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Abstract

High-speed punching presses are widely applied in modern manufacture industry, especially in occasions of automatic production with a large or a super large quality. To obtain a better production efficiency and a wilder application in industries, it is meaningful to develop the high-speed punching presses with a large press force and a long press stroke. Considering the limitations of mechanical and servo high-speed punching presses, hydraulic technology is the most possible way to develop the high-speed punching press with a large press force and a long press stroke so far. This article has proposed a hydraulic operating principle based on the hydraulic vibration technology for the development of the high-speed punching press with a press force of 1250 kN, a press stroke of 50 mm, and a press velocity of 1000 hpm. With a simulation model built under the AMESim environment, performances of the system are obtained and analyzed. Then, parameter optimization is also carried out and discussed. Finally, the operating principle and the parameters are refined. The obtained results have proved that the design meets the aimed performances of the high-speed punching press.

Keywords

Punching press hydraulic vibration control valve optimization AMESim

Introduction

High-speed punching presses are one of the important machines in modern manufacture industry. With the advantages of high product efficiency, high material utilizing rate, low machining cost, and good reproduction accuracy, they are widely applied in fields of electronics, telecommunication, motor, household appliances, and automobiles. At present, it is the most practical and feasible way to achieve the automatic production with a large or a super large quality.

Press force, press stroke, and press velocity are the three import indices to evaluate the performances of the high-speed press. Here, press force refers to the force acting on the workpiece during the press forming, press stroke refers to the displacement which the punching press can achieve, and press velocity refers to the number of press cycles that the punching press can achieve in a unit time, which is usually counted in a minute and obtain a unit of hits per minute (hpm). It is easy to understand that both a large press force and a fast press velocity have benefits to improve the production efficiency. For example, a large press force makes it possible to produce more workpieces in one press forming, while it is much evident in the relationship between the press velocity and the production efficiency. Meanwhile, a long press stroke is also important for the punching press. To some kind of machining workpiece, it needs a longer press stroke to complete the press forming. For example, a can may need a press stroke of 10–20 mm, while the silicon sheet only needs a press stroke of <1 mm. All in all, to obtain a better production efficiency and a wilder application in modern industries, it is meaningful to develop the high-speed punching presses with a large press force and a long press stroke.

According to the operation principles, the high-speed punching presses can be divided into three types:¹ mechanical ones, hydraulic ones, and servo ones. Considering the poor production flexibility and serious power loss, there are few new types of mechanical high-speed punching presses developed in recent years. Also, because of the limitations of servo motors, such as the small power capacity and the long response time under heavy load,² hydraulic technology is the most possible way to develop the high-speed punching presses with a large press force and a long press stroke so far.

Hydraulic high-speed punching press usually utilizes a main control valve to make the piston of a cylinder move in a reciprocation linear way. Researches are always focused on the performance improvements of the punching press, such as increasing press velocity, reducing power loss, suppressing vibration and noise. Brahmer³ proposed a method with over-control position set points around the upper and lower dead points. Through this way, the openings of main control valve would be comparatively large, and the piston movement is carried out more quickly, especially in the area near the upper and lower dead points, which leads to a rapid press velocity. To reduce the noise and cavitation caused by cutting impact phenomenon, Brahmer⁴ used a pressure-controlled throttle valve in the return flow line. Schaber⁵ developed a multi-pressure operation principle to control the oil pressure based on the pressing stages, which could reduce the energy requirement considerably. What is more, a new-styled main control valve was proposed, which is a slide valve with a mechanical displacement feedback from the piston and is placed inside the piston. Ostini⁶ invented an electro-hydraulic hybrid system for the high-speed punching presses. The main control valve of the system is a slide valve with the spool driven by the oil pressure, which is controlled by a small piston driven by a servo motor–cam system. This system has combined the powerful press force of hydraulics and the precision of electronic control at the same time. Essentially, the above control valves are servo valves and are always working at the middle position, which leads to a relatively high throttling power loss and a rapid temperature rise of the oil. Therefore, servo valves are not suitable for the development of the high-speed punching press with a great power, which will cause a severe problem of high oil temperature. To solve the problems caused by servo valves, Tang⁷ proposed a two-stage pilot-operated cartridge valve for the main control valve. In this valve, two high-speed on-off valves first control a direction valve and then the direction valve controls a cartridge valve. It is hoped to combine the advantages of the high-speed on-off valve and the cartridge valve, in order to meet the need of the high-speed punching press with a great power. However, the service life of the cartridge valve will not be long enough in such a high frequency on-off working condition. And there is no report about the successful use of the valve until now. At the same time, other researches^8–12 expected to solve the throttling power loss problems by the displacement-control systems. Considering the rather low operating frequency, they are obviously not working for the high-speed punching press.

