Keywords

1 Introduction

According to Thomas et al. [1], a hydraulic press is a machine press which uses hydraulic cylinder to generate a compressive force to perform various pressing operations such as metal forging, punching, stam**, etc. The press provides an efficient means of pushing and pulling, rotating, thrusting and controlling load [2, 3]. Some hydraulic press applications include compression moulding, injection moulding, drawing, forging, blanking, coining, clam**, compacting, Forming, pad forming, potting, punching, and stacking, bending, stam** and trimming [4, 5]. The use of hydraulic cylinders for controls boasts of cost-effectiveness, high rate of production, positive response to changes, ease of control of parameters and primarily suitable when a heavy workpiece is to be machined [3, 6, 7]. Other advantages include tonnage adjustment and cycle time maximisation [8, 9]. According to Maneetham and Afzulpurkar [10], Hydraulic Servo Systems (HSS) have been used in many modern industrial applications by their small size to power ratios and their ability to apply considerable force and torque.

On the other hand, by using a simulation model for hydraulic systems, the dynamic performance of these systems may be validated in the absence of actual hardware, which is accomplished via the use of specialised modelling and simulation tools [11,12,13,14]. In addition, the bending forces and moment can easily be predicted using simulation tools to determine the magnitude of stain, buckling and distortion [15,16,17]. This will enhance the use of hydraulic cylinders with sufficient clam** force that ensures adequate strength without distortion. This study aims to design a reconfigurable press brake assembly with hydraulic cylinders for holding a workpiece and adjusting the ram height during machining operations on a press brake. This is to enhance adequate clam** and precision during manufacturing operations. Despite productivity gains achieved through automation of design routines and manufacturing tasks, the authors Kumar et al. [18] and Ulah et al. [19] report that nearly 85% of all fixture processes and design plans are still performed manually, and detailed optimisation plans are rarely created. The interchangeability of parts is critical to the successful operation of any mass production facility because it allows for quick assembly and lower unit costs. Mass production methods demand fast and easy positioning for accurate operations [20, 21]. When designing jigs and fixtures, the strength of the clamp should be sufficient to hold the workpiece firmly in place and to withstand the strain of the cutting tool without springing [22, 23]. When producing large quantities of different materials on a large scale, a significant amount of time is spent setting up the device and clam** it [24, 25]. According to Pachbhai and Raut [26] as well as Daniyan et al. [27], hydraulic cylinders instead of manual adjustment are characterised by quick and automatic adjustment, greater accuracy, high productivity, consistent performance clam** force, and repeatable clamp location. Computer-aided design, modelling, and simulation tools have been used to improve the development of fixtures. For instance, Ruksar et al. [28] carried out the FEA and optimisation of machine fixtures, while Wang [29] applied a polynomial fit-based simulation method in a hydraulic actuator control system. Shrikant and Raut [30] employed computer-aided design for fixture development. It is necessary to reconfigure existing machines to have efficient work holding capacity to increase overall productivity, location accuracy, and surface finish quality of the finished product. The study aims to design the locating, supporting and clam** methods for a reconfigurable press brake using hydraulic controls.

2 Methodology

This paper proposes a six-cylinder automated hydraulic brake press. The press brake is constructed with a balanced frame, conveyor rollers and support, belt, chuck, and six hydraulic cylinders that are bolted and nutted together. Solidworks was used to design and model the fixture. According to Khurmi and Gupta [31], the maximum distortion energy theory for yielding is expressed as Eq. 1.

