Journal of Cleaner Production 428 (2023) 139341
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Journal of Cleaner Production
journal homepage: www.elsevier.com/locate/jclepro
Design and optimization of press slider with steel-aluminum composite
bionic sandwich structure for energy saving
Feng Liang a, b, c, Rui Wang a, b, c, Qiu Pang d, **, Zhili Hu a, b, c, *
a
Hubei Longzhong Laboratory, Xiangyang, 441000, Hubei, China
Hubei Key Laboratory of Advanced Technology for Automotive Components, Wuhan University of Technology, Wuhan, 430070, China
c
Hubei Engineering Research Center for Green & Precision Material Forming, Wuhan University of Technology, Wuhan, 430070, China
d
School of Machinery and Automation, Wuhan University of Science and Technology, Wuhan, 430081, China
b
A R T I C L E I N F O
A B S T R A C T
Handling editor: Panos Seferlis
During the operation of a power press, the movement of moving parts consumes large amounts of energy.
Minimizing the mass of moving parts through lightweight design is considered an effective strategy to reduce this
energy consumption. The slider has the largest mass among moving parts and is the best object for lightweight
design. Therefore, this paper proposes a lightweight design and structural optimization method for the slider to
reduce the energy consumption of the power press. The reduction in mass of the slider is achieved through the
application of a novel steel-aluminum composite (steel face sheets, high-strength aluminum alloy core) bionic
sandwich structure (SACBSS). Afterward, the performance of the SACBSS is evaluated, and the SACBSS slider is
designed. The energy consumption model and the finite element model of the slider are established, respectively.
Finally, the response surface method is used to establish the optimization model by comprehensively considering
energy consumption and structural performance. The results show that compared to the referenced steel slider,
the optimized SACBSS slider reduces the mass by 18.9% and the energy consumption by 6.1%. The proposed
design and optimization method of the slider can effectively reduce energy consumption while ensuring per­
formance, which is important for the realization of green production in factories.
Keywords:
Energy consumption
Steel-aluminum composite bionic sandwich
structure
Slider
Lightweight design
Structural optimization
1. Introduction
Electricity used in manufacturing is mainly derived from fossil fuels,
which results in significant greenhouse gas emissions and creates an
environmental burden (Panagiotopoulou et al., 2022). According to
statistics, the manufacturing sector consumes 26% of the energy
consumed in the United States (Triebe et al., 2021) and 59% of the
energy consumed in China (Li et al., 2019). As indispensable equipment
for industrial production, power presses (see Fig. 1(a)) are extensively
employed in metal material forming. However, their high energy con­
sumption (Yin et al., 2021) and low energy efficiency (Huang et al.,
2019) result in significant carbon emissions. With the promotion of
sustainable manufacturing (He et al., 2016), reducing the energy con­
sumption of power presses has become an urgent task.
Much research has been conducted on structural design, system
improvement, and process optimization to reduce the energy con­
sumption of power presses. Strano et al. (2013) proposed a press frame
optimization model including energy consumption, stress, and strain to
guide the frame design with the lowest energy consumption. Li et al.
(2017a, b) proposed an energy-efficient hydraulic system with a shared
double-actuator for paired hydraulic presses. Gao et al. (2018) devel­
oped an energy consumption model for the deep drawing process. As is
the case for cars (Poulikidou et al., 2015) and airplanes (Huang et al.,
2016), lightweighting is also an important way to reduce the energy
consumption of power presses. Zhao et al. (2016) used topology opti­
mization to lighten the frame of a 12,000 kN fine blanking press,
resulting in a 13.66% reduction in frame weight. Zhao et al. (2020) used
size optimization to reduce the mass of the fine blanking press frame by
12.94% based on sensitivity analysis. The lightweighting objects in the
above studies are all power press frames, which are beneficial for
reducing energy consumption during manufacturing and transportation.
However, the energy consumption of the power press is mainly
concentrated in its operation, accounting for about 90% of the total
lifecycle energy consumption (Yu et al., 2013). The movement of mov­
ing parts consumes a lot of energy during the operation of the power
* Corresponding author. Hubei Longzhong laboratory, Xiangyang, 441000, Hubei, China.
** Corresponding author.
E-mail addresses: Pqiu@wust.edu.cn (Q. Pang), zhilihuhit@163.com (Z. Hu).
https://doi.org/10.1016/j.jclepro.2023.139341
Received 27 May 2023; Received in revised form 13 September 2023; Accepted 15 October 2023
Available online 16 October 2023
0959-6526/© 2023 Elsevier Ltd. All rights reserved.
F. Liang et al.
Journal of Cleaner Production 428 (2023) 139341
Nomenclature
CEF
D
E
EI
F
CE
h
I
k
K
L
M
S
SF
q
P
T
U
V
θ
ω
σ
b
c
cp
deb
dep
elec
ep
es
fc
fs
m
n
opt
p
r
rec
s
sp
t
ts
tran
w
wc
ys
Carbon emission factor
Deflection, mm
Energy, J
Energy intensity
Force, kN
Carbon emissions, kg CO2
Height of center of mass, mm
Rotational inertia, kgm2
Power factor
Stiffness
Length, mm
Mass, kg
Displacement of slider, mm
Safety factor
Stamping cycle
Power of motor, kW
Torque, Nm
Deformation of slider, mm
Speed of slider, m/s
Angle, rad
Angular speed, rad/s
Stress, MPa
Bending
Crank
Compression
Deformation of blank
Deformation of power press
Electricity
Edge pressing
Empty stroke
Friction during working stroke
Fatigue strength
Material
Nominal force
Operation of power press
Power press
Connecting rod
Material recycling
Slider
Span between lower face sheet constraints
Torsion
Tensile strength
Transportion of slider
Width of sandwich structure
One working cycle
Yield strength
Subscripts
acq
Raw material acquisition
Fig. 1. Schematic diagram of the power press and slider.
Fig. 2. Microstructure structure of the beetle elytron.
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F. Liang et al.
Journal of Cleaner Production 428 (2023) 139341
above studies, and the structural properties of GBEPs with multiple
materials have yet to be studied. In addition, the four sides of the core
unit of GBEP are equal in length, which makes it have a certain partic­
ularity. Therefore, the rectangular grid beetle elytron plate (RGBEP) is
proposed in this paper.
The mass of the slider comes from welded steel plates. To reduce the
mass of the slider, it is necessary to decrease the use of steel plates. The
substitution of aluminum for steel has become a significant approach to
structural lightweighting due to the high structural index (the ratio of
the cubic root of Young’s modulus to density) of aluminum alloys
(Aggogeri et al., 2010) and their mature manufacturing processes
(Aamir et al., 2020). Although it leads to increased energy consumption
and carbon emissions during the product manufacturing phase, exten­
sive research has demonstrated that the use of aluminum alloys signif­
icantly diminishes energy use and greenhouse gas emissions over the
entire lifecycle (He et al., 2020; Kelly et al., 2015; Kim and Wallington,
2013). In addition, there have been some attempts to use aluminum
alloys in mechanical equipment. Aggogeri et al. (2017) evaluated the
feasibility of making a z-axis ram for a milling machine with the Al
corrugated sandwich. They found through experiments that the ram
made of Al corrugated sandwich has better vibration suppression per­
formance. Dietmair et al. (2010) designed an energy-efficient milling
machine in which aluminum alloy was chosen as the milling head ma­
terial, resulting in a 27% mass reduction in the milling head but at the
expense of some stiffness. However, for the slider, simply using
aluminum alloy is likely to cause insufficient stiffness, and using more
steel plates will lead to excessive mass. One of the advantages of sand­
wich structure is that the face sheets and core structure can be selected
from different materials to achieve different performance requirements.
