Simulation Analysis and Key Performance Index for Experimental Verification of a New Type of Press Transmission Mechanism

Abstract

In this paper, the characteristics of special stamping process requirements and the shortcomings of existing application technology are studied. Through repeated motion simulation analysis and size optimization calculation, a kind of inertia-controlled molding press is proposed to make up for this technical vacancy. This paper first describes the working principle and structure scheme of the main transmission mechanism of the inertia-controlled molding press. With the help of Adams-View x64 2013, the main transmission mechanism is simulated and analyzed from the aspects of kinematics and dynamics, and a physical prototype is made to test the key reliability indexes of the main transmission mechanism. According to the test data, the static strength, stiffness, vibration characteristics, and dynamic characteristics of the 200 t inertia-controlled molding press are evaluated, which provides a reference for the design of this kind of machine tool.

1. Introduction

At present, there is a trend of designing the stamping parts of automobiles and home appliances with less height and weight to meet energy-saving and cost-reduction requirements, which promotes the thinning and lightening of some stamping components related to energy consumption. To make these parts thinner while also guaranteeing their strength, using high-strength steel plates to form these components becomes an inevitable choice. However, when molding a high-strength steel plate, there is a high risk of spring-back deformation after the bottom has been stamped. It is necessary to restrain this spring back deformation to ensure qualified parts. For this type of stamping process, it is very unsuitable to use an ordinary crank press. It is essential to keep the slider under pressure for a while after the bottom has been stamped to suppress the rebound deformation so that the parts can release the spring-back potential energy in the mold to allow full formation [1].

Firstly, we analyzed the working principle of the transmission mechanism and simulated and analyzed the motion characteristics of the transmission mechanism using ADAMS software to verify the design results [2]. Based on theoretical analysis, a physical prototype was designed and manufactured, and the relevant performance indicators of the physical prototype were experimentally verified.

In state-of-the-art technology, there are usually three ways to meet the technical requirements of the slider for slow stamping and pressure retention near the bottom stop: The first method is to use an oil hydraulic press for formation, which can cause the slider to maintain pressure for a period of time after the stamping and then return to its position. However, this method has two shortcomings. First, due to the compressibility of the hydraulics, the oil cylinder cannot push the slider to the bottom stop as fully and thoroughly as the mechanical press while the hydraulic press maintains pressure at the bottom stop. Thus, deficient deformation of the high-strength steel plate is inevitable at the bottom stop when molding high-strength steel parts. Second, the low efficiency of the hydraulic press will greatly reduce the production efficiency of the entire production line. The second method is to use a crank press and a complex multi-link mechanism to achieve the characteristics of the pressure-maintaining molding of the slider at the bottom stop [3]. This kind of mechanism has a good slow stamping effect, but the duration of continuously maintained pressure is very short. Sometimes, in order to guarantee sufficient holding time, the production efficiency has to be sacrificed, and the number of strokes has to be reduced to increase the cycle time and the pressure-maintaining time. Due to the more complicated design of this type of mechanism, more geometric tolerance, more difficult processing, and larger integrated cumulative clearance of the transmission chain, the geometric requirements for assembly will be higher. It is hard to guarantee assembly accuracy, so manufacturing is relatively less economical with higher costs [4]. Furthermore, it is extremely difficult to maintain. The third method is to use a direct-drive servo press to control the movement curve of the slider through programming to achieve the process requirements of slow bottom-stop stamping and pressure retention. However, in most cases, for the mass production of parts, the slider working curve is almost fixed. There is no need to change the process curve frequently [5]. The use of a direct-drive servo press will increase the unit manufacturing cost of the parts and become less economical. Therefore, it is very hard to promote and popularize the use of this type of machine tool. Moreover, it is necessary to use the motor stalling torque to achieve the slow stamping and pressure-maintaining molding process near the bottom stop, which has a requirement of a long peak torque duration for the servo motor. Continuous stalling work will cause the motor to heat up easily and thus shorten its service life significantly, which is neither economical nor reliable [6].

The appropriate rebound compensation coefficient can be identified through a stamping simulation and combined with the programmable control of the servo press to set the most suitable process curve that suppresses rebound and ensures shape and size precision. To sum up, previous methods have corresponding deficiencies in this kind of special process requirement and cannot be well promoted and applied. After repeated motion simulation analysis and size optimization calculation, this paper proposes an inertia-controlled molding press to make up for this technical gap [7].

