Power Drive Architectures for Industrial Hydraulic Axes: Energy-Efficiency-Based Comparative Analysis

Abstract

In hydraulic systems, energy dissipation can be significant. The pressure losses that can occur in the hydraulic circuit, which are influenced by the adopted drive architecture, result in power consumption that is often significantly higher than that required by the mechanical system. This paper presents a comparative study of the energy efficiency of five common drive architectures in industrial hydraulic axes. The analysis is applied to a variable speed and force hydraulic blanking press, a fairly common industrial system, e.g., in the manufacture of semi-finished brass products. Standard, regenerative, high–low, variable-displacement pumps and variable speed drive configurations for a fixed-displacement pump were analyzed and compared. In each case, an appropriate and optimized sizing of the different components of the system was performed, and then the energy consumption was estimated for a load cycle common to all the considered cases. The results show that the choice of the power generation architecture of the hydraulic system has a very significant impact on the energy efficiency and consequently on the operating costs and the carbon footprint. The performed quantification of the potential energy efficiency of the considered drive architectures can be very useful in helping to make energy-conscious decisions.

Full text

Citation: Tiboni, M. Power Drive
Architectures for Industrial
Hydraulic Axes: Energy-
Efficiency-Based Comparative
Analysis. Appl. Sci. 2023,13, 10066.
https://doi.org/10.3390/
app131810066
Academic Editors: Nicu¸sor Baroiu,
Cˇ
atˇ
alin Dumitrescu and Paweł
´
Sliwi´nski
Received: 17 August 2023
Revised: 4 September 2023
Accepted: 5 September 2023
Published: 6 September 2023
Copyright: © 2023 by the author.
Licensee MDPI, Basel, Switzerland.
This article is an open access article
distributed under the terms and
conditions of the Creative Commons
Attribution (CC BY) license (https://
creativecommons.org/licenses/by/
4.0/).
applied
sciences
Article
Power Drive Architectures for Industrial Hydraulic Axes:
Energy-Efficiency-Based Comparative Analysis
Monica Tiboni
Department of Mechanical and Industrial Engineering, University of Brescia, Via Branze 38, 25123 Brescia, Italy;
monica.tiboni@unibs.it
Abstract:
In hydraulic systems, energy dissipation can be significant. The pressure losses that can
occur in the hydraulic circuit, which are influenced by the adopted drive architecture, result in power
consumption that is often significantly higher than that required by the mechanical system. This paper
presents a comparative study of the energy efficiency of five common drive architectures in industrial
hydraulic axes. The analysis is applied to a variable speed and force hydraulic blanking press, a
fairly common industrial system, e.g., in the manufacture of semi-finished brass products. Standard,
regenerative, high–low, variable-displacement pumps and variable speed drive configurations for a
fixed-displacement pump were analyzed and compared. In each case, an appropriate and optimized
sizing of the different components of the system was performed, and then the energy consumption
was estimated for a load cycle common to all the considered cases. The results show that the choice
of the power generation architecture of the hydraulic system has a very significant impact on the
energy efficiency and consequently on the operating costs and the carbon footprint. The performed
quantification of the potential energy efficiency of the considered drive architectures can be very
useful in helping to make energy-conscious decisions.
Keywords:
energy efficiency; energy savings; industrial hydraulics; power drive architecture; CO
2
emissions reduction
1. Introduction
The global energy market has been experiencing a considerable growth in demand
in recent years. This has been accompanied by a significant increase in sales and delivery
prices and it is very likely that this trend will continue in the coming years [
1
]. The growing
demand for energy is most evident in the industrial sector [2], where processes of various
kinds are very often carried out with high or very high energy requirements [
3
]. The
higher the power requirements of a process, the more important the optimal selection
of the drive technology of the machine used to perform the task. Typically, hydraulic
technology is characterized by a high power density, far exceeding that of pneumatic or
electric drives [
4
]. Therefore, hydraulic technology is the preferred choice when the energy
input in an industrial plant is high. In these circumstances, the economic savings that
result from pursuing energy efficiency goals is a business priority. Energy dissipation in
hydraulic systems can be significant [
5
]. The distributed pressure losses that may occur in
the piping network or in the components of the hydraulic circuit determine upstream a
pressure higher than the one actually required by the actuators, resulting in an increase in
power consumed [6,7].
In addition, the architecture chosen for power generation has a very large impact on
the energy efficiency of the system. For example, in many systems a significant amount
of oil-hydraulic energy is dissipated between the inlet and outlet ports of the pressure
relief valve when dissipative speed adjustments of the actuators are achieved [
8
]. This
peculiarity of hydraulic technology is clearly a contributing factor to some degree of
inherent inefficiency, but it also determines a significant potential for improvement that
Appl. Sci. 2023,13, 10066. https://doi.org/10.3390/app131810066 https://www.mdpi.com/journal/applsci

Appl. Sci. 2023,13, 10066 2 of 33
can be achieved through innovation. Therefore, hydraulic technology intended for the
actuation and control of industrial machinery, is fertile ground for the development of
design solutions characterized by an ever-increasing level of energy efficiency.
Control systems with low energy consumption and architectures that optimize overall
efficiency have been studied and developed for hydraulic applications in industrial or
mobile fields [
4
,
5
,
9
–
17
]. Numerous circuits for recovering energy for hydraulic devices
have been studied. The most studied energy recovery circuits are based on accumulators,
gas cylinders, gravity energy, flywheels, generator supercapacitors, or generator–battery
circuits. The hydraulic energy regeneration system (ERS) with accumulator is ideal for
machines that need to be started and stopped frequently, according to Tianliang et al. [
18
],
who discussed various forms of ERS used in hydraulic construction equipment.
Ho and Le presented, in [
19
], a high efficiency hydraulic system with high efficiency
that saves about 20% energy compared to systems without energy recovery. A high-
pressure hydraulic accumulator, a relief valve, an AC servomotor, a fixed-displacement
hydraulic pump, a hydraulic cylinder, seven directional control valves, and a tank are the
main components of the system. The method uses a check valve to boost the pump and a
low-pressure accumulator as the tank. The direction of flow to the pump is controlled by
two check valves, and the the speed of the hydraulic cylinder is controlled by the speed
regulation of the motor. The high-pressure accumulator boosts the pump and is used
to store recovery energy. Due to the switching algorithm of the valves, this system can
be operated in different configurations, so that it can be flexibly adapted to the desired
operating requirements, e.g., high speed, or high torque, but also high effectiveness of
energy recovery from the load.
Niu et al., in [
11
], propose an innovative multifunctional energy-saving electro-
hydraulic servo system. A servomotor actuates the pump with a pressure control, two
proportional valves, and four switches, the settings of which allow the selection of three dif-
ferent control modes: independent single-valve control (SI), separate meter in and separate
meter out control (SMISMO), and parallel dual valve control (DP). The supply pressure is
controlled by a disturbance observer and the supply flow is controlled by a gray predictor.
Experiments have shown that supply pressure and flow losses can be reduced, resulting in
energy savings.
In [
20
], Schmidt and Hansen present the concept of connecting multi-cylinder/motor
drives as a variable-speed drive network. Electrically and hydraulically interconnected
variable-speed displacement units allow the complete elimination of throttle elements, and
the sharing of auxiliary functions and fluid reservoirs, as well as hydraulic and electri-
cal power.
Xu et al., presented, in [
4
], an analytical technique based on a mathematical energy
dissipation model of hydraulic components to calculate the energy dissipation of the system
by building a high-precision simulation model. They applied the method to a 10,000 kN
fine blanking press.
To determine the cause of the low energy efficiency in large and medium-sized hy-
draulic presses, Zhao et al. [
21
] suggested an analytical method to quantify the energy flow
in the system. Starting from the basic formula of energy consumption, they determined a
limited number of unknown coefficients whose values can be estimated by experiments.
Based on a study of the energy flow characteristics of the hydraulic system,
Li et al. [15]
propose an energy-saving strategy by balancing the load of the press operations. The
method is to share the motor pumps of the drive system at different times with a unit con-
sisting of two hydraulic presses to minimize the energy loss during unloading operations.
These two presses are also coupled, and during certain operations, the excess energy from
one press can be used as input energy for the other to increase the efficiency of the drive
system. In addition, the potential energy can also be used directly. To supply energy to both
hydraulic presses, the drive systems of the two presses are combined into one combined
drive system (CDS). Presses 1 and 2 receive energy from each motor-pump of the combined
drive system at specific times, with each motor-pump being controlled separately.

Appl. Sci. 2023,13, 10066 3 of 33
A direct-drive pump-controlled hydraulic (DCP) system, characterized by a closed
circuit type, speed-controlled electric servomotor, and fixed-displacement pump, was
studied by Koitto et al. [
14
] for a stationary industrial material-handling application that
sequentially lifts and lowers a fixed mass. Energy savings ranged from 53 to 87% compared
to a conventional valve-controlled approach. Although the actuator tended to vibrate
when reaching the desired position and the system pressures varied greatly as the cylinder
moved, the dynamics of the system did not quite meet the stated requirements.
Some recent studies are concerned with the theoretical or experimental determination
of the efficiency of hydraulic components or systems. Zhang et al. [
22
] propose a new
method for calculating the volumetric efficiency and hydraulic efficiency of centrifugal
pumps based on the principle of energy balance. Using a low-specific-speed centrifugal
pump pumping media with different viscosities, two efficiencies are calculated at the points
with the best efficiency and compared with those of two existing methods.
Beni´c et al. [23]
performed a detailed analysis of energy efficiency between a direct-drive hydraulic system
(DDH) and a proportional electro-hydraulic system. The analysis was based on experi-
mental results. They concluded that the efficiency of the DDH system with a fully loaded
cylinder was 28%, while the efficiency of the proportional electro-hydraulic system was
4%.
Rana et al. [9]
studied the effect of the bypass valve in a conventional lifting system
with a vertical servohydraulic cylinder and experimentally compared the improvement in
energy efficiency. The study found that the energy efficiency is higher at low pressure and
gradually decreases when the operating pressure becomes high. The energy efficiency is
inversely proportional to the operating pressure for the larger load. When the weight of the
cylinder load is doubled, the energy efficiency also increases. Maximum energy efficiency
can be achieved by keeping the pressure as low as possible to overcome gravity, and the
required speed can be achieved by maintaining the maximum load on the cylinder.
Studies have also been conducted in the industrial sector to increase the energy
efficiency of hydraulic injection molding machines (HIMMs). Clamping force control (FC)
requires a high pressure but a low flow rate, so only a small amount of energy is needed.
Widely used hydraulic valve-controlled cylinder systems, in which the maximum supply
pressure is set with a relief valve and constant-displacement pumps, offer good servo
control performance but poor energy efficiency. To increase energy efficiency, energy-
saving control methods based on an electro-hydraulic variable-displacement pump system
(EHVDPS), such as load-sensing control (LSC) and constant supply pressure control (CSPC)
are used. From the studies of Chiang et al. [
24
], it emerged that the energy consumptions
of (FC + LSC) and (FC + CSP) are 17.9 and 67.4% of the input energy of the EHVDPS
method, respectively.
Table 1summarizes solutions presented in the literature for energy saving in industrial
hydraulic applications.
Table 1. Energy savings obtained in the literature with different methods.
Ref. Method Field Pump Motor Energy Saving
[11]
Control of an electro-hydraulic servo
system, with two proportional directional
valves, and four switch valves
Industrial Fixed
displacement Servo motor Not declared
[4]Low–high pressure system
with accumulator
Industrial (fine
blanking press)
Fixed
displacement Fixed speed 50%
[13] Multiple motor pumps—system control Industrial
(hydraulic press)
Variable
displacement Fixed speed 26.97%
[14] Direct-driven hydraulic system Industrial
(hydraulic press)
Fixed
displacement Servo motor 53–87%
[15] Two presses combination Industrial (hydraulic
forming presses)
Variable
displacement Fixed speed 36%