Li et al. (2020) derived the equivalent bending stiffness of the sandwich
structure and found that the face sheets contribution accounted for
95.6%. Therefore, steel for the face sheets and aluminum alloy for the
core structure can achieve high stiffness and lightweight design
requirements.
Based on the above analysis, to utilize the energy-saving advantages
brought by lightweighting, this paper proposes a novel steel-aluminum
composite bionic sandwich structure (SACBSS) for the lightweight
design of the slider. The reduction in mass usually leads to a loss of
stiffness, so an optimization method for the slider that comprehensively
considers energy consumption and structural performance is proposed
to maximize stiffness and minimize energy consumption. The remainder
of the paper is organized as follows. The performance of the SACBSS is
evaluated and introduced into the structural design of the slider in
Section 2. The energy-saving effect and structural performance of the
SACBSS slider are studied in Section 3. Optimization modeling and
comparative analysis are described in Section 4. Section 5 provides a
discussion and analysis of the research results. Section 6 concludes the
paper.
Fig. 3. Dimensions, shapes and internal structures of sandwich structures.
Table 1
Material parameters for Q235 and 6061-T6.
Q235 (Ji et al., 2016)
Density
Young’s modulus
Poisson’s ratio
Yield strength
Tensile strength
7850 kg/m
210 GPa
0.3
235 MPa
433 MPa
3
6061-T6 (Abundez et al., 2016)
2700 kg/m3
69 GPa
0.33
249 MPa
281 MPa
press. Therefore, the lightweight design of moving parts in the power
press has significant energy-saving benefits.
The slider (see Fig. 1(b)), as the terminal executing component of the
stamping process, is related to the forming quality of the workpiece. In
the research by Xu et al. (2018), the reduction in the mass of the slider
resulted in a decrease in the vibration quantity of the power press.
Moreover, the lightweight design of the slider does not affect the
stamping force. In hydraulic presses, the movement of the piston in the
hydraulic cylinder generates a pressure difference, resulting in forming
force (Xu et al., 2020), while in mechanical presses, the stamping force
exerted on the blank by the slider is derived from the energy stored in
the flywheel (Osakada et al., 2011). For the currently popular servo
presses, the stamping force originates from the torque output of the
servo motor (Halicioglu et al., 2016). On the contrary, an increase in the
mass of the slider results in an elevated energy requirement to overcome
the inertia force of the slider. Lastly, the slider has the largest mass share
in the transmission mechanism. These factors collectively render it an
ideal choice for the lightweight design of moving parts. The traditional
slider structure is similar to a closed sandwich structure (see Fig. 1(c))
(Li et al., 2017a, b), which needs to ensure structural stiffness while
reducing weight. Bionic design is an innovative approach to lightweight
design that draws inspiration from natural biological structures that
have evolved over millions of years to achieve optimal performance with
minimal material usage and space allocation (Yang et al., 2021). As one
of the oldest species on Earth, the beetle’s unique elytron structure
protects the body from injury while maintaining a low flying weight,
which is considered a high-stiffness and lightweight biological structure.
In the previous studies by Chen et al. (2012a, 2012b), the beetle elytron
is a typical sandwich structure, which consists of an upper and lower
skin with a trabecular-honeycomb structure (see Fig. 2). On this basis,
Chen et al. developed the grid beetle elytron plate (GBEP), which was
tested to have excellent resistance to compression (Hao et al., 2022),
bending (Chen et al., 2021) and shear (Chen et al., 2022). The above
studies show that GBEPs have good prospects for engineering applica­
tions. However, a single material has been chosen for GBEPs in the
2. Design of the SACBSS slider
2.1. Steel-aluminum composite rectangular grid beetle elytron plate
When the power press works, the slider has two working conditions:
uniform load condition and offset load condition. The uniform load
condition needs to consider compression and bending, and the offset
load condition needs to consider torsion. In order to avoid the blind
introduction of slider designs, this section evaluates the compressive,
bending, and torsional resistance of the RGBEP. Since the bottom surface
of the slider is usually rectangular, a sandwich structure model was
designed in accordance with the GB/T 1456–2021 standard (GB/T
1456-2021, 2021). To reduce repetitive modeling, the same model was
used for the evaluation of all three properties.
2.1.1. Geometric configuration
Fig. 3(a) shows the geometric shape and core configuration of the
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Journal of Cleaner Production 428 (2023) 139341
Fig. 4. Bending, torsion and compression of the sandwich structure.
Fig. 5. FE stress results for sandwich structures under bending condition.
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F. Liang et al.
Journal of Cleaner Production 428 (2023) 139341
Fig. 6. FE stress results for sandwich structures under compression condition.
RGBEP. The external dimensions of RGBEP are 180 mm × 60 mm × 20
mm. The thickness of both upper and lower face sheets is 2 mm, and the
height of the core is 16 mm. The RGBEP core unit is shown in Fig. 3(b),
where the length L1 is 18 mm, the width L2 is 15 mm, the outer radius r
of the hollow cylinder is 3 mm, and the wall thickness t is 1 mm (the
hollow cylinder has the same wall thickness as the grid). For compari­
son, a honeycomb plate (HP) and a rectangular grid plate (RGP) with
equal wall thickness are added, and the specific parameters are shown in
Fig. 3(c) and (d). In these sandwich structures, the materials of the core
structure and face sheets are steel (Q235) and aluminum alloy (6061T6), respectively. The specific material parameters are shown in Table 1.
was applied to the sandwich plate.
Compression: As described in Fig. 4(d), a uniform load with a total
value of 100 kN was applied to the upper face sheet, and the displace­
ment of the lower face sheet in the Y direction was constrained.
In order to select a suitable core structure to be introduced into the
structural design of the slider, the stiffness under each working condi­
tion is taken as the performance evaluation indicator. As illustrated in
Figs. 5–7, the maximum stress values in all FE results do not exceed the
material yield strength (235 MPa). Therefore, the simulation results are
within the elastic deformation range of sandwich structures. More
detailed data are shown in Table 2.
As shown in Table 2, RGBEP exhibits minimal bending deflection,
compression displacement, and torsion angle for equal wall thickness,
with more excellent resistance to deformation. The bending stiffness Kb
(Wang et al., 2018), compressive stiffness Kcp (Wu et al., 2014), and
torsional stiffness Kt (Karimipour et al., 2022) of the three types of
sandwich structures are calculated on the basis of the data in Table 2.
2.1.2. Performance analysis
The finite element (FE) commercial software ABAQUS 2020 is used
to analyze the mechanical properties of sandwich structures. In FE
models, eight-node solid elements (C3D8R) were used for the face sheets
and the core. The mesh size of the face sheets was 1 × 1 mm, and the
mesh size of the core was 0.5 × 0.5 mm. “Tie” constraints were used
between the face sheets and the core. The specific FE settings for each
mechanical property are as follows.
Bending: As shown in Fig. 4(a), a uniform load with a total value of 5
kN was applied to the middle region of the upper face sheet. All trans­
lational degrees of freedom (DOFs) at one end of the lower face sheet
were constrained, and the Y-directional DOFs at the other end were
constrained (Pirouzfar and Zeinedini, 2021) (see Fig. 4(b)).