2. Working Principle and Scheme Structure

2.1. Working Principle

Figure 1 shows the schematic diagram of the main transmission mechanism of the inertia-controlled molding press. The working principle is as follows: the transmission part drives the rotation of the eccentric wheel following the swinging of the pendulum. Meanwhile, the center of mass of the pendulum is always up and down, reciprocating linearly. The other end of the pendulum is connected to the slider through the mandrel. The slider follows the pendulum to move in the same way [8]. Its structural characteristics are as follows: the eccentricity of the eccentric wheel is e, the length of the swing rod is L, and the stroke of the press is H = 2e, where e < L must be satisfied to establish the motion relationship. On the premise of satisfying the rod length relationship, different slider motion curves can be obtained according to the design of different lengths of the connecting rod L; that is, different sliding process characteristic curves can be obtained.

Changing the length of the pendulum rod can greatly slow down the speed of the slider during downward punching and maintain the pressure for a period of time at the bottom point so that this type of mechanism has the characteristics for perfect slow stamping and pressure-maintaining processes [9].

2.2. Scheme Structure

In Figure 2a, the main transmission mechanism of the inertia-controlled molding press is shown. In order to ensure the balance of force and smooth transmission of the mechanism, the layout of the whole machine adopts the symmetrical design structure of an eccentric gear and eccentric wheel. The main transmission part mainly includes a large eccentric gear, an eccentric wheel, a swing rod, a mandrel, a connecting rod, a pinion shaft, a flywheel, a clutch, a synchronous belt, a small synchronous wheel, and a low-power permanent magnet synchronous servo motor [10].

The structural layout is as follows: the small synchronous wheel is fixed on the output shaft of the servo motor. The small synchronous wheel and the flywheel are driven by a synchronous belt. The flywheel is installed on the pinion shaft, which is engaged with the eccentric gear. The eccentric gear is installed on the fuselage through the fuselage tile and is rotationally connected to the pin shaft. The pin shaft is fixedly connected to the swing rod [11], which is rotationally connected to the mandrel. The mandrel is fixedly connected to the connecting rod. The slider is fixedly connected to the connecting rod, and the connecting part of the slider is designed with a hydraulic overload protection structure.

It should be noted that the structural characteristics of the eccentric gear and the eccentric wheel in the mechanism are unique. Unlike the conventional eccentric gear press, the rotation center of the eccentric wheel part and the large gear are concentric, as shown in Figure 2b.

During the operation process, the servo motor transmits power from the synchronous wheel to the flywheel through the synchronous belt. The on/off status of the power transmission between the flywheel and gear shaft is controlled by the clutch. The pinion shaft drives the large eccentric gear to rotate, which in turn drives the swing lever to swing and performs an up-and-down reciprocating linear motion. The mandrel follows the pendulum to also perform an up-and-down reciprocating motion. Then, the connecting rod fixed on the mandrel follows the mandrel to move in an up-and-down reciprocating motion, which causes the slider fixed with the connecting rod to move in the same way [12].

The slider moves downward to achieve stamping and molding. During the working process, due to the work required by the stamping workpiece, the flywheel speed is forced to decrease from ${\mathsf{\omega }}_{\mathrm{t}0}$ to ${\mathsf{\omega }}_{\mathrm{t}1}$, and the work size is $\mathrm{A}=\frac{1}{2}\mathrm{J}({{\mathsf{\omega }}_{\mathrm{t}0}}^2-{{\mathsf{\omega }}_{\mathrm{t}1}}^2)$, where J is the flywheel inertia; ${\mathsf{\omega }}_{\mathrm{t}0}$ is the initial angular velocity of the flywheel, and ${\mathsf{\omega }}_{\mathrm{t}1}$ is the angular velocity of the flywheel after doing work. The conventional press achieves a speed reduction via the pulley slipping, which releases rotational kinetic energy. The transmission system described in this paper enables a synchronous servo motor to drive the flywheel through the synchronous belt wheel. Unlike a conventional press, it is unable to realize speed reduction via pulley slipping, and thereby unable to function.