Appl. Sci. 2023,13, 10066 4 of 33
The high power-to-weight ratio and increasing miniaturization of components make
hydraulics increasingly suitable for applications in robotics: in exoskeletons [
25
,
26
], in
humanoid, bipedal, or quadrupedal robots [
27
,
28
], and in industrial and collaborative
robots [
29
]. Energy efficiency is fundamental in these systems because their autonomy
must be maximized. There are numerous studies in the literature that address the energy
efficiency of these systems and propose innovative implementation solutions [30–32].
System power consumption can also be an indicator of operating conditions. In
particular, application-independent variations in power consumption can be associated
with hydraulic system malfunctions [
33
–
38
]. Condition monitoring techniques based on
artificial intelligence algorithms can be used to prevent system malfunction.
The following is an overview of the reviews available in the literature on hydraulic
systems and their energy efficiency.
Some reviews on the energy efficiency of hydraulic systems energy can be found in
the literature. Mahato et Goshal [
6
] provide an overview of energy-saving approaches for
hydraulic drive systems. They divide the energy-saving approaches into four categories:
hybridization, control algorithms, energy recovery, and energy loss reduction, and for each
category they provide a detailed literature review.
Quan et al. [
39
] investigate direct pump control technology, that eliminates the throttle
losses in the main power line and achieves high energy efficiency. Their work focuses on
the system structure, control system, and derived energy recovery system.
A classification and overview of pump-controlled differential cylinder drives is given
by Ketelsen et al. [
40
], starting with systems based on variable-displacement hydraulic
pumps and vented reservoirs to more modern hydraulic system architectures based on
variable-speed electric motors and sealed reservoirs. They classify the architectures dis-
cussed in the literature into some basic classes, and highlight the advantages and disadvan-
tages of each class.
Shen et al. [
41
] provide an overview of common pressure rail-based hydraulic systems,
that can be widely used for construction equipment. Energy savings can be achieved
through the ability to maintain a high power density of the hydraulic system, elimination
of throttling losses, and energy recovery from actuators.
Xu et al. [
42
] review the development of electro-hydraulic control valves, industry
4.0 oriented, and their related technologies. The three topics covered in this review paper
are: condition sensing with sensors or indirect sensing; control approaches with digital
controllers and innovative valves; and online condition monitoring through data interaction
and problem detection.
Schmidt and Hansen [
20
] introduce the idea of electro-hydraulic variable-speed drive
networks, with interconnected multi cylinder/motor drives that do not have throttle control.
They discuss design considerations, integration of hydraulic accumulators, compactness of
the drive, and control concepts.
An analysis of the literature shows that various solutions have been proposed in terms
of plant architecture or control systems to increase the energy efficiency of hydraulic sys-
tems. However, a timely comparison of the energy consumption of the different solutions
has not been investigated so far.
Analytical modeling of all phenomena involved in the operation of an oil-hydraulic
drive architecture to estimate overall efficiency is complex [
43
]. Distributed and concen-
trated pressure drops, temperature-induced variations in fluid viscosity, the influence of
fluid viscosity, pressure, and speed on pump volumetric hydromechanical efficiency, valve
switching delay, active and reactive electric current components absorbed by the power
supply (influenced by speed), and dynamic friction between the piston and cylinder body
are some of the many factors that affect the operation and efficiency of a hydraulic system.
However, an estimate of energy efficiency and the impact of energy consumption on
operating costs for different system architectures is information of considerable value for
the design of these systems. In order to fill this gap, in the present work a comparative
study is carried out in terms of the energy efficiency of different architectural solutions for

Appl. Sci. 2023,13, 10066 5 of 33
the power section of a variable speed and force hydraulic blanking press, a very common
industrial process, for example, in the production of semi-finished brass products. In a
variable speed and force process, the flow rate and pressure of a fluid change over time,
under these operating conditions, the different technical solutions that can be used for the
power section have significantly different energy efficiency levels.
The energy losses that distinguish one solution from another were examined in the
analysis, while the contributions common to all solutions were ignored.
The analysis is applied to a specific industrial process, namely, the pressing of semi-
finished brass products. Different hydraulic architectures are presented, all equally capable
of providing the required performance, but characterized by increasingly higher energy
efficiency. The different configurations analyzed, starting from a traditional hydraulic archi-
tecture with low energy efficiency, are characterized by the introduction of the regenerative
technique and variations in the unit for converting electrical energy into hydraulic energy
carried by the pressurized fluid.
Five configurations, whose architecture is described below and summarized in Table 2,
have been analyzed and compared.
1.
Standard configuration: traditional architecture consisting of a round-case induction
electric motor, a fixed-displacement vane pump, a pressure relief valve, and a standard
directional control valve (DCV).
2.
Regenerative configuration: standard configuration with DCV with regeneration
A—hybrid.
3.
High–low configuration: regenerative A—hybrid configuration in which the single
fixed-displacement vane pump is replaced by two pumps of the same type, one
with high flow rates and low maximum pressures, one with low flow rates and high
maximum pressures.
4.
VDP (variable-displacement pump) configuration: regenerative A—hybrid configu-
ration in which the single fixed-displacement vane pump is replaced by a variable-
displacement axial piston pump.
5.
DCP (drive-controlled pump) configuration: regenerative A—hybrid configuration in
which rotational speed of the asynchronous motor is continuously adjustable through
an AC frequency converter.
Table 2.
Considered and compared power drive architectures for a press for semi-finished brass products.
Configuration Electric Drive Pump Directional Valve
Standard
Round-case induction electric motor
Fixed-displacement vane pump Standard DCV
Regenerative
Round-case induction electric motor
Fixed-displacement vane pump DCV with regeneration A—hybrid
High–low
Round-case induction electric motor
Two fixed-displacement
vane pumps DCV with regeneration A—hybrid
VDP
Round-case induction electric motor
Variable-displacement axial
piston pump DCV with regeneration A—hybrid
DCP AC frequency converter +
round-case induction electric motor Fixed-displacement vane pump DCV with regeneration A—hybrid
For a quantitative comparison of the energy efficiency of the studied configurations,
it is necessary to perform an optimized dimensioning of the main components for each
of them.
For configurations 1 to 4, the sizing can be achieved by neglecting the angular ac-
celeration or deceleration ramps that characterize the transition from one speed range to
another, since for architectures that use an asynchronous motor fed directly from the grid,
the inertial torques resulting from the acceleration or deceleration ramps can be neglected in
a first approximation. On the other hand, the contribution of inertial torques becomes more
relevant when the motor is fed by a frequency converter, since in this mode of operation the

Appl. Sci. 2023,13, 10066 6 of 33
motor is subjected to speed variations of very large amplitude, even in very short periods,
which leads to non-negligible inertia torques. For architecture 5, both an approximate
sizing, in which the inertial contribution was neglected, and a more accurate sizing, in
which the inertial contribution, was also considered, were performed.
The remainder of the paper is organized as follows. Section 2is devoted to the
description of the variable speed and force hydraulic blanking press, the selection of the
hydraulic cylinder and fluid, the theoretical analysis of a valve-controlled hydraulic system,
and the description of all architectural configurations considered. In Section 3, the results
are presented, in terms of estimated absorbed power (active and reactive), mean active
electric power on the working cycle, and hourly cost of energy absorbed by the electricity
grid, for each configuration considered. Furthermore, the results of the experimental
measurements in the plant where the architecture with DCP was implemented are reported.
Section 4provides an analysis of the results. Section 5summarizes the main findings of
the work.
2. Materials and Methods: Architectural Configurations for the Case Study
This section is devoted to the description of the characteristics of the solutions consid-
ered, applied to a case study with reference to a real application. In detail, the following are
presented: the structure and parameters of the process considered as a case study, the tech-
nical requirements, the elements common to all the cases analyzed, i.e., actuator and fluid,
the theoretical analysis applied to estimate physical quantities required to the dimensions
of the system, the procedures applied to select and dimension the circuit components, and,
finally, all the final configurations of the architectures considered, as reported in Table 2.
2.1. The Hydraulic System
In this section, the main characteristics of the hydraulic system under study are
described. First, the structure of the system is described and the technical specifications
considered for performing the quantitative analysis are identified, then the technical data
of the circuit elements common to all the configurations considered are given. Appendix A
contains a graphical representation of the main components of each configuration considered.
2.1.1. System Structure and Technical Requirements
The main technical requirements for the actuator during operation are two: the mini-
mum thrust force that the rod must exert during the blanking of the semi-finished product
and the maximum duration of the entire working cycle of the press. The system considered
as a case study for pressing semi-finished brass products consists of a hydraulic press with
a system for feeding/unloading parts that includes a rotary table with four stations, a robot
manipulator, and a conveyor belt, as shown in Figure 1.
The typical blanking work cycle is divided into two macro-phases. The first includes
the rod quick descent, the blanking, the expulsion of the piece, and the rode quick rise. In
this first macro-phase the cylinder is actively involved in the work cycle. The second, on
the other hand, involves only the rotation of the table for loading the parts to be machined.
In this case the cylinder rod does not perform any movement and remains in a waiting state.
The duration of this second macro-phase is usually set by the end user of the machine and is
0.3 s in the process considered as a case study. It is, therefore, obvious that the constraint on
the maximum duration of the blanking cycle determines a limit on the maximum duration
of the first macro-phase, in which the cylinder is in motion. In addition to the specifications
for the minimum force that must be generated during operation and for the maximum
duration of the working cycle, a third limitation is established. This defines the limit mass
of the tool that can be anchored to the free end of the actuator rod. The numerical values of
the above parameters are summarized in Table 3.