Torsion: As depicted shown in Fig. 4(c), all DOFs were constrained at
one end of the sandwich plate, while at the other end, a torque of 80 Nm
Kb =
Fb Lsp 3
48Lw Db
(1)
Fcp
Lcp
(2)
T
θt
(3)
Kcp =
Kt =
where Fb is the bending load, Lsp is the span between lower face sheet
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Journal of Cleaner Production 428 (2023) 139341
Fig. 7. FE stress results for sandwich structures under torsional condition.
Table 2
FE simulation results of mechanical properties.
Bending
Compression
Torsion
Deflection (mm)
Maximum stress (MPa)
Displacement ( × 10− 2 mm)
Maximum stress (MPa)
Torsion angle ( × 10− 2 rad)
Maximum stress (MPa)
HP
RGP
RGBEP
0.210
174.6
2.296
144.2
1.425
191.7
0.153
117.8
1.705
94.38
0.850
132.9
0.144
115.0
1.542
86.94
0.793
124.6
constraints, Db is the bending deflection, Lw is the width of the sandwich
structure, Fcp is the compression load, Lcp is the compression displace­
ment, T is the torque, and θt is the torsion angle.
In order to eliminate the effect of mass, the calculated stiffness is
divided by the model mass, so the stiffness described later refers to the
stiffness per unit mass. As shown in Fig. 8, it is clear that both RGBEP
and RGP exhibit excellent bending, compression, and torsional resis­
tance compared to HP. Among them, RGBEP performs the best, with a
37.9% improvement in bending stiffness, 40.7% in compressive stiff­
ness, and 69.8% in torsional stiffness. Compared to RGP, RGBEP’s
bending, compressive, and torsional stiffness are 3.9%, 8.1%, and 4.6%
higher, respectively. With the same face sheet parameters and structural
materials, the performance improvement comes from the core structure.
Therefore, the inclusion of hollow cylinders enhances the load-bearing
capacity of conventional RGP. This result is in good agreement with
the conclusion of Li et al. (2020), which found that HPs with hollow
Fig. 8. Performance comparison of sandwich structures.
cylinders enhance the load-bearing capacity of conventional HPs by
promoting more cells to bear the load. Therefore, RGBEP is chosen to
design the slider.
2.2. Design of the SACBSS slider
To ensure sufficient stiffness, manufacturers tend to design sliders
with greater mass. However, too much mass can cause excessive inertia
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Journal of Cleaner Production 428 (2023) 139341
Fig. 9. The shape, internal structure and dimensions of the original slider and the SACBSS slider.
Fig. 10. Manufacturing of the SACBSS slider.
forces, increasing motor operating loads and energy consumption. Take
the 2500 kN servo press produced by Xuzhou metalforming machine
factory as an example. The original slider structure is shown in Fig. 9(a)
and (b), which consists of the guide rail, top plate, bottom plate, side
plates, connector seat plate, and support plate. The length of the bottom
plate is 2600 mm, the width is 1400 mm, and the thickness is 80 mm.
The whole structure is made of welded Q235 steel plates.
In the previous section, the excellent mechanical properties of
RGBEP have been demonstrated. In terms of manufacturability, the
difficulty of fabricating a sandwich structure typically depends on the
materials used and the core structure (Queheillalt et al., 2008; Wang
et al., 2019a, b). Due to its regular shape and simple configuration, the
core structure of RGBEP allows multiple processing methods, such as
additive manufacturing (Chen et al., 2021), mold fabrication (Hao et al.,
2023), and wire cutting. This contributes to the favorable manufactur­
ability of RGBEP. Therefore, RGBEP is introduced into the slider design,
as shown in Fig. 9(c). The support plate of the SACBSS slider is made of
6061-T6 aluminum alloy, and the rest is welded with Q235 steel plates.
Fig. 9(d) illustrates the structural parameters of the support plates for
the two types of sliders. The mass of the original slider is 6509.4 kg, and
the mass of the SACBSS slider is 5665.1 kg. The use of the SACBSS
reduces the mass of the slider by 12.97% without changing the structural
shape and size.
Mechanical connections and welding are common methods for
achieving steel-aluminum connections. However, the welding of steel
and aluminum results in the formation of brittle intermetallic com­
pounds (Stavropoulos et al., 2022). The presence of these compounds
leads to a reduction in mechanical properties and affects fatigue life.
Moreover, the combination of dissimilar metals in hybrid structures
formed by welding leads to galvanic corrosion in the joint, which re­
duces the reliability of the joint (Wang et al., 2021). To solve this
problem, a mechanical joining method has been developed, which
contributes to reducing the processing difficulty and cost. As shown in
Fig. 10, the support plate is machined by casting or wire cutting. The top
and bottom plates are CNC milled to machine grooves for assembly with
the support plate. Subsequently, the adhesive is employed in the as­
sembly of the support plate with the top and bottom plates. It not only
enhances the connection stiffness but also prevents galvanic corrosion
resulting from the close contact of steel and aluminum (Jiang et al.,
2021). The fabrication of the SACBSS slider is then completed by
welding the top plate and bottom plate with other steel plates.
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Journal of Cleaner Production 428 (2023) 139341
Table 3
Technology indexes of the 2500 kN servo crank press.
Name
Value
Nominal force (Fn )
Nominal force stroke (Ln )
Nominal pressure angle (θn )
Friction equivalent force arm (Lfc )
Vertical stiffness of press (Kp )
Slider stoke (2Lc )
Connecting rod length (Lr )
Crank rotational speed (ωc )
Connecting rod rotational inertia (Ir )
Crank rotational inertia (Ic )
Crank mass (Mc )
Connecting rod mass (Mr )
2500 kN
8 mm
30◦
12.5 mm
1200 kN/mm
120 mm
600 mm
2π rad/s
5.7 kgm2
14.2 kgm2
57 kg
119 kg
1
1 Fn 1
2500
= 2406 (J)
Edep = Fn Dp = Fn = × 2500 ×
2
2 Kp 2
1200
where Fn is the nominal force; Ln is the nominal force stroke; Lc is the
length of the crank; Lfc is the friction equivalent force arm; θn is the
nominal pressure angle; Dp is the vertical deflection of the power press;
Kp is the vertical stiffness of the power press.
The energy consumption of the servo press during the empty stroke is
equal to the movement consumption of the transmission mechanism, i.
e., the change of mechanical energy of the transmission mechanism.
According to the energy conservation law, the energy consumption
during the descent of the empty stroke is calculated by Eq. (9).
⃒
⃒
⃒
⃒1
1
1
Ees− down = ⃒⃒ (Ms + Mr )ΔV 2 + Ir Δω2r + Ic Δω2c + (Ms + Mr )gΔhs + Mc gΔhc ⃒⃒
2
2
2
(9)
3. Energy consumption and performance analysis of the slider
where Ms , Mr and Mc are the mass of slider, connecting rod and crank,
respectively; ΔV is the change in speed of slider at the beginning and end
of the descent; Δωr and Δωc are the change in angular speed of con­
necting rod and crank before and after the descent; Ir and Ic are the
rotational inertia of connecting rod and crank, respectively; Δhs and Δhc
are the change in center of gravity of slider and crank before and after
the descent.
The same method can be used to find the energy consumption Ees− up
in the ascent phase of the empty stroke. For the crank curve, the energy
consumption during the ascent phase of the empty stroke is equal to the
energy consumption during the descent phase of the empty stroke, i.e.,
3.1. Energy consumption
In order to understand the energy-saving effect of the SACBSS slider
on a power press, the energy consumption Ewc in one working cycle of a
servo press is calculated in detail by Eq. (4) (Zhao, 2019).