$$(\sigma t_{1} )^{2} + (\sigma t_{2} )^{2} - 2\sigma t_{1} \times \sigma t_{2} = \left( {\frac{{\sigma t_{1} }}{FS}} \right)$$
(1)

where: σt1 is the maximum principal stress (N/m2); σt2 is the minimum principal stress (N/m2), and σyt is the stress at yield point (N/m2); F.S is the factor of safety. The maximum and minimum principal stress calculated from Von mises stress analysis is given as 2.39365 × 105 N/m2 and 5.44655 × 107 N/m2 respectively. The volumetric parameters for the entire model are given as mass: 442.634 kg, volume: 0.056748 m3, density: 7800 kg/m3, and weight: 4337.82 N. Buckling is a possibility in the lower beam, which is the area where the fixture is loaded. The analytical results are compared to those obtained from the FEA simulation to determine the likelihood of buckling. With a length of 150 mm, the support for the tested section flexural rigidity equals 6.6 × 10–6. Nm−2. This is calculated as the modulus of elasticity and moment of inertia for the section under consideration. As a result, the buckling force is represented by Eq. 2.

$$F_{b} = \frac{{El \cdot \pi^{2} }}{{L_{c}^{2} }}.$$
(2)

Fb is the buckling force (N), EI is the flexural rigidity 6.6 × 106 Nm−2, and Lc is the effective length (m). Since both ends are pinned, the effective length equals the actual length. Hence,

$$L_{c} = L$$
(3)

The model of the designed fixture assembly and the assembly drawing are shown in Fig. 1.

Fig. 1.
figure 1

The model assembly drawing of the fixture.

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From Eq. 2, the buckling force is calculated as 2.8958 × 109 N. Due to the fact that the section only has to support a resultant load of 499.716 N, the buckling force exceeds the resultant force, thus, giving no chances for buckling. The design was based on the maximum tonnage of the press brake, which is determined by the material type, thickness, length, and method of bending and clam**. When performing Von Mises stress analysis, failure or yielding occurs at a point in a member where the distortion strain energy is most significant [31, 32]. Furthermore, according to the results of a simple tension test, the shear strain energy per unit volume in a bi-axial stress system reaches the limiting distortion energy at the yield point per unit volume at the yield point.

The area of the piston is expressed by Eq. 4.

$$A = \frac{{\pi D^{2} }}{4}$$
(4)
$$A = \frac{{3.142D^{2} }}{4} = 0.7854\ d^{2} m^{2}$$

where ‘D’ is the internal diameter of the piston-cylinder (m). The stress-induced is expressed by Eq. 5.

$$\sigma = \frac{F}{A}$$
(5)

‘F’ is the force applied (N), and ‘A’ is the piston cross-sectional area (m2). Introducing the maximum stress given as 2.47903 × 105 N/m2, reaction force 69.4426 N calculated from Von mises stress analysis and cross-sectional area calculated from Eq. 2 as 0.7854 d2m2 into Eq. 3; we have

$$\begin{gathered} 2.47903 \times 10^{5} = \frac{69.4426}{{0.7854\ d^{2} }} \hfill \\ d = 0.0188\,{\text{m}}\,\,\,\,\,or\,\,18.8\,{\text{mm}} \hfill \\ \end{gathered}$$

Using a safety factor of 2 and correcting to the nearest standard size, the piston diameter is calculated as 40 mm; therefore, the area is calculated as

$$0.7854 \times 0.04^{2} = 1.26 \times 10^{ - 3} {\text{m}}^{{2}}$$

The piston will be subjected to shear stress; hence its thickness should be sufficient to resist failure by shearing. The minimum thickness of the piston required to resist shearing is given by Eq. 6.

$$t = \frac{pd}{{2\sigma }}$$
(6)

where: d is the internal diameter of the piston-cylinder is 0.04 m, σ is the maximum allowable stress (7.23826 × 108 N/m2) and ρ is the pressure in the cylinder (2.47903 × 105 N/m2), and thickness is calculated as 0.006849 m. Using a safety factor of 2, the thickness is calculated as 0.015 m to the nearest standard thickness. The volumetric properties of the hydraulic cylinder are as follow; mass: 0.618573kg; volume: 8.03342e-005 m3; density: 7700 kg/m3 and weight: 6.06202 N.