As mentioned above, the authors choose the hydraulic vibration system as the operation principle of the high-speed punching press with a large press force and a long press stroke. Here, to describe briefly, the high-speed punching press refers to the high-speed punching press with a large press force and a long press stroke in the following part of this article. The hydraulic vibration system utilizes the fluid as the working medium, converts the pressure energy of the fluid to the kinetic energy of the piston, and outputs the energy with the vibration movement, which is very convenient to complete the high-speed punching movement. However, both the modelling difficulties of the hydraulic vibration system^13–19 and the nonlinearities of hydraulics^20–27 make it very hard to study the performances of such a hydraulic vibration system such as the high-speed punching press. To the best of authors’ knowledge, there is no report about this kind of research until now.

With the aim to develop a high-speed punching press with a press force of 1250 kN, a press stroke of 50 mm, and a press velocity of 1000 hpm, a hydraulic operating principle based on the hydraulic vibration technology is proposed first. Then, with the roughly determined parameters, a simulation model of the hydraulic system is built under the AMESim environment. Based on the analysis of performances, some optimizations are carried out. Finally, the operating principle and the parameters are refined. The obtained results have proved that the design meets the aimed performances.

Operating principle based on the hydraulic vibration technology

Based on the hydraulic vibration technology, the operating principle for the high-speed punching press is illustrated by the system shown in Figure 1. The features of this system are as follows:

A dual-pressure circuit is designed. The high pressure, P_s, is provided by the pump, and the low pressure, P_a, is provided by the accumulator. Here, since the pressure P_s is much higher than the pressure P_a, we assume the fluid pressure of the accumulator remains constant as P_a during the punching process.

A differential cylinder with three pressure chambers is designed. The three pressure chambers are labeled as CV₁, CV₂, and CV₃. Here, CV₃ keeps in connection to the pump.

Three control valves are utilized. Valve V₁ controls the fluid flowing from the pump into CV₂ or from CV₂ into the tank, valve V₂ controls the fluid flowing between the accumulator and CV₁, and valve V₃ controls the fluid flowing from the pump into CV₁. In order to obtain a predesigned punching movement, the three control valves must operate according to a certain time sequence, which is carried out by the control unit.

A relief valve connects the accumulator to the tank.

Figure 1.

Operating principle for the high-speed punching press.

A punching cycle can be divided into three stages: extending, punching, and retracting. Here, extending stage makes the piston move quickly from the upper dead point to the position very close to the workpiece, punching stage completes the press forming action, and retracting stage makes the piston return quickly from the lower dead point to the upper dead point. And the hydraulic schematic diagrams of the above three stages can be demonstrated in Figure 2 in detail:

Extending stage. Valves V₁, V₂, and V₃ work at the lower position. Therefore, CV₁ is connected to the accumulator, while CV₂ is connected to the pump. Here, the cylinder works as a differential cylinder. With the high pressure acting on CV₂ and CV₃ and the gravity of the piston, the piston extends quickly toward the workpiece. The accumulator should provide enough fluid into CV₁ to prevent the cavitation.

Punching stage. Valve V₁ works at the lower position, while valves V₂ and V₃ work at the upper position. Then, CV₁ and CV₂ are connected to the pump, which makes the piston have the largest press force to punch the workpiece.

Retracting stage. Valve V₁ works at the upper position, while valves V₂ and V₃ work at the lower position. Then, CV₁ is connected to the accumulator, and CV₂ is connected to the tank, which switches the movement direction of the piston under the high pressure in CV₃.

Figure 2.

Three stages of one punching cycle: (a) extending, (b) punching, and (c) retracting.