Ewc = Edeb + Eep + Efc + Edep + Ees
(4)
where Edeb is the energy consumption for the blank deformation; Eep is
the energy consumed for the edge pressing during the drawing process;
Efc is the energy consumed by friction during the working stroke; Edep is
the energy consumed by the elastic deformation of the servo press; Ees is
the energy consumed by the servo press during the empty stroke.
As an example, the following is a detailed analysis of a 2500 kN servo
crank press (technical indexes are shown in Table 3). In the analysis, the
process curve is the crank curve (Osakada et al., 2011). Fig. 11 shows a
complete working cycle of a servo crank press. The detailed calculation
is as follows (Zhou, 2016).
The deformation energy of the blank depends on the forming pro­
cess. In cold stamping, the deformation energy of the blank can be
calculated by Eq. (5).
1
1
Edeb = Fn Ln = × 2500 × 8 = 10000 (J)
2
2
(5)
The energy consumed by the edge pressing during the stamping
process is calculated by Eq. (6).
Eep =
Fn 2Lc 2500 × 120
= 8333 (J)
=
36
36
(6)
The frictional energy consumption of the crank slider mechanism
during the working stroke of the power press can be calculated by Eq.
(7).
1
1
1
Efc = Lfc Fn θn = × 12.5 × 2500 × π = 8177 (J)
2
2
6
(8)
(7)
When the power press works, it is subjected to load to produce elastic
deformation and accumulate elastic deformation energy.
Fig. 12. Sketch of the crank-slider mechanism.
Fig. 11. One working cycle of servo crank press.
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Journal of Cleaner Production 428 (2023) 139341
in the rotation speed of the connecting rod when the servo press de­
scends on an empty stroke. Eq. (20) is the amount of change in the center
of gravity of the slider and crank.
{
ΔV = 0.204
(19)
Δωr = − 5.7
Δhs = Δhr = 0.112
(20)
Substitute Eqs. (19) and (20) into Eq. (9) to obtain the energy con­
sumption of the servo press when the slider is descending with an empty
stroke.
Ees− down = 1.1Ms − 38
(21)
Therefore, the energy consumption in one working cycle of the servo
press can be obtained by the above calculation.
Ewc = Edeb + Eep + Efc + Edep + Ees
= 10000 + 8333 + 8177 + 2406 + 2.2Ms − 76
= 2.2Ms + 28840
Based on the parameters in Table 3, the relationship between the
mass of the slider and the energy consumption of the servo press is
clarified by calculating the energy consumption of the 2500 kN servo
crank press. As shown in Fig. 13, the energy consumption increases
linearly with the increase in the mass of the slider for the defined
technical parameters of the power press. Substituting the mass of the
slider in section 2.2 into Eq. (22), the energy consumption for one
working cycle of the servo crank press using the original slider is 43.16
kJ, and after replacing it with a SACBSS slider, the energy consumption
is 41.30 kJ, a reduction of 4.3%.
A brief calculation of the required motor power for the power press
was conducted to better quantify the energy-saving benefits of the
lightweight design of the slider. The motor power is calculated from Eq.
(23) (Gao et al., 2022).
Fig. 13. The relationship between mass of the slider and energy consumption
of the power press.
Ees− up = Ees− down . Therefore, Ees can be found by Eq. (10).
Ees = Ees− up + Ees− down = 2Ees− down
(10)
Fig. 12 shows a sketch of the motion of the crank-slider mechanism.
The position of the slider at the bottom dead center is the starting point
of the displacement of the slider, and the rotational speed of the crank is
ωc . Then the relationship between the displacement of the slider S and
the crank angle θc can be expressed by Eq. (11).
S = Lc (1 − cos θc ) + Lr (1 − cos θr )
(11)
Let λ = Lc /Lr , then Eqs. (12) and (13) can be obtained from the
geometric relationship.
sin θr = λ sin θc
(12)
√̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅
cos θr = 1 − λ2 sin2 θc
(13)
P=
(15)
(
)
1
V = ωc Lc sin θc + λ sin 2θc
2
(16)
When the crank rotates, the motion of the connecting rod can be
decomposed into linear motion with the slider and rotational motion
around point C. Let the angular velocity of rotation of the connecting rod
around point C be ωr . Then, ωr can be calculated from the geometric
relationship.
ωr cos θr = λωc cos θc
(23)
3.2. Static performance
Static performance refers to the structural characteristics of the slider
under the nominal force and is evaluated in terms of static stiffness. The
relationship between the structural stiffness and the applied load is
shown in Eq. (24), so the deformation can be used to assess the structural
stiffness.
(17)
Since λ is usually much less than 0.3 (He, 1981; Zheng et al., 2012),
then cos θr ≈ 1 and Eq. (17) can be simplified to Eq. (18).
ωr = λωc cos θc
kEwc
q
where k is the power factor, k = 1.2; q is the stamping cycle, q = 1/spm,
spm = 60/min.
It is calculated that the motor power required for the power press
with the original slider is 51.79 kW and a 55 kW motor should be used,
while the motor power required for the power press with the SACBSS
slider is 49.56 kW and a 50 kW motor should be used. If each machine
operates for 10 h a day, a factory with 50 power presses can save 2500
kWh of electricity per day. In 2022, the average electricity price for
industrial use in China was $ 0.0852/kWh (Zhao and Hu, 2020). The
electrical cost of a factory would be saved by $ 77 thousand for a year
(365 days). Therefore, SACBSS sliders can bring significant
energy-saving benefits.
The lightweight design of the slider for the purpose of reducing en­
ergy consumption should be carried out on the basis of ensuring stiff­
ness. Therefore, the evaluation indicators for the structural performance
of the slider are introduced subsequently.
Using Taylor’s formula to expand Eq. (13) to obtain Eq. (14), the
displacement can be converted to Eq. (15). Derivative of the displace­
ment gives the velocity of the slider.
√̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅̅
1
cos θr = 1 − λ2 sin2 θc ≈ 1 − λ2 sin2 θc
(14)
2
]
[
λ
S = Lc (1 − cos θc ) + (1 − cos 2θc )
4
(22)
(18)
[K]{U} = {F}
Since the process curve is a crank curve, the speed of the crank is a
constant value, then Δωc = 0. Substitute the data in Table 3 into Eqs.
(16) and (18) to obtain the change in speed of the slider and the change
(24)
where [K] is the stiffness matrix, {U} is the deformation vector, {F} and
is the load vector.
According to the slider production standard (JB/T 1647.2-2012,
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F. Liang et al.
Journal of Cleaner Production 428 (2023) 139341
Fig. 14. Sketch of slider under uniform load.
Fig. 15. Simulation results of the original slider.
2012), the ratio of the maximum deformation of the bottom plate to the
length of the bottom plate needs to meet Eq. (25). In this paper, the
length of the bottom plate is 2600 mm, so the permissible maximum
deformation [U] needs to meet Eq. (26).
where, {σ} is the stress vector, [D] is the elastic matrix, [B] and is the
strain matrix.
)
(
(28)
σ fs = 0.27 σ ys + σts
Umax
1
1
∼
≤
8000 6000
L
(25)
where, σ ys is the material yield strength, and σ ts is the material tensile
strength.
[U] ≤ 0.325
(26)
3.3. Dynamic performance
It should be noted that when the structural stiffness meets the re­
quirements, the strength also meets them. However, the stress concen­
tration is prone to fatigue failure, and the maximum stress should be
avoided to exceed the fatigue strength. From Eq. (27), it can be observed
that stress is related to deformation. Meanwhile, fatigue strength σ fs be
obtained through Eq. (28) (Zhao et al., 2016). Based on the material
parameters in Table 2, the fatigue strength of Q235 is calculated to be
180 MPa, while that of 6061-T6 is 143 MPa. According to Eq. (28), it can
be seen that the fatigue strength is related to the yield strength and
tensile strength of the material, which is the reason why high-strength
aluminum alloy is chosen as the core instead of ordinary aluminum
alloy, i.e., to avoid structural fatigue failure.