2.1 Computer Aided Modelling and Simulation

The modelling and simulation for the two components under investigation (hydraulic cylinder and assembly fixture) were carried out in the Solidworks 2018 environment. The study type is to investigate the stress, displacement and strain of the hydraulic cylinder and fixture analysis. The the linear elastic isotropic model type and the Von mises failure criterion was used to determine the stresses induced in the component member, resultant loads and the corresponding displacements. The general standard static analysis of the finite element modelling was set up for the model analysis. From the material database, the mechanical properties of the materials selected for the hydraulic cylinder and assembly fixture (stainless steel 304 and cast alloy steel, ASTM A216) were selected. This was followed by the free body model of the components and the assignment of the loading conditions vis-à-vis the service requirements. Next is the dicretisation of the model. This is to mesh the developed models into finite elements and the application of the mesh control. The properties of stainless steel 304 employed for the design of the hydraulic cylinder are presented in Table 1 while Table 2 presents the mechanical properties of the cast alloy steel employed for the design of the fixture model.

Table 1. Mechanical properties of stainless steel 304 [33]
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Table 2. Properties of cast alloy steel (ASTM A216) for the fixture model [34].
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The linear elastic isotropic model type was selected and the Von mises failure criterion was employed for the failure analysis. A mesh size of 2 mm was employed in the Solidworks environment to mesh model into finite elements.

Using a mesh interval of 0.2 mm, it was observed that the computational time decreases with an increase in the mesh size up to 2.0 mm for the hydraulic cylinder. Further increase in the mesh size up to 2.6 mm resulted in a slight increase in the computational time. Hence, the mesh size of 2.0 mm which produced the least computational time (152 s) was selected for the hydraulic cylinder. For the fixture, it was observed that the computational time decreases with an increase in the mesh size up to 2.2 mm. Further increase in the mesh size up to 2.6 mm resulted in a slight increase in the computational time. Hence, the mesh size of 2.0 mm which produced the least computational time (367 s) was selected for the fixture (Fig. 2).

Fig. 2.
figure 2

The mesh time and the corresponding computational time.

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2.2 Model Control of Triplet Cylinder

The category of the technical properties used to control the component variables can be assigned to other variables in the “read” or “write” mode for sending or receiving control signals. The driving force is an assignable variable to apply a driving force to the component. If there is not enough pressure, this force will drive the rod-piston assembly. The curve defining the external driving force is expressed in terms of the percentage of the cylinder position. From 0% to100%, the force is applied during the extension of the cylinder until the cylinder reaches the end of its stroke. Once the end of stroke is reached, the curve used will be in the −100% to 0% quadrant. Between 0% and 100%, if the read value is positive and there is not enough pressure to oppose, the rod-piston assembly will retract; inversely, but if the force is opposing, it will extend. Between −100% and 0%, if the value of the force is positive and there is not enough negative pressure, the piston-rod assembly will retract and extend if negative. This curve is of null value by default—external mass assignment variable of mass to allow the dynamic change of the mass during simulation. The default unit of the variable is the kilogram. The resistive force assignable variable is to apply a resistive force to the component. This force is resistive and will oppose the displacement of the piston-rod assembly. The curve that defines this force is expressed in terms of the percentage of the cylinder position. When the cylinder is extending, the curve is read in the 0–100% quadrant; inversely, the force will be read between −100 to 0% quadrant when the cylinder retracts. The value of this force can only be positive by convention. Figure 3 presents the mechanical working principles of the double-acting triplet cylinder.

Fig. 3.
figure 3

Mechanical working principles of the double-acting triplet cylinder.

