700 Bar
Filter Sofima
Displacement sensor Micro-Epsilon P60-SR
Controller Delta Systems RMC151E
28,85
16
461,60
167,00
8
1.336,00
16,50
1
16,50
388,00
8
3.104,00
5.400,00
1
5.400,00
Electric motor Siemens
250,00
1
250,00
Variable speed drive General Electric VAT200
696,00
1
696,00
1.559,00
1
1.559,00
1.000,00
1
1.000,00
Accessories (refilling and protection caps, oil level
indicators etc.)
500,00
1
500,00
Oil Shell Tellus T 68 209L
502,00
1
502,00
Distribution board (includes:)
Hydraulics
Power and
Control
230V supply line (power)
24V supply line (control)
Enclosure structure IP66
Emergency stop and signalization
Circuit breakers, relays, switch and wires
Thermostat, grid and air filtration
Structure and installation
TOTAL
Table 6: Multi-outlet pump system quotation
77
27.433,30
Additional
euros
8 UNITS SINGLE PUMP EQUIPED
Description
Price/Unit
Radial piston pump Rexroth PR4 2,50
Quantity
Total
1.010,00
8
8.080,00
4/3 Directional valve solenoid oper. Bieri WV700
465,00
8
3.720,00
Proportional pressure relief valve Bieri PDV700
334,00
8
2.672,00
Pressure relief valve Bieri DV700
73,00
8
584,00
Sub plate mounting port NG6
131,00
8
1.048,00
Pressure transducer Wika S-10
303,00
8
2.424,00
Pressure gauge Wika 0-1000/63
75,90
8
607,20
Couplings Enerpac C604
Shut-off valve Bieri 700 Bar
Filter Sofima
Displacement sensor Micro-Epsilon P60-SR
Controller Delta Systems RMC151E
28,85
16
461,60
167,00
8
1.336,00
16,50
8
132,00
388,00
8
3.104,00
5.400,00
1
5.400,00
Electric motor Siemens
200,00
8
1.600,00
Variable speed drive General Electric VAT200
696,00
8
5.568,00
1.259,00
8
10.072,00
Structure and installation
1.500,00
1
1.500,00
Accessories (refilling and protection caps, oil level
indicators etc.)
1.000,00
1
1.000,00
502,00
1
502,00
Distribution board (includes:)
Hydraulics
Power and
Control
230V supply line (power)
24V supply line (control)
Enclosure structure IP66
Emergency stop and signalization
Circuit breakers, relays, switch and wires
Thermostat, grid and air filtration
Oil Shell Tellus T 68 209L
TOTAL
Table 7: 8 units single pump equipped quotation
78
49.810,80
Additional
euros
From the tables analysis, the price difference is denotative, being the second equipment cost
dependence more related with power and control components, that almost with hydraulic hardware.
Both constructive differences are preferred by different technical opinions and working
techniques approach, for which some inquired professionals underline pros and cons. This chapter
aim is to rate their price and underline differences in order to allow investors to decide, based not only
on systems technical features. Even if costs change abruptly from one system type to another, both
can be valid if the overall profit of the equipments application justifies their acquisition.
For this, a financial forecast is made for a 5 year time interval based on revenues and
expenditures. The following forecasts represent simple possibilities and mainly aim to identify the
magnitude order of the equipments costs.
Once quotations obtained, are from several national and international suppliers, negotiations
for long-term payment conditions with sellers become impossible. Considering so a bank loan for
preventing an excessive financial effort:
INITIAL INVESTMENT
Type
Loan
Own Resources
Origin
80%
20%
100%
LOAN
Duration
Interest Rate
3 years
8% / year
Table 8: Investment characteristics
The cash flow for the first equipment is:
CASH FLOW 1
TOTAL
Year 0
Year 1
Year 2
Year 3
Year 4
Operating Activities
Equiment acquisition costs
Certification consulting*
-27433
-900
0
0
0
0
0
0
0
0
-27433
-900
Specialized manpower
Maintenance
-10000
-1000
-10000
-1000
-10000
-1000
-10000
-1000
-10000
-1000
-50000
-5000
Financing Activities
Income
Amortization
Resources
10000
-1756
-14539
50000
-1159
-4255
50000
-579
-8639
50000
0
0
50000
0
0
210000
-3494
-27433
21946
0
0
0
0
-23682
33586
29782
39000
39000
21946
117.687 €
Investment Activities
Loan
TOTAL
Table 9: Multi-outlet pump system cash flow
79
And respectively for the second:
CASH FLOW 2
TOTAL
Year 0
Year 1
Year 2
Year 3
Year 4
Operational Activities
Equiment acquisition costs
Certification consulting*
Specialized manpower
Maintenance
-49810
-900
-10000
-1000
0
0
-10000
-1000
0
0
-10000
-1000
0
0
-10000
-1000
0
0
-10000
-1000
-49810
-900
-50000
-5000
Finantial Activities
Income
Amortization
Resources
10000
-3188
-26399
50000
-2104
-7726
50000
-1052
-15685
50000
0
0
50000
0
0
210000
-6344
-49810
39848
-41449
0
29170
0
22263
0
39000
0
39000
39848
87.984 €
Investment Activities
Loan
TOTAL
Table 10: 8 units single pump equipped cash flow
* Includes consulting for: Security approval (directiva máquinas DL 50/2005) and CE Marking)
Note: Incomings admitted represent annual values for the liquid profits taken on activities carried out with the
equipment.