Appl. Sci. 2023,13, 10066 7 of 33
HYDRAULIC CYLINDER
ROBOTIC MANIPULATOR
CONVEYOR BELT
4 STATIONS ROTATING TABLE
Figure 1.
Pressing system considered as a case study and main constituent elements. The vertical
hydraulic cylinder of the press is located in the dashed square.
Table 3. Technical requirements for the hydraulic system.
Parameter Value Unit
Minimum blanking force 294,300 (N)
Rotation phase duration for the rotary table 0.3 (s)
Maximum duration of the entire processing cycle 1.4 (s)
Maximum duration of operation for the actuator 1.1 (s)
Maximum mass of the tooling anchored to the rod 80 (kg)
The entire process consists of five individual phases: quick descent (QD), blanking (B),
expulsion (E), quick rise (QR), and table rotation (TR). During the blanking phase, pressure
and velocity have a rather complex trend, that should be determined either experimentally
or by simulating the process. For the purposes of this work, i.e., the functional comparison
between the different system architectures in terms of energy efficiency, the impact of
the variation in these two quantities during the blanking phase on the overall energy
consumption values is limited. The discussion is, therefore, made under the simplified
assumption that the speed of the load is uniform, and corresponds to the average value
during the cutting operation and that the oil pressure is constant and corresponds to the
maximum value reached during blanking.
First of all, it is necessary to define the main project specifications, which include
the translation speed of the cylinder rod, the stroke, and the delivery pressure for each
working phase. The choice of values for the translation speed is arbitrary and limited only
by mechanical or other constraints. The stroke is defined as a function of the thickness of
the semi-finished product to be sheared in the shearing phase and of the characteristics of
the system in the other phases. For the feeding pressure, a pressure level is chosen that
is considered sufficient to maintain the control of the valves, to compensate for the load
losses distributed in the tubes, and to overcome the resistance of any load at the free end of

Appl. Sci. 2023,13, 10066 8 of 33
the cylinder rod. In Table 4, the values of these parameters are given for each phase of the
blanking cycle. Figure 2shows the cyclogram of the working phases, obtained considering
a periodic cyclic load with negligible dynamics on the actuator. For each phase, the gray
rectangle represents its duration, while the white rectangle indicates the total duration of
the previous phases.
Table 4.
Rod stroke, piston translation speed, and maximum pressure on the delivery in the processing
cycle phases.
Parameter Unit Quick Descent—QD Blanking—B Expulsion—E Quick Rise—QR Table Rotation—TR
Stroke (s) (mm) 40 20 20 80 -
Velocity (v)
(mm/s)
250 67 67 250 -
Max. pressure (P) (bar) 50 200 70 70 10
1.08
0.76
0.46
0.16
1.38
0.3
0.32
0.3
0.3
0.16
0 0.2 0.4 0.6 0.8 1 1.2 1.4
WORK CYCLE
TABLE ROTATION
QUICK RISE
EXPULSION
BLANKING
QUICK DESCENT
Time [s]
Time accumulated at the previous stage Single phase duration
Figure 2. Cyclogram with periodic cyclic loading at negligible dynamics.
The dynamic effects at the shaft of the electric motor are much more relevant. In
architectures where the three-phase induction motor is directly connected to the electrical
network, the motor maintains a speed that is close to the rated speed. The only deviations
from the rated speed, excluding the start and stop phases, are caused by the phenomenon
of motion irregularity due to the periodicity of the load. The amplitude of the speed
deviation is very small, so the inertia effects on the crankshaft can be neglected. For the
DCP architecture, where an inverter is used, given the total cycle time of 1.4 s, it is obvious
that the acceleration and angular deceleration ramps cause moments of inertia on the motor
shaft that must be taken into account. In summary, the inertial aspects of the crankshaft can
be neglected in the design of standard, high–low, and VDP architectures. However, they
must be taken into account when dimensioning the architecture with DCP.
2.1.2. Hydraulic Cylinder and Fluid
Figure 3shows the flowchart of the cylinder sizing process. A family of cylinders is
selected that is characterized by a maximum pressure. The maximum working pressure is
chosen; therefore, considering the most critical phase and the respective maximum force
that can be developed, the minimum bore of the actuator is calculated and a normalized
value is chosen. Starting from the cylinder mounting, the running factor (f) is selected, the
basic length (Lb) is calculated, and the rod diameter in the catalog is selected based on Lb
and the maximum thrust.

Appl. Sci. 2023,13, 10066 9 of 33
➢Maximum pressure for the chosen cylinder
type: 250 bar
➢Maximum selected pressure (pmax): 200 bar
➢Minimum required push force : 294400 N
➢More critical phase: THRUST
DESIGN DATA
Selection of a standard bore diameter higher than
minimum bore size from cylinder catalogue table
of Thrust Force as a function of bore diameter
and pressure (f, p)
▪f = 0.5 run factor (Front flange with feet and
fixed load with rigid guide)
▪s = 300 mm (stroke)
▪Lb = f * s =150 mm (base length)
▪Rod diameter selected from catalogue diagram
of rod as a function of base length and thrust
Cylinder bore sizing
( fc = 140 mm)
Cylinder rod sizing
(fr = 63 mm)
Figure 3. Flowchart of cylinder sizing.
Table 5lists the designations and associated technical characteristics of the selected
cylinder and fluid, which are common to all configurations.
Table 5. Fluid and cylinder: designations and technical data.
Circuit Element Parameter Value
Fluid
Code Castrol serie HYSPIN ZZ 46
Fluid type Mineral oil
Viscosity ISO VG 46
Cylinder
Code Parker 140MF3MMAXRN23M300M1133AOAO
Type Double acting
Bore 140 (mm)
Rod diameter 100 (mm)
Stroke 300 (mm)
Fixing Front flange with feet and fixed load with rigid guide
Area Ratio 1.96
2.1.3. Theoretical Analysis of a Valve-Controlled Circuit
System sizing requires actual forces and velocities and consequently pressures and flow
rates. This section contains the theoretical analysis required to calculate these quantities.
The simplest control circuit that can be used for a linear actuator combines the differ-
ential cylinder with an area ratio of 2:1 with a standard 1:1:1:1 directional control valve.
Let us adopt the nomenclature from Figure 4: the flow rate through the rear chamber and
flow rate through the front chamber of the actuator are defined as
QA
and
QB
, respectively.
pA
and
pB
denote the relative pressures occurring in the rear and front chambers of the
cylinder, respectively.
pP
is the pressure occurring at the outlet of the pumping unit. The
fluid pressure in the reservoir and pressure tank is denoted by
pT
. Since the oil pressure
in the tank is generally equal to atmospheric pressure, the pressure
pT
is assumed to be
zero as a value relative to atmospheric pressure. For a differential cylinder, the area ratio
α

Appl. Sci. 2023,13, 10066 10 of 33
is defined as
α=AP
AT
and the flow rates for the two chambers in the exit stroke have the
values given in Equation (1).
For the 4-way directional valve used (P = pressure, T = discharge, A = first port;
B = second port
), 1:1:1:1 means that for the paths AT:PA:PB:BT the pressure drop is equal
to
∆pn
when 100% of the nominal flow
∆Qn
passes through it. The value 1 stands for 100%,
another value means another percentage, e.g., 0.5 for 50%.
In the following, the pressure drops at the DCV, the pressures generated in the rear
and front chambers of the cylinder, the thrust and traction forces that can be exerted by
the rod, and the maximum speed it can reach during the outward and return phases are
determined at the theoretical level. In the analysis, both the pressure and the translational
velocities of the rod are assumed to be constants with respect to the ejection and the return.
To distinguish the parameters related to the outward stroke from those related to the inward
stroke, the superscripts “o” (outward) and “i” (inward) are used.
Figure 4. Symbols for theoretical analysis of the system.
Qo
A=vAPQo
B=Qo
A
α(1)
The relationship between the flow rate traversing a DCV path and the pressure drop
occurring on the path itself is expressed by Equation (2).
Qx=Qns∆px
∆pn(2)
Consequently, the pressure drops on the PA and BT paths can be expressed using
Equations (3) and (4).
∆pPA =pP−po
A=∆pnQo
A
Qn2
=∆pnvA p
Qn2
(3)
∆pBT =po
B−pT=po
B=∆pnQo
B
Qn2
=∆pnvAP
αQn2
(4)

Appl. Sci. 2023,13, 10066 11 of 33
For a cylinder with area ratio 2, the ratio between the pressure drops between ports P
and A and B and T is
∆pPA
∆pBT =α2=
4. From Equations (3) and (4), the pressures in the rear
and front chambers of the actuator can be derived (Equations (5) and (6)).
po
A=pP−∆pnvA p
Qn2
(5)
po
B=∆pnvAP
αQn2
(6)
The thrust force created by the rod is expressed by Equation (7).
Fo=po
AAP−po
B
AP
α=pPAP−AP∆pnvAP
Qn2
−AP
α
∆pnvAP
αQn2
(7)
For inward motion, a similar analysis procedure is used as for outward motion, with
flow rates, pressure drops on the valve paths, and pressures in the cylinder chambers
determined as in Equations (8)–(12).
Qi
B=vAT=vAP
αQi
A=αQi
B=vAP(8)
∆pPB =pP−pi
B=∆pn Qi
B
Qn!2
=∆pnvA p
αQn2
(9)
∆pAT =pi
A−pT=pi
A=∆pn Qi
A
Qn!2
=∆pnvAP
Qn2
(10)
pi
B=pP−∆pnvA p
αQn2
(11)
pi
A=∆pnvAP
Qn2
(12)
In the inward phase, the ratio between the pressure drops between ports P and B and
A and T is ∆pBT
∆pAT =1
α2=1
4.
During the retraction phase, the stem exerts a pulling action. This force is expressed
by Equation (13).
Fi=pi
BAT−pi
AAP=pi
B
AP
α−pi
AAP=pPAP
α−AP
α
∆pnvAP
αQn2
−AP∆pnvAP
Qn2
(13)
With
v
and
pp
being equal, the thrust force is greater than the pulling force. Both in
the exit and retraction phases, the maximum force value is obtained for
v=
0. In this case,
the ratio between Foand Fiis equal to α.
Fo
Fiv=0
=α(14)
If
QP
is the effective flow rate delivered by the pumping group, the maximum trans-
lation speed of the rod during the output stroke is the minimum value between the one
obtained from Equation (1) setting
Qo
A=QP
and the one that derives from setting
Fo
equal
to zero.
vo
max =min




QP
AP
,v|Fo=0=v
u
u
t
pP
∆pnA2
p
Q2
n1+1
α3(15)

Appl. Sci. 2023,13, 10066 12 of 33
Similarly, the maximum translational velocity of the rod during the return stroke is
the minimum value between the value obtained from Equation (8) when
Qi
B=QP
is set
and the value obtained when the force Fiis set equal to zero.
vi
max =min






QP
AT
,v|Fi=0=v
u
u
u
t
pP
α
∆pnA2
p
Q2
n1+1
α3(16)
Regeneration usually occurs during rod extension and consists in directing the oil flow,
which exits the front chamber of the actuator as a result of the translational motion of the
piston, into the sleeve of the rear chamber cylinder. This results in a sensitive energy saving,
which is why architectures that use DCVs with regeneration are called economizer circuits.
In the proposed study, the point where the fluid is re-supplied is downstream of the
PA path, i.e., towards the user, therefore, the regeneration is of type A, and A—hybrid.
The configuration with regeneration A—hybrid integrates in the valve body the
functions of a check valve in the line BA and an on–off valve in the connection between
point A and the tank. These two functions allow it to go from a configuration with internal
regeneration—A to a standard configuration during the discharge stroke. In this way,
it is possible to take advantage of both the benefits of internal regeneration—A (energy
efficiency) and the benefits of the standard configuration (high forces), if necessary.
The analysis of flow rates, pressures, velocities, and forces for the outward stroke in
the case of regeneration A—internal follows Equations (17)–(21), in a similar manner as for
the standard configuration.
∆pPA =pP−po
A=∆pnQP
Qn2
po
A=pP−∆pnQP
Qn2
(17)
∆pBA =po
B−po
A=∆pnQo
B
Qn2
=∆pnvAP
αQn2
po
B=pP−∆pnQP
Qn2
+∆pnvAP
αQn2
(18)
Fo=pPAP1−1
α−AP∆pn1−1
αQP
Qn2
−AP
α
∆pnvAP
αQn2
(19)
Fo=pPAP1−1
α−AP∆pnvAP
αQn2
(20)
v|Fo=0=v
u
u
u
u
t
pP1−1
α
∆pnA2
p
Q2
nα2
(21)
With regeneration, the maximum thrust that can be exerted when the rod is extended
is only half that which can be generated with a standard circuit. In terms of design, the
DCV used for regeneration—A hybrid integrates in its upper part the functions of the check
valve and the on–off valve. The connections between the DCV and the branches of the
hydraulic circuit, located in the lower part of the valve body, remain unchanged compared
to those of a standard DCV. This allows modernizing an existing hydraulic actuator and
increasing its energy efficiency by introducing regeneration technology, simply by replacing
the body of a conventional directional control valve.
2.2. Standard Configuration
The main elements of the standard architecture are the actuator, a 3-phase 4-pole
asynchronous motor with a round case connected directly to the grid, a fixed-displacement
vane pump with a maximum pressure at the pressure port of 320 bar, a 1:1:1:1 DCV standard