{σ} = [D][B]{U}
As the load acting on the slider is not constant, its dynamic perfor­
mance will affect the accuracy of the equipment and workpiece forming
quality, especially for moving parts like the slider that are in direct
contact with the workpiece (Stavropoulos et al., 2023). Therefore, even
if the slider under nominal force meets the stiffness requirement, it still
cannot be judged to meet the design requirements. For the slider, the
dynamic performance indicator is the dynamic stiffness, which is related
to the load frequency and is usually assessed by the structural modal
frequency (Zhao et al., 2020).
⃒
⃒
⃒[K] − f 2 [M]⃒ = 0
(29)
i
(27)
where [M] is the mass matrix, and fi is the ith-order natural frequency.
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Journal of Cleaner Production 428 (2023) 139341
Fig. 16. Simulation results of the SACBSS slider.
Table 4 shows the performance comparison between the original
slider and the SACBSS slider. As expected, the maximum deformation of
the SACBSS slider under nominal force increases compared to the
original slider due to the replacement of Q235 with 6061-T6 with a
lower modulus of elasticity. However, the decrease in stiffness does not
mean the design criteria are unmet. According to Eq. (26), the maximum
deformation of the SACBSS slider is still within the required range and
meets the stiffness requirement. The maximum stress of both sliders does
not exceed the fatigue strength, meeting the strength requirement. In
addition, when the power press works, the rotation frequency of the
motor is 25 Hz, and the first order natural frequency of the two sliders is
far away from the excitation frequency, so there will be no resonance.
Therefore, the SACBSS is appropriate for the lightweight design of the
slider, which also helps eliminate equipment manufacturers’ concerns
about the lack of structural stiffness due to the low modulus of elasticity
of the aluminum alloy.
Table 4
Performance comparison between original slider and SACBSS slider.
Energy consumption (kJ)
Maximum deformation (mm)
Maximum stress (MPa)
First order natural frequency (Hz)
Mass (kg)
Original slider
SACBSS slider
43.16
0.1103
59.5
101.14
6509.4
41.30
0.2074
51.4
100.25
5665.1
3.4. Performance comparison
During stamping, the stamping force increases gradually from 0 as
the slider contacts the blank and reaches the nominal force at 8 mm
(nominal force stroke) above the bottom dead center. When the nominal
force is reached, the slider’s velocity is low, and the impact is minimal.
Therefore, the finite element analysis is carried out under static loading
conditions (Li et al., 2017a, b; Liu et al., 2019). The slider is usually in
the uniform load condition when the power press is in normal operation.
The FE commercial software ANSYS Workbench 2022R1 is used to
analyze the structural performance of the slider in a uniform load con­
dition. Since the mold is not in contact with the entire slider bottom
surface during stamping (see Fig. 14(b)), the uniform load covers 80% of
the bottom surface. As shown in Fig. 14(a) and (c), the translational
DOFs of the slider in the Z direction and X direction are constrained by
slide rails. In addition, the connector limits the motion of the slider in
the Y direction. Figs. 15 and 16 show the FE results of the two sliders for
the uniform load condition.
4. Optimization for the SACBSS slider
The SACBSS slider in the previous section only changed the structure
and material of the support plate but did not change the structural pa­
rameters of the slider. Therefore, there is still room for optimization of
the SACBSS slider. In addition, the impact of the structural parameters
for the slider on energy consumption and structural performance has not
yet been understood. Therefore, this section is structured as follows.
Firstly, the structural parameters are selected based on sensitivity
analysis. Secondly, the linkage between structural parameters and
evaluation indicators is established by response surface methodology
Fig. 17. Structural parameters of the slider.
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Journal of Cleaner Production 428 (2023) 139341
Table 5
Initial values and ranges of structural parameters.
Design variable
Initial value (mm)
Range (mm)
x1
x2
x3
x4
x5
x6
x7
x8
x9
x10
x11
x12
x13
x14
x15
30
30
30
150
450
200
600
1000
30
30
30
30
30
120
30
20~40
10~40
20~40
130~170
425~465
150~250
500~700
900~1100
20~40
20~40
20~40
20~40
20~40
110~180
20~40
Fig. 19. The pareto frontier for multi-objective optimization.
Table 8
Structural design variables.
Origin
Before optimization
After optimization
Table 6
Analysis of variance for each indicator.
M
f1
Umax
σmax
P value
R2
Adjusted R2
Adequate precision
<0.0001
100%
100%
46130.651
<0.0001
97.64%
94.45%
20.943
<0.0001
99.73%
99.36%
69.529
<0.0001
95.72%
89.96%
17.273
Table 7
Parameters used in NSGA-II.
Parameter
Value
Population size
Crossover probability
Mutation probability
Maximum number of iteration generations
150
0.8
0.05
500
x2
x3
x7
x9
x14
x15
30
30
13
30
30
25
600
600
607
30
30
40
120
120
180
30
30
40
constant. The top plate and side plates are made of steel, and their
thickness significantly influences the mass of the slider, rendering them
the initial design parameters. The core structural parameters within the
SACBSS slider influence structural performance (Chen et al., 2021,
2022). Therefore, the parameters of the support plate have also been
selected as initial design parameters. As shown in Fig. 17, there are 15
initially identified structural parameters, each of which has a different
degree of influence on the evaluation indicators. Considering them all as
design variables will significantly increase the computational
complexity, so deleting the less influential parameters is necessary to
reduce the computation time.
Spearman correlation analysis is a common method to study the
correlation between two variables. The correlation range is [− 1, 1], and
the closer the result is to +1 or − 1, the stronger the correlation between
the two variables (Ge et al., 2022). In ANSYS 2022R1, the Design
Exploration module integrates parameter correlation analysis capabil­
ities. It conducts experimental design based on the number of design
parameters and objective functions, as well as the range of values for the
design parameters. Subsequently, simulation analyses are performed for
each experimental group, leading to the eventual determination of the
correlation between the design parameters and the objective functions
(Reh et al., 2006). The initial range of structural parameters is shown in
Table 5. The evaluation indicators of the slider have been given in the
previous section: first order natural frequency, maximum deformation,
maximum stress, and energy consumption. The mass of the slider is
positively correlated with the energy consumption of the power press.
Thus, mass can replace energy consumption as one of the evaluation
indicators. To ascertain the final parameter design variables, Spearman
correlation analysis is performed on each design parameter using the
Design Exploration module of ANSYS 2022 R1.
Fig. 18 shows the parameters significantly affecting the first order
natural frequency, maximum deformation, stress, and structural mass. It
can be clearly seen that the front and rear side plates thickness x1 have
the greatest effect on the first order natural frequency; the top plate
thickness x3 , grid wall spacing x7 , and hollow cylinder outer diameter
Fig. 18. Sensitivity of design parameters to the objective function.
Project
x1
30
30
25
(RSM). Finally, the SACBSS slider is optimized, and the influence of the
design parameters on the evaluated indicators is discussed.
4.1. Design variable selection
To ensure that the installation dimensions of the slider are un­
changed, the dimensions of the slide rail and bottom plate are kept
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F. Liang et al.