On a five-year basis, both systems seem to be valid investments, even the second equipment
that has larger acquisition costs. The fact that both are quality investments along with the works
(quantity and type) to be developed will decide whether to choose on between the options presented.
80
7 Conclusions and future work
This work comes from a potential opportunity in a business area that arose with the need to
develop more modern systems. The presented device is not only able to do more complex heavy
lifting and structural displacement operations, but also to raise standards of quality in these types of
operations.
Hydraulic synchronism is an old problem of hydraulics that has an enormous ensemble of
solutions developed across the years, for which several types of equipments were designed. Even for
synchronism in high pressure EH systems (over 200 bar), there is a huge number of hypothesis and
tactics that can be explored, specifically on valve control (discrete/continuous actuation); valve type
(proportional/servo); pressure control (meter in/out, pressure-relief valve, etc); flow control (pump
displacement, flow control valve, metering cylinder) and so on. There is not a consensus for the best
solution and even for heavy lifting applications there are various types of solutions, starting from
pressure levels where SV’s are available, till 200-350 bar, to others till 600-700 bar. Reasons to
choose ones and others have to do with the following aspects:
1. Economical reasons, translated into what would be the frequency of equipment´s use: (a)
would it be for a single application; (b) for a construction sector company; (c) for a specialized
technical works company; (d) produced by hydraulic products sellers with profit objective on the
equipments selling and not on the operations; (e) for instance if the equipment is destined to be
rented; or a combination of more than one reason.
2.
Actuators fleet.
3.
The forces and movement types involved.
After a selection of control and components for system’s core, two major types of equipments
were adopted as possibilities:
The first with a single construction body, being significantly cheaper, less complex on
maintenance procedures and less exposed in the working environment. However it needs hoses with
greater lengths and is heavier to place in the working field position, using a small crane or similar
mandatorily.
Referring to the second equipment, its physical construction is split in smaller units which can
be easily positioned with a pair of assembled rollers by a single person. Being more portable, also
81
allows fitting constrained spaces closer to actuators working positions. This also reduces the required
hose quantity. As drawbacks this second approach has higher price, mainly due to an increased
number of motors, pumps and distribution boards. In addition the divided structure requires more
maintenance and superior testing procedures.
The present study aims to identify advantages and disadvantages of both equipments,
underlining facts to select the most convenient option.
7.1 Simulations
The software applied to hydraulics proved to be a powerful tool for testing hypothesis during
this work development. Simulations performed on Automation Studio revealed to be efficient in
predicting system behaviors, simplifying kinematic analysis and identifying constraints, while
SimHydraulics toolbox from Simulink made possible to build up the systems control due to its large
possibilities and hydraulic blocks available.
Simulink is also an interesting option for developing rapid prototyping or to build a real model
allowing to directly connect it to the computer in order to make essays. Physical components
simulated virtually and linked to the control algorithm allow to enhance its development, leaving
beyond the option to directly connect real hydraulic hardware to Simulink by data acquisition boards.
For this objective it would be recommended to connect a complete modular axis from the 8 axis
overall, including pressure and displacement sensors.
7.2 Prospect improvements and adaptation possibilities to Innovation
The present work aims to guarantee that the equipment’s design assure defined requirements.
It also allows taking advantage of the acquired knowhow, to be used in future and similar applications.
Regarding the future improvements of the evaluated equipment, variable speed drives for electric
motors were already included on the quotations. The objective is allowing further developments on the
control algorithm for flow speed, leaving the possibility to render the system more versatile and more
energy efficient.
Concerning adaptation possibilities, the system features and metering capabilities of this
equipment could easily be adapted to improve quality standards on other civil engineering technical
operations. Taking the example of prestressed concrete applications, it is possible to measure applied
forces and displacements in the steel strands on-time and with accuracy.
For future and innovative applications, the system can be dimensioned for applications in
processes such as:
 Vertical construction methods
 Dynamic formworks
 Modular construction systems
82
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Appendix
86
A - Hydraulic circuit diagram
87
B - Electrical diagram
88
C - Oil temperature dependence of viscosity
89
D – Valves technical data
Bieri WV700
Bieri PDV700
90
E – Simulink Model
91
92