Obviously, the system could complete the extending stage quickly with a rather small power by a proper design of the area difference between CV₂ and CV₃. Specifically, the extending stage requires valve V₂ must have a sufficient throughflow capability in order to reduce or eliminate the cavitation occurred in CV₁, which is caused by the rapid movement of the piston. At the meantime, the relief valve connected to the accumulator is also different from the conventional one. During the extending stage, it is assumed that the fluid volume into CV₁ is Vol₁ with a pressure of P_a. Then, in the punching stage, the fluid pressure in CV₁ is increased from P_a to P_s under the squeezing action of the high-pressure fluid from the pump through valve V₃, which obtains a considerable energy saving compared to the conventional method with the high-pressure fluid working during the extending and punching stages. Here, it is assumed that the fluid volume into CV₁ is Vol₂ with a pressure of P_s. As we all know, during the retracting stage, the added fluid volumes in CV₁, Vol₁ with a pressure of P_a and Vol₂ with a pressure of P_s are pushed back into the accumulator. Clearly, to the accumulator, the fluid of Vol₂ with a pressure of P_s is the added part and is needed to be discharged into the tank through the relief valve, which could avoid the pressure rise of the accumulator and then slow down the retracting speed of the piston. Therefore, the relief valve works at a high-speed on-off state. Apparently, the three control valves and the relief valve in the system are all high-speed on-off valves, which make the design of the valves very different from those in the conventional hydraulic punching press.

Modeling and analysis

In order to obtain the performances of the system, a model is built with the help of AMESim software, which is demonstrated in Figure 3. Here, a constant pressure source is used to simulate the hydraulic power unit, and a force is used to simulate the load generated during the punching stage.

Figure 3.

Simulation model of the high-speed punching press.

Calculation of basic parameters

The basic parameters of the system can be determined according to the above-aimed performance, which is a press force of 1250 kN, a press stroke of 50 mm, and a press velocity of 1000 hpm, with the assumption of an operating pressure of 210 bar and a piston mass of 200 kg.

Cylinder

The maximum punching force occurs in the punching stage, which can be demonstrated by equation (1). And according to the extending and retracting stages, equations (2) and (3) can be obtained. Here, d₁ is the piston diameter, d₂ is the rod diameter in CV₃, d₃ is the rod diameter in CV₁ and CV₂, P_a is the fluid pressure of the accumulator, P_s is the operating pressure of the system, and G_p is the piston gravity (see Figure 1)

\frac{π}{4} \times d_{2}^{2} \times P_{s} = F_{p}

(1)

\frac{π}{4} \times (d_{2}^{2} - d_{3}^{2}) \times P_{s} + \frac{π}{4} \times d_{3}^{2} \times P_{a} + G_{p} = m_{p} \times a_{f}

(2)

\frac{π}{4} \times (d_{1}^{2} - d_{2}^{2}) \times P_{s} - \frac{π}{4} \times d_{3}^{2} \times P_{a} - G_{p} = m_{p} \times a_{r}

(3)

It is assumed that the three stages have the same amount of time, which is 20 ms. Then, with equations (1)–(3), the parameters of the cylinder can be obtained: d₁ = 281.1 mm, d₂ = 275.4 mm, and d₃ = 270.1 mm.

Control valves

According to the above-mentioned analysis, valve V₂ has the largest throughflow capability. To obtain a proper set value of valve flow rate, the mean flow rate into or out the cylinder should be calculated. The mean velocity of the piston, $\bar{v}$ , can be obtained by equation (4)

\bar{v} = \frac{s}{t}

(4)

Here, s is the distance completed by the piston during a punching cycle and t is the corresponding time.

Then, the mean flow rate into or out the cylinder can be obtained by equation (5)

\bar{Q} = A_{3} \bar{v}

(5)

Here, A₃ is the cross-sectional area of the rod in CV₁.

Since the piston should complete 1000 cycles in 1 min and the distance of one cycle is double the press stroke, the mean flow rate into or out the cylinder can be obtained by combining equations (4) and (5), the value of which is about 5968.8 L/min. Therefore, the mean flow rate of valve V₂ is about 5968.8 L/min. With respect to the instantaneous flow rate, it is clear that the value is much larger. Therefore, it is assumed that the throughflow capability of the valves V₂ has a flow rate of 10,000 L/min under a corresponding pressure drop of 1 bar.