Journal of Cleaner Production 428 (2023) 139341
Fig. 20. Simulation results of the SACBSS slider after multi-objective optimization.
is less than 0.0001, which suggests that the fitted model is significant
(Xie et al., 2022). In addition, R2 is greater than 95%, Adequate preci­
sion is greater than 4, and adjusted R2 is greater than 85%, which in­
dicates that the RSM models have sufficient prediction accuracy (Ye
et al., 2018).
Table 9
Comparison of evaluation indicators.
Origin
Before optimization
After optimization
M (kg)
Umax (mm)
f1 (Hz)
σmax (MPa)
E (kJ)
6509.4
5665.1
5278.4
0.1103
0.2074
0.1773
101.14
100.25
83.87
59.5
51.4
43.2
43.16
41.30
40.45
4.3. Optimization result
x14 have the most significant effect on the maximum stress; the grid wall
thickness x9 and hollow cylinder wall thickness x15 have the greatest
effect on the maximum deformation; the front and rear side plates
thickness x1 , left and right side plates thickness x2 and top plate thick­
ness x3 have the greatest effect on the structural mass. Therefore, these
seven resultant parameters were selected as design variables.
Based on the above study, a multi-objective optimization model of
the slider is established to reduce the structural mass as much as possible
and improve the structural stiffness of the slider while ensuring the
structural strength and first order natural frequency. Based on the above
requirements, the multi-objective optimization problem can be
expressed by Eq. (31).
4.2. Response surface optimization model
Minimize M(X), U(X)
X = (x1 , x2 , x3 , x7 , x9 , x14 , x15 )
⎧
xi min ≤ xi ≤ xi max
⎪
⎪
⎪
⎪
⎪
⎨ σ max ≤ [σ]
Subject to
⎪
⎪
⎪ Umax ≤ [U]
⎪
⎪
⎩
75 ≤ f1
RSM is a statistical method for solving multiple design variables and
finding optimal design parameters by fitting complex response re­
lationships between design variables and objective functions in test data
sets through multiple quadratic regression equations. It provides a wellfitting model between the input parameters and the target response
without taking into account the complex interactions of the variables.
The second-order polynomial response model has high solution accuracy
and computational efficiency, and the expression is Eq. (30) (Benkhelifa
et al., 2022).
∑k
∑k
∑k
y = β0 +
β xi +
β x2 +
β xi xj + ε
(30)
i=1 i
i=1 ii i
i<j ij
(31)
where xi max and xi min are the maximum and minimum values of each
structural parameter; The safety factor SF is set to 1.5 (Wang et al.,
2019a, b; Zhou et al., 2023), then the permissible maximum stress can be
expressed as [σ ] = σ fs /SF = 95.
The Non-Dominated Sorting Genetic Algorithm (NSGA-II) is a fast
non-dominated sorting algorithm with an elite strategy that effectively
reduces the computational complexity while expanding the sampling
space and preserving locally dominant individuals. In this paper, NSGAII is used to solve the above multi-objective optimization model. The
parameter settings of NSGA-II in optimization are shown in Table 7, and
the obtained Pareto frontier is shown in Fig. 19.
The pareto frontier provides optimal solution sets that satisfy the
multi-objective requirements. It can be seen from Fig. 19 that reducing
mass and increasing stiffness are two conflicting indicators, and an in­
crease in one indicator leads to a decrease in the other. In order to select
an optimal target value from them, the minimum distance selection
method (TMDSM) is used to determine the optimal design point, and this
method can be expressed by Eq. (32) (Meng et al., 2020; Peng et al.,
2017). As shown in Fig. 19, the Knee point is obtained and marked by
where y is the evaluation indicator, ε is the residual error, xi is the
structural design parameter, β0 , βi , βii and βij are the coefficients of the
polynomial, respectively.
To reduce the test cost and improve the design efficiency, the BoxBehnken design (BBD) method is used to establish the test sample
data. The test sample points are determined by the software Design
Expert 8.0.6, and the evaluation indicator data are obtained by ANSYS.
The specific sample data are shown in Table B1. The second-order
polynomial model was used to fit each objective function based on the
sample data, and the fitting coefficients of each objective function are
shown in Table B2.
The analysis of variance (ANOVA) is performed for each indicator
model, and the results are shown in Table 6. The P value of each function
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Journal of Cleaner Production 428 (2023) 139341
Fig. 21. Impact of the structural parameters on evaluation indicators. a x1 , b x2 , c x3 .
using TMDSM. The detailed structural parameters of the Knee point are
shown in Table 8. Fig. 20 illustrates the FE results of the optimized
SACBSS slider.
)
(
i
∑
(fcn − min(fn (x)))2
4.4. Impact of structural parameters on evaluation indicators
Analyzing the influence of design parameters on the evaluation in­
dicators can better guide the slider design. The variation curves of the
evaluation indicators with the design parameters are plotted by keeping
the other design parameters constant and changing a certain design
parameter.
Fig. 21(a) shows the effect of design parameter x1 on the four eval­
uation indicators. The mass and stiffness of the slider increase as x1
increases, leading to an increase in energy consumption and first order
natural frequency. As x1 increases, the stiffness increases while the
maximum nominal force remains the same, which leads to a decrease in
deformation and stress. In addition, the first order vibration shape of the
slider is shown as the deformation of the front and rear side plates, as
seen in Figs. 15 and 16. According to Eq. (29), as x1 increases, the in­
crease in stiffness of the front and rear side plates is greater than the
increase in mass, bringing about an increase in the first order natural
frequency.
As shown in Fig. 21(b), when x2 increases, the mass of the slider
increases, leading to a rise in energy consumption and a decrease in
deformation. When x2 is too small, the stiffness of the front and rear side
plates is greater than that of the left and right side plates, which makes
1
2
Dmin =
(32)
n=1
where i represents the number of objective functions; fcn denotes the nth
objective value in the cth Pareto solution; and D denotes the distance
from the Knee point to the Utopia point.
As shown in Table 9, the multi-objective optimization reduces the
mass of the SACBSS slider from 5665.1 to 5278.4 kg (6.8% reduction),
the maximum deformation from 0.2074 to 0.1773 mm (14.5% reduc­
tion), the structural stiffness was improved, and the energy consumption
was reduced from 41.30 to 40.45 kJ. The reduction in the thickness of
the front and rear side plates (as shown in Table 8) results in a small drop
in the first order natural frequency, but still far away from the excitation
frequency. Compared with the conventional steel welded slider, the
SACBSS slider, after multi-objective optimization, reduces the mass by
18.9% and energy consumption by 6.1%, with the obvious effects of
lightweighting and energy savings.
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Journal of Cleaner Production 428 (2023) 139341
Fig. 22. Impact of the structural parameters on evaluation indicators. a x7 , b x9 , c x14 , d x15 .
the first order vibration shape manifest as deformation of the left and
right side plates. Therefore, as x2 increases, the first order natural fre­
quency first increases and then remains constant. The change in x2 does
not affect the maximum stress on the slider.
From Fig. 21(c), with the increase of x3 , the energy consumption and
the first order natural frequency always increase. As the thickness of the
top plate increases, the stiffness of the sandwich structure increases, and
likewise, the stiffness of the slider increases, making the deformation
decrease. From Figs. 15 and 16, it can be seen that the maximum stress in
the slider is concentrated at the seat plate, and as x3 increases, the
deformation of the top plate decreases, making the maximum stress
decrease.