The implementation principle of the work structure of this scheme is that the flywheel will decelerate instantly when the machine tool is working. Meanwhile, by controlling the servo motor, the flywheel’s synchronous belt reversely causes the servo motor to slow down instantly to function correctly. If continuous functioning is required, then extending the deceleration time is simply needed. During the return stroke of the slider, the synchronous servo motor will accelerate to rotate, causing the flywheel’s angular velocity to reach the initial value and replenish the lost rotational kinetic energy of the flywheel, so that the stamping function is realized cyclically [13]. In the above process, the servo motor needs to be accelerated and decelerated and the speed of the flywheel needs to be changed to control the inert kinetic energy.

The servo motor is used to change the speed of the slider during operation by controlling the movement of the flywheel and the passive inertia components through the synchronous belt. It can optimize the slider motion curve and special process requirements such as lower stamping speed, longer pressure retention time, and shorter return time, etc. The implementation method is the built-in PLC, which generates a smooth speed curve under the speed control mode and solidifies its motion curve into the servo drive. It prevents the servo motor from resisting the over-current alarm by controlling the PID and limiting the torque to achieve machine tool motion control. Figure 3 is a comparison chart of the slider motion curve. The slider speed can be reduced by 20–50% at the beginning of the stamping process and is increased by 30–60% during the return stroke to change the slider motion curve. Thus, an optimal slider motion curve can be achieved to meet the needs of the process characteristics in a better way.

Because the mechanism retains the energy storage part of the flywheel, it does not need to rely entirely on the servo motor to provide energy when doing work. Therefore, the power of the servo motor will be greatly reduced compared to the power of the direct drive servo press, so the manufacturing cost of the machine tool will be significantly reduced. At the same time, the motor is designed to be air-cooled. A special embedded magnetic steel structure is used to achieve a three-time torque overload. The similar slip difference of the motor can reach 30% to 40%, which can achieve greater energy release and stretch molding work requirements. At the same time, in the non-working area of the press, the braking kinetic energy generated during acceleration and deceleration, and part of the flywheel inertial potential energy can be converted into electric energy and stored in the capacitor. When the flywheel kinetic energy needs to be supplemented at the beginning of the next cycle, a part of the electric energy required by the motor can be provided by the capacitor which reduces the amount of electricity drawn from the power network, thereby saving the energy. In addition, the synchronous servo motor does not require the excitation current like the asynchronous motor, resulting in a 20% to 30% decrease in power compared to the conventional press with an asynchronous motor. This makes it more energy-efficient, with an overall improvement in energy efficiency of 25% to 40% [14].

To ensure that the slider follows the planned motion curve, the motor needs to control the flywheel through the synchronous belt to run at the specified rotation angle and planned speed. To achieve the above control effect, the matching design of the rotor inertia of the motor and the equivalent inertia of the passive components, such as the flywheel, is critical and must satisfy the relationship of Equation (1):

$${\mathrm{J}}_{\mathrm{MR}}=(1.5~2){\mathrm{J}}_{\mathrm{E}}$$

(1)

where ${\mathrm{J}}_{\mathrm{MR}}$ is the rotor inertia of the motor and ${\mathrm{J}}_{\mathrm{E}}$ is the equivalent inertia of the flywheel and passive components.

$${\mathrm{J}}_{\mathrm{E}}={\mathrm{J}}_1+{\mathrm{J}}_2\frac{1}{{{\mathrm{i}}_1}^2}+{\mathrm{J}}_3\frac{1}{{({\mathrm{i}}_1{\mathrm{i}}_2)}^2}$$

(2)

In Equation (2), ${\mathrm{J}}_1$ is the inertia of the synchronous wheel of the motor; ${\mathrm{J}}_2$ is the sum of the inertia of the flywheel and the pinion shaft; ${\mathrm{J}}_3$ is the sum of the inertia of the eccentric gear, eccentric wheel, pendulum rod, and slider; ${\mathrm{i}}_1$ is the transmission ratio of the synchronous belt; and ${\mathrm{i}}_2$ is the gear transmission ratio [15].