Appl. Sci. 2023,13, 10066 13 of 33
directional control valve, and a pilot-operated proportional pressure relief valve. The power
source part of the standard configuration is shown in Figure 5.
Downstream
circuit
3-Phase
Asinchronous
Motor
Proportional
control pressure
relief valve
Fixed displacement
Vane Pump
P
Figure 5. Standard architecture power source scheme.
A flowchart of the system sizing procedure for the standard configuration is shown in
Figure 6. For the other configurations, a similar sizing approach is followed.
Pump sizing is based on pump displacement, maximum speed Nmax−pum p, and max-
imum pressure
pmax
, while motor sizing is based on rated power
Wn
. After selecting
the motor and pump types (in this case, a 4-pole 3-phase induction motor and a rotary
vane pump), an initial design attempt is performed for the electric motor and pump. The
theoretical flow rate of the
Qp−th
pump must be higher than the flow rate
Q
required by
the load in the most critical condition. The rated power of the electric motor must be higher
than the average power required in the cycle by its load, in this case the pump. In the
case of an asynchronous drive directly connected to the mains, the motor speed is very
close to the rated speed, and in the case of direct connection between the motor and the
pump, the speeds of the motor and the pump are the same. Knowing the speed
Npum p
and
the theoretical flow rate of the pump
Qp-th
, it is possible to calculate the minimum pump
volume
Vpump
and select the pump (considering the minimum flow rate, maximum speed,
and maximum pressure limits). After determining the effective flow rate of the pump
Qp−e f f
(by subtracting the leakage losses from the theoretical flow rate), it is necessary to
verify that it is sufficient under all operating conditions. If this is verified, it is possible to
proceed, otherwise another pump size must be selected. Then, the directional valve and the
maximum pressure valve are sized: the first one based on the effective flow
Qp−e f f
and the
pressure losses between the paths depending on the opening degree; the second one based
on the estimated control pressure depending on the characteristic curve (opening pressure
popening
, flow
Qf low
). Finally, the torque to be supplied by the motor must be verified. The
nominal torque of the motor
Mn−motor
must be greater than the root mean square of the
torque
MRMS−load
required in the cycle. In estimating
MRMS−last
many factors must be
considered, e.g., flow rates, leakage, pressures, the regulated pressure of the relief valve,
and the percentage of opening of the DCV. Therefore, the estimation of pressures and flow
rates (according to the theoretical approach in Section 2.1.3) is required for the application
of the method.

Appl. Sci. 2023,13, 10066 14 of 33
➢Qp-th > max Q required by load (in the worst
case)
➢Motor Wn> Wmean required to the pump on the
cycle
➢Nn-motor = Nn-pump (4 poles asynchronous motor)
➢Vpump > Qp-th/Nn-pump
ELECTRIC MOTOR and PUMP SIZING
Estimate of effective pump flow rate – Qp_eff
➢Qp-eff = Qp-th - leakage losses in the pump
Pump sizing based on:
Vpump, Nmax-pump, pmax
Motor sizing based
on: Wn
Qp_eff > Qload
For every phase of the
work cycle
NOT VERIFIED
YES
➢Cursor choice based on Qp-eff and on
nominal pressure losses
DIRECTIONAL VALVE SIZING
➢Estimation of the regulated opening pressure
pREG from curves correlating popening - Qdischarged
PROPORTIONAL PRESSURE RELIEF VALVE SIZING
Mn-motor > MRMS-load
NOT VERIFIED
YES
COMPONENTS CHOICE and
SYSTEM SIZING CONCLUDED
To estimate MRMS_load :
➢pressures estimation based on force
balancing on the piston
➢discharge pressure
➢flow rates
➢flow leaks
➢pREG of the pressure relief valve
➢DCV valve opening percentage
Figure 6. Standard architecture: system sizing flowchart.
Table 6shows the main technical data of the selected components; Figure A1 shows
the main components; and Figure 7shows the active electric power absorbed in the cycle
for the standard configuration.

Appl. Sci. 2023,13, 10066 15 of 33
Table 6. Designations and technical data of the main elements of the standard architecture.
Circuit Element Parameter Value
Electric motor
Code Parker Mod. MR4P05500
Type 4 poles—round case
Nominal power 55 (kW)
Nominal speed 1480 (RPM)
Nominal torque 354.87 (Nm)
Efficiency 0.935
Pump
Code Parker Mod. T7E 050
Type Fixed-displacement vane pump
Nominal displacement 158.5 (mL/min)
Speed range 600–2200 (RPM)
Max continuous pressure 210 (bar)
Directional Valve
Code Parker Mod. D41FEE02FC1NB70
Type Pilot-operated proportional DCV 1:1:1:1
Nominal flow rate 200 (L/min)
Nominal pressure drop 5 (bar)
Pressure Relief Valve
Code Parker Mod. R4V03535P0PM10VA1
Type Pilot-operated proportional pressure relief valve
Nominal flow rate 250 (L/min)
Maximum pressure 350 (bar)
STANDARD CONFIGURATION
+
+
Mod. MR4P05500
Round case 4 poles Asynchronous
Motor
Mod. T7E 050
Fixed Displacement Vane Pump
Mod. D41FEE02FC1NB70
Pilot operated Proportional DCV
1:1:1:1
Wel.a.mean
Wel.r.mean
40.48 kW
29.06 kW
(71.79 % of Wel.a.mean ) → Rate: 0.0323 €/kvarh
Γ
En.el.a
Γ
En.el.r
4.86 €/hour
+
0.94 €/hour
(40.48 kW ∙ 0.12 €/kWh)
(29.06 kW ∙ 0.0323 €/kvarh)
17254.2 €/year
+
3334.1 €/year
(40.48 kW ∙ 0.12 €/kWh ∙ 16 h/day ∙ 222 days/year)
(29.06 kW ∙ 0.0323 €/kvarh ∙ 16 h/day ∙ 222 days/year)
(PENALTY)
Γ
En.el.TOT
5.80 €/hour
20588.3 €/year
0
10
20
30
40
50
60
70
80
90
100
0 0.2 0.4 0.6 0.8 1 1.2 1.4
Power [kW]
Time [s]
Potenza elettrica attiva
Termine / Inizio della Fase
Potenza elettrica attiva media
Active Electric Power
Phase Stop/Start
Mean Active Electric Power
Figure 7.
Standard configuration. Active electrical power absorbed over the work cycle: instanta-
neous value and average value.
2.3. Regenerative Configuration
The regenerative configuration differs from the standard configuration by introducing
the regeneration function in the fast descent phase, i.e., when a higher flow rate is required.
The DCV valve used allows a new path for the oil flow leaving the chamber on the actuator
stem side, the BA path, and this internal path allows regeneration. Therefore, a valve
with regeneration A—hybrid is chosen, which allows the introduction of regeneration—A
(Figure 8) during the phase in which the spindle requires a high translation speed at the
output, and to go to the standard configuration when the spindle must realize a considerable
thrust. Table 7shows the characteristics of the components used in this configuration. The
regenerative function allows a significant reduction in the flow rate that must be delivered
by the pump during the rapid descent phase: with regeneration, the value is 117 L/min,
compared to the 230.8 L/min required without regeneration. Consequently, the motor
and electric and pump sizes are significantly reduced, as can be seen from the comparison
between data in Tables 6and 7.

Appl. Sci. 2023,13, 10066 16 of 33
PT
AB
Qreg
Qpump
Figure 8. Active regeneration for high-speed movement.
Table 7shows the most significant technical data of the chosen components, Figure A2
the main components, and Figure 9the active electrical power absorbed in the cycle for the
regenerative A—hybrid configuration.
Table 7. Designations and technical data of the main elements of the regenerative architecture.
Circuit Element Parameter Value
Electric motor
Code Parker Mod. MR4P03000
Type 4 poles—round case
Nominal power 30 (kW)
Nominal speed 1460 (RPM)
Nominal torque 196.22 (Nm)
Efficiency 0.923
Pump
Code Parker Mod. Mod. T7D B24
Type Fixed-displacement vane pump
Nominal displacement 81.1 (mL/min)
Speed range 600–3000 (RPM)
Max continuous pressure 250 (bar)
Directional Valve
Code Parker Mod. D41FEZ32FC1NB70
Type Pilot-operated proportional DCV with regeneration—A hybrid
Nominal flow rate 200 (L/min)
Nominal pressure drop 5 (bar)
Pressure Relief Valve
Code Parker Mod. R4V03535P0PM10VA1
Type Pilot-operated proportional pressure relief valve
Nominal flow rate 250 (L/min)
Maximum pressure 350 (bar)

Appl. Sci. 2023,13, 10066 17 of 33
REGENERATIVE CONFIGURATION
+
+
Mod. MR4P03000
Round case 4 poles Asynchronous
Motor
Mod. T7D B24
Fixed Displacement Vane Pump
Mod. D41FEZ32FC1NB70
Pilot operated Proportional DDCV
with Regeneration - A Hybrid
Wel.a.mean
Wel.r.mean
20.05 kW
15.60 kW
(77.81 % of Wel.a.mean ) → Rate: 0.0421 €/kvarh
Γ
En.el.a
Γ
En.el.r
2.41 €/hour
+
0.66 €/hour
(20.05 kW ∙ 0.12 €/kWh)
(15.60 kW ∙ 0.0421 €/kvarh)
8546.1 €/anno
+
2332.8 €/year
(20.05 kW ∙ 0.12 €/kWh ∙ 16 h/day ∙ 222 days/year)
(15.60 kW ∙ 0.0421 €/kvarh ∙ 16 h/day ∙ 222 days/year)
(PENALTY)
Γ
En.el.TOT
3.07 €/hour
10878.9 €/year
0
10
20
30
40
50
60
70
80
90
100
0 0.2 0.4 0.6 0.8 1 1.2 1.4
Power [kW]
Time [s]
Potenza elettrica attiva
Termine / Inizio della Fase
Potenza elettrica attiva media
Active Electric Power
Phase Stop/Start
Mean Active Electric Power
Figure 9.
Regenerative A—hybrid configuration. Active electrical power absorbed over the work
cycle: instantaneous value and average value.
2.4. High–Low Configuration
The high–low architecture allows time-varying generation of hydraulic power. The
generation of hydraulic power is divided into two distinct phases, such as the initial stroke
of the rod, and meets the variable needs of the user rather than considering only the
heaviest operating condition. The first phase of power generation takes place during the
rapid descent, and the second during the release. Thus, the generation of hydraulic power
is divided into a period in which a high oil flow rate is required at relatively low pressure,
and a period in which the fluid flow rate is reduced, accompanied by a pressure increase.
The designation high–low is due to this change in operating conditions to which the power
generation must adapt. This architecture is very suitable for the punching operation carried
out by the press. In Figure 10, the circuit diagram of the power generation part for the high–
low architecture is reproduced. The system consists of two fixed-displacement hydraulic
vane pumps. The pumps are driven by a single round-case asynchronous motor connected
directly to the power supply system and have opposite hydraulic characteristics. One pump
has a relatively low maximum operating pressure limit and high displacement capacity
and is referred to as LPHD (low pressure high displacement). The other pump features a
high maximum operating pressure and relatively low displacement and is referred to as
HPLD (high pressure low displacement). The effectiveness and proper functioning of this
architecture depends largely on the choice of maximum pressure limits that can be achieved
when pumping the two pumps and on the displacement of the pumps. The high–low
configuration also differs from the standard architecture by the presence of two additional
valves: an unloading valve and a check valve. Table 8shows the characteristics of the
components chosen with the high–low architecture for the case study.
From the analysis of the operation of the system, it can be deduced that the shear
phase is the only phase in which the flow reaching the actuator is supplied only by the
HPLD pump, while in the other phases both pumps contribute to the supply of the cylinder.
The values for the opening and closing pressure of the drain valve must be 97.2 bar and
70 bar, respectively. The high–low configuration allows a significant reduction in the size
of the electric motor (as seen in a comparison between Tables 6–8), and of the maximum
pressure range, since the maximum pressure values are no longer required simultaneously.
Table 8shows the most significant technical data of the chosen components; Figure A3
the main components; and Figure 11 the active electrical power absorbed in the cycle for
the high–low configuration.