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2.3 The Operational Model of the Triplet Cylinder

The hydraulic press brake utilises mechanically connected cylinders that operate in parallel. Linear actuators are devices that convert fluid energy to mechanical energy. As the name implies, the linear actuators will deliver the powers straight. In fluid power systems, linear actuators are often available with various components attached to the end of the rod. Mechanical linkage, levers or cables can be attached to the cylinder to transform the force in the type of movement wanted—technical modelling category of the properties that affect the components simulation model. The drop-down list options allow to edit other parameters or enable/disable the performance curve modelling. For the operating condition, the category of the properties relates to the components operating conditions, especially those that describe its operation limits. Most of these properties assess a faulty component and automatically trigger a failure, thus, activating the respective simulation option as “Automatic Failures”. The maximum force that can be applied to the component, or the maximum force range that the directional valve command can apply in proportional operation mode can be selected. The maximum pressure supported by the component supposes the option Monitor Faulty Components is activated in the simulation options. In that case, a visual warning will be displayed next to the component to inform the user that the value is exceeded during the simulation. If the option “Automatic Failure Trigger” is activated in the troubleshooting branch, the user will trigger a failure when this maximum value is reached during simulation. The failure must first be declared and selected, which will only be triggered with this property. The maximum distance travelled by piston per unit time is the distance moved by the piston from one end of the cylinder to the other end. Suppose the option “Monitor Faulty Components” is activated in the simulation options, a visual warning will be displayed next to a component to inform user that the value is exceeded during simulation. If the option “Automatic Failure Trigger” is activated in the troubleshooting branch, the user will trigger a failure when this maximum value is reached during simulation. The failure must first be declared and selected. It will then only be triggered with this property.

Figure 4 shows the model design of the automatic and manually operated triple cylinders.

Fig. 4.
figure 4

Model design of the automatic and manually operated triple cylinders.

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3 Results and Discussion

The result of the simulation of the hydraulic cylinder using Solidworks and the linear elastic isotropic model type and the Von mises failure criterion is presented in Tables 3 and 4 as well as Fig. 5. While Table 3 summarises the reaction forces and moments, Table 4 and Fig. 5 present the strain, stress and displacement analysis. The resultant force from Table is 69.4426 N with the highest reaction experienced along the vertical axis (Y-axis).

Table 3. Reaction forces and moments.
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It can be seen in Table 4 that the deformation per unit length is negligible, if not completely non-existent. It indicates that the clam** force is sufficient to prevent distortion in this particular instance. Furthermore, the stress-induced is minimal, and the cylinder will not yield to the applied force due to this stress.

Table 4. Strain, stress and displacement analysis for the hydraulic cylinder.
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Fig. 5.
figure 5

(a) Strain analysis of the hydraulic cylinder (b) Von mises stress analysis of the hydraulic cylinder (c) Displacement analysis.

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Figure 5 (b) shows the modelling result of the stress-induced in the hydraulic cylinder due to machining. The maximum stress induced is 2.47903 × 105 N/m2 while the minimum is 4.48733 × 10–7 N/m2. From Fig. 5(c), the maximum relative displacement of the cylinder from its mean position is 0.000114313 mm. Comparing the magnitude of the maximum stress induced in the cylinder to the yield strength of the material (2.40 × 108 N/m2), then it can be concluded that the material is not likely to fail under the required service condition.

Table 5 presents the summary of the reaction forces and moments for the fixture. The resultant reaction force obtained from Solidworks simulation is 499.716 N. This force is insufficient to produce any bending as the resultant bending moment is zero.

Table 5. Reaction forces and moment.
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The summary of results of the simulations for the strain, stress and displacement analyses are presented in Table 6 and Fig. 6 for the fixture. The maximum and minimum strains were found to be 5.58621 × 10–7 and 2.26322 × 10–16 respectively. Both are negligibly insignificant. In this case, the fixture orientation does not change significantly while the bending operation is being performed.