3. Simulation Analysis of the Main Transmission Mechanism

This paper uses the ADAMS simulation analysis software to build a rigid body simulation analysis model for the main transmission mechanism of the inertia-controlled molding press. The modeling method utilizes the main transmission scheme structure of the inertia-controlled molding press shown in Figure 2a, where the main technical parameters of the machine tool are given. On the premise of ensuring that the simulation analysis is not distorted, for the convenience of analysis, the main transmission mechanism is simplified and modeled. In ADAMS, only a simplified model of the swing rod, eccentric wheel, and the slider of the main transmission mechanism in Figure 2a were built. The three-dimensional modeling of the physical prototype is shown in the Figure 4. It can be clearly seen that the scheme structure of the main transmission mechanism is the primary synchronous belt transmission, primary gear transmission, and the sliding block is connected by a swing rod. Its main characteristic parameters are machine tool nominal force, ${\mathrm{P}}_{\mathrm{g}}=2000\mathrm{KN}$; stroke number, $\mathrm{N}=60\mathrm{spm}$; eccentric wheel eccentricity, $\mathrm{e}=150\mathrm{mm}$; and swing bar length, $\mathrm{L}=225\mathrm{mm}$ [12].

Figure 5 is the motion characteristic curve of the main transmission mechanism obtained by ADAMS simulation. From the curve, we can intuitively see that the cycle time of the slider operation is $\mathrm{t}=1\mathrm{s}$, that is, the number of slider strokes is $\mathrm{N}=60\mathrm{spm}$, the slider stroke is $\mathrm{S}=300\mathrm{mm}$, the maximum linear velocity of the slider is ${\mathrm{V}}_{\max }=1156\mathrm{mm}/\mathrm{s}$, the maximum linear acceleration of the slider is ${\mathrm{a}}_{\max }=9870\mathrm{mm}/{\mathrm{s}}^2$, and the linear acceleration of the slider to the bottom stop is ${\mathrm{a}}_{0.5}=1975\mathrm{mm}/{\mathrm{s}}^2$. When the slider runs from 0.474 s to 0.528 s, the slider runs through the bottom stop at a very slow linear speed, i.e., $0~30\mathrm{mm}/\mathrm{s}$, and the slider can be considered to be in a holding state; the holding time is 0.054 s. The maintained pressure angle near the bottom stop of the slider is 170° to 190°. If stretches start forming at a distance of 15 mm from the bottom stop, the eccentric gear rotation angle is 138° and the linear velocity of the slider is ${\mathrm{V}}_{\mathrm{s}}=280\mathrm{mm}/\mathrm{s}$. The linear velocity of the slider should be at the maximum allowable linear velocity of the material only within a certain range to ensure the quality of stretch-forming [16]. The maximum allowable stamping speed of various materials is different, as shown in Table 1. When the slider reaches a distance of 7 mm from the bottom stop, which is the nominal force stroke ${\mathrm{S}}_{{\mathrm{p}}_{\mathrm{g}}}$, the slider pressure reaches the maximum ${\mathrm{P}}_{\max }=2000\mathrm{KN}$. At this point, the eccentric gear rotation angle is 150°, and the slider linear speed is ${\mathrm{V}}_{\mathrm{tp}}=179\mathrm{mm}/\mathrm{s}$. The linear acceleration of the slider is ${\mathrm{a}}_{\mathrm{t}}=2620\mathrm{mm}/{\mathrm{s}}^2$. If the slider starts to punch a 2 mm sheet, the eccentric gear rotation angle is 165°, the slider linear velocity is ${\mathrm{V}}_{\mathrm{t}}=85\mathrm{mm}/\mathrm{s}$, and the slider linear acceleration is ${\mathrm{a}}_{\mathrm{t}}=2136\mathrm{mm}/{\mathrm{s}}^2$.

From the above simulation analysis results, it can be seen that when the slider reaches the bottom stop, both the linear velocity and linear acceleration gradually decrease. This results in slow stamping and pressure retention molding at the bottom stop. The materials with lower drawing speed in Table 1, such as stainless steel and duralumin, are very suitable for the stamping process due to their characteristics. In addition, they have excellent adaptability to the stamping process of high-strength steel plates that require pressure molding at the bottom stop [17].

To study the interference of the moving parts and the movement law of the pendulum, the displacement curve of the pendulum in the X and Y directions is obtained by simulation. As shown in Figure 5, the displacement space in the X direction is −75 mm~+75 mm, and the displacement space in the Y direction is −300 mm~0 mm. It can be clearly seen from the above data that the pendulum swings in a working cycle while moving in reciprocating up-and-down straight lines.