Appl. Sci. 2023,13, 10066 18 of 33
Unloading valve
with venting
Non-return
valve
3-Phase
Induction
Motor
Proportional
control pressure
relief valve
Fixed displacement
Vane Pumps
LPHDHPLD
P
Figure 10. Circuit of the actuation architecture for the high–low configuration.
Table 8. Designations and technical data of the main elements of the high-low architecture.
Circuit Element Parameter Value
Electric motor
Code Parker Mod. MR4P01850
Type 4 poles—round case
Nominal power 18.5 (kW)
Nominal speed 1460 (RPM)
Nominal torque 121.32 (Nm)
Efficiency 0.914
HPLD pump
Code Parker Mod. T7B B14
Type Fixed-displacement vane pump
Nominal displacement 45 (mL/min)
Speed range 600–3000 (RPM)
Max continuous pressure 275 (bar)
LPHD Pump
Code Parker Mod. T7ASW B40
Type Fixed-displacement vane pump
Nominal displacement 40 (mL/min)
Speed range 600–3000 (RPM)
Max continuous pressure 240 (bar)
Directional Valve
Code Parker Mod. D41FEZ32FC1NB70
Type Pilot-operated proportional DCV with regeneration—A hybrid
Nominal flow rate 200 (L/min)
Nominal pressure drop 5 (bar)
Pressure Relief Valve
Code Parker Mod. R4V03535P0PM10VA1
Type Pilot-operated proportional pressure relief valve
Nominal flow rate 250 (L/min)
Maximum pressure 350 (bar)
Unloading Valve
Code Parker Mod. R4U03—533
Type Pilot-operated unloading valve with venting
Nominal flow rate 150 (L/min)
Nominal open–close pressure drop 28% of the setting pressure

Appl. Sci. 2023,13, 10066 19 of 33
HIGH - LOW CONFIGURATION
+
+
+
Mod. MR4P01850
Round case 4 poles
Asynchronous Motor
Mod. T7B B14
Fixed Displacement
Vane Pump (HPLD)
Mod. T7ASW B40
Fixed Displacement
Vane Pump (LPHD)
Mod.
D41FEZ32FC1NB70
Pilot operated
Proportional DCV with
Regeneration - A Hybrid
Wel.a.mean
Wel.r.mean
15.77 kW
11.72 kW
(74.32 % of Wel.a.mean ) → Rate: 0.0323 €/kvarh
Γ
En.el.a
Γ
En.el.r
1.89 €/hour
+
0.38 €/hour
(15.77 kW ∙ 0.12 €/kWh)
(11.72 kW ∙ 0.0323 €/kvarh)
6721.8 €/year
+
1344.6 €/year
(15.77 kW ∙ 0.12 €/kWh ∙ 16 h/day ∙ 222 days/year)
(11.72 kW ∙ 0.0323 €/kvarh ∙ 16 h/day ∙ 222 days/year)
(PENALTY)
Γ
En.el.TOT
2.27 €/hour
8066.4 €/year
0
10
20
30
40
50
60
70
80
90
100
0 0.2 0.4 0.6 0.8 1 1.2 1.4
Power [kW]
Time [s]
Potenza elettrica attiva
Termine / Inizio della Fase
Potenza elettrica attiva media
Active Electric Power
Phase Stop/Start
Mean Active Electric Power
Figure 11.
High–low configuration. Active electrical power absorbed over the work cycle: instanta-
neous value and average value.
2.5. VDP Configuration
The VDP configuration has the same elements as the regenerative configuration, except
for the pump, which is of the variable-displacement type. In the case study, an axial piston
pump is considered. There are several control configurations for a variable-displacement
piston pump that allow the theoretical flow rate of the pump to be adjusted during opera-
tion. Among the available configurations, the one most suitable for direct comparison with
the configurations considered in this analysis must be selected. The discussion will refer to
a proportional-displacement closed-loop pressure control (Figure 12).
Downstream
circuit
3-Phase
Asinchronous
Motor
Fixed displacement
Vane Pump
Displacement
adjustment
cursor
Flow Meter
Pressure
Gauge
Pressure
compensation
cursor
Figure 12. VDP configuration: proportional-displacement control with closed-loop pressure control.
The main components of the VDP configuration are listed in Table 9, with their
main features.
Figure A4 shows the principal components, and Figure 13 shows the active electric
power absorbed in the cycle for the VDP configuration.

Appl. Sci. 2023,13, 10066 20 of 33
Table 9. Designations and technical data of the main elements of the VDP architecture.
Circuit Element Parameter Value
Electric motor
Code Parker Mod. MR4P01850
Type 4 poles—round case
Nominal power 18.5 (kW)
Nominal speed 1460 (RPM)
Nominal torque 121.32 (Nm)
Efficiency 0.914
Pump
Code Parker Mod. PV 092
Type Axial piston pump—variable displacement
Max displacement 92 (mL/min)
Speed range 400–2500 (RPM)
Nominal pressure 350 (bar)
Directional Valve
Code Parker Mod. D41FEZ32FC1NB70
Type Pilot-operated proportional DCV with regeneration—A hybrid
Nominal flow rate 200 (L/min)
Nominal pressure drop 5 (bar)
Pressure Relief Valve
Code Parker Mod. R4V03535P0PM10VA1
Type Pilot-operated proportional pressure relief valve
Nominal flow rate 250 (L/min)
Maximum pressure 350 (bar)
VDP CONFIGURATION
+
+
Mod. MR4P01850
Round case 4 poles Asynchronous
Motor
Mod. PV 092
Axial Piston Pump - variable
displacement
Mod. D41FEZ32FC1NB70
Pilot operated Proportional DCV
with Regeneration - A Hybrid
Wel.a.mean
Wel.r.mean
16.67 kW
12.71 kW
(76.24 % of Wel.a.mean ) → Rate: 0.0421 €/kvarh
Γ
En.el.a
Γ
En.el.r
2.00 €/hour
+
0.54 €/hour
(16.67 kW ∙ 0.12 €/kWh)
(12.71 kW ∙ 0.0421 €/kvarh)
7105.4 €/year
+
1900.7 €/year
(16.67 kW ∙ 0.12 €/kWh ∙ 16 h/day ∙ 222 days/year)
(12.71 kW ∙ 0.0421 €/kvarh ∙ 16 h/day ∙ 222 days/year)
(PENALTY)
Γ
En.el.TOT
2.54 €/hour
9006.1 €/year
0
10
20
30
40
50
60
70
80
90
100
0 0.2 0.4 0.6 0.8 1 1.2 1.4
Power [kW]
Time [s]
Potenza elettrica attiva
Termine / Inizio della Fase
Potenza elettrica attiva media
Active Electric Power
Phase Stop/Start
Mean Active Electric Power
Figure 13.
VDP configuration. Active electrical power absorbed over the work cycle: instantaneous
value and average value.
2.6. DCP Configuration
In this configuration, a square-case three-phase asynchronous motor is connected to
the power supply network through an AC frequency converter (inverter), and the motor
speed is controlled using vector control, while the pump is a fixed-displacement pump
(
Figure 14
). As with the VDP, it is also possible with the DCP configuration to adjust the
flow rate delivered by the pump to the needs of the drive. In this case, however, the flow
rate is fixed while the speed is varied. While in the previous configurations the speed of
the electric motor varies little around the nominal speed (there are only oscillations caused
by the phenomenon of motion irregularity associated with cyclic and periodic loading), in
this architecture the asynchronous motor varies its own speed considerably. The transition
from one speed to another occurs with acceleration or deceleration ramps that generate
non-negligible inertia torques and consequently the profile of the flow rate required by
the actuator and the desired pressure at the pump outlet are very different from what
characterizes the architectures considered previously. Therefore, two sizing procedures are
carried out: in a first phase the inertial actions are neglected, then they are included in the

Appl. Sci. 2023,13, 10066 21 of 33
calculations, to verify how much their real contribution is and whether they have a very
large impact on the consumption.
DC/AC
Converter
(Inverter)
P_eff
N_eff
P_ref
Q_ref
Figure 14. Direct pump control architecture: power source scheme.
2.6.1. DCP Configuration—Static Sizing
The pump selected shall be a fixed-displacement vane pump suitable for variable
speed drives and shall be sized so that the allowable speed range will provide the minimum
and maximum flow required by the cylinder. Table 10 contains the technical data of the
system components for the DCP configuration. For the asynchronous motor in this case, it
is advisable to use a different type than in the previous configurations, i.e., an asynchronous
motor with a square housing, because although there are no significant differences at the
electromechanical level compared to an asynchronous motor with a round housing, the
inertia is significantly lower.
Table 10 shows the main technical data of the selected components; Figure A5 the
main components; and Figure 15 the active electric power absorbed in the cycle for the
static DCP configuration.
DCP CONFIGURATION (Inertial actions neglected)
+
+
+
Mod. 31V –4G0045– B
AC30V frequency
converter in alternating
current (inverter)
Mod. MS54133KFA
4 Poles Square case
asynchronous motor
Mod. T7B E12
Fixed Displacement Vane
Pump
Mod.
D41FEZ32FC1NB70
Pilot operated
Proportional DCV with
Regeneration - A Hybrid
The capacitor bank installed inside the frequency converter
generates a reactive (capacitive) electrical power component
that compensates for the reactive (inductive) electrical power
required by the motor. The compensation is such as to bring
the total reactive electrical power absorbed below 50% of the
active component. As a consequence, the power factor cosϕ
is reported above 0.9. In these operating conditions, the
economic penalties charged to the user are reduced to zero.
Wel.a.mean
14.43 kW
Γ
En.el.a
Γ
En.el.r
1.73 €/hour
+
≅ 0 €/hour
(14.43 kW ∙ 0.12 €/kWh)
(PENALTY)
6150.6 €/year
+
≅ 0 €/year
(14.43 kW ∙ 0.12 €/kWh ∙ 16 h/day ∙ 222 days/year)
(PENALTY)
Γ
En.el.TOT
1.73 €/hour
6150.6 €/year
0
10
20
30
40
50
60
70
80
90
100
0 0.2 0.4 0.6 0.8 1 1.2 1.4
Power [kW]
Time [s]
Potenza elettrica attiva
Termine / Inizio della Fase
Potenza elettrica attiva media
Active Electric Power
Phase Stop/Start
Mean Active Electric Power
Figure 15.
DCP static configuration. Active electrical power absorbed over the work cycle: instanta-
neous value and average value.