The throughflow capability of valve V₃ can be determined by its operation principle. As mentioned above, the fluid volume through valve V₃ into CV₁ makes the pressure of CV₁ raised from P_a to P_s. Then, equation (6) can be obtained

Vo l_{2} = \frac{P_{s} - P_{a}}{E} Vo l_{CV 1}

(6)

Here, Vol_CV₁ is the volume of CV₁ and E is the bulk modulus of hydraulic oil with a value of 1000 MPa.

And

Vo l_{CV 1} = \frac{π}{4} d_{3}^{2} \times \frac{s}{2}

(7)

q_{3} = \frac{Vo l_{CV 1}}{t_{2}}

(8)

Here, q₃ is the mean flow rate of valve V₃ and t₂ is the operating time of the punching stage.

Equations (6)–(8) can derive that q₃ has a value of about 187.5 L/min. Then, it is assumed that the throughflow capability of the valves V₃ has a flow rate of 250 L/min under a corresponding pressure drop of 1 bar.

Similarly, the throughflow capability of valve V₁ can be determined by the retracting stage. Equation (9) can be derived. Here, q₁ is the mean flow rate of valve V₁, t₃ is the operating time of the retracting stage, and Vol_CV₃ is the volume of CV₃

q_{1} = \frac{Vo {l_{CV}}_{3}}{t_{3}}

(9)

And

Vo {l_{CV}}_{3} = \frac{π}{4} (d_{1}^{2} - d_{3}^{2}) \times \frac{s}{2}

(10)

Equations (9) and (10) derive that q₁ has a value of about 713.9 L/min. Then, it is assumed that the throughflow capability of valve V₁ has a flow rate of 1000 L/min under a corresponding pressure drop of 1 bar. With respect to the relief valve, we just assume that it has the same throughflow capability with valve V₁, which is 1000 L/min under a corresponding pressure drop of 1 bar.

Operation timing of the control valves

As mentioned above, the three stages have the same working time, which is 20 ms. Then, based on the simulation model in Figure 3, the operation timing of the control valves is illustrated in Figure 4. Here, it is assumed that the control valves could open or close in a very short time.

Figure 4.

Operation timing of the control valves.

Accumulator

The accumulator should provide CV₁ with enough fluid in the extending stage to avoid the cavitation. Then, the needed fluid volume, Vol_a, can be calculated by equation (11), which has a result of 2.98 L. Therefore, the accumulator should provide about 3 L fluid for CV₁ in the extending stage

Vo l_{a} = \frac{π}{4} \times d_{3}^{2} \times \frac{s}{2}

(11)

Because the volume change of the accumulator is rather quick in the extending stage, it can be assumed as an adiabatic process. Then, equation (12) is obtained

{(P_{g 1} + 1)}^{1.4} \times Vo l_{g 1} = {(P_{g 0} + 1)}^{1.4} \times Vo l_{g 0}

(12)

Here, P_g₀ is the gas precharge pressure, P_g₁ is the operating pressure, Vol_g₀ is the gas volume at P_g₀, and Vol_g₁ is the gas volume at P_g₁. And

Vo l_{g 0} = Vo l_{g 1} + 3 L

(13)

Here, the units of both Vol_g₀ and Vol_g₁ are liter.

To obtain a longer service life,²⁸ equations (14) and (15) should be satisfied

Vo l_{g 0} \geq \frac{1}{4} Vo l_{g 1}

(14)

P_{g 1} \geq \frac{1}{4} P_{g 0}

(15)

In order to obtain the above parameters of the accumulator, the value of Vol_g₁ is assumed as 1 L and pg0 is assumed as one-fourth of P_g₁. Then, equations (12)–(15) can yield: P_g₀ = 1.3 bar, P_g₁ = 5.2 bar, and Vol_g₀ = 4 L. Here, the relief pressure of the relief valve is set as 5.2 bar.

Simulation results and analysis

With the obtained basic parameters, the simulation results of the system can be obtained. Here, the load force during the punching stage is modeled by a ramped force, which varies from 0 to 1250 kN at the time interval from 20 to 40 ms and is demonstrated in Figure 5.

Figure 5.

Assumed load force during the punching stage.

The piston displacement is demonstrated in Figure 6. The piston could arrive at the maximum displacement with a value of 50 mm at the moment of 17.5 ms. However, the piston cannot overcome the load at the moment of 22 ms, which can be verified in Figure 7. The pressure in CV₁ is quite smaller than the preset operating pressure of 210 bar. At the moment of 40 ms, the retracting stage starts, and the piston does not return to the origin at the end of a cycle, which has a distance of about 8 mm. Obviously, with the above parameters, the system for high-speed punching press cannot complete the aimed performance, and the optimization will be carried out in the next section.