In order to study the force change rule of the pin shaft and the core shaft in the main transmission mechanism, the vertical force of ${\mathrm{P}}_{\mathrm{g}}=2000\mathrm{KN}$ is loaded on the bottom surface of the slider. Figure 6 shows the oscillating pin in the main transmission mechanism calculated by the ADAMS simulation. The curve of the torque and the bending moment of the connecting rod mandrel change with the rotation angle of the eccentric gear. As can be seen from the curve in Figure 7, the bending moment of the connecting rod mandrel is ${\mathrm{M}}_{\mathrm{cs}}=48,000\mathrm{N}.\mathrm{m}$, and the value is independent of the rotation angle of the eccentric gear. The torque of the swing pin is related to the rotation angle of the eccentric gear. As the slider descends to the bottom stop, as it approaches 180°, the torque gradually decreases. When the slider runs to the nominal force stroke of 7 mm, the maximum torque of the pin is ${\mathrm{T}}_{\mathrm{smax}}=65,900\mathrm{N}.\mathrm{m}$. The above data serve as a theoretical basis for designing and checking the working capacity of the machine tool and the strength of the pin and mandrel [18].

The curves shown in Figure 5 represent the angle timeline of the large gear (shown as a red straight line), the displacement time curve of the slider’s online motion (shown as a blue solid curve), the velocity time curve of the slider (shown as a pink dashed curve), and the acceleration time curve of the slider (shown as a dark blue dashed curve). From the curve, it is evident that the speed of the slider begins to slow down around 160 degrees and maintains pressure at around 180 degrees.

The curves shown in Figure 6 represent the displacement curves of the pendulum in the X and Y directions, respectively. The orange curve represents the displacement curve in the X direction, while the blue dashed curve represents the displacement curve in the Y direction. From the curve, it can be observed that the pendulum reaches its maximum displacement in the X direction at 90 degrees and 270 degrees, while the maximum displacement in the Y direction is at 180 degrees.

The curves shown in Figure 7 are as follows: the red straight line represents the torque curve of the pendulum pin shaft; the blue dashed curve represents the bending moment curve of the connecting rod core shaft; and the orange dashed curve represents the position curve of the slider. It can be seen that the bending moment of the connecting rod core shaft remains constant, whereas the torque of the swing rod pin shaft changes with the position of the slider. It gradually decreases from a maximum value of 90 degrees to a minimum of 270 degrees. As the slider gradually approaches the bottom dead center (160–180 degrees), the torque of the pin shaft gradually decreases. This range is the force area, and the torsional strength needs to be verified based on the actual force position.

4. Key Reliability Index Detection of Physical Prototype

In ensure that the key reliability indexes of the inertia-controlled molding press are in line with the technical parameters in Figure 4 above, a physical prototype was designed and manufactured, and the static and dynamic tests and vibration tests were carried out on the physical prototype. The objective of these tests is to evaluate the static strength, stiffness, vibration characteristics, and dynamic characteristics of the 200 t inertia-controlled molding press.

4.1. Introduction of Physical Prototype

4.1.1. Mechanical System

According to the geometric parameters of the main transmission mechanism of the inertia-controlled molding press shown in Figure 2a, a physical prototype was made, as shown in Figure 8. The mechanical transmission system mainly includes a servo motor, a synchronous wheel, an eccentric gear, an eccentric wheel, and a swing rod.

4.1.2. Control System

For the physical prototype shown in Figure 8, the servo motor was taken as the control object, the eccentric gear angle was taken as the control value, and the embedded industrial PC was used as the controller, belonging to the semi-closed loop servo control. The pure motion control scheme with the position loop on the servo driver side was adopted. The control system structure is shown in Figure 9.

4.2. Test Content

The strength and stiffness of the body of the physical prototype and the dynamic characteristics of the machine tool are tested.

4.2.1. Static Test

The slider of the machine tool is located at the lower starting point. The static stress of the machine tool, as well as the deformation of the slider and the workbench under the upper loads, are tested.

4.2.2. Dynamic Test

During the piling of the machine tool, the dynamic stress of the high-stress measurement point in the static test is tested. The specific working conditions are as follows:

The machine tool moves continuously without load, and the acceleration of the sliding block is tested; when the piling is loaded, the acceleration of the workbench is tested.

4.3. Testing Instrument

(1)

The YE2539 (high-speed static strain testing system, Sichuan Haochi Instrument Co., Ltd., Deyang, China) high-speed static strain gauge has a test precision of 1 με and a test range of 0~±19,999 με. It is used for static stress–strain detection at various measurement points.