As shown in Fig. 22(a), the variation of the grid wall spacing x7 does
not change the mass of the slider, and therefore the energy consumption
remains constant. As the grid wall gets closer to the edge of the
connector, the bottom plate is more evenly stressed, so the deformation
decreases and then increases as x7 increases. The mass and stiffness of
the side plates remain unchanged, so the first order natural frequency
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Journal of Cleaner Production 428 (2023) 139341
remains unchanged. As shown in Fig. 22(b)–(d), the trend of the first
order nature frequency is the same as that in Fig. 22(a). And the trends in
energy consumption and deformation are consistent with Fig. 21. The
reasons are not repeated. In addition, the maximum stress shows a
decreasing trend as x7 and x9 increase, while there are inflection points
in the stress curve as x14 and x15 increase.
The above analysis shows that the design parameters affect not only
the structural performance but also the energy consumption. Generally,
as the structural parameters increase, the mass and stiffness of the slider
increase, resulting in increased energy consumption and reduced
deformation. For the slider in operating condition, the first order mode
shape is manifested by the swing of the side plates, so the thickness of
the side plates (x1 , x2 ) is critical for the low order modal frequency of the
slider. In addition, within the range of parameter changes mentioned
above, the maximum stress is much smaller than the fatigue strength and
yield strength. Therefore, if there is a trade-off between strength and
deformation during design, priority should be given to reducing
deformation.
energy saving is proposed to reduce the energy consumption of the
power press in operation. To identify the relationship between the mass
of the slider and the energy consumption, the energy consumption of the
power press for one working cycle was analyzed. This analysis indicated
that the lower the mass, the lower the energy consumption. Therefore, a
novel steel-aluminum composite bionic sandwich structure was
designed for the lightweight design of the slider. On the one hand, the
use of lightweight material effectively reduces mass. On the other hand,
the core structure has proven to have excellent performance. On this
basis, the designed SACBSS slider was optimized by considering energy
consumption and structural performance. By analyzing the optimization
results, we conclude that the SACBSS slider can effectively reduce the
energy consumption of the power press while meeting the press slider
design standard.
Based on this study, the manufacturing process optimization of the
SACBSS slider is one of the expansion directions of the research. The use
of multiple materials and the complex support plate structure lead to
complexity in the production of the SACBSS slider. It increases the
machining process, leading to higher manufacturing costs. Though the
use of the SACBSS slider reduces the cost of electricity, which helps
offset the rise in manufacturing costs, the increase in the manufacturing
process leads to more carbon emissions during the manufacturing phase.
Therefore, optimization of manufacturing processes (e.g., green pro­
duction of high-strength aluminum alloys, changes in welding methods,
etc.) is necessary to achieve cleaner production.
5. Discussion
The proposed method aims to reduce the energy consumption of the
power press in its operation by reducing the slider’s mass. In this study,
the SACBSS effectively reduced the mass of the slider (by 18.6%),
resulting in a reduction in the operational energy consumption of the
power press (by 6.1%). This reduction in energy consumption is
particularly significant for heavy-duty presses, where the sliders tend to
be bulky. Due to the long service life of power presses, the reduction in
energy consumption during operation is long-term, which reduces
electricity costs and contributes to cleaner production.
However, compared with conventional sliders made of welded steel
plates, the SACBSS slider is more complex to manufacture. The use of
aluminum alloys and the complex support plate structure greatly in­
crease the machining process, which leads to more time and production
costs. In this situation, efficient component machining techniques and
the mass production of products contribute to reducing production cy­
cles and costs.
In addition, the use of aluminum alloys increases carbon emissions
during the manufacturing phase of the slider (see Appendix A). As
shown in Table A3, solely in terms of raw material acquisition, the
SACBSS slider emits 3685.5 kg CO2 more than the steel-welded slider.
However, the emission reduction advantage of aluminum alloys lies in
the transportation, operation of the power press, and recycling stages. In
these three stages, the carbon emissions generated by the SACBSS slider
are lower than those of the steel-welded slider. Meanwhile, over the
entire lifecycle, the use of the SACBSS slider reduces emissions by
511,574 kg CO2. Therefore, the SACBSS slider mitigates the environ­
mental pollution caused by power press production.
The research on the SACBSS slider is still ongoing. The
manufacturing process and its optimization require further investiga­
tion, as they are closely related to production costs and carbon emissions
during the manufacturing phase.
CRediT authorship contribution statement
Feng Liang: Methodology, Formal analysis, Writing – original draft,
Data curation. Rui Wang: Investigation, Software, Writing – review &
editing. Qiu Pang: Validation, Resources. Zhili Hu: Conceptualization,
Funding acquisition, Project administration, Resources, Supervision,
Writing – review & editing.
Declaration of competing interest
The authors declare that they have no known competing financial
interests or personal relationships that could have appeared to influence
the work reported in this paper.
Data availability
Data will be made available on request.
Acknowledgements
This study was funded by the National Key Research and Develop­
ment Program of China (grant number 2022YFB3706900); the National
Natural Science Foundation of China (grant numbers 52075400,
52275368); Independent Innovation Projects of the Hubei Longzhong
Laboratory (2022ZZ-04); the 111 Project (grant number B17034), and
the Key Research and Development Program of Hubei Province (grant
number 2021BAA200).
6. Conclusion
In this work, a design and optimization method of the slider for
Appendix A
The carbon emissions generated from raw material acquisition can be determined by Eq. (A.1) (Zhang et al., 2016).
(A.1)
CEacq = Mm × CEFacq
where Mm represents the mass of the material and CEFacq stands for the carbon emission factor of the material.
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F. Liang et al.
Journal of Cleaner Production 428 (2023) 139341
The power press is transported from the manufacturer to the customer using diesel trucks. Assuming that the transportation distance is 500 km, the
carbon emissions generated by the slider during the transportation of the power press can be calculated using Eq. (A.2) (Li et al., 2021).
(A.2)
CEtran = Ms × CEFfuel × EIengine
where CEFfuel is the fuel emission coefficient and EIengine is the energy intensity of the diesel engine.
Electricity is the energy source used during the operation of the power press. Therefore, the carbon emissions during this operational process can be
derived from electricity consumption. Taking the 2500 kN press in this paper as an example, assuming a lifespan of 30 years (365 days per year) with
daily operation for 10 h, the carbon emissions during the operational phase can be obtained from Eq. (A.3).
(A.3)
CEopt = P × t × CEFelec
where P is the motor power, t is the time, and CEFelec is the electricity consumption coefficient.
The carbon emissions during the recycling process can be obtained by Eq. (A.4) (Li et al., 2015).
(A.4)
CErec = Mm × CEFrec
where CEFrec represents the carbon emission factor during material recycling.
In the optimized SACBSS slider, the mass of aluminum alloy is 520.6 kg, and the mass of steel is 4757.8 kg. Tables A1 and A2 present the numerical
values of the parameters used in the calculations. Due to the primary focus of this study on reducing energy consumption during the operation of the
power press, carbon emissions from processes such as material heat treatment and material processing have been overlooked. These aspects will be the
focus of future research. The carbon emission data for the four stages are presented in Table A3.