Figure 6.

Piston displacement in one punching cycle.

Figure 7.

CV₁ pressure in one punching cycle.

Optimization design

According to Figure 6, the system has an enough fast extending stage but a less powerful punching stage and a slower retracting stage. Many factors could influence the operation performances of the punching press, such as the operating pressure of the accumulator, the rod diameter of the piston, the piston mass. In this section, the optimization of the operation timing of the control valves is mainly discussed.

Operation timing of the control valves

In section “Calculation of basic parameters,” the operation timing of the control valves is set in a rather rough way, which just divides one cycle into three equal parts. According to Figure 6, this timing set is obviously not working for the aimed performance of high-speed punching press. To improve the performance, some points need to be considered in the timing optimization:

It is easy to understand that a short working time of high-pressure fluid is helpful for reducing the system power. Then, the punching time should be as short as possible.

The working time of valve V₂ should match the extending velocity of the piston. Otherwise, it may lead to the waste of time. The piston has completed the 50 mm displacement, while valve V₂ is still working, which makes a time interval of about 3 ms wasted (see Figure 6).

It should set a longer time for the retracting stage in order to complete the cycle on time.

In summary, it should decrease the extending and punching time and increase the retracting time. Therefore, according to the performance demonstrated in Figure 6, the extending time is set from 0 to 17 ms, the punching time is set from 18 to 32 ms, and the retracting time is set from 33 to 60 ms. The corresponding operation timing of the control valves is shown in Figure 8. And the working time of the load force is the same with the punching time, which will not be described for avoiding repetition in the following sections.

Figure 8.

New operation timing of the control valves.

The improved piston displacement is demonstrated in Figure 9. Clearly, the piston could achieve the aimed stroke in the extending stage and return to the origin in the retracting stage. However, it still cannot overcome the load in the punching stage. What is more, the retracting velocity is a little faster, which makes the piston arrive at the origin with about 3 ms earlier.

Figure 9.

Improved piston displacement in one punching cycle.

Figure 10 shows the flow rate performance of valve V₂. The maximum flow rate of valve V₂ occurs at the moment of 14 ms, which has a value of about 14,828 L/min and then the accumulator cannot provide any fluid into valve V₂. Then, at the punching stage, a reverse flow rate peak with a value of about 18,590 L/min occurs at the moment of 32.2 ms. The reason for this reverse flow rate peak is that valve V₃ has not closed completely, while valve V₂ starts to open at this moment. Because the throughflow capability of valve V₂ is much larger than that of valve V₃, it makes the fluid through valve V₃ flow through valve V₂ and relief valve into the tank, while not flow into CV₁ and work for the punching stage, which makes not only a waste of power but also a reduction in punching capability of the system. Therefore, in the determination of operation timing of the control valves, it should make sure that valve V₃ starts connecting to the CV₁ after the moment that valve V₂ has disconnected from the CV₁.

Figure 10.

Flow rate of valve V₂ in one punching cycle.

Above all, to decrease the extending and punching time and increase the retracting time, it can make the piston return to the origin in time. But the piston still cannot overcome the load in the punching stage. Also, since it needs some time to open or close the valves in practice, the existing operation timing of the control valves does not isolate valve V₃ to valve V₂ and then leads to a large reverse flow rate through valves V₃ to valve V₂ into the tank.

Time interval between valves V₂ and V₃

As mentioned above, a time interval between valves V₂ and V₃ needs to be set to make sure that these two valves do not work at the same time. The new operation timing of valves V₂ and V₃ is demonstrated in Figure 11. Valve V₂ has closed before the starting of valve V₃ with an interval of ΔT, and valve V₃ has closed ahead the starting of valve V₂ with an interval of ΔT. Here, the value of ΔT is set as 1 ms.

Figure 11.

Time interval between valves V₂ and V_3.

Then, the improved flow rate performance of valve V₂ is demonstrated in Figure 12, which has no large reverse flow rate peak in Figure 10. The time interval of 1 ms has made some improvements (see Figure 13). However, the improvement still cannot make the punching stage work well.

Figure 12.