(2)

The test accuracy of dial indicator is 0.01 mm, and the test range is 10 mm. It is used for detecting the deformation displacement of the bottom surface of the slider relative to the workbench under load conditions.

(3)

The CA-YD-132 (GST pressure sensor in the United States, USA) piezoelectric acceleration sensor has a sensitivity of 104.7 pc/m/s², a maximum allowable acceleration of 103 m/s², and a test frequency range of 0.2~2.5 KHz [19]. It is used for detecting vibration waveforms and spectra at various measurement points under various working conditions.

(4)

The CRAS V6.2 (Anzheng Software, Nanjing, China) vibration and dynamic signal acquisition and analysis system has an acceleration acquisition accuracy of 0.0001 m/s². It is used for collecting vibration and dynamic signals under various working conditions.

(5)

The Br120-2AA (resistance strain gauge, Juhang Technology, Nanjing, China) resistance strain gauge 32 is used for stress and strain detection by sticking to measurement points.

(6)

The AZ216R (data collection system, Keinz, Osaka, Japan) data acquisition system is used for collecting detection data.

(7)

The WE-300 (universal material testing machine, VABOTEST Huabo Instrument, Dongguan, China) material universal testing machine, 0~3000 kN, is used for testing the mechanical performance parameters of experimental materials.

4.4. Test Plan

4.4.1. Static Test

The key parts of the press are mainly affected by the reaction force produced by stamping, which is the unidirectional tension (compression) stress.

4.4.2. Dynamic Stress Test

In the dynamic test, the measurement points with larger stress are selected and connected to the bridge box according to the full bridge circuit, and the dynamic strain waveforms are collected under various working conditions [20].

4.4.3. Deformation Test

The displacement of each typical position of the press under static loading was measured by a dial indicator. It provides a reliable basis for the rigidity performance of machine tools.

4.4.4. Acceleration Test

The acceleration responses of measurement points’ positions under different working conditions are tested. The acceleration signal is analyzed, and the vibration frequency of the machine tool is obtained.

4.5. Test Data

4.5.1. Static Stress Test

(1)

Distribution of static stress measurement points.

The stress measurement points of the body of the physical prototype are shown in Figure 10 below.

(2)

Static stress test data

The static strains of each measurement point are shown in Table 2 below (unit: µε).

The static stresses of each measurement point are shown in Table 3 below (unit: MPa).

4.5.2. Dynamic Stress Test

Stress measurement points to test the stress change of pile driving were selected. The stress measurement points selected were right 1, right 6, right 10, and connecting rod, etc., and the magnitude of the impact load on the iron pile was tested [21].

From Figure 11a, it can be seen that the dynamic strain at point 1 on the right is 201.6 με, and from Figure 11b, it can be seen that the dynamic strain at point 6 on the right is 395 με. From Figure 11c, it can be seen that the dynamic stress at point 10 on the right is 40.32 MPa, while from Figure 11d, it can be seen that the dynamic stress of the connecting rod is 39.5 MPa.

4.5.3. Deformation Test

The deformation test diagram is shown in Figure 12. Strain gauges are attached to areas with significant deformation on the machine tool. Then, loading equipment is installed on the workbench, and the bottom surface of the slider is loaded with different tonnages to detect the stress and strain, as well as the relative deformation displacement, between the workbench and the bottom surface of the slider.

Table 4 below shows the displacement of the measurement points relative to the subsequent table (unit: mm).

4.5.4. Dynamic Performance Test of the Machine Tool

Figure 13 shows the schematic diagram of the acceleration measurement points under two working conditions.

Figure 14 shows the vibration waveform and frequency spectrum of each measurement point under various working conditions.

4.6. Analysis and Conclusion of Experimental Results

4.6.1. Static Stress Test

Larger static stress points on the 200 t loaded machine tool can be found in Table 5.

4.6.2. Stress Analysis

(1)

According to the test results, the stress on the main body of the press is low, with most measurement points being less than 40 MPa. However, the local position of the fuselage has a stress level above 110 MPa. Attention should be paid to the effects of stress concentration [22].

(2)

The average dynamic stress and static stress of measurement point 1 are 110.4 MPa, 87.25 MPa, 65.75 MPa, and 101.9 MPa, respectively.