Appl. Sci. 2023,13, 10066 22 of 33
Table 10.
Designations and technical data of the main elements of the DCP architecture, in the
static case.
Circuit Element Parameter Value
Electric Motor
Code Parker Mod. MS54133KFA
Type 4 poles—square case
Nominal power 16 (kW)
Nominal speed 1500 (RPM)
Nominal torque 102 (Nm)
Efficiency 0.85
Frequency Converter
Code Parker Mod. 31V—4G0045
Nominal electric power 22 (kW)
Nominal current 45 (A)
Maximum current 49.5 (A)
Efficiency 0.98
Pump
Code Parker Mod. T7B E12
Type Fixed-displacement vane pump for variable speed drives
Max displacement 41 (mL/min)
Speed range 300–3000 (RPM)
Nominal pressure 275 (bar)
Directional Valve
Code Parker Mod. D41FEZ32FC1NB70
Type Pilot-operated proportional DCV with regeneration—A hybrid
Nominal flow rate 200 (L/min)
Nominal pressure drop 5 (bar)
Pressure Relief Valve
Code Parker Mod. R4V03535P0PM10VA1
Type Pilot-operated proportional pressure relief valve
Nominal flow rate 250 (L/min)
Maximum pressure 350 (bar)
2.6.2. DCP Configuration—Dynamic Sizing
If the contribution of the moments of inertia is also considered in the sizing, the size of
the pump and the inverter (Table 11) changes compared to the simplified case (
Table 10
)
in which they were neglected. The size of the electric motor, on the other hand, does
not change.
Figure A6 shows the main components, and Figure 16 the active electrical power
absorbed in the cycle for the DCP dynamic configuration.
DCP CONFIGURATION (Inertial actions included)
+
+
+
Mod. 31V –4G0060– B
AC30V frequency
converter in alternating
current (inverter)
Mod. MS54133KFA
4 Poles Square case
asynchronous motor
Mod. T7B E15
Fixed Displacement Vane
Pump
Mod.
D41FEZ32FC1NB70
Pilot operated
Proportional DCV with
Regeneration - A Hybrid
The capacitor bank installed inside the frequency converter
generates a reactive (capacitive) electrical power component
that compensates for the reactive (inductive) electrical power
required by the motor. The compensation is such as to bring
the total reactive electrical power absorbed below 50% of the
active component. As a consequence, the power factor cosϕ
is reported above 0.9. In these operating conditions, the
economic penalties charged to the user are reduced to zero.
Wel.a.mean
16.06 kW
Γ
En.el.a
Γ
En.el.r
1.93 €/hour
+
≅ 0 €/hour
(16.06 kW ∙ 0.12 €/kWh)
(PENALTY)
6845.4 €/year
+
≅ 0 €/year
(16.06 kW ∙ 0.12 €/kWh ∙ 16 h/day ∙ 222 days/year)
(PENALTY)
Γ
En.el.TOT
1.93 €/hour
6845.4 €/year
0
10
20
30
40
50
60
70
80
90
100
0 0.2 0.4 0.6 0.8 1 1.2 1.4
Power [kW]
Time [s]
Potenza elettrica attiva
Termine / Inizio della Fase
Potenza elettrica attiva media
Active Electric Power
Phase Stop/Start
Mean Active Electric Power
Figure 16.
DCP dynamic configuration. Active electrical power absorbed over the work cycle:
instantaneous value and average value.

Appl. Sci. 2023,13, 10066 23 of 33
Table 11.
Designations and technical data of the main elements of the DCP architecture, in the
dynamic case.
Circuit Element Parameter Value
Electric Motor
Code Parker Mod. MS54133KFA
Type 4 poles—square case
Nominal power 16 (kW)
Nominal speed 1500 (RPM)
Nominal torque 102 (Nm)
Efficiency 0.85
Frequency Converter
Code Parker Mod. 31V—4G0060
Nominal electric power 30 (kW)
Nominal current 60 (A)
Maximum current 66 (A)
Efficiency 0.98
Pump
Code Parker Mod. Mod. T7B E15
Type Fixed-displacement vane pump for variable speed drives
Max displacement 50 (mL/min)
Speed range 300–2700 (RPM)
Nominal pressure 240 (bar)
Directional Valve
Code Parker Mod. D41FEZ32FC1NB70
Type Pilot-operated proportional DCV with regeneration—A hybrid
Nominal flow rate 200 (L/min)
Nominal pressure drop 5 (bar)
Pressure Relief Valve
Code Parker Mod. R4V03535P0PM10VA1
Type Pilot-operated proportional pressure relief valve
Nominal flow rate 250 (L/min)
Maximum pressure 350 (bar)
3. Results
3.1. Estimated Absorbed Power for the Considered Architectures
For standard, regenerative, high–low, and VPD plant architectures, the inertial effects
can usually be neglected; however, this is not true for the architecture with DCP. For
the first four architectures, the angular acceleration or deceleration ramps experienced
by the asynchronous motor during operation are caused by the phenomenon of motion
irregularity. As a result of the cyclic and periodic loading (the motion of the pumping unit),
the speed of the electric motor fluctuates in a typically very small range, which is around
the rated speed. The resulting moments of inertia generally do not significantly change the
electrical active power consumption data. Table 12 summarizes the absorbed active and
reactive power in each phase and their values in the cycle for each architecture considered.
They have been estimated on the basis of the previously performed dimensioning and the
mechanical and electrical analysis of the system.
In Figure 17, the average values of the active power absorbed in the working cycle
for each considered architecture are drawn with histograms for a more effective visual
comparison. For the regenerative configuration, the white part of the bar shows the energy
savings compared to the standard solution. For the other configurations, the portion of
energy saved is also visually represented, but the regenerative configuration is used as a
reference. The analysis was performed for both the static and dynamic cases for the DCP
architecture, and the estimated data of active electrical power consumption with respect to
the architectures with DCP show that the estimated average active power consumption
is 1.63 kW higher when dynamics are taken into account. This obviously non-negligible
difference is caused by the presence of very strong dynamic effects during the operation of
the architecture with DCP. In this system architecture, the fluctuations of the motor speed
are very large (even on the order of 1000 RPM) and typically occur in times on the order of
fractions of a second in the considered application. In contrast, for the architectures that do

Appl. Sci. 2023,13, 10066 24 of 33
not use DCP, the motor speed fluctuations are on the order of 10 RPM. In the case of the
DCP architecture, therefore, dynamic sizing that takes inertial effects into account provides
a more reliable estimate of the energy savings than static sizing.
Table 12. Active and reactive power absorbed in the cycle phases.
Configuration Cycle Phase Active Power
(kW)
Reactive Power
(VAr)
Mean Active
Power (kW)
Mean Reactive
Power (VAr)
Standard
QD 24.32 21.92
40.48 29.06
B 93.36 52.92
E 32.69 23.44
QR 32.69 23.44
TR 12.33 20.63
Regenerative
QD 12.4 12.44
20.05 15.60
B 48.39 26.14
E 16.59 13.07
QR 16.59 13.07
TR 3.07 11.97
High–Low
QD 12.37 9.81
15.77 11.72
B 27.82 16.97
E 16.99 11.23
QR 16.99 11.23
TR 2.99 8.52
VDP
QD 14.85 10.5
16.67 12.71
B 33.56 21.87
E 11.53 9.62
QR 19.04 12.03
TR 3.35 8.55
DCP Static
QD 12.86 24.54
14.43 16.49
B 29.18 22.29
E 9.83 9.29
QR 16.91 28.05
TR 0.44 1.27
DCP Dynamic
QD Acc1 2.34 1.6
16.06 21.26
QD Acc2 35.25 48.47
QD Unif 12.9 20.63
QD Dec1 −25 41.62
QD Dec2 11.44 11.58
QD Acc2 35.25 48.47
B 37.58 46.75
E 12.41 11.74
QR Acc1 21.66 16.89
QR Acc2 29.9 44.49
QR Unif 16.91 22.76
QR Dec1 −24.22 34.42
QR Dec2 −3.19 2.15
TR 0.44 1.05
Figure 18 reports a comparison between the considered architectures, similarly struc-
tured to the one in Figure 17, referring in thiscase to the hourly cost of the electricity absorbed.
The matrix in Table 13 shows the percentage variation between the average active
power required by one architecture (in the rows) versus another (in the columns). In
Table 14, a similar comparison is reported for absorbed reactive power, and in Table 15 for
the hourly cost of absorbing electricity from the power supply grid.

Appl. Sci. 2023,13, 10066 25 of 33
0 5 10 15 20 25 30 35 40 45
Standard Configuration
Regenerative configuration
High - Low configuration
VDP configuration
DCP configuration - static sizing
DCP configuration - dynamic sizing
40.48
20.05
15.77
16.67
14.43
16.06
20.43
4.28
3.38
5.62
3.99
Mean Active Electric Power on the working cycle [kW]
Figure 17.
Comparison between the different plant architectures in terms of average active electrical
power absorption during the work cycle. For the regenerative configuration, the white parts of the
bars represent the power saving with respect to the standard configuration. For the other configura-
tions, white parts of the bars represent power savings with respect to the regenerative configuration.
Table 13.
Comparison between the different configurations in terms of absorbed active power. Values
in the matrix are the percentage saving/increase of the configuration in the row, with respect to the
one in the column.
Standard Regenerative High–Low VDP DCP Static
Regenerative −50.5%
High–Low −61.1% −21.3%
VDP −58.8% −16.8% +5.7%
DCP Static −64.3% −28.1% −8.4% −13.4%
DCP Dynamic −60.3% −19.9% +1.8% −3.6% +11.2%
Table 14.
Comparison in percentage terms of reactive power absorbed with the different configura-
tions. Values in the matrix are the percentage saving/increase of the configuration in the row, with
respect to the one in the column.
Standard Regenerative High–Low VDP DCP Static
Regenerative −46.3%
High–Low −59.7% −24.8%
VDP −56.3% −18.5% +8.4%
DCP Static −43.2% +5.7% +40.6% +29.6%
DCP Dynamic −26.8% +36.2% +81.3% +67.1% +28.9%

Appl. Sci. 2023,13, 10066 26 of 33
0 1 2 3 4 5 6
Standard Configuration
Regenerative configuration
High - Low configuration
VDP configuration
DCP configuration - static sizing
DCP configuration - dynamic sizing
5.8
3.07
2.27
2.54
1.73
1.93
2.73
0.8
0.53
1.34
1.14
Hourly cost of energy absorbed by the electricity grid [€/h]
Figure 18.
Comparison between the different system architectures in terms of the hourly cost of
electricity absorbed (active and reactive) during the work cycle. For regenerative configuration,
the green parts of the bars represent the hourly energy cost reduction with respect to the standard
configuration. For the other configurations, white parts of the bars represent the hourly energy cost
reduction with respect to the regenerative configuration.
Table 15.
Comparison in percentage terms of the hourly cost of absorbing electricity from the power
supply grid between the different configurations (based on Italian energy market rules). Values in the
matrix are the percentage saving/increase of the configuration in the row, with respect to the one in
the column.
Standard Regenerative High–Low VDP DCP Static
Regenerative −47.1%
High–Low −60.8% −26.1%
VDP −56.2% −17.3% +11.9%
DCP Static −70.1% −43.7% −23.8% −31.9%
DCP Dynamic −66.7% −37.1% −15.1% −24% +11.5%
3.2. Experimental Results
Figure 19 reports the results of experimental measurements conducted on the plant in
which the architecture with DCP was implemented. More specifically, Figure 19a reports
the profile of the absorbed current by phase; Figure 19b reports the electrical active power
absorbed by the frequency converter. The average value of the experimentally measured
active electric power is 14.85 kW.