Table 6. Strain, stress and displacement analysis for the fixture.
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From Fig. 6a the maximum strain is 5.5862 × 10–7 while the minimum is 2.2633 × 10–16. For the entire fixture model, the maximum stress induced is 2.26322 × 105 N/m2 while the minimum is 5.44655 × 10–7 N/m2 (Fig. 6b). From Fig. 6c, the maximum relative displacement of the cylinder from its mean position is 8.77685 mm. As shown in Fig. 6c, the front beam has a larger displacement while the beam along the neutral plane has a smaller displacement. This is due to the fact that at the neutral plane, the beam is not under the influence of any stress either compressional or tensional stress. The Von Mises analysis also revealed that the maximum and minimum stress-induced are 2.394 × 103 N/m2 and 5.447 × 10–7 N/m2 respectively. The maximum stress induced is lower than the yield strength (5.56 × 108 N/m2) of the cast alloy from which the fixture was designed (ASTM A216). As a result, the stress induced by bending may not cause the cast alloy to yield. The tensile strength is also sufficient to withstand bending forces without displacement or distortion as the maximum value of displacement is 8.77695 mm.

Fig. 6.
figure 6

(a) Strain analysis of the entire fixture (b) Stress analysis of the entire fixture (c) Displacement analysis of the entire fixture

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Figure 7 shows the hydraulic circuit comprising a configurable 3/n way valve with three connections and negligible hydraulic resistance and the 4/n way valve with four connections. It also comprises a double-acting cylinder with a shock absorber at the stroke end. The connected pressure loads control the cylinder piston while the shock absorber can be adjusted using two adjustable screws. The piston of the cylinders contains a permanent solenoid that can be used to operate a proximity switch. The diameter of the piston is 20 mm with a maximum stroke length of 200 mm. The tank is a part of the pump unit and is integrated into it. To reduce the risk of damaging the component, the filter with negligible hydraulic resistance limits the amount of contamination in the fluid. The pump unit delivers a volumetric flow, with the operating pressure being limited by an internal pressure relief valve within the pump units housing. There are two tank connections on the pump. In addition, the relief valve is included in the circuit, which is closed in the normal position. Assume that the operating pressure has been reached at one of the end openings, the other opening opens when the pressure falls below the current level. The valve closes with the pilot pressure generated by the input pressure, resulting in the valve being closed again. It also has a pilot stage and the main stage; when the pilot stage is open. Thus, there is less volumetric flow through it than when the main stage is closed.

Fig. 7.
figure 7

The hydraulic circuit

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Figure 8 shows the variation in force as the piston position varies. The maximum force at a piston position of 46 mm is 0.58 kN. The magnitude of the force applied decreases with an increase in the piston position. At a maximum stroke length of 200 mm, the magnitude of the force becomes negligibly small. The simulation result from the AUTOMATION STUDIO Fluid sim is in agreement with the FEA simulation, which calculates the resultant reaction force as 0.499716 kN. This force is insufficient to produce any bending, strain or displacement, which confirms conclusively that the designed hydraulic brake press has sufficient strength to withstand bending stresses and forces without distortion.

Fig. 8.
figure 8

Change in force with piston position.

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4 Conclusion

A reconfigurable hydraulic press brake was designed using Solidworks and simulated on a hydraulic AUTOMATION STUDIO Fluidsim. The maximum strain, stress, and displacement values obtained from manual Solidworks simulation and Von mises stress analysis were found to be 8.29 × 10–7 N/m2, 2.48 × 103N/m2 and 0.000114 mm respectively. The hydraulic cylinder boasts greater efficiency than manual means or the use of a jack. It facilitates quick adjustment and greater accuracy in equipment and workpiece setting as the ram retreats automatically, and the machine is quickly returned without any waste of time. The results indicate that the hydraulic cylinder actuator can sufficiently withstand the machining forces while providing sufficient strength and rigidity during machining operations. Hence, the reconfigured actuator system possesses efficient work holding capacity without sacrificing rigidity and stiffness. Results obtained from FEA simulation when compared with the mechanical properties of the hydraulic press brake indicate that the reconfigurable hydraulic press brake possesses adequate strength to prevent buckling, strain, distortion, and displacement. Future work can consider the development of the designed hydraulic press brake.