Flow rate of valve V₂ in one punching cycle with a ΔT of 1 ms.

Figure 13.

Difference between the piston displacements in one punching cycle caused by a ΔT of 1 ms.

Other parameters

Clearly, it is hard to achieve the aimed performance of the high-speed punching press without changing other parameters. Through varying those parameters one by one, it can obtain the corresponding results from the simulation models and derive the influence of the parameters on the operation performances of the punching system. Here, since the analysis processes are very similar, they are omitted, and only the results are given as follows:

The accumulator with a large air volume and a small gas precharge pressure is good to increase the extending velocity.

Decreasing d₃ could increase the extending velocity and increasing d₂ leads to a quicker retracting velocity. A large value of d₁ is meaningful not only to increase the extending velocity but also slow down the retracting velocity.

Both a small piston mass and a high operation pressure cause a quick extending velocity.

A large throughflow capability of valve V₃ improves the performance of the punching stage.

Final model and analysis

Based on the previous analysis, the final operating principle of the high-speed punching press is demonstrated in Figure 14.²⁹ Here, a special tank, labeled by upper tank, has taken the place of the accumulator and the relief valve in Figure 1, which is different from other two tanks and placed just on the top of the cylinder in practice. And valve V₃ is changed to a three-port valve with three positions. The advantages of these improvements are as follows:

The system is much simple, which leads to a small volume and a lightweight.

The previous accumulator and relief valve are always working at a very high frequency. Compared to the conventional components, it is much difficult to design and manufacture those two components, and the cost is very high. Therefore, removing those two components is good to lower the design difficulty and reduce the cost.

The upper tank can be seen as an accumulator with an infinitely large gas volume and an infinitely low gas precharge pressure, which is helpful to improve the extending and retracting performances.

Just at the start of the retracting stage, valve V₃ works at the middle position, which makes CV₁ connected to the tank. Then, the CV₁ pressure is quickly reduced to a very small value, which makes the retracting velocity much quick.

Figure 14.

Final operating principle for the high-speed punching press.

The improved system has four stages: extending stage, punching stage, unloading stage, and retracting stage (see Figure 15). Compared to the previous system, the main difference is the added unloading stage. Thus, only the unloading stage is described in detail. Valve V₃ works at the lower position in the unloading stage and then CV₁ is connected to the tank, which makes the CV₁ pressure rapidly reduced (see Figure 15(c)). Clearly, it decreases the pressure load of the piston and makes a quick movement of the piston in the retracting stage.

Figure 15.

Four stages of one punching cycle: (a) extending stage, (b) punching stage, (c) unloading stage, and (d) retracting stage.

With the help of AMESim, the simulation model is shown in Figure 16. Here, valve V₂ is a rotational valve, and its flow rate performance can be obtained by a computational fluid dynamics (CFD) simulation method, which is the flow rate value array under the corresponding pressure when the valve opening varies from zero to its full opening. Since the focus of this article is not valve V₂, the details of the design and analysis of valve V₂ will not be discussed here. And the system parameters are listed in Table 1.

Figure 16.

Final simulation model of the high-speed punching press.

Table 1.

System parameters of the high-speed punching press.

Component	Parameter	Value
Cylinder	Upper rod diameter	276.3 mm
	Piston diameter	282.3 mm
	Lower rod diameter	279.2 mm
	Piston mass	200 kg
Valve V₃	Throughflow capability	100 L/min/bar
Valve V₁	Throughflow capability	1000 L/min/bar
Pump	Displacement	210 cm³/rev
Pump	Rotation speed	1000 rev/min
Relief valve	Relief pressure	250 bar
Accumulator	Gas volume	200 L
	Gas precharge pressure	180 bar
	Operating pressure	240 bar

rev: revolution.

The operation timing of the control valves is demonstrated in Figure 17. In one cycle, the extending stage works from 0 to 26 ms, the punching stage works from 27 to 32 ms, the unloading stage works for 1 ms, and the extending stage works from 34 to 60 ms. Here, because the unloading stage works for a rather short time, it is seemed as a part of the punching stage and not labeled in Figure 17. Moreover, the time interval between valves V₂ and V₃ is 1 ms at the start of the punching stage, which is not labeled in Figure 17 either.

Figure 17.

Final operation timing of the control valves.