Appl. Sci. 2023,13, 10066 27 of 33
(a)
(b)
Power [kW] Current [A]
Figure 19.
Active and active electrical power absorbed by the real blanking press with DCP power
architecture. (
a
) Active electrical current absorbed [A] by the architecture with DCP over time [s]. The
measurement is made upstream of the frequency converter and is related to a single phase. To get the
total current just multiply by
√3
. (
b
) Active electrical power absorbed [kW] by the architecture with
DCP over time [s]. The measurement is made upstream of the frequency converter.
4. Discussion
From the analysis performed on the basis of the system modeling, it appears that
the difference in the absorbed active power is underestimated by 1.63 kW in the case of
the DCP architecture neglecting the dynamic aspects, compared to the case in which the
dynamic effects are taken into account. This difference is significant, which is why the
following considerations take into account the dynamic DCP architecture.
The introduction of regeneration in the directly controlled valve leads to a reduction in
the average active power consumption by 50.5%. In architectures 3, 4, and 5, regeneration
is kept as a common element, so the differences in active power consumed are determined
only by the differences in the configurations.
From the results summarized in Table 13, it appears that the high–low, VDP, and
DCP solutions achieve higher energy efficiency for the plant than the purely renewable
architecture. In fact, in all three cases there is a more or less pronounced decrease in the
active electric power absorbed from the grid.
The solutions that achieve the greatest savings in terms of active power consumption
are, in order, the high–low configuration, followed by DCP, and finally VDP. The difference
between the high–low configuration and DCP is very small (1.8%).
The particular efficiency of the high–low architecture for this application is due to
the nature of the variable speed industrial process. In fact, the high–low architecture
is particularly well suited for operation with phases in which high oil flow rates at low
pressure alternate with phases of low liquid flow rates at high pressure. This is typically
the case with punching. The lower energy efficiency of the solution with VDP is caused by
the additional control oil flow that must be continuously ensured to maintain the flow rate
control of the axial piston pump.
Comparing the VDP configuration with the DCP configuration, it can be seen that
the DCP configuration provides savings in power consumption, even if the savings are
small in the case considered. It can be concluded that the DCP solution is slightly more
advantageous than the VDP solution from the point of view of absorbing the average active
electric power during the duty cycle, or that the VDP solution is in no way better than the
DCP solution.

Appl. Sci. 2023,13, 10066 28 of 33
From the data referring to the active electrical power parameter (Figure 13 and
Table 13
), the solution with DCP realized a marked advantage over the standard architec-
ture, a slight advantage over the VDP configuration, and a comparable, or even slightly
lower, result than the high–low solution.
If, on the other hand, the comparison is made using the cost parameter for electrical
energy (based on Italian energy prices), the DCP architecture is significantly cheaper, even
compared to the high–low and VDP solutions. This is due to the reactive power component
of the electric power (Table 14), for which there are economic disadvantages when there is
a high absorption of reactive power from the power grid. In the DCP architecture, these
economic disadvantages do not exist. This is ensured by the effect of the capacitor bank
present in the drive, which brings the absorbed electrical reactive energy (inductive) below
50% of the active energy or lowers the power factor below the value of 0.9. Therefore,
taking into account the contributions of both active and reactive power, the DCP solution is
the most advantageous among the solutions studied in terms of energy operating costs,
which also results in a saving of 15% compared to the high–low architecture.
For the case of the architecture with DCP, a comparison can be made between the
theoretical absorption of the active electric power input (16.06 kW) and the absorption
actually measured on the system in one cycle (14.85 kW). An overestimation of the actual
value of about 8% is found. This value is quite low. Therefore, it can be assumed that the
performed analysis reflects the reality sufficiently well.
5. Conclusions
The choice of power generation architecture of a hydraulic system has a very large
impact on the energy efficiency and operating costs.
All the architectures considered provide a significant reduction in active power con-
sumption compared to the standard architecture, ranging from 50.5% to 64.3%. The hourly
cost of electricity absorption is also significantly reduced, with percentages ranging from
47.1% to 70.1%. This shows the importance of the architecture of the system. Moreover, the
comparative study of the most common hydraulic architectures used to drive an industrial
process with time-varying speeds and work forces has highlighted the differences between
the various solutions in quantitative terms and provided objective comparative data that
was not available in the literature.
Comparing the values of energy savings with those reported in the literature for differ-
ent approaches, it can be observed that the values obtained for the considered architectures
are consistent with the best ones reported in Table 1. Moreover, it can be deduced that even
with a simple and low-cost solution such as regeneration, significant and larger energy
savings can be achieved than with the more complex solutions reported in the literature.
The modern hydraulic architecture with drive-controlled pump and DCV with
regeneration—A hybrid, limited to the category of industrial processes, is a very efficient
solution in terms of energy consumption and economy.
It has been shown that the dynamic phenomena, often overlooked in the sizing, have
a significant impact on the total energy absorption.
The oleo-hydraulic architectures studied represent a significant part of the state of the
art in hydraulic technology for the actuation of industrial machinery, but not its entirety.
Other solutions characterized by high energy efficiency include architectures with high-
efficiency square-case induction motors or architectures with a brushless synchronous
electric motor and an internally geared pump.
Another oleo-hydraulic architecture that could be explored, but which is a poor fit
for the industrial process example, involves the use of an accumulator installed in parallel
with a fixed-displacement pump.
Furthermore, from an economic point of view, the analysis could be complemented by
an analysis of the initial investment costs required for each architecture considered.
However, it can be concluded that energy-conscious decisions for the power architec-
ture of a hydraulic axis can benefit from the presented analysis.

Appl. Sci. 2023,13, 10066 29 of 33
Funding: This research received no external funding.
Acknowledgments:
Acknowledgement of the valuable technical contribution of the engineers Davide
Ceppelli, Alessandro Pennacchio, and Felice Lanzetti.
Conflicts of Interest: The authors declare no conflict of interest.
Abbreviations
The following abbreviations are used in this manuscript:
ERS Energy regeneration system
SI Single-valve independent control
SMISMO Separate meter in and separate meter out control
DP Dual valve parallel control
DCP Direct-driven pump-controlled hydraulic system
DCV Directional control valve
CDS Combined drive system
HIMMs Hydraulic injection molding machines
FC Force control
EHVDPS Electro-hydraulic variable-displacement pump system
LSC Load-sensing control
CSPC Constant supply pressure control
VDP Variable-displacement pump
QD Quick descent
B Blanking
E Expulsion
QR Quick rise
TR Table rotation
Nomenclature
The following mathematical symbols are used in this manuscript:
sStroke
vVelocity
φcCylinder bore
φrRod bore
APPiston area in the rear chamber side
ATPiston area in the front chamber side
αCylinder area ratio
Qo
AOutward flow rate from the rear chamber
Qo
BOutward flow rate from the front chamber
Qi
AInward flow rate into the rear chamber
Qi
BInward flow rate into the front chamber
pPPressure at the delivery of the pump
pTPressure inside the tank
QnNominal flow rate
QxGeneric flow rate
∆pnNominal pressure drop
∆pxGeneric pressure drop
∆pPA Pressure drop on the PA way
∆pAT Pressure drop on the AT way
∆pPB Pressure drop on the PB way
∆pBT Pressure drop on the BT way
po
BPressure in the front chamber for the outward stroke
po
APressure in the rear chamber for the outward stroke
pi
BPressure in the front chamber for the inward stroke
pi
APressure in the rear chamber for the inward stroke
FoThrust force

Appl. Sci. 2023,13, 10066 30 of 33
FiPull force
vo
max Maximum rod speed during the outward stroke
vi
max Maximum rod speed during the inward stroke
Qp−th Theoretical pump flow rate
Qp−e f f Effective pump flow rate
WnElectric motor nominal power
VPPump displacement
Nmax−pu mp Maximum pump speed
pmax Maximum pump pressure
Qloa d Flow rate required by the load
popening Pressure relief valve opening pressure
pREG Regulated pressure of the pressure relief valve
Qdisch arged Flow rate discharged through the pressure relief valve
Mn−motor Nominal torque of the electric motor
MRMS−loa d Root mean square torque required by the load
Appendix A. Main Components for Each Considered Configuration
In this appendix, the chosen main components for each considered configuration
are shown.
STANDARD CONFIGURATION
Mod. MR4P05500
Round case 4 poles Asynchronous
Motor
+
Mod. T7E 050
Fixed Displacement Vane Pump
+
Mod. D41FEE02FC1NB0
Pilot operated Proportional DCV
1:1:1:1
Figure A1. Main components of the standard configuration.
REGENERATIVE CONFIGURATION
Mod. MR4P03000
Round case 4 poles Asynchronous
Motor
+
Mod. T7D B24
Fixed Displacement Vane Pump
+
Mod. D41FEZ32FC1NB70
Pilot operated Proportional DCV
with Regeneration - A Hybrid
Figure A2. Main components of the regenerative A—hybrid configuration.

Appl. Sci. 2023,13, 10066 31 of 33
HIGH –LOW CONFIGURATION
Mod. MR4P01850
Round case 4 poles
Asynchronous Motor
+
Mod. T7B B14
Fixed Displacement
Vane Pump (HPLD)
+
Mod. T7ASW B40
Fixed Displacement
Vane Pump (LPHD)
+
Mod. D41FEZ32FC1NB70
Pilot operated Proportional DCV
with Regeneration - A Hybrid
Figure A3. Main components of high–low configuration.
VDP CONFIGURATION
Mod. MR4P01850
Round case 4 poles
Asynchronous Motor
+
Mod. PV 092
Axial Piston Pump - variable
displacement
+
Mod. D41FEZ32FC1NB70
Pilot operated Proportional DCV
with Regeneration - A Hybrid
Figure A4. Main components of VDP configuration.
DCP CONFIGURATION (Inertial actions neglected)
Mod. 31V–4G0045–B
AC30V frequency converter
in alternating current
(inverter)
+
Mod. MS54133KFA
4 Poles Square case
asynchronous motor
+
Mod. T7B E12
Fixed Displacement
Vane Pump
+
Mod. D41FEZ32FC1NB70
Pilot operated Proportional DCV
with Regeneration - A Hybrid
Figure A5. Main components of DCP static configuration.
DCP CONFIGURATION (Inertial actions included)
Mod. 31V–4G0060–B
AC30V frequency converter
in alternating current
(inverter)
+
Mod. MS54133KFA
4 Poles Square case
asynchronous motor
+
Mod. T7B E15
Fixed Displacement
Vane Pump
+
Mod. D41FEZ32FC1NB70
Pilot operated Proportional DCV
with Regeneration - A Hybrid
Figure A6. Main components of DCP dynamic configuration.