Table A.1
The carbon emission factor of the material (Li et al., 2015)
Material
Steel
Aluminum
Carbon emission factors of material acquisition (kg CO2/kg)
Carbon emission factor of material recycling (kg CO2/kg)
2.69
0.361
16.13
0.256
Table A.2
The numerical values of the parameters in the study
Value
Energy intensity of the diesel engine (L/t⋅km)
Fuel emission coefficient (kg CO2/L)
Electricity consumption coefficient (kg CO2/kWh)
0.065 (He et al., 2015)
2.76 (Li et al., 2021)
0.94 (Li et al., 2021)
Table A.3
Comparison of carbon emissions between the SACBSS slider and the steel-welded slider
Raw material acquisition (kg CO2)
Transportation (kg CO2)
Operation of the power press (kg CO2)
Material recycling (kg CO2)
Total (kg CO2)
Steel-welded slider
SACBSS slider
17,510.2
583.9
5,661,150
2349.9
5,681,594
21,195.7
473.5
5,146,500
1850.8
5,170,020
Appendix B
Table B.1
Dataset of the BBD experimental design
Name
x1
x2
x3
x7
x9
x14
x15
M
Umax
f1
σmax
1
2
3
4
5
6
7
8
9
10
30
25
25
35
35
25
25
35
35
30
20
10
10
10
10
30
30
30
30
10
30
30
30
30
30
30
30
30
30
25
600
600
600
600
600
600
600
600
600
520
30
30
30
30
30
30
30
30
30
30
145
110
180
110
180
110
180
110
180
145
30
30
30
30
30
30
30
30
30
30
5540.1
5194.0
5247.7
5544.5
5598.1
5482.3
5535.9
5832.7
5886.4
5277.6
0.204
0.216
0.204
0.212
0.199
0.212
0.200
0.208
0.196
0.227
100.24
72.95
72.94
72.95
72.95
85.79
85.92
113.67
113.88
72.46
49.3
56.8
46.0
52.1
46.1
57.4
45.3
54.8
44.4
65.0
(continued on next page)
17
F. Liang et al.
Journal of Cleaner Production 428 (2023) 139341
Table B.1 (continued )
Name
x1
x2
x3
x7
x9
x14
x15
M
Umax
f1
σmax
11
12
13
14
15
16
17
18
19
20
21
22
23
24
25
26
27
28
29
30
31
32
33
34
35
36
37
38
39
40
41
42
43
44
45
46
47
48
49
50
51
52
53
54
55
56
30
30
30
30
30
30
30
25
25
35
35
25
25
35
35
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
25
25
25
25
35
35
35
35
30
30
30
30
30
30
30
10
30
30
10
10
30
30
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
10
10
10
10
30
30
30
25
25
25
35
35
35
35
25
25
25
25
35
35
35
35
30
30
30
30
30
30
30
30
25
25
25
25
35
35
35
35
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
680
520
680
520
680
520
680
600
600
600
600
600
600
600
600
520
680
520
680
520
680
520
680
600
600
600
600
600
600
600
600
520
680
520
680
520
680
520
680
600
600
600
600
600
600
600
30
30
30
30
30
30
30
20
40
20
40
20
40
20
40
20
20
40
40
20
20
40
40
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
20
40
20
40
20
40
20
145
145
145
145
145
145
145
145
145
145
145
145
145
145
145
110
110
110
110
180
180
180
180
110
110
180
180
110
110
180
180
145
145
145
145
145
145
145
145
145
145
145
145
145
145
145
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
30
20
40
20
40
20
40
20
40
20
20
40
40
20
20
40
40
20
20
40
40
20
20
40
5277.6
5565.8
5565.8
5514.4
5514.4
5802.6
5802.6
5215.1
5278.2
5565.5
5628.6
5451.9
5515.0
5802.3
5865.4
5477.9
5477.9
5549.2
5549.2
5539.6
5539.6
5594.6
5594.6
5352.2
5423.5
5356.0
5527.0
5589.0
5660.3
5592.8
5763.8
5297.2
5297.2
5418.3
5418.3
5647.6
5647.6
5768.8
5768.8
5296.9
5360.0
5418.0
5481.1
5585.1
5648.2
5706.2
0.216
0.222
0.213
0.212
0.204
0.208
0.201
0.238
0.195
0.233
0.190
0.226
0.184
0.220
0.180
0.257
0.244
0.208
0.191
0.228
0.227
0.194
0.187
0.232
0.209
0.235
0.187
0.219
0.198
0.221
0.177
0.243
0.230
0.206
0.200
0.237
0.225
0.201
0.194
0.255
0.210
0.216
0.178
0.251
0.206
0.213
72.46
98.04
97.60
73.47
73.46
102.70
102.43
83.73
83.75
110.86
110.89
87.69
87.72
116.27
116.26
100.32
99.95
100.35
99.98
100.49
100.13
100.51
100.14
97.58
97.76
97.61
98.08
102.34
102.51
102.34
102.74
85.82
85.60
85.97
85.73
113.56
113.09
113.99
113.53
72.94
72.95
72.93
72.96
100.12
100.13
100.40
49.9
61.2
49.5
57.2
46.5
53.7
42.5
56.3
52.6
53.4
49.5
50.8
46.9
49.8
44.3
81.7
47.1
54.8
47.8
60.0
45.2
50.0
43.4
59.5
55.6
52.4
45.9
50.0
51.5
47.2
42.4
60.8
50.2
65.5
50.1
60.1
49.9
64.3
41.5
58.3
51.9
52.1
45.9
55.8
49.7
54.7
Table B.2
Fitting coefficient of indicators
Term
y1 (M)
y2 (f1 )
y3 (Umax )
y4 (σmax )
x1
x2
x3
x7
x9
x14
x15
x21
x22
x23
x27
x29
x214
x215
x1 x2
x1 x3
x1 x7
x1 x9
x1 x14
x1 x15
x2 x3
175.21
144.13
118.41
0
31.56
26.81
60.56
− 1.410E-003
− 1.410E-003
− 1.410E-003
2.819E-003
0.12
0.094
− 7.13
0
0
0
0
0
0
0
11.59
13.58
2.05
− 0.15
9.614E-003
0.069
0.12
− 0.44
− 13.50
− 0.13
0.055
− 0.051
− 0.032
− 0.095
6.97
0.37
− 0.027
7.850E-003
0.020
0.070
0.94
2.451E-003
− 1.964E-003
− 6.132E-003
− 4.388E-003
− 0.022
− 6.022E-003
− 0.017
4.024E-004
6.093E-004
3.215E-004
7.811E-003
3.464E-003
6.729E-004
4.621E-003
− 7.948E-005
7.871E-005
− 5.159E-005
2.188E-004
1.783E-005
7.935E-005
8.593E-005
− 0.49
− 0.16
− 3.43
− 8.15
− 3.35
− 3.84
− 2.44
0.052
0.20
0.33
3.30
0.39
− 0.15
0.24
− 0.10
− 0.69
− 0.31
0.45
− 0.032
1.13
0.18
(continued on next page)
18
F. Liang et al.
Journal of Cleaner Production 428 (2023) 139341
Table B.2 (continued )
Term
y1 (M)
y2 (f1 )
y3 (Umax )
y4 (σmax )
x2 x7
x2 x9
x2 x14
x2 x15
x3 x7
x3 x9
x3 x14
x3 x15
x7 x9
x7 x14
x7 x15
x9 x14
x9 x15
x14 x15
0
0
0
0
0
0
0
0
0
0
0
− 4.07
0
24.94
5540.11
− 0.088
− 2.621E-003
0.040
0.070
0.046
3.603E-003
− 9.177E-003
− 0.011
− 4.855E-004
− 2.546E-003
− 3.644E-003
− 1.332E-003
− 5.828E-004
0.064
100.26
3.039E-004
− 1.506E-005
6.553E-005
1.548E-004
6.281E-004
6.512E-004
1.925E-004
6.515E-004
− 3.935E-004
1.796E-003
1.506E-003
4.104E-003
1.721E-003
− 5.968E-003
0.20
− 0.17
− 0.30
0.29
− 0.66
0.70
− 0.021
0.76
0.015
4.02
2.83
− 3.45
1.44
0.086
− 0.75
50.02
ε
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