Then, the displacement performance of the high-speed punching press can be obtained. The piston has completed the aimed stroke, 50 mm, at the moment of 26.2 ms (see Figure 18). Then, the piston could overcome the load force, 1250 kN, to keep at the position of 50 mm for about 5 ms. The CV₁ pressure remains at 240 bar in the punching stage, which makes the piston generate a punching force of 1250 kN (see Figure 19). At the moment of 32.4 ms, the piston starts to retract and has returned to the origin at the moment of 59.5 ms. Therefore, the system could complete the aimed performances of the high-speed punching press. Figure 20 shows the piston displacement in three cycles.

Figure 18.

Final piston displacement in one punching cycle.

Figure 19.

CV₁ pressure in one punching cycle.

Figure 20.

Piston displacement in three punching cycles.

The flow rate performances of the control valves are demonstrated in Figures 21 –23. Valve V₂ has the maximum flow rate of 11,957 L/min at the retracting stage and a second largest flow rate of 7270.2 L/min at the extending stage (see Figure 21). As demonstrated in Figure 22, valve V₃ has a flow rate peak of 1562.2 L/min at the moment of 26.7 ms. At this moment, valve V₃ works at the upper position (Figure 15(b)) and a rather large pressure difference between the inlet and outlet of valve V₃ leads to this flow rate peak. Another flow rate peak occurs at the moment of 31.3 ms with a value of 300 L/min. At this point, valve V₃ works at the lower position (Figure 15(c)), which makes a part of fluid in CV₁ flow through valve V₃ into the tank. In Figure 23, valve V₁ has a maximum flow rate of 599.2 L/min at the extending stage and a maximum flow rate of 524.6 L/min at the retracting stage.

Figure 21.

Flow rate of valve V₂ in one punching cycle.

Figure 22.

Flow rate of valve V₃ in one punching cycle.

Figure 23.

Flow rate of valve V₁ in one punching cycle.

Conclusion

In this article, a hydraulic operation principle based on hydraulic vibration technology is proposed for the high-speed punching press with a press force of 1250 kN, a press stroke of 50 mm, and a press velocity of 1000 hpm. By the simulation model built under the AMESim environment, results are obtained and analyzed. Then, optimization design has been carried out, which leads to the final operation principle. The new hydraulic system has a simple structure and a low cost. Moreover, it utilizes a unload stage to raise the retracting velocity. According to the simulation results, the design could achieve the aimed performances of the high-speed punching press. Meantime, the flow rate performances of the control valves are obtained, which is important for the following design and manufacturing of the control valves. Obviously, the control valves are the key part of the design of the high-speed punching press. However, considering the length limitation of this article and the focus of this article, the design of the control valves³⁰ is not included here and will be described in detail in another article. At the same time, the prototype of the high-speed punching press is in the process of manufacturing, and the experimental research will be carried out in the near future.

Footnotes

Academic Editor: Jun Li

Declaration of conflicting interests

The author(s) declared no potential conflicts of interest with respect to the research, authorship, and/or publication of this article.

Funding

The author(s) disclosed receipt of the following financial support for the research, authorship, and/or publication of this article: This research was supported by the Project of Key Innovation Team of Zhejiang Province (grant no. 2013TD14) and the Project of Major Science and Technology of Zhejiang Province (grant no. 2014C01061).

References

Song

YZ.

Manufacturing technology of precision stamping part for motors, electrical devices and electronics. Beijing, China: Machinery Industry Press, 2013.

Zhao

Chen

Cui

. The developing status and trend of the servo press. Forg Stamp Technol 2015; 40: 1–7.

Brahmer

Method for operating a hydraulic drive device and method of operating the same. US20120233990A1 Patent, 2011.

Brahmer

Hydraulic drive, in particular for machine tools, and method for controlling the hydraulic drive. US20100132349A1 Patent, 2008.

Schaber

Hydraulic drive device having two pressure chambers and method for operating a hydraulic drive device having two pressure chambers. US9121419B2 Patent, 2015.

Ostini

Fluid distributor apparatus and punching method. US8485087B2 Patent, 2008.

Tang

CM.

Research on the cartridge valve controlled by high-speed on-off valves and its application in the hydraulic punching press. Changsha, China: Central South University, 2009.

Shimotomai

Hydraulic drive unit of a press machine and a swash plate type variable capacity axial piston pump to use for said device. US5868555 Patent, 1999.