Appl. Sci. 2023,13, 10066 32 of 33
References
1.
Yagi, M.; Managi, S. The spillover effects of rising energy prices following 2022 Russian invasion of Ukraine. Econ. Anal. Policy
2023,77, 680–695. [CrossRef]
2.
Hu, Y.; Man, Y. Energy consumption and carbon emissions forecasting for industrial processes: Status, challenges and perspectives.
Renew. Sustain. Energy Rev. 2023,182, 113405. [CrossRef]
3. Li, B. Modeling Energy Consumption in the Production Processes of Industrial Units Based on Load Response Programs in the
Energy Market. Energy Eng. J. Assoc. Energy Eng. 2023,120, 461–481. [CrossRef]
4.
Xu, Z.; Liu, Y.; Hua, L.; Zhao, X.; Guo, W. Energy analysis and optimization of main hydraulic system in 10,000 kN fine blanking
press with simulation and experimental methods. Energy Convers. Manag. 2019,181, 143–158. [CrossRef]
5.
Yan, X.; Chen, B.; Yin, F.; Ji, H.; Ma, Z.; Nie, S. Energy optimization of main hydraulic system in a forging press by simulation and
experimental methods. Energy 2023,277, 127620. [CrossRef]
6.
Mahato, A.C.; Ghoshal, S.K. An Overview of Energy Savings Approaches on Hydraulic Drive Systems. Int. J. Fluid Power
2020
,
21, 81–118. [CrossRef]
7.
Schmidt, C.; Li, W.; Thiede, S.; Kara, S.; Herrmann, C. A Methodology for Customized Prediction of Energy Consumption in
Manufacturing Industries. Int. J. Precis. Eng.-Manuf.-Green Technol. 2015,2, 163–172. [CrossRef]
8.
Lyu, L.; Chen, Z.; Member, S.; Yao, B.; Member, S. Energy Saving Motion Control of Independent Metering Valves and Pump
Combined Hydraulic System. IEEE Asme Trans. Mechatron. 2019,24, 1909–1920. [CrossRef]
9.
Rana, M.; Nahian, S.A.; Chowdhury, H. Energy efficiency improvement of a vertical hydraulic system Energy Efficiency
Improvement of a Vertical Hydraulic System. In AIP Conference Proceedings; AIP Publishing: New York, NY, USA, 2022;
Volume 2681.
10.
Helian, B.; Chen, Z.; Yao, B. Energy-saving and accurate motion control of a hydraulic actuator with uncertain negative loads.
Chin. J. Aeronaut. 2021,34, 253–264. [CrossRef]
11.
Niu, S.; Wang, J.; Zhao, J.; Shen, W.; Yang, J. A Novel Multifunctional Energy-saving Electro-hydraulic Servo System. In
Proceedings of the 2021 IEEE/ASME International Conference on Advanced Intelligent Mechatronics (AIM), Delft, Netherlands,
12–16 July 2021; IEEE: Toulouse, France, 2021; pp. 314–319. [CrossRef]
12.
Dindorf, R.; Wos, P. Energy-Saving Hot Open Die Forging Process of Heavy Steel Forgings on an Industrial Hydraulic. Energies
2020,13, 1620. [CrossRef]
13.
Huang, H.; Zou, X.; Li, L.; Li, X.; Liu, Z. Energy-Saving Design Method for Hydraulic Press Drive System with Multi
Energy—Saving
Design Method for Hydraulic Press Drive System with Multi Motor—Pumps. Int. J. Precis. Eng.-Manuf.-Green
Technol. 2019,6, 223–234. [CrossRef]
14.
Koitto, T.; Calonius, O.; Kauranne, H.; Minav, T.; Pietola, M. Enhanced energy efficiency of industrial application by direct driven
hydraulic unit. In Proceedings of the 2018 Global Fluid Power Society PhD Symposium (GFPS), Samara, Russia, 18–20 July 2018;
IEEE: Toulouse, France, 2018; pp. 1–6. [CrossRef]
15.
Li, L.; Huang, H.; Zhao, F.; Sutherland, J.W.; Liu, Z. An Energy-Saving Method by Balancing the Load of Operations for Hydraulic
Press. IEEE Asme Trans. Mechatron. 2017,22, 2673–2683. [CrossRef]
16.
Wang, T.; Wang, Q. An Energy-Saving Pressure-Compensated Hydraulic System With Electrical Approach. IEEE/Asme Trans.
Mechatron. 2014,19, 570–578.
17.
Duflou, J.R.; Sutherland, J.W.; Dornfeld, D.; Herrmann, C.; Jeswiet, J.; Kara, S.; Hauschild, M.; Kellens, K. CIRP Annals—Manufacturing
Technology Towards energy and resource efficient manufacturing: A processes and systems approach. CIRP Ann.—Manuf.
Technol. 2012,61, 587–609. [CrossRef]
18.
Lin, T.; Chen, Q.; Ren, H.; Huang, W.; Chen, Q.; Fu, S. Review of boom potential energy regeneration technology for hydraulic
construction machinery. Renew. Sustain. Energy Rev. 2017,79, 358–371. [CrossRef]
19.
Ho, T.H.; Le, T.D. Development and Evaluation of Energy-Saving Electro-Hydraulic Actuator. Actuators
2021
,10, 302. [CrossRef]
20.
Schmidt, L.; Hansen, K.V. Electro-Hydraulic Variable-Speed Drive Networks;Idea, Perspectives, and Energy Saving Potentials.
Energies 2022,15, 1228. [CrossRef]
21.
Zhao, K.; Liu, Z.; Yu, S.; Li, X.; Huang, H.; Li, B. Analytical energy dissipation in large and medium-sized hydraulic press. J. Clean.
Prod. 2015,103, 908–915. [CrossRef]
22.
Zhang, Y.L.; Li, W.G. A precisive calculation method of volumetric and hydraulic efficiency of centrifugal pumps. Phys. Fluids
2023,35, 077104. [CrossRef]
23.
Beni´c, J.; Brezak, D.; Željko, Š.; Kolar, D.; Lisjak, D. Detailed experimental comparison of energy efficiency between proportional
and direct driven hydraulic system. Results Eng. 2023,18, 101178. [CrossRef]
24.
Chiang, M.H.; Yeh, Y.P.; Yang, F.L.; Chen, Y.N. Integrated control of clamping force and energy-saving in hydraulic injection
moulding machines using decoupling fuzzy sliding-mode control. Int. J. Adv. Manuf. Technol. 2005,27, 2673–2683. [CrossRef]
25.
Tiboni, M.; Borboni, A.; Vérité, F.; Bregoli, C.; Amici, C. Sensors and Actuation Technologies in Exoskeletons: A Review. Sensors
2022,22, 884. [CrossRef] [PubMed]
26.
Tiboni, M.; Amici, C. Soft Gloves: A Review on Recent Developments in Actuation, Sensing, Control and Applications. Actuators
2022,11, 232. [CrossRef]
27.
Cho, B.; Kim, S.W.; Shin, S.; Oh, J.H.; Park, H.S.; Park, H.W. Energy-Efficient Hydraulic Pump Control for Legged Robots Using
Model Predictive Control. IEEE ASME Trans. Mechatron. 2023,28, 3–14. [CrossRef]

Appl. Sci. 2023,13, 10066 33 of 33
28.
Yang, K.; Rong, X.; Zhou, L.; Li, Y. Modeling and Analysis on Energy Consumption of Hydraulic Quadruped Robot for Optimal
Trot Motion Control. Appl. Sci. 2019,9, 1771. [CrossRef]
29.
Ntsiyin, T.; Markus, E.D.; Masheane, L. A Survey of Energy-Efficient Electro-hydraulic Control System for Collaborative
Humanoid Robots. In Proceedings of the IOT with Smart Systems; Senjyu, T., Mahalle, P., Perumal, T., Joshi, A., Eds.; Springer
Nature: Singapore, 2022; pp. 65–77.
30.
Fang, D.; Shang, J.; Xue, Y.; Yang, J.; Wang, Z. Energy-Saving Control of a Novel Hydraulic Drive System for Field Walking Robot.
IOP Conf. Ser. Mater. Sci. Eng. 2018,301, 12141. [CrossRef]
31.
Moshayedi, A.J.; Shuvam Roy, A.; Sambo, S.K.; Zhong, Y.; Liao, L. Review On: The Service Robot Mathematical Model. EAI
Endorsed Trans. Robot. 2022,1, 8. [CrossRef]
32.
Zhang, J.; Xue, Y.; Yang, J.H.; Wang, Z.; Shang, J.Z. Mobile Robot Hydraulic Drive Energy-Saving Circuit Simulation Based on the
High-Frequency Switching Valve. Appl. Mech. Mater. 2013,278, 568–575. [CrossRef]
33.
Teguede, A.; Kamsu-foguem, B.; Houe, R. Advances in Engineering Software Health condition monitoring of a complex hydraulic
system using Deep Neural Network and DeepSHAP explainable XAI. Adv. Eng. Softw. 2023,175, 103339. [CrossRef]
34.
Tiboni, M.; Remino, C.; Bussola, R.; Amici, C. A Review on Vibration-Based Condition Monitoring of Rotating Machinery. Appl.
Sci. 2022,12, 972. [CrossRef]
35.
Tiboni, M.; Incerti, G.; Remino, C.; Lancini, M. Comparison of signal processing techniques for condition monitoring based on
artificial neural networks. Appl. Cond. Monit. 2019,15, 179–188. [CrossRef]
36.
Gu, L.; Shi, Y. Online Monitoring Technique of Power Condition for Inverter-Fed Motor Driven Hydraulic System. Math. Probl.
Eng. 2019,2019, 4908942. [CrossRef]
37.
Tiboni, M.; Remino, C. Condition monitoring of a mechanical indexing system with artificial neural networks. In Proceedings of
the WCCM 2017—1st World Congress on Condition Monitoring, London, UK, 13–16 June 2017; British Institute of Non-Destructive
Testing: Northampton, UK, 2017.
38.
Makansi, F.; Schmitz, K. Data-Driven Condition Monitoring of a Hydraulic Press Using Supervised Learning and Neural
Networks. Energies 2022,15, 6217. [CrossRef]
39.
Quan, Z.; Quan, L.; Zhang, J. Review of energy ef fi cient direct pump controlled cylinder electro-hydraulic technology. Renew.
Sustain. Energy Rev. 2014,35, 336–346. [CrossRef]
40.
Ketelsen, S.; Padovani, D.; Andersen, T.O.; Ebbesen, M.K.; Schmidt, L. Classification and Review of Pump-Controlled Differential
Cylinder Drives. Energies 2019,12, 1293. [CrossRef]
41.
Shen, W.; Huang, H.; Pang, Y.; Su, X. Review of the Energy Saving Hydraulic System Based on Common Pressure Rail. IEEE
Access 2017,5, 655–669. [CrossRef]
42.
Xu, B.; Shen, J.; Liu, S.; Su, Q.; Zhang, J. Research and Development of Electro—Hydraulic Control Valves Oriented to Industry
4.0: A Review. Chin. J. Mech. Eng. 2020,33, 29. [CrossRef]
43.
Rydberg, K.E. Energy Efficient Hydraulics—System solutions for loss minimization. In Proceedings of the National Conference
on Fluid Power, Linköping University, Linköping, Sweden, 16–17 March 2015; pp. 16–17.
Disclaimer/Publisher’s Note:
The statements, opinions and data contained in all publications are solely those of the individual
author(s) and contributor(s) and not of MDPI and/or the editor(s). MDPI and/or the editor(s) disclaim responsibility for any injury to
people or property resulting from any ideas, methods, instructions or products referred to in the content.