4.6.3. Dynamic Stress Test

(1)

Dynamic strain calibration

The hydraulic universal material testing machine is used to calibrate the dynamic strain of the iron pile. The load of the testing machine is 100 t, and the dynamic strain waveform is shown in Figure 15. When the machine tool is running, the dynamic strain waveform of the stress measurement point on the iron pile is shown in Figure 16 below [23,24].

According to the above dynamic strain values, the driving force can be calibrated.

$$\begin{array}{cc}\hfill \mathrm{Nominal}\mathrm{force}\left(\mathrm{t}\right)& \\ & =\mathrm{Peak}\mathrm{value}\mathrm{of}\mathrm{dynamic}\mathrm{strain}\mathrm{during}\mathrm{pile}\mathrm{driving}\hfill \\ & /\mathrm{Peak}\mathrm{value}\mathrm{of}\mathrm{dynamic}\mathrm{strain}\mathrm{during}\mathrm{calibration}\times 100\hfill \end{array}$$

By using the above formula, the force of the press at that time point is 125.5 t.

(2)

Calculation of dynamic stress

The dynamic strain and stress of each measurement point under 125.5 t nominal force are shown in Table 6 below:

When converted to a full load, the dynamic stress of each measurement point is shown in Table 7 below:

The dynamic strain amplitude of typical measurement points when the machine tool is running is converted into the full load. Subsequently, when compared with the full static load condition, the dynamic stress of the machine tool is slightly less than the static stress of the machine tool [25].

4.6.4. Acceleration Test

The vibration frequencies of the machine tool under various working conditions are shown in the following tables.

(1) The frequencies of each order under the slider when the machine tool is continuously unloaded are shown in Table 8.

(2) Table 9 shows the frequency of each step of the workbench when the machine tool is running continuously.

(3) The stroke times of the press are relatively low, and the strike frequency is far lower than the first natural frequency of the fuselage. The natural frequency of the fuselage is much greater than the excitation frequency, which indicates that the fuselage has a good dynamic stiffness [26].

4.6.5. Deformation Test

The deformation of measurement points is shown in Table 10 (unit: mm).

4.6.6. Deformation Analysis

(1) This paper mainly tests the deformation of the bottom surface of the slider. The tests indicate that the displacement of the bottom surface of the slider relative to the lower workbench is largest under a full load. When the maximum load is 200 t, the deformation of the bottom surface of the slider is 1.45 mm, with the values of the corresponding two measurement points staying consistent [27,28].

(2) The deformation of the bottom surface of the slider directly reflects the deformation of the crankshaft.

4.6.7. Table Vibration Test

The amplitude changes in the vertical direction of the workbench when the speed is increased by 100 are shown in Figure 17. When the stamping times are SPM = 50, the maximum amplitude is 0.052 mm and the minimum value is −0.076 mm; when the stamping times are SPM = 100, the maximum amplitude value is 0.064 mm and the minimum value is −0.082 mm [29].

4.6.8. Table Vibration Test

(1)

The vibration test data of the workbench show that the amplitude of the workbench does not increase significantly with the increase in the rotating speed, and it basically remains at the level of 0.075 mm, which indicates that the press can operate normally under the working conditions of 50–100 revolutions per minute [30].

(2)

It can be seen from the test results that the vibration amplitude of the press is within a safe range when the speed is 50–100 rpm, suggesting that there is potential for further speed improvement [31,32].

5. Conclusions

(1)

The motion characteristics of the active transmission mechanisms of this kind of press were designed and analyzed using ADAMS software. The displacement, velocity and acceleration curves of the slider; the motion track of the swing bar; and the bending moment and torque curve of the key stressed parts are given in detail. Through the study of these curves, it can be seen that the main transmission mechanism of this kind of inertia-controlled molding press can obtain different sliding block characteristics by changing the length and eccentricity of the swing rod. The sliding block moves slowly near the bottom stop, and the full load working area is long. To sum up, this kind of mechanism has obvious process characteristics of increasing force, slow stamping, and a constant pressure near the bottom stop. Thus, this kind of press is especially suitable for low-speed molding processes or molding processes that maintain the bottom stop pressure.

(2)

In this paper, the key reliability indexes of the physical prototype were comprehensively tested. Additionally, the evaluations of the static strength, stiffness, and vibration characteristics of the 200 t inertia-controlled molding press, together with the simulation analysis and test data in this paper, provide a reference for the design of this kind of machine.