Schaber

Hydraulic drive with rapid stroke and load stroke. US20160084276A1 Patent, 2015.

10.

Junginger

Hydraulic drive with rapid stroke and load stroke. US20160084278A1 Patent, 2015.

11.

Brahmer

Hydraulic drive with rapid stroke and load stroke. US20160084279A1 Patent, 2015.

12.

Maier

Hydraulic drive with rapid stroke and load stroke. US20160084280A1 Patent, 2015.

13.

Wang

Guo

YQ.

Foreign hydraulic rock drill. Beijing, China: Coal Industry Press, 1987.

14.

Astashev

Tresvyatskij

AN.

Hydraulic self-sustained vibrations exciter for vibrating technological devices. Problemy Prochnosti 1993; 6: 52–59.

15.

Takahashi

Sasaki

Koike

. Fundamental properties of oil hydraulic multiple function cylinder. Shigen-to-Sozai 1993; 109: 885–889.

16.

Ding

Application of electro-hydraulic vibration system in the file chopping machine. China Mech Eng 1999; 8: 866–868.

17.

Margolis

Shim

Instability due to interacting hydraulic and mechanical dynamics in backhoes. In: Proceedings of ASME 2002 international mechanical engineering congress and exposition, New Orleans, LA, 17–22 November 2002, pp.81–86. New York: ASME.

18.

Dokukova

Konon

PN.

Equivalence of the impedance method and the method of amplitude-frequency characteristics for investigations of vibrations in hydraulically powered supports. Inzhenerno-Fizicheskii Zhurnal 2003; 76: 174–176.

19.

Rahmfeld

Ivantysynova

An overview about active oscillation damping of mobile machine structure. Int J Fluid Power 2004; 5: 5–24.

20.

Yao

Jiao

. High-accuracy tracking control of hydraulic rotary actuators with modelling uncertainties. IEEE/ASME T Mech 2014; 19: 633–641.

21.

Yao

Jiao

DW.

Extended-state-observer-based output feedback nonlinear robust control of hydraulic systems with backstepping. IEEE T Ind Electron 2014; 61: 6285–6293.

22.

Chen

Yao

Wang

Accurate motion control of linear motors with adaptive robust compensation of nonlinear electromagnetic field effect. IEEE/ASME T Mech 2013; 18: 1122–1129.

23.

Chen

Pan

Integrated adaptive robust control for multilateral teleoperation systems under arbitrary time delays. Int J Robust Nonlin 2016; 26: 2708–2728.

24.

Chen

Yao

Wang

. µ-Synthesis based adaptive robust control of linear motor driven stages with high-frequency dynamics: a case study with comparative experiments. IEEE/ASME T Mech 2015; 20: 1482–1490.

25.

Yuan

Chen

Yao

. Time optimal contouring control of industrial biaxial gantry: a high-efficient analytical solution of trajectory planning. IEEE/ASME T Mech. Epub ahead of print 18 July 2016. DOI: 10.1109/TMECH.2016.2592518.

26.

Sun

Gao

Kaynak

Finite frequency H_∞ control for vehicle active suspension systems. IEEE T Contr Syst T 2011; 19: 416–422.

27.

Sun

Gao

Kaynak

Vibration isolation for active suspensions with performance constraints and actuator saturation. IEEE T Mech 2015; 20: 675–683.

28.

CX.

Hydraulic components. Beijing, China: Machinery Industry Press, 1982.

29.

Ding

Song

Liu

. Hydraulic drive system of big stroke high velocity ram machine of heavy load. Patent CN205167648U, 2015.

30.

Ding

Song

Liu

. Low-power-consumption and large-flow high-speed switching valve. Patent CN105114386A, China, 2015.

Investigation on the design of hydraulic high-speed punching press with a large press force and a long press stroke

Abstract

Keywords

Introduction

Operating principle based on the hydraulic vibration technology

Modeling and analysis

Calculation of basic parameters

Cylinder

Control valves

Operation timing of the control valves

Accumulator

Simulation results and analysis

Optimization design

Operation timing of the control valves

Time interval between valves V2 and V3

Other parameters

Final model and analysis

Conclusion

Footnotes

Declaration of conflicting interests

Funding

References

Time interval between valves V₂